IC ENGINE FUNDAMENTALS SAE BOOK.pdf

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Internal Combustion Engine Handbook

6606_Book.indb 1

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Other SAE International Books of Interest Modern Engine Technology from A to Z By Richard van Basshuysen and Fred Schäfer (Product Code R-373)

Introduction to Internal Combustion Engines, Fourth Edition By Richard Stone (Product Code R-391)

Laser Diagnostics and Optical Measurement Techniques in Internal Combustion Engines By Hua Zhao (Product Code R-406)

Engine Combustion: Pressure Measurement and Analysis By David R. Rogers (Product Code R-388)

An Introduction to Engine Testing and Development By Richard D. Atkins (Product Code R-344)

Automotive Fuels Reference Book, Third Edition Paul Richards (Product Code R-297)

Fuel Engine Interactions By Gautam Kalghatgi (Product Code R-409)

For more information or to order a book, contact: SAE INTERNATIONAL 400 Commonwealth Drive Warrendale, PA 15096 Phone: +1.877.606.7323 (U.S. and Canada only) Or +1.724.776.4970 (outside U.S. and Canada) Fax: +1.724.776.0790 Email: [email protected] Website: books.sae.org

6606_Book.indb 2

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Internal Combustion Engine Handbook 2nd Edition

Basics, Components, System, and Perspectives Edited by Richard van Basshuysen and Fred Schäfer Translated by TechTrans

Warrendale, Pennsylvania USA

Copyright © 2016 SAE International

6606_Book.indb 3

eISBN: 978-0-7680-8287-6

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

400 Commonwealth Drive Warrendale, PA 15096-0001 USA E-mail: [email protected] Phone: +1 877-606-7323 (inside USA and Canada) +1 724-776-4970 (outside USA) Fax: +1 724-776-0790 Copyright © 2016 SAE International. All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted, in any form by any means, without the prior written permission of SAE International. For permission and licensing requests, contact: SAE Permissions, 400 Commonwealth Drive, Warrendale, PA 15096-0001 USA; e-mail: [email protected]; phone: +1-724772-4028; fax: +1-724-772-4891. Printed in the United States of America SAE Order No. R-434 http://dx.doi.org/10.4271/r-434 Handbuch Verbrennungsmotor: English Internal combustion engine handbook: basics, components, systems, and perspectives / edited by Richard van Basshuysen and Fred Schäfer.   1152 p.  20.32 cm. Includes bibliographical references and index. ISBN 978-0-7680-8024-7 1. Internal combustion engines. 2. I. van Basshuysen, Richard, 1932– II. Schäfer, Fred, 1948– III.  Title. Library of Congress Control Number: 2015959682 Information contained in this work has been obtained by SAE International from sources believed to be reliable. However, neither SAE International not its authors guarantee the accuracy or completeness of any information published herein and neither SAE International nor its authors shall be responsible for any errors, omissions, or damages arising out of use of this information. This work is published with the understanding that SAE International and its authors are supplying information, but are not attempting to render engineering or other professional services. If such services are required, the assistance of an appropriate professional should be sought. ISBN-print 978-0-7680-8024-7 ISBN-PDF 978-0-7680-8287-6 ISBN-epub 978-0-7680-8289-0 ISBN-prc 978-0-7680-8288-3 To purchase bulk quantities, please contact: SAE Customer Service E-mail: [email protected] Phone: +1-877-606-7323 (inside USA and Canada) +1-724-776-4970 (outside USA) Fax: +1-724-776-0790 Visit the SAE Bookstore at BOOKS.SAE.ORG Translation from the German language edition: Handbuch Verbrennungsmotor By Richard van Basshuysen and Fred Schäfer Copyright © 2012 Vieweg+Teubner Verlag Vieweg+Teubner Verlag is a part of Springer Science+Business Media All Rights Reserved

6606_Book.indb 4

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Foreword to the Second English Edition The complexity of a modern internal combustion engine is certainly one of the reasons why one person is no longer able to comprehensively present all the important interplays in their full depth. Perhaps it is also one of the reasons why there has been no complete work on this subject to date anywhere. Although a large number of technical books deal with certain aspects of the internal combustion engine, there has been no publication until now that covers all of the major aspects of diesel and SI engines. The more than 100-year development of the internal combustion engine has resulted in an enormous amount of important information and detailed knowledge on the different demands, the large number of components, and their interactions. With an updated and expanded volume of more than 1200 pages, 1793 illustrations and more than 1300 bibliographical references, all essential technical aspects of the internal combustion engine are represented. It was, therefore, a particular endeavor of the publishers to place emphasis in all the right places and thus to present a work that closes significant gaps in the technical literature. Of particular note is the fact that this revision and expansion was produced in a very short time and, therefore, effectively reflects the current high status of the present-day technical development and permits a glance into the future. The editors were extremely keen to present theory and practice in a balanced ratio. This was achieved, in particular, by harnessing the cooperation of more than 120 authors from science and industry. With their help, a publication has been created that is a unique source of information and advice in the day-to-day work of education, research, and practice. It is aimed, in particular, at specialists involved in science and practice in the automotive, engine, mineral oil, and accessories industry and at students for whom it is designed to provide

valuable help throughout their studies. Furthermore, it is intended to be a useful advisor for patent lawyers, the motor vehicle trade, government offices, journalists, and interested members of the general public. The question of the future of the internal combustion engine is reflected in many new approaches to the solution of the problems concerning, for example, fuel consumption and environmental compatibility. Particularly under these aspects, by comparison with the alternatives, it is not difficult to predict that the reciprocating piston engine and its use in mobility will probably remain with us in its fundamental elements for many years to come. New drive systems always face the problem of having to compete with more than 100 years of development with enormous development capacities worldwide. This surely applies to the electric drive of cars—contrary to the euphoria created by the body politic. In addition to the presentation of the present-day status of motor development, it is important to answer the questions: In what direction is the internal combustion engine developing? How do we assess its potential regarding fuel consumption, cost optimization and environmental impact after more than one hundred years of development? What options are offered by future alternative fuels? Are there competing systems that could replace it in the coming decades? This book tries to give conclusive answers to these questions on the basis of present knowledge. Even though the main focus of the book is on the car engine, certain basic aspects also relate to the commercial vehicle engine. It is also innovative that different aspects of the gasoline engine as compared with the diesel engine are illustrated in this book. Will there be any fundamental difference between the gasoline and the diesel engine in a few years? We have only to look at the growing approximation

Internal Combustion Engine Handbook | v

6606_Book.indb 5

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Foreword to the Second English Edition

between gasoline and diesel engines: Gasoline engines with direct injection—in the future perhaps diesel engines with homogeneous combustion. Our special thanks go to all the authors for their constructive and disciplined collaboration and their understanding of the difficult task it is to coordinate the contributions of so many participants. Of particular mention is the punctuality of the authors that enabled the revised and expanded book to be published in a timely and thus current fashion—an event that deserves special mention in our opinion. Following the success of the first five issues—between 2002 and 2011, more than 20,000 copies have been printed in the German and English languages—we have updated the content for this sixth issue with a careful revision of, amongst others, the chapter “Fuel consumption.” The growing importance of the discussions about greenhouse gases and CO2 has been

taken into account and the influences of the engine application on CO2 emission has been demonstrated. At other places, the content has been updated to the current state of the art, where required, and the bibliographical references have been completed. More than 1,300 bibliographical references are now provided which will add even more benefits for the reader. Thanks to Vieweg+Teubner Verlag and the editors Ewald Schmitt and Gabriele McLemore in particular, for their constructive and proactive collaboration. Last but not least we thank IAV GmbH for the technical and material support in the creation of this work, without whose cooperation this book could never have been published. Bad Wimpfen/Hamm, 2011 Richard van Basshuysen, VDI Fred Schäfer, VDI/SAE

vi | Internal Combustion Engine Handbook

6606_Book.indb 6

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapters, Articles, and Authors Chapter

Article

Author

Chapter

Article

1

Historical Review

Prof. Dr. Ing. Stefan Zima (†)/ Prof. Dr.-Ing. Claus Breuer

5.5

Energy balance in the engine

6

Crank gears

2

Definition and Classification of Reciprocating Piston Engines

Dr.-Ing. Hanns Erhard Heinze Prof. Dr.-Ing Helmut Tschöke

6.1

Crankshaft Drive

Prof. Dr.-Ing. Stefan Zima (†)

6.2

Rotational Oscillations

Prof. Dr.-Ing. Claus Breuer

6.3

Variability of compression and swept volume

Prof. Dr.-Ing. Fred Schäfer

2.1

Definitions

2.2

Potentials for Classification

3

Characteristics

Prof. Dr.-Ing. Ulrich Spicher/

3.1

Piston displacement

Dr.-Ing. Sören Bernhardt

3.2

Compression Ratio

3.3

Rotational Speed and Piston Speed

3.4

Torque and Power

3.5

Fuel Consumption

3.6

Gas Work and Mean Pressure

3.7

Efficiency

3.8

Air Throughput and Cylinder Charge

3.9

Air-Fuel Ratio

4

Curves

Dr.-Ing. Peter Wolters/

4.1

Consumption curves

Dipl.-Ing. Bernd Haake

4.2

Emission maps

4.3

Ignition and injection maps

4.4

Exhaust gas temperature maps

5

Thermodynamic fundamentals

5.1

Cyclical processes

5.2

Comparative processes

5.3

Open comparative processes

5.4

Efficiency

Prof. Dr.-Ing. Fred Schäfer

Author

7

Engine components

7.1

Pistons/Wristpins/Wristpin Circlips

Dr.-Ing. Uwe Mohr/ Dr. Wolfgang Issler

7.2

Connecting rod

Philippe Damour

7.3

Piston rings

Prof. Dr.-Ing. Claus Breuer/ Dipl.-Ing. Frank Münchow

7.4

Engine block

Dipl.-Ing. Günter Helsper/ Dipl.-Ing. Karl B. Langlois/ Dr. Michael Wagner

7.5

Cylinders

Prof. Dr.-Ing. Claus Breuer/ Dipl.-Ing. Frank Münchow

7.6

Oil pan

Dipl.-Ing. Günter Helsper/ Dipl.-Ing. Karl B. Langlois/ Dr. Michael Wagner

7.7

Crankcase venting

Dr.-Ing. Uwe Meinig

7.8

Cylinder head

Prof. Dr.-Ing. Wilhelm Hannibal/ Dipl.-Ing. Johann Schopp

7.9

Crankshafts

Dr. mont. Leopold Kniewallner

7.10

Valve train components

Dipl.-Ing. Michael Haas

7.11

Valves

Dr.-Ing. Klaus Gebauer/

7.12

Valve springs

Dr.-Ing. Rudolf Bonse

7.13

Valve seat rings

Dr.-Ing. Gerd Krüger

7.14

Valve guides

7.15

Oil pump

Dr. Olaf Josef

Dr. Christof Lamparski/

Internal Combustion Engine Handbook | vii

6606_Book.indb 7

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapters, Articles, and Authors

Chapter

Article

Author

Chapter

Article

Author

7.16– 7.16.9

Camshaft

Dipl.-Ing. Hermann Hoffmann/ Dr.-Ing. Martin Lechner/ Dipl.-GwL. Falk Schneider/ Dipl.-Ing. Markus Lettmann/ Dipl.-Ing. Rolf Kirschner

10.4

Variable Valve Actuation

Prof. Dr.-Ing. Wilhelm Hannibal/ Dipl.-Ing. Andreas Knecht/ Dipl.-Ing. Wolfgang Stephan

10.4.4

Perspectives for the variable Prof. Dr.-Ing. Rudolf Flierl/ valve train Prof. Dr.-Ing. Wilhelm Hannibal

10.5– 10.5.5

Impulse Supercharging and Load Control of Reciprocating Piston Engines by utilizing an Air Stroke Valve

Dr.-Ing. Alfred Elsäßer/ Dipl.-Ing. René Dingelstadt/ Dipl.-Ing. Tobias Neubrand

10.5.6

Pulse Turbocharging with Controllable Aspiration Valves

Dipl.-Ing. Werner Wallrafen

11

Supercharging of Internal Combustion Engines

11.1

Mechanical Supercharging

Prof. Dr.-Ing. Hans Zellbeck

11.2

Exhaust Gas Turbocharging

Dr.-Ing. Tilo Roß

11.3

Intercooling

11.4

Interaction between Engine and Compressor

11.5

Dynamic Behavior

11.6

Additional Measures for Supercharged Internal Combustion Engines

11.7

Performance explosion by Register and Two-stage supercharging for Private Vehicles (Supercharging)

Dipl.-Ing. Marc Sens/ Dipl.-Ing. Guido Lautrich

11.8

Ascertaining turbocharger characteristic maps on turbocharger test benches

Dipl.-Ing. Marc Sens Dr. Panagiotis Grigoriadis

7.16.10

Camshaft shifter systems

Andreas Strauss

7.17

Chain drive

Dr.-Ing. Peter Bauer

7.18

Belt drives

Dipl.-Ing. Ralf Walter/ Dipl.-Ing. Wolfgang Körfer/ Dipl.-Ing. Michael Neu/ Dipl.-Ing. Franz Fusenig

7.19

Bearings in combustion engines

Dipl.-Ing. Dr. techn. Rainer Aufischer

7.20

Intake systems

Kay Brodesser

7.20.1

Thermodynamics in air intake systems

Dr.-Ing. Stephan Wild

7.20.2

Acoustics

Dipl.-Ing. Matthias Alex

7.21

Sealing systems

7.21.1

Cylinder head sealing systems

Dipl.-Ing. Armin Diez

7.21.2

Special seals

Dipl.-Ing. Wilhelm Kullen/ Dr.-Ing. Oliver Göb

7.21.3

Elastomer sealing systems

Dipl.-Ing. Eberhard Griesinger

7.21.4

Development methods

Dipl.-Ing. Uwe Georg Klump/ Dr. rer. nat. Hans-Peter Werner

7.22

Threaded connectors at the engine

Dipl.-Ing. Siegfried Jende

7.23

Exhaust Manifold

Dipl.-Ing. Hubert Neumaier

12

7.24

Coolant Pumps for Combustion Engines

Dipl.-Ing. Peter Amm/ Dipl.-Ing. Franz Pawellek

Mixture Formation and Related Systems

12.1

Internal Mixture Formation

7.25

Control Mechanisms for

Dr.-Ing. Uwe Meinig Two-Stroke Cycle Engines

12.2

External Mixture Formation

12.3

8

Engines

Mixture Formation in Diesel Engines

8.1

Engine concepts

12.3.8.1

8.2

Current engines

Intake Manifold Injection System

Prof. Dr.-Ing. Helmut Tschöke

8.3

Motorcycle engines/Special engines

12.3.8.2

Systems for Direct Injection

Dipl.-Ing. Achim Koch

12.4

8.4

Rotary piston engine/ Wankel engine

Mixture Formation in Diesel Engines

Prof. Dr.-Ing Helmut Tschöke

12.4.3

9

Tribology

Systems with a Central Pressure Reservoir

Dipl.-Ing. Wolfgang Bloching/ Dr. Klaus Wenzlawski

9.1

Friction

Dr.-Ing. Franz Maassen

12.4.4

9.2

Lubrication

Prof. Dr. Stefan Zima (†)

Injection Nozzles and Nozzle-Holder Assemblies

Prof. Dr.-Ing. Helmut Tschöke

10

Charge Cycle

12.4.5

10.1

Gas Exchange Devices in 4-Stroke Engines

Prof. Dr.-Ing. Ulrich Spicher/

Adapting the Injection System to the Engine

12.5

Fuel Supply Systems

10.2

Calculating Charge Cycle

Dr.-Ing. Sören Bernhardt

12.5.1

Fuel Tanks

10.3

The Charge Cycle in TwoStroke Engines

Dr.-Ing. Uwe Meinig

12.5.2

The Tank Venting System

12.5.3

Requirements for the Fuel Supply system

Prof. Dr.-Ing. Fred Schäfer Andreas Bilek

Prof. Dr.-Ing. Fred Schäfer

Dr.-Ing. Thomas Zapp

Dipl.-Ing. Holger Dilchert/ Dipl.-Ing. Bernd Jäger/ Dipl.-Ing. Frank Kühnel/ Dipl.-Ing. Ralph Schröder

viii | Internal Combustion Engine Handbook

6606_Book.indb 8

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapters, Articles, and Authors

Chapter

Article

Author

Chapter

Article

Author

12.5.4

The Filling Level Measuring

Dipl.-Ing. Knut Schröter

18.1

Temperature sensors

Dr.-Ing. Bernd Last

13

Ignition

Dr. Manfred Adolf/

18.2

Fuel level sensors

13.1

Gasoline Engines

Dipl.-Ing. Heinz-Georg Schmitz

18.3

Knock sensors

18.4

Exhaust gas sensors

18.5

Pressure sensors

18.6

Air mass sensor

18.7

Speed sensors

18.8

Combustion chamber pressure sensors for diesel engines

13.2

Diesel Engine

14

Combustion

Univ. Prof. Dr.-Ing. habil.

14.1

Fuels and Fuel Chemistry

Günter P. Merker/

14.2

Oxidation from hydrocarbons

Dr.-Ing. Peter Eckert

14.3

Autoignition

14.4

Flame Propagation

19

Actuators

14.5

Model Development and Simulation

19.1

Drive

15

Combustion systems

19.2

Throttle valve actuators

15.1

Diesel engines

19.3

Swirl and tumble plates, resonance charging

19.4

Turbocharger with variable turbine geometry

Dipl.-Ing. Marc Sens Dipl.-Ing. Reinhold Bals/ Dipl.-Ing. Ralf Waschek Dipl.-Ing. Michael Riess

19.5

Exhaust Gas Recirculation Valves

19.6

Evaporation emissions, components

Dr.-Ing. Uwe Meinig

20

Cooling of Internal Combustion Engines

Dipl.-Ing. Matthias Banzhaf/

20.1

General information

Dr.-Ing. Wolfgang Kramer

20.2

Demands on the Cooling System

20.3

Principles for Calculation and Simulation Tools

20.4

Engine Cooling Subsystems

20.5

Cooling modules

20.6

Overall Engine Cooling System

21

Exhaust emissions

21.1

Legal requirements

ao. Univ.-Prof. Dipl.-Ing. Dr. techn.

21.2

Exhaust measuring technology

Ernst Pucher

21.3

Pollutants and Their Origin

21.4

Reduction in pollutants

21.5

Exhaust Treatment for gasoline Engines

21.5.1

Catalytic converter design and chemical reactions

Dipl.-Ing.Stefan Brandt

21.5.2

Catalytic converter concepts of engines with stoichiometric operation

Dr. Stephan Siemund/ Dr.-Ing. Susanne Stiebels

21.5.3

Catalytic converter concept for lean-burn engines

Dipl.-Ing. Stefan Brandt/ Dipl.-Ing. Uwe Dahle

21.5.4

Metallic catalytic converter substrate

Dr. Andrée Bergmann

21.6

Exhaust Treatment in Diesel Engines

21.6.1

Diesel oxidation catalytic converters

Dr. rer. nat. Peter Scherm

21.6.2

NOx Adsorbers for Diesel Passenger Cars

Dr. rer. nat. Tilman Beutel

15.2

Spark-injection engines

Prof. Dr.-Ing Helmut Tschöke/ Dr.-Ing. Detlef Hieber/

15.3

Two-stroke diesel engines

15.4

Two-stroke SI engines

16

Electronics and mechanics for engine and shift control transmission management

16.1

Environmental requirements Dipl.-Ing. Rainer Riecke/

16.2

Stand-alone products

Dipl.-Ing. Karl Smirra/

16.3

Connection systems

Dr. rer. Nat.-Phys. Thomas Riepl/

16.4

Stand-alone products Integrated Products (MTM = Mechatronic Transmission Module)

Dipl.-Ing. Gerwin Höreth

16.5

Electronic design, structures and components

16.6

Electronics in the electronic control unit

16.7

Software structures

Dr.-Ing. Robert Rehbold

16.8

Torque-Based Functional Structure for Engine Management

Dipl.-Ing. Karl Smirra

16.9

Functions

17

The powertrain

17.1

Powertrain Architecture

17.2

The Motor-Vehicle’s Longitudinal Dynamics

17.3

Transmission types

17.4

Power level and signalprocessing level

17.5

Transmission management

17.6

Integrated powertrain management (IPM®)

17.7

Components for powertrain electricification

Dipl.-Ing. Uwe Möhrstädt

18

Sensors

Dr.-Ing. Anton Grabmaier/

Dr.-Ing. Michael Ulm

Dipl.-Ing. Friedrich Graf

Dipl.-Ing. Stefan Klöckner

Dipl.-Wirt.-Ing. Axel Tuschik

Internal Combustion Engine Handbook | ix

6606_Book.indb 9

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapters, Articles, and Authors

Chapter

Article

Author

Chapter

Article

21.6.3

Particles/Particle Filters

Dr. h.c. Dipl.-Ing. Andreas C. R. Mayer/ Dr. Markus Kasper/ Prof. Dr. Heinz Burtscher

27.9

Sound Engineering

27.10

Simulation Tools

27.11

Antinoise Systems: Noise Reduction using Antinoise

Dipl.-Ing. Alfred Punke

28

Engine Measuring Technology

Dipl.-Ing. Dr. techn. Christian Beidl/ Dipl.-Ing. Dr. techn. KlausChristoph Harms/ Dr. Christoph R. Weidinger

Author

21.6.3.16

Catalytic particle filter

22

Operating fluids

22.1

Fuels

Wolfgang Dörmer

22.1.2.3

Alternative gasolines

Norbert Neumann/

22.2

Lubricants

Dr. Ulrich Pfisterer/

29

Hybrid Drive Systems

Prof. Dr.-Ing. Fred Schäfer

22.3

Coolant

Dr. Oliver Busch/ Martin Redzanowski

29.1

History

Carsten von Essen

29.2

Classification of Hybrid Drive Systems (General Overview)

29.3

Classification of Hybrid Drive Systems

29.4

Electrical Hybrid Drive Systems

29.5

Energy Storage Device

29.6

Hybrid Drive System Transmissions

23

Filtration of operating fluids

23.1

Air filter

Dr.-Ing. Manfred Tumbrink/ Dr.-Ing. Pius Trautmann

23.2

Fuel filters

Jochen Reyinger

23.3

Engine oil filtration

Markus Kolzcyk/ Dr.-Ing. Pius Trautmann

24

Calculation and simulation

24.1

Strength and vibration calculation

Dr.-Ing. Werner Dirschmid/ Dr.-Ing. Erich Blümcke

29.7

Energy Management System

24.1.3

Piston calculations

Dr.-Ing. Uwe Lehmann/ Dr. Ralf Meske

29.8

Operating Strategies

29.9

Current Hybrid Vehicles

Dr.-Ing. Werner Dirschmid/ Dr.-Ing. Erich Blümcke

29.10

Future Development

30

Alternative Vehicle Drives and APUs (Auxiliary Power Units)

24.2

Flow calculation

25

Combustion diagnostics— indication and visualization in combustion development

30.1

Reasons Behind Alternatives

25.1

Discussion

30.2

Hybrid Vehicles

25.2

Indicating

30.3

Electric Drive System

25.3

Visualization

30.4

Energy Storage Devices

26

Fuel Consumption

Prof. Dr.-Ing. Peter Steinberg/

30.5

Stirling Engine

30.6

Gas Turbine

Dipl.-Ing. Dirk Goßlau

30.7

The Fuel Cell as a Vehicle Drive System

30.8

Synoptic Evaluation of Alternative Energies and Drive Systems

30.9

Generating Electricity using an Auxiliary Power Unit = APU

31

Energy Management in the Engine and Vehicle

31.1

Losses During Energy Conversion

26.1

General Influencing Factors

26.2

Engine Modifications

26.3

Transmission Ratios

26.4

Driver Behavior

Dr. Ernst Winklhofer/ Dr. Walter F. Piock/ Dr. Rüdiger Teichmann

26.5

CO2 Emissions

27

Noise Emission

Dr.-Ing. Hans-Walter Wodtke/

27.1

Basic Physical Principles and Terms

Prof. Dipl.-Ing. Dr. techn.

27.2

Legal Provisions Concerning Emitted Noise

Hartmut Bathelt/

27.3

Sources of Emitted Noise

Dipl.-Ing. Andreas Gruber

27.4

Emitted Noise-Reduction Provisions

31.2

Requirement-Based Energy Management

27.5

Engine Noise in the Vehicle Interior

31.3

Generating Electricity in Vehicles

27.6

Acoustic Guidelines for the Engine Designer

31.4

Heat Management

32

Forecast

27.7

Measuring and Analytical Methods

27.8

Psychoacoustics

Prof. Dr.-Ing. Ulrich Seiffert/ Dipl.-Ing. Wilfried Nietschke

Prof. Dr.-Ing. Fred Schäfer/ Dr.-Ing. E.h. Johannes Liebl

Dr.-Ing. E.h. Richard van Basshuysen

x | Internal Combustion Engine Handbook

6606_Book.indb 10

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents Note to the reader: Bibliographic references that are included at the end of a chapter but are not marked in the text, indicate additional literature. This may serve to provide the reader with more in-depth information on the material covered in the respective chapter.

Foreword to the second English edition . . . . . . . . . . . . . . v Chapters, Articles, and Authors . . . . . . . . . . . . . . . . . . . . vii

1 Historical Review . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8

2 Definition and Classification of Reciprocating Piston Engines . . . . . . . . . . . . . 9 2.1 Definitions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 2.2 Potentials for Classification . . . . . . . . . . . . . . . . . . . . . . . . 10 2.2.1 Combustion Processes. . . . . . . . . . . . . . . . . . . . . . . . 10 2.2.2 Fuel. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10 2.2.3 Working Cycles. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.4 Mixture Generation. . . . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.5 Gas Exchange Control . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.6 Supercharging. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.7 Configuration. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11 2.2.8 Ignition. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 2.2.9 Cooling. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 2.2.10 Load Adjustment. . . . . . . . . . . . . . . . . . . . . . . . . . . . 13 2.2.11 Applications. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14 2.2.12 Speed and Output Graduations. . . . . . . . . . . . . . . 14 Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14

3 Characteristics . . . . . . . . . . . . . . . . . . . . . . . . . 15 3.1 Piston Displacement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15 3.1.1 Calculation of Stroke and Piston Displacement from the Crankshaft Position (Figure 3.1) . . . . . . . . . . . . 15

3.2 Compression Ratio. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.3 Rotational Speed and Piston Speed. . . . . . . . . . . . . . . . . . 3.3.1 Rotational Speed. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.3.2 Angular Velocity. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.3.3 Piston Speed. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.3.4 Mean Piston Speed. . . . . . . . . . . . . . . . . . . . . . . . . . . 3.3.5 Maximum Piston Speed. . . . . . . . . . . . . . . . . . . . . . . 3.4 Torque and Power . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.5 Fuel Consumption. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.6 Gas Work and Mean Pressure . . . . . . . . . . . . . . . . . . . . . . 3.6.1 Indicated Mean Pressure. . . . . . . . . . . . . . . . . . . . . . 3.6.2 Effective Mean Pressure. . . . . . . . . . . . . . . . . . . . . . . 3.6.3 Friction Mean Pressure. . . . . . . . . . . . . . . . . . . . . . . . 3.7 Efficiency. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.7.1 Indicated Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . 3.7.2 Effective Efficiency. . . . . . . . . . . . . . . . . . . . . . . . . . . 3.7.3 Mechanical Efficiency. . . . . . . . . . . . . . . . . . . . . . . . . 3.8 Air Throughput and Cylinder Charge . . . . . . . . . . . . . . . 3.8.1 Air Expenditure. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.8.2 Volumetric Efficiency. . . . . . . . . . . . . . . . . . . . . . . . . 3.9 Air–Fuel Ratio. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

16 17 17 17 17 18 18 18 19 20 20 21 22 22 22 22 22 23 23 23 24 25

4 Curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27 4.1 Consumption Curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.2 Emission Maps. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.3 Ignition and Injection Maps. . . . . . . . . . . . . . . . . . . . . . . . 4.4 Exhaust Gas Temperature Maps . . . . . . . . . . . . . . . . . . . .

28 29 32 33

Internal Combustion Engine Handbook | xi

6606_Book.indb 11

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

5 Fundamentals of Thermodynamics . . . . . 35 5.1 Cyclical Processes. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.2 Comparative Processes. . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.2.1 Simple Model Processes. . . . . . . . . . . . . . . . . . . . . . . 5.2.2 Exergy Losses. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.3 Open Comparative Processes. . . . . . . . . . . . . . . . . . . . . . . 5.3.1 Work Cycle of the Perfect Engine. . . . . . . . . . . . . . . 5.3.2 Approximation of the Real Working Cycle. . . . . . . 5.4 Efficiency. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.5  Energy Balance in the Engine. . . . . . . . . . . . . . . . . . . . . . 5.5.1 Balance Equation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

35 37 37 39 40 40 42 45 46 46 47

6 Crank Gears . . . . . . . . . . . . . . . . . . . . . . . . . . . . 49 6.1 Crankshaft Drive. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.1 Design and Function. . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.2 Forces Acting on the Crankshaft Drive. . . . . . . . . . 6.1.3 Tangential Force Characteristic and Average Tangential Force. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.4 Inertial Forces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.5 Mass Balancing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.6 Internal Torques. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1.7 Throw and Firing Sequences. . . . . . . . . . . . . . . . . . . 6.2 Rotational Oscillations . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.2.1 Fundamentals . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.2.2 Reduction of the Machine System . . . . . . . . . . . . . . 6.2.3 Natural Frequencies and Modes of Natural Vibration. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.2.4 Exciter Forces, Work, and Amplitudes. . . . . . . . . . . 6.2.5 Measures to Reduce Crankshaft Excursions. . . . . . 6.2.6 Two-Mass Flywheels . . . . . . . . . . . . . . . . . . . . . . . . . 6.3 Variability of Compression and Swept Volume. . . . . . . . 6.3.1 Variable Swept Volume . . . . . . . . . . . . . . . . . . . . . . . 6.3.2 Variable Compression. . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

49 49 52 58 60 66 70 71 72 72 73 73 75 76 77 78 78 79 82

7 Engine Components . . . . . . . . . . . . . . . . . . . . 83 7.1 Pistons/Wristpins/Wristpin Circlips. . . . . . . . . . . . . . . . 83 7.1.1 Piston . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83 7.1.2 Wristpins. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94 7.1.3 Wristpin Circlips. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94 7.2 Connecting Rod. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95 7.2.1 Connecting Rod Design. . . . . . . . . . . . . . . . . . . . . . . 95 7.2.2 Loading. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 96 7.2.3 Conrod Bolts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 97 7.2.4 Design. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 97 7.2.5 Conrod Manufacture . . . . . . . . . . . . . . . . . . . . . . . . . 98 7.2.6 Conrod Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . 100 7.3 Piston Rings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 101 7.3.1 Designs. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 102 7.3.2 Ring Sets. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 104 7.3.3 Characterizing Features. . . . . . . . . . . . . . . . . . . . . . 105 7.3.4 Manufacturing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 107 7.3.5 Loading, Damage, Wear, and Friction. . . . . . . . . . 109

7.4 Engine Block. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.4.1 Tasks and Functions. . . . . . . . . . . . . . . . . . . . . . . . . 7.4.2 Engine Block Design. . . . . . . . . . . . . . . . . . . . . . . . . 7.4.3 Optimizing Acoustic Properties. . . . . . . . . . . . . . . 7.4.4 Minimizing Engine Block Mass . . . . . . . . . . . . . . . 7.4.5 Casting Process for Engine Blocks. . . . . . . . . . . . . 7.5 Cylinders. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.5.1 Cylinder Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.5.2 Machining Cylinder Running Surfaces. . . . . . . . . 7.5.3 Cylinder Cooling. . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.6 Oil Pan. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.6.1 Oil Pan Design. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.7 Crankcase Venting. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.7.1 Regulatory Marginal Conditions . . . . . . . . . . . . . . 7.7.2 Technical Requirements. . . . . . . . . . . . . . . . . . . . . . 7.7.3 System Structure of Current Crankcase Venting Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.7.4 Oil Mist Separation. . . . . . . . . . . . . . . . . . . . . . . . . . 7.7.5 Crankcase Pressure Control. . . . . . . . . . . . . . . . . . . 7.7.6 Modules and Valve Bonnet Integration. . . . . . . . . 7.8 Cylinder Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.8.1 Basic Design for the Cylinder Head. . . . . . . . . . . . 7.8.2 Cylinder Head Design. . . . . . . . . . . . . . . . . . . . . . . 7.8.3 Casting Processes . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.8.4 Model and Mold Manufacturing . . . . . . . . . . . . . . 7.8.5 Machining and Quality Assurance. . . . . . . . . . . . . 7.8.6 Shapes Implemented for Cylinder Heads. . . . . . . 7.8.7 Perspectives in Cylinder Head Technology . . . . . 7.9 Crankshafts. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.9.1 Function in the Vehicle. . . . . . . . . . . . . . . . . . . . . . . 7.9.2 Manufacturing and Properties . . . . . . . . . . . . . . . . 7.9.3 Lightweight Construction and Processes for Increasing Strength . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.9.4 Calculating Crankshafts. . . . . . . . . . . . . . . . . . . . . . 7.10 Valve Train Components. . . . . . . . . . . . . . . . . . . . . . . . . 7.10.1 Standard Valve Train . . . . . . . . . . . . . . . . . . . . . . . 7.10.2 Belt Tensioning Systems, Idler, and Deflection Pulleys . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.10.3 Chain Tensioning and Guide Systems. . . . . . . . . 7.11 Valves. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.11.1 Functions and Explanation of Terms and Concepts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.11.2 Types of Valves and Manufacturing Techniques. . . . . . . . . . . . . . . . . . . . . . . . 7.11.3 Embodiments. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.11.4 Valve Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.11.5 Special Valve Designs. . . . . . . . . . . . . . . . . . . . . . . 7.11.6 Valve Keepers. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.11.7 Valve Rotation Devices. . . . . . . . . . . . . . . . . . . . . . 7.12 Valve Springs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.12.1 Determining Strain under Load. . . . . . . . . . . . . . 7.13 Valve Seat Rings. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.13.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.13.2 Demands Made on Valve Seat Inserts. . . . . . . . . 7.14 Valve Guides. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.14.1 Requirements for Valve Guides . . . . . . . . . . . . . .

110 110 112 116 117 118 119 120 123 124 125 126 126 126 127 129 129 134 136 137 137 139 147 151 152 153 158 159 159 160 162 163 164 164 174 177 179 179 180 181 183 183 184 185 186 187 189 189 190 196 196

xii | Internal Combustion Engine Handbook

6606_Book.indb 12

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

7.14.2 Materials and Properties . . . . . . . . . . . . . . . . . . . . 7.14.3 Geometry of the Valve Guide. . . . . . . . . . . . . . . . 7.14.4 Installing in the Cylinder Head . . . . . . . . . . . . . . 7.15 Oil Pump. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.15.1 Overview of Oil Pump Systems. . . . . . . . . . . . . . 7.15.2 Regulating Principles. . . . . . . . . . . . . . . . . . . . . . . 7.15.3 Flow Rate Regulating Pumps. . . . . . . . . . . . . . . . 7.15.4 Consumption Savings in NEFZ Cycle. . . . . . . . . 7.15.5 Engineering Principles. . . . . . . . . . . . . . . . . . . . . . 7.15.6 Cavitation and Noise Emission. . . . . . . . . . . . . . . 7.15.7 Calculation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.16 Camshaft. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.16.1 Camshaft Function. . . . . . . . . . . . . . . . . . . . . . . . . 7.16.2 Valve Train Configurations. . . . . . . . . . . . . . . . . . 7.16.3 Structure of a Camshaft. . . . . . . . . . . . . . . . . . . . . 7.16.4 Technologies and Materials. . . . . . . . . . . . . . . . . . 7.16.5 Reduction of Mass. . . . . . . . . . . . . . . . . . . . . . . . . . 7.16.6 Factors Influencing Camshaft Loading. . . . . . . . 7.16.7 Design of Cam Profiles. . . . . . . . . . . . . . . . . . . . . . 7.16.8 Kinematics Calculation . . . . . . . . . . . . . . . . . . . . . 7.16.9 Dynamics Calculations. . . . . . . . . . . . . . . . . . . . . . 7.16.10 Camshaft Shifter Systems . . . . . . . . . . . . . . . . . . 7.17 Chain Drive. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.17.1 Chain Designs. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.17.2 Typical Chain Values . . . . . . . . . . . . . . . . . . . . . . . 7.17.3 Sprockets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.17.4 Chain Guide Elements. . . . . . . . . . . . . . . . . . . . . . 7.18 Belt Drives. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.18.1 Belt Drives Used to Drive Camshafts. . . . . . . . . . 7.18.2 Toothed V-Belt Drive to Power Auxiliary Units. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19 Bearings in Combustion Engines. . . . . . . . . . . . . . . . . . 7.19.1 Principles. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19.2 Calculating and Dimensioning Engine Bearings. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19.3 Bearing Materials . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19.4 Bearing Versions—Structure, Load-Handling Capacity, Application . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19.5 Bearing Failure. . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.19.6 Prospects. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.20 Intake Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.20.1 Thermodynamics in Air Intake Systems. . . . . . . 7.20.2 Acoustics. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.21 Sealing Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.21.1 Cylinder Head Sealing Systems. . . . . . . . . . . . . . 7.21.2 Special Seals. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.21.3 Elastomer Sealing Systems . . . . . . . . . . . . . . . . . . 7.21.4 Development Methods. . . . . . . . . . . . . . . . . . . . . . 7.22 Threaded Connectors at the Engine . . . . . . . . . . . . . . . 7.22.1 High-Strength Threaded Connectors. . . . . . . . . . 7.22.2 Quality Requirements. . . . . . . . . . . . . . . . . . . . . . 7.22.3 Threaded Connectors. . . . . . . . . . . . . . . . . . . . . . . 7.22.4 Threaded Connections in Magnesium Components. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.22.5 Screw Tightening Process. . . . . . . . . . . . . . . . . . . 7.23 Exhaust Manifold. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

198 201 202 202 203 208 211 214 216 220 222 224 225 225 226 227 232 233 234 235 236 236 238 240 241 241 242 243 243 249 251 252 254 258 262 267 268 268 269 271 275 275 280 284 287 291 291 291 292 297 298 299

7.23.1 Manifold Development Process. . . . . . . . . . . . . . 7.23.2 Manifolds as Individual Components. . . . . . . . . 7.23.3 The Manifold as a Submodule . . . . . . . . . . . . . . . 7.23.4 Manifold Components. . . . . . . . . . . . . . . . . . . . . . 7.24 Coolant Pumps for Combustion Engines. . . . . . . . . . . 7.24.1 Requirements, Models, and Constructive Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.24.2 Impeller and Spiral Port. . . . . . . . . . . . . . . . . . . . . 7.24.3 Coolant-Side Sealing. . . . . . . . . . . . . . . . . . . . . . . . 7.24.4 Map and Similarity Relations of the Coolant Pump. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.24.5 Cavitation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7.24.6 Electric Coolant Pump and Switchable Mechanical Coolant Pump. . . . . . . . . . . . . . . . . . . . . . . . 7.25 Control Mechanisms for Two-Stroke Engines. . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

300 301 303 304 304 304 306 307 308 309 311 312 315

8 Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 321 8.1 Engine Concepts. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 321 8.1.1 Engine Models. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 323 8.1.2 Differentiating Features of Engine Concepts Regarding the Basic Engine . . . . . . . . . . . . . . . . . . . . . . . 325 8.1.3 Further Concept Criteria. . . . . . . . . . . . . . . . . . . . . 327 8.1.4 Engine Arrangement Concepts in the Vehicle . . . 327 8.2 Current Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 328 8.2.1 V6 Diesel Engine by Mercedes-Benz. . . . . . . . . . . 328 8.2.2 4-Liter V8 Diesel Engine by Mercedes-Benz. . . . . 328 8.2.3 V10 FSI engine by Audi. . . . . . . . . . . . . . . . . . . . . . 329 8.2.4 1.6-Liter V8 Turbocharged Gasoline Engine by GM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 329 8.2.5 2-Liter 4V-TDI with Common-Rail by Volkswagen. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 330 8.2.6 6-Liter V12 TDI Engine by Audi. . . . . . . . . . . . . . . 331 8.2.7 4.8-Liter V8 SI Engine by Porsche (Turbo and Naturally Aspirated Engine) . . . . . . . . . . . . . . . . . . . . . . 333 8.2.8 Stratified Combustion Process for Four- and SixCylinder Intake Gasoline Engines by BMW. . . . . . . . . 334 8.2.9 Four-Cylinder Diesel Engine by Mercedes-Benz. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 336 8.2.10 SI Engines with Direct Injection and Dual Charging by VW. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 337 8.3 Motorcycle Engines/Special Engines. . . . . . . . . . . . . . . 339 8.3.1 Motorcycles for the Road (On-Road). . . . . . . . . . . 339 8.3.2 Off-Road Motorcycles. . . . . . . . . . . . . . . . . . . . . . . . 353 8.3.3 Legislation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 360 8.3.4 Racing Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 369 8.3.5 Special Applications. . . . . . . . . . . . . . . . . . . . . . . . . 375 8.4 Rotary Piston Engine/Wankel Engine . . . . . . . . . . . . . . 382 8.4.1 History . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 382 8.4.2 General Functionality of a Rotary Piston Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 383 8.4.3 Four-Stroke Principle. . . . . . . . . . . . . . . . . . . . . . . . 384 8.4.4 The Rotary Piston Engine of the Passenger car Renesis. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 384 8.4.5 Hydrogen Rotary Piston Engine. . . . . . . . . . . . . . . 386 Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 387

Internal Combustion Engine Handbook | xiii

6606_Book.indb 13

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

9 Tribology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 389 9.1 Friction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.1.1 Characterizing Features. . . . . . . . . . . . . . . . . . . . . . 9.1.2 Friction States. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.1.3 Methods of Measuring Friction. . . . . . . . . . . . . . . . 9.1.4 Influence of the Operating State and the Boundary Conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.1.5 Influence of Friction on the Fuel Consumption. . . 9.1.6 Friction Behavior of Previously Introduced Combustion Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.1.7 Method of Calculating the Friction on the Example of the Piston Group. . . . . . . . . . . . . . . . . . . . . . 9.2 Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.2.1 Tribological Principles . . . . . . . . . . . . . . . . . . . . . . . 9.2.2 Lubrication System. . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

389 389 389 390 391 393 394 403 404 404 406 412

10 Charge Cycle . . . . . . . . . . . . . . . . . . . . . . . . . 415 10.1 Gas Exchange Devices in Four-Stroke Engines. . . . . . 10.1.1 Valve Gear Designs. . . . . . . . . . . . . . . . . . . . . . . . . 10.1.2 Components of the Valve Train. . . . . . . . . . . . . . . 10.1.3 Kinematics and Dynamics of the Valve Train. . . 10.1.4 Design of Gas Exchange Devices in Four-Stroke Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.2 Calculating Charge Cycle. . . . . . . . . . . . . . . . . . . . . . . . 10.2.1 The Filling and Emptying Method. . . . . . . . . . . . 10.3 The Charge Cycle in Two-Stroke Engines . . . . . . . . . . 10.3.1 Scavenging. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.3.2 Gas Exchange Organs. . . . . . . . . . . . . . . . . . . . . . . 10.3.3 Scavenging Air Supply. . . . . . . . . . . . . . . . . . . . . . 10.4 Variable Valve Actuation. . . . . . . . . . . . . . . . . . . . . . . . . 10.4.1 Camshaft Timing Devices . . . . . . . . . . . . . . . . . . . 10.4.2 Systems With Stepped Variation of the Valve Stroke or Opening Time . . . . . . . . . . . . . . . . . . . . . . . . . . 10.4.3 Infinitely Variable Valve Actuation. . . . . . . . . . . 10.4.4 Perspectives for the Variable Valve Train . . . . . . 10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.5.1 Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.5.2 Technology Description. . . . . . . . . . . . . . . . . . . . . 10.5.3 Construction Principle and Boundary Conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10.5.4 Thermodynamic Potential. . . . . . . . . . . . . . . . . . . 10.5.5 Evaluation Summary. . . . . . . . . . . . . . . . . . . . . . . 10.5.6 Pulse Turbocharging With Controllable Aspiration Valves. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

416 416 418 423 425 436 437 439 439 441 443 444 446 453 456 469

471 471 471 472 472 478 479 483

11 Supercharging of Internal Combustion Engines . . . . . . . . . . . . . . . . . . . . 487 11.1 Mechanical Supercharging. . . . . . . . . . . . . . . . . . . . . . . 488 11.2 Exhaust Gas Turbocharging. . . . . . . . . . . . . . . . . . . . . . 489 11.2.1 Constant Pressure Turbocharging. . . . . . . . . . . . . 489

11.2.2 Ram Induction Turbocharging or Impulse Turbocharging. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.3 Intercooling. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.4 Interaction Between Engine and Compressor. . . . . . . 11.4.1 Four-Stroke Engine in Compressor Map. . . . . . . 11.4.2 Mechanical Supercharging. . . . . . . . . . . . . . . . . . . 11.4.3 Exhaust Gas Turbocharging. . . . . . . . . . . . . . . . . . 11.5 Dynamic Behavior. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.5.1 Improvement Measures. . . . . . . . . . . . . . . . . . . . . 11.5.2 Active Residual Gas Discharge. . . . . . . . . . . . . . . 11.5.3 Electric Support for Exhaust Gas Turbocharging. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.5.4 Mechanical Auxiliary Compressor. . . . . . . . . . . . 11.6 Additional Measures for Supercharged Internal Combustion Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.6.1 SI Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.6.2 Diesel Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.7 Performance Explosion by Register and Two-Stage Supercharging for Private Vehicles . . . . . . . . . . . . . . . . . . . 11.7.1 The History and Evolution of Two-Staged Turbocharger Processes . . . . . . . . . . . . . . . . . . . . . . . . . . 11.7.2 Thermodynamics of Two-Staged Superchargers. . . . . . . . . . . . . . . . . . . . . . . . 11.7.3 Sequential Turbocharger and Two-Stage Supercharger Concept/Supercharger Systems. . . . . . . 11.7.4 Applications Areas . . . . . . . . . . . . . . . . . . . . . . . . . 11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.8.1 Basic Design of a Turbocharger Test Rig. . . . . . . 11.8.2 Compressor and Turbine Maps. . . . . . . . . . . . . . . 11.8.3 Special Features When Using Turbocharger Engine Maps in the Engine Process Simulation . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

489 490 491 491 492 493 498 500 500 500 501 501 501 502 502 502 503 505 507 507 508 508 511 512

12 Mixture Format Ion and Related Systems . . . . . . . . . . . . . . . . . . . . . . . . 515 12.1 Internal Mixture Formation. . . . . . . . . . . . . . . . . . . . . . 12.2 External Mixture Formation. . . . . . . . . . . . . . . . . . . . . . 12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection) . . . . . . . . . . . . . . . . . . . . . . 12.3.1 Mode of Operation of the Carburetor . . . . . . . . . 12.3.2 Designs. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.3.3 Important Systems on Carburetors . . . . . . . . . . . 12.3.4 Electronically Controlled Carburetors. . . . . . . . . 12.3.5 Constant Vacuum Carburetor. . . . . . . . . . . . . . . . 12.3.6 Operating Behavior. . . . . . . . . . . . . . . . . . . . . . . . . 12.3.7 Lambda Closed-Loop Control. . . . . . . . . . . . . . . . 12.3.8 Mixture Formation by Utilizing Gasoline Injection. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.4 Mixture Formation in Diesel Engines. . . . . . . . . . . . . . 12.4.1 Injection Systems—An Overview. . . . . . . . . . . . . 12.4.2 Systems with Injection-Synchronous Pressure Generation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.4.3 Systems with a Central Pressure Reservoir. . . . .

515 515 516 516 517 518 521 522 522 523 523 532 534 537 545

xiv | Internal Combustion Engine Handbook

6606_Book.indb 14

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

12.4.4 Injection Nozzles and Nozzle-Holder Assemblies. . . . . . . . . . . . . . . . . . . . . . . . 12.4.5 Adapting the Injection System to the Engine. . . 12.5 Fuel Supply Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.5.1 Fuel Tanks. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12.5.2 The Tank Venting System. . . . . . . . . . . . . . . . . . . . 12.5.3 Requirements for the Fuel Supply System . . . . . 12.5.4 The Filling-Level Measuring. . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

551 555 557 557 560 560 566 567

13 Ignition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 569 13.1 Gasoline Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.1.1 Introduction to Ignition. . . . . . . . . . . . . . . . . . . . . 13.1.2 Requirements for the Ignition Systems. . . . . . . . 13.1.3 Minimum Ignition Energy. . . . . . . . . . . . . . . . . . . 13.1.4 Fundamentals of Spark Ignition. . . . . . . . . . . . . . 13.1.5 Coil Ignition System (Inductive). . . . . . . . . . . . . . 13.1.6 Other Ignition Systems. . . . . . . . . . . . . . . . . . . . . . 13.1.7 Summary/Outlook. . . . . . . . . . . . . . . . . . . . . . . . . 13.1.8 Spark Plug . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.2 Diesel Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.2.1 Autoignition and Combustion . . . . . . . . . . . . . . . 13.2.2 Cold Start Diesel Engine . . . . . . . . . . . . . . . . . . . . 13.2.2.1 Important Influence Parameters . . . . . . . . . . . . 13.2.3 Components for Supporting Cold Starts. . . . . . . 13.2.4 Outlook. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

569 569 569 569 569 570 572 573 573 578 578 579 579 581 585 586

14 Combustion . . . . . . . . . . . . . . . . . . . . . . . . . . 589 14.1 Fuels and Fuel Chemistry. . . . . . . . . . . . . . . . . . . . . . . . 14.2 Oxidation of Hydrocarbons. . . . . . . . . . . . . . . . . . . . . . 14.3 Autoignition. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.3.1 The H2–O2 System. . . . . . . . . . . . . . . . . . . . . . . . . . 14.3.2 Ignition of Hydrocarbons. . . . . . . . . . . . . . . . . . . 14.3.3 Rapid Compression Machines . . . . . . . . . . . . . . . 14.3.4 Diesel Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.3.5 HCCI Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.3.6 Engine Knock. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.3.7 Modeling the Autoignition . . . . . . . . . . . . . . . . . . 14.4 Flame Propagation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.4.1 Turbulent Scales . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.4.2 Flame Types. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.5 Model Development and Simulation. . . . . . . . . . . . . . 14.5.1 Classifications for Combustion Models. . . . . . . . 14.5.2 0D Models . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14.5.3 Phenomenological Models. . . . . . . . . . . . . . . . . . . 14.5.4 3D-CFD Models. . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

589 591 592 593 593 594 594 594 595 595 596 596 596 598 599 599 601 603 604

15 Combustion Systems . . . . . . . . . . . . . . . . . 607 15.1 Diesel Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.1.1 Diesel Combustion. . . . . . . . . . . . . . . . . . . . . . . . . 15.1.2 Diesel Four-Stroke Combustion Systems . . . . . . 15.2 Spark-Injection Engines. . . . . . . . . . . . . . . . . . . . . . . . .

607 607 612 623

15.2.1 Combustion Processes in Port Fuel Injection Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.2.2 Combustion Process of Direct Injection Spark Ignition Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.3 Combustion Process (Two-Stroke Engine). . . . . . . . . . 15.4 Two-Stroke SI Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

623 633 649 650 653

16 Electronics and Mechanics for Engine and Shift Control Transmission Management . . . . . . . . . . . . . . 657 16.1 Environmental Requirements . . . . . . . . . . . . . . . . . . . . 16.1.1 Classes of Installation. . . . . . . . . . . . . . . . . . . . . . . 16.1.2 Thermal Management . . . . . . . . . . . . . . . . . . . . . . 16.2 Stand-Alone Products. . . . . . . . . . . . . . . . . . . . . . . . . . . 16.3 Connection Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.4 Integrated Products (MTM). . . . . . . . . . . . . . . . . . . . . . 16.5 Electronic Design, Structures, and Components. . . . . 16.5.1 Basic Structure. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.5.2 Electronic Components . . . . . . . . . . . . . . . . . . . . . 16.6 Electronics in the Electronic Control Unit . . . . . . . . . . 16.6.1 General Description . . . . . . . . . . . . . . . . . . . . . . . . 16.6.2 Signal Processing. . . . . . . . . . . . . . . . . . . . . . . . . . . 16.6.3 Signal Evaluation . . . . . . . . . . . . . . . . . . . . . . . . . . 16.6.4 Signal Output. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.6.5 Power Supply . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.6.6 Interfaces . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.6.7 Electronics for Transmission ECUs. . . . . . . . . . . . 16.7 Software Structures. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.7.1 Task of the Software in Controlling Engines. . . . 16.7.2 Demands on the Software. . . . . . . . . . . . . . . . . . . 16.7.3 The Layer Approach to Software . . . . . . . . . . . . . 16.7.4 The Software Development Process. . . . . . . . . . . 16.8 Torque-Based Functional Structure for Engine Management. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.8.1 Model-Based Functions Using the Example of the Intake Manifold Model . . . . . . . . . . . . . . . . . . . . . 16.9 Functions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.9.1 λ Regulation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16.9.2 Anti-Jerk Function. . . . . . . . . . . . . . . . . . . . . . . . . . 16.9.3 Throttle Valve Vontrol . . . . . . . . . . . . . . . . . . . . . . 16.9.4 Knocking Control. . . . . . . . . . . . . . . . . . . . . . . . . . 16.9.5 On-Board Diagnosis. . . . . . . . . . . . . . . . . . . . . . . . 16.9.6 Safety Strategy. . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

657 657 658 661 663 664 665 665 666 672 672 672 672 672 673 673 674 676 676 676 677 678 678 681 683 683 685 687 688 690 692 694

17 The Powertrain . . . . . . . . . . . . . . . . . . . . . . . 695 17.1 Powertrain Architecture. . . . . . . . . . . . . . . . . . . . . . . . . 17.2 The Longitudinal Dynamics of the Vehicle . . . . . . . . . 17.3 Transmission Types. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.4 Power Level and Signal-Processing Level. . . . . . . . . . 17.4.1 Power Level (Figure 17.6) . . . . . . . . . . . . . . . . . . . 17.4.2 Signal Level (Figure 17.6). . . . . . . . . . . . . . . . . . . . 17.4.3 Links. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.5 Transmission Management. . . . . . . . . . . . . . . . . . . . . . .

695 696 696 698 698 698 698 698

Internal Combustion Engine Handbook | xv

6606_Book.indb 15

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

17.5.1 Functions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.6 Integrated Powertrain Management (IPM®). . . . . . . . 17.7 Components for Powertrain Electricification . . . . . . . 17.7.1 Overview . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.7.2 Hybrid and Electric Drive Variants . . . . . . . . . . . 17.7.3 Components. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.7.4 Power Electronics. . . . . . . . . . . . . . . . . . . . . . . . . . 17.7.5 Electric Motor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17.7.6 Energy Storage Systems. . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

698 701 702 702 703 704 704 705 706 709

18 Sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 711 18.1 Temperature Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.2 Fuel-Level Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.3 Knock Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.4 Exhaust Gas Sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.4.1 Lambda Sensors . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.4.2 NOx Sensor. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.5 Pressure Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.5.1 Normal Pressure Sensors. . . . . . . . . . . . . . . . . . . . 18.5.2 Medium-Pressure Sensors. . . . . . . . . . . . . . . . . . . 18.5.3 High-Pressure Sensors. . . . . . . . . . . . . . . . . . . . . . 18.5.4 Pressure Switch. . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.6 Air Mass Sensor. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.6.1 Measuring Principle. . . . . . . . . . . . . . . . . . . . . . . . 18.6.2 Mass Air-Flow Sensor. . . . . . . . . . . . . . . . . . . . . . . 18.6.3 Secondary Air Mass Sensors (SAF). . . . . . . . . . . 18.7 Speed Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.7.1 Passive Speed Sensors . . . . . . . . . . . . . . . . . . . . . . 18.7.2 Active Sensors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18.8 Combustion Chamber Pressure Sensors for Diesel Engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

711 712 712 713 713 714 715 715 716 716 716 717 717 717 718 718 718 718 719 720

19 Actuators . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 721 19.1 Drive. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19.1.1 Pneumatic Drives . . . . . . . . . . . . . . . . . . . . . . . . . . 19.1.2 Electric Drives. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19.1.3 Communication with Engine Control Electronics. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19.1.4 Reset/Default Position. . . . . . . . . . . . . . . . . . . . . . 19.2 Throttle Valve Actuators. . . . . . . . . . . . . . . . . . . . . . . . . 19.2.1 Key Function in SI Engines. . . . . . . . . . . . . . . . . . 19.2.2 Key Function in Diesel Engines . . . . . . . . . . . . . . 19.2.3 Additional Functions . . . . . . . . . . . . . . . . . . . . . . . 19.2.4 “Drive-by-Wire”/E-Gas. . . . . . . . . . . . . . . . . . . . . 19.2.5 Waste Gate Function. . . . . . . . . . . . . . . . . . . . . . . . 19.2.6 Vacuum/Prethrottle Actuators. . . . . . . . . . . . . . . 19.3 Swirl and Tumble Plates, Resonance Charging . . . . . 19.3.1 Port Deactivation. . . . . . . . . . . . . . . . . . . . . . . . . . . 19.3.2 Stratified Charge. . . . . . . . . . . . . . . . . . . . . . . . . . . 19.4 Turbocharger with Variable Turbine Geometry. . . . . . 19.5 Exhaust Gas Recirculation Valves. . . . . . . . . . . . . . . . . 19.6 Evaporation Emissions, Components. . . . . . . . . . . . . . 19.6.1 Tank Ventilation Valves . . . . . . . . . . . . . . . . . . . . .

721 721 722 722 723 723 723 723 724 725 726 726 726 726 727 728 728 730 730

19.6.2 Evaporative Emissions Diagnostics. . . . . . . . . . . 731 Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 733

20 Cooling of Internal Combustion Engines . . . . . . . . . . . . . . . . . . . . 735 20.1 General Information . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20.2 Demands on the Cooling System. . . . . . . . . . . . . . . . . . 20.3 Principles for Calculation and Simulation Tools. . . . . 20.4 Engine Cooling Subsystems. . . . . . . . . . . . . . . . . . . . . . 20.4.1 Cooling by Coolant. . . . . . . . . . . . . . . . . . . . . . . . . 20.4.2 Intercooling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20.4.3 Exhaust Gas Cooling . . . . . . . . . . . . . . . . . . . . . . . 20.4.4 Oil Cooling. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20.4.5 Fans and Fan Drives. . . . . . . . . . . . . . . . . . . . . . . . 20.5 Cooling Modules. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20.6 Overall Engine Cooling System. . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

735 735 736 737 737 740 740 741 741 742 743 744

21 Exhaust Emissions . . . . . . . . . . . . . . . . . . . . 745 21.1 Legal Requirements. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.1.1 Europe. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.1.2 Japan. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.1.3 Harmonization of the Exhaust Gas Regulations. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.2 Exhaust Measuring Technology. . . . . . . . . . . . . . . . . . . 21.2.1 Measuring Technology for the Certification of Motor Vehicles. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.2.2 Measuring Technology for Engine Development. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.3 Pollutants and Their Origin. . . . . . . . . . . . . . . . . . . . . . 21.3.1 Gasoline Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.3.2 Diesel Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.4 Reduction in Pollutants. . . . . . . . . . . . . . . . . . . . . . . . . . 21.4.1 Engine-Related Approaches . . . . . . . . . . . . . . . . . 21.5 Exhaust Treatment for Gasoline Engines . . . . . . . . . . . 21.5.1 Catalytic Converter Designand Chemical Reactions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.5.2 Catalytic Converter Concepts of Engines with Stoichiometric Operation. . . . . . . . . . . . . . . . . . . . . 21.5.3 Catalytic Converter Concept for Lean-Burn Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21.5.4 Metallic Catalytic Converter Substrate . . . . . . . . 21.6 Exhaust Treatment in Diesel Engines. . . . . . . . . . . . . . 21.6.1 Diesel Oxidation Catalytic Converters . . . . . . . . 21.6.2 NOx Adsorbers for Diesel Passenger Cars. . . . . . 21.6.3 Particles/Particle Filters. . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

745 745 746 746 746 746 748 752 753 754 756 756 760 760 761 767 775 782 782 786 788 809

22 Operating Fluids . . . . . . . . . . . . . . . . . . . . . 813 22.1 Fuels. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.1.1 Diesel Fuel (DK) . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.1.2 Gasoline . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.2 Lubricants. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.2.1 Lubricant Types. . . . . . . . . . . . . . . . . . . . . . . . . . . .

813 814 824 842 842

xvi | Internal Combustion Engine Handbook

6606_Book.indb 16

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

22.2.2 Tasks of Lubrication. . . . . . . . . . . . . . . . . . . . . . . . 22.2.3 Lubrication Types. . . . . . . . . . . . . . . . . . . . . . . . . . 22.2.4 Lubrication Requirements. . . . . . . . . . . . . . . . . . . 22.2.5 Viscosity/viscosity Index (V.I.). . . . . . . . . . . . . . . 22.2.6 Basic Liquids. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.2.7 Additives for Lubricants . . . . . . . . . . . . . . . . . . . . 22.2.8 Engine Oils for Four-Stroke Engines. . . . . . . . . . 22.2.9 Engine Oils for Two-Stroke Engines. . . . . . . . . . . 22.3 Coolant. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.3.1 Frost Protection. . . . . . . . . . . . . . . . . . . . . . . . . . . . 22.3.2 Corrosion Protection. . . . . . . . . . . . . . . . . . . . . . . . 22.3.3 Specifications. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

842 842 842 843 845 846 848 858 860 860 860 861 861

23 Filtration of Operating Fluids . . . . . . . . . 863 23.1 Air Filter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.1.1 The Importance of Air Filtration for Internal Combustion Engines. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.1.2 Impurities in Engine Intake Air . . . . . . . . . . . . . . 23.1.3 Data for Air Filter Medium Assessment . . . . . . . 23.1.4 Measuring Methods and Analysis. . . . . . . . . . . . 23.1.5 Demands on Modern Air Filter Systems. . . . . . . 23.1.6 Design Criteria for Engine Air Filter Elements. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.1.7 Filter Housings . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.2 Fuel Filters. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.2.1 Gasoline Filters . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.2.2 Diesel Fuel Filters. . . . . . . . . . . . . . . . . . . . . . . . . . 23.2.3 Performance Data of Fuel Filters . . . . . . . . . . . . . 23.3 Engine Oil Filtration . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23.3.1 Wear and Filtration. . . . . . . . . . . . . . . . . . . . . . . . . 23.3.2 Full-Flow Oil Filters. . . . . . . . . . . . . . . . . . . . . . . . 23.3.3 Removal Efficiency and Filter Fineness. . . . . . . . 23.3.4 Bypass Oil Filtration. . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

863 863 863 863 864 865 865 866 866 866 867 870 870 870 871 872 873 874

24 Calculation and Simulation . . . . . . . . . . . 875 24.1 Strength and Vibration Calculation. . . . . . . . . . . . . . . . 24.1.1 Procedures and Methods. . . . . . . . . . . . . . . . . . . . 24.1.2 Selected Examples of Application . . . . . . . . . . . . 24.1.3 Piston Calculations. . . . . . . . . . . . . . . . . . . . . . . . . 24.2 Flow Calculation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24.2.1 One- or Quasi-Dimensional Methods . . . . . . . . . 24.2.2 3-D Flow Calculation. . . . . . . . . . . . . . . . . . . . . . . 24.2.3 Selected Examples of Application . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

875 875 877 879 889 889 891 893 898

25 Combustion Diagnostics— Indication and Visualization in Combustion Development . . . . . . . . . . . . . . . 901 25.1 Discussion. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.2 Indication. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.2.1 Measuring Systems. . . . . . . . . . . . . . . . . . . . . . . . . 25.2.2 Quality Criteria. . . . . . . . . . . . . . . . . . . . . . . . . . . .

901 901 902 904

25.2.3 Indicating—Prospects. . . . . . . . . . . . . . . . . . . . . . . 25.2.4 Cycle-Precise Signal- and Model-Based Engine Control. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.3 Visualization. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.3.1 Functions and Discussion . . . . . . . . . . . . . . . . . . . 25.3.2 Visualization Methods for Real Engine Operation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25.3.3 Visualization of Combustion in Real Engine Operation by the Flame’s Intrinsic Luminescence . . . . 25.3.4 Visualization of Illuminated Processes . . . . . . . . 25.3.5 Visualization: The Future. . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

905 905 906 906 906 908 914 915 916

26 Fuel Consumption . . . . . . . . . . . . . . . . . . . . 917 26.1 General Influencing Factors. . . . . . . . . . . . . . . . . . . . . . 26.1.1 Aerodynamic Drag. . . . . . . . . . . . . . . . . . . . . . . . . 26.1.2 Weight. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.1.3 Wheel Resistance. . . . . . . . . . . . . . . . . . . . . . . . . . . 26.1.4 Fuel Consumption . . . . . . . . . . . . . . . . . . . . . . . . . 26.2 Engine Modifications. . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.1 Downsizing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.2 Downspeeding. . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.3 Diesel Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.4 Gasoline Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.5 Combustion Process HCCI . . . . . . . . . . . . . . . . . . 26.2.6 Variable Valve Train . . . . . . . . . . . . . . . . . . . . . . . . 26.2.7 Cylinder Shutoff . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.8 Auxiliary Units . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.2.9 Heat Management Methods for Achieving Fuel Consumption Reductions. . . . . . . . . . . . . . . . . . . . . 26.2.10 Hybrid Concepts. . . . . . . . . . . . . . . . . . . . . . . . . . 26.3 Transmission Ratios. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.3.1 Selection of Direct Drive . . . . . . . . . . . . . . . . . . . . 26.3.2 Selection of Overall Transmission Ratio in the Highest Gear. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.4 Driver Behavior . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.5 CO2 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.5.1 CO2 Emissions and Fuel Consumption. . . . . . . . 26.5.2 The Influence of Engine Use on CO2 Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26.5.3 The Trend in Global CO2 Emissions. . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

918 918 919 920 921 922 922 924 925 926 928 929 931 932 934 934 936 937 937 939 940 941 943 943 943

27 Noise Emissions . . . . . . . . . . . . . . . . . . . . . . 945 27.1 Basic Physical Principles and Terms. . . . . . . . . . . . . . . 27.2 Legal Provisions Concerning Emitted Noise. . . . . . . . 27.2.1 Emitted Noise Measuring Procedure. . . . . . . . . . 27.2.2 Critical Evaluation of the Meaningfulness of the Existing Emission Noise Measuring Procedure. . . 27.2.3 Future Emission Noise Measuring Procedure and Limit Values . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.3 Sources of Emitted Noise . . . . . . . . . . . . . . . . . . . . . . . . 27.4 Emitted Noise-Reduction Provisions . . . . . . . . . . . . . . 27.4.1 Engine Approaches. . . . . . . . . . . . . . . . . . . . . . . . . 27.4.2 Vehicle Approaches. . . . . . . . . . . . . . . . . . . . . . . . .

945 948 948 948 948 949 950 950 950

Internal Combustion Engine Handbook | xvii

6606_Book.indb 17

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

27.5 Engine Noise in the Vehicle Interior. . . . . . . . . . . . . . . 27.6 Acoustic Guidelines for the Engine Designer . . . . . . . 27.7 Measuring and Analytical Methods. . . . . . . . . . . . . . . 27.8 Psychoacoustics. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.9 Sound Engineering . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.10 Simulation Tools. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27.11 Antinoise Systems: Noise Reduction using Antinoise. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

952 953 954 957 957 958 959 959

28 Engine Measuring Technology . . . . . . . . 961 28.1 Measuring Technology in the Test Bench Overall System. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.1.1 Dynamometers . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.2 Mechanical Measurands. . . . . . . . . . . . . . . . . . . . . . . . . 28.3 Thermodynamic:Measurands . . . . . . . . . . . . . . . . . . . . 28.4 Flow Measuring Technology . . . . . . . . . . . . . . . . . . . . . 28.5 Intake Air Volumetric Flow Measurement. . . . . . . . . . 28.6 Fuel Consumption Measurement . . . . . . . . . . . . . . . . . 28.7 Fuel Conditioning . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.8 Oil Consumption Measurement Technology. . . . . . . . 28.9 Blow-By Measuring Technology. . . . . . . . . . . . . . . . . . 28.10 Urea Consumption Measuring Technology. . . . . . . . 28.11 Direct Exhaust Quantity Measurement. . . . . . . . . . . . 28.12 Exhaust Gas Measuring Technology. . . . . . . . . . . . . . 28.13 Measurement of the Volumetric Concentration of Gaseous Exhaust Gas Components. . . . . . . . . . . . . . . . . . . . 28.14 Pollutant Mass Determination at Undiluted Exhaust Gas. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.15 Pollutant Mass Determination at Diluted Exhaust Gas. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.16 Exhaust Gas Particulate Measuring Technology. . . . 28.16.1 Gravimetric Particulate Measurement. . . . . . . . 28.16.2 Dynamic Particulate Measurement . . . . . . . . . . 28.16.3 Measurement of the Soot or Smoke Emission. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.16.4 Smoke Emission Value Measurement with “Smoke Meter”. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.16.5 Opacity Measurement Using “Opacimeters”. . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28.16.6 Photoacoustics (PASS) . . . . . . . . . . . . . . . . . . . . . 28.16.7 Laser-Induced Incandescence. . . . . . . . . . . . . . . 28.16.8 Scattered Light Measurement. . . . . . . . . . . . . . . 28.16.9 Particle Group and Particle Total. . . . . . . . . . . . 28.16.10 Measurement Data Processing and Evaluation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

961 962 963 964 965 965 965 968 968 969 970 972 972 973 973 974 974 974 975 975 975 975 975 976 976 976 976 977

29 Hybrid Drive Systems . . . . . . . . . . . . . . . . 979 29.1 History. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29.2 Fundamentals of Hybrid Drive Systems (General Overview). . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29.2.1 Principle. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29.2.2 Components. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

979 983 983 983

29.2.3 Functions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 984 29.3 Classification of Hybrid Drive Systems. . . . . . . . . . . . 986 29.3.1 Types. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 986 29.3.2 Power Classification. . . . . . . . . . . . . . . . . . . . . . . . 987 29.4 Electrical Hybrid Drive Systems. . . . . . . . . . . . . . . . . . 988 29.4.1 Electrical Machine. . . . . . . . . . . . . . . . . . . . . . . . . . 988 29.4.2 Power Range . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 995 29.4.3 Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 995 29.4.4 Power Electronics. . . . . . . . . . . . . . . . . . . . . . . . . . 995 29.4.5 Current Converter. . . . . . . . . . . . . . . . . . . . . . . . . . 995 29.5 Energy Storage Devices. . . . . . . . . . . . . . . . . . . . . . . . . . 996 29.5.1 Lead-Acid Battery. . . . . . . . . . . . . . . . . . . . . . . . . . 998 29.5.2 Nickel–Metal Hydride Battery. . . . . . . . . . . . . . . 998 29.5.3 Sodium–Nickel-Chloride Battery. . . . . . . . . . . . . 999 29.5.4 Lithium-Ion Battery . . . . . . . . . . . . . . . . . . . . . . . 1000 29.5.5 SuperCaps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1001 29.5.6 Battery Management . . . . . . . . . . . . . . . . . . . . . . 1003 29.6 Hybrid Drive System Transmissions. . . . . . . . . . . . . . 1005 29.6.1 Transmissions Without Integrated Electrical Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1006 29.6.2 Transmissions with Integrated Electrical Machine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1007 29.6.3 Special Transmission Designs. . . . . . . . . . . . . . . 1008 29.7 Energy Management System. . . . . . . . . . . . . . . . . . . . 1010 29.7.1 Start/Stop. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1011 29.7.2 Regulating the Generator. . . . . . . . . . . . . . . . . . . 1012 29.7.3 Energy Recuperation . . . . . . . . . . . . . . . . . . . . . . 1012 29.7.4 SOC Regulation. . . . . . . . . . . . . . . . . . . . . . . . . . . 1013 29.7.5 Energy Distribution Management . . . . . . . . . . . 1013 29.7.6 On-Board Power Supply System . . . . . . . . . . . . 1014 29.8 Operating Strategies . . . . . . . . . . . . . . . . . . . . . . . . . . . 1014 29.8.1 Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1014 29.8.2 Energy Balance . . . . . . . . . . . . . . . . . . . . . . . . . . . 1016 29.8.3 Fuel Consumption . . . . . . . . . . . . . . . . . . . . . . . . 1016 29.8.4 Exhaust Emissions . . . . . . . . . . . . . . . . . . . . . . . . 1016 29.8.5 Driving Performance . . . . . . . . . . . . . . . . . . . . . . 1017 29.8.6 Approaches for Determining an Operating Strategy. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1017 29.9 Current Hybrid Vehicles. . . . . . . . . . . . . . . . . . . . . . . . 1017 29.9.1 Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1018 29.9.2 Vehicle Design. . . . . . . . . . . . . . . . . . . . . . . . . . . . 1021 29.10 Future Development. . . . . . . . . . . . . . . . . . . . . . . . . . 1024 29.10.1 Gasoline Hybrid Drive System. . . . . . . . . . . . . 1024 29.10.2 Diesel Hybrid Drive System. . . . . . . . . . . . . . . 1024 29.10.3 Electric-Only Drive. . . . . . . . . . . . . . . . . . . . . . . 1025 Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1025

30 Alternative Vehicle Drives and APUs (Auxiliary Power Units) . . . . . . . . . . . . . . . . . 1029 30.1 Reasons Behind Alternatives. . . . . . . . . . . . . . . . . . . . 1029 30.2 Hybrid Vehicles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1030 30.2.1 Fuel Consumptions in the Official Test Mode . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1033

xviii | Internal Combustion Engine Handbook

6606_Book.indb 18

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Contents

30.2.2 Practical Fuel Consumption Standard Test Route: AMS. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.2.3 Plug-In Hybrid Vehicles. . . . . . . . . . . . . . . . . . . . 30.3 Electric Drive System. . . . . . . . . . . . . . . . . . . . . . . . . . . 30.4 Energy Storage Devices. . . . . . . . . . . . . . . . . . . . . . . . . 30.5 Stirling Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.6 Gas Turbine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.7 The Fuel Cell as a Vehicle Drive System. . . . . . . . . . . 30.7.1 Design of the PEM Fuel Cell. . . . . . . . . . . . . . . . 30.7.2 The Fuel Cell in the Vehicle. . . . . . . . . . . . . . . . . 30.7.3 Assessment of the Fuel Cell in Comparison to other Drive Systems . . . . . . . . . . . . . . . . . . . . . . . . . . 30.8 Synoptic Evaluation of Alternative Energies and Drive Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.9 Generating Electricity using an Auxiliary Power Unit = APU. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.9.1 Thermoelectrics. . . . . . . . . . . . . . . . . . . . . . . . . . . 30.9.2 The Fuel Cell as an APU. . . . . . . . . . . . . . . . . . . 30.9.3 Internal Combustion Engine in Combination with a Linear Generator (Free-Piston Linear Generator). . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Further Literature on the Subject of Hybrids and Fuel Cells . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1033 1036 1036 1039 1040 1041 1042 1042 1043 1046 1047 1047 1047 1047

31 Energy Management in the Engine and Vehicle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1053 31.1 Losses During Energy Conversion . . . . . . . . . . . . . . . 31.2 Requirement-Based Energy Management. . . . . . . . . 31.3 Generating Electricity in Vehicles . . . . . . . . . . . . . . . . 31.3.1 Thermoelectric Generator (TEG) . . . . . . . . . . . . 31.4 Heat Management. . . . . . . . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1055 1055 1056 1057 1059 1060

32 Forecast . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1061 32.1 Gasoline Engine. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32.2 Diesel Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32.3 Concluding Observations. . . . . . . . . . . . . . . . . . . . . . . Bibliography. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1061 1062 1063 1064

Color Section . . . . . . . . . . . . . . . . . . . . . . . . . . . 1065 Index . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1109

1050 1050

About the Editors . . . . . . . . . . . . . . . . . . . . . . 1129

1051

Internal Combustion Engine Handbook | xix

6606_Book.indb 19

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 20

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

1 Historical Review Motor vehicles have been built for more than a century. The advancements in vehicle appearance, even to the technical layman, are astonishing. Advancements in basic engine appearance, on the other hand, have been relatively minimal: the similarity in dimensions and layout (and a few other details) between engines of the past and current models hide just how much has also been done in engine technology over the years (Figure 1.1). The origins of motor vehicle engines lie ultimately in the needs of the craftsmen and small traders who could not afford the expensive and complex steam engines as power generators. The costly steam engines were subject to strict

DaimlerPhoenix-Engine

regulations and were primarily owned by larger companies who could afford them. Thus, the first internal combustion engines (gasoline-powered stationary motors for driving machines of all kinds) were produced because of the need for an affordable and simple source of power. Work on such drive systems had been done in various parts of the world. In 1876, Nikolaus August Otto successfully implemented the four-stroke process patented by the Frenchman Beau de Rochas. This engine had a decisive advantage that when compared with the gasoline engines already being built by the Frenchman Jean Joseph Etienne Lenoir, it utilized precompression of the mixture. The British

1899

1914 4 cyl. gasoline d = 100 mm s = 140 mm n = 660 1/min P = 8,8 kW

De Dion

4 cyl. gasoline d = 110 mm s = 150 mm

 Figure 1.1 Engines 1899–1998 [1-10]. 

(continues)

Internal Combustion Engine Handbook | 1

6606_Book.indb 1

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 1 Historical Review

Wanderer W 24

1937

1958 MercedesBenz

6 cyl. gasoline d = 80 mm s = 72,8 mm n = 4800 1/min P = 70 kW

4 cyl. gasoline d = 75 mm s = 100 mm n = 3500 1/min P = 31 kW 

1984

1998

Mercedes-Benz M 102

Opel 1,8 l

4 cyl. gasoline d = 89,0 mm s = 80,25 mm n = 5200 1/min P = 77 kW

4 cyl. gasoline d = 80,5 mm s = 888,2 mm n = 5400 1/min P = 85 kW 

Figure 1.1 Engines 1899–1998 [1-10]. (Continued)

2 | Internal Combustion Engine Handbook

6606_Book.indb 2

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Historical Review

engineer Dougald Clerk “shortened” the four-stroke process to the two-stroke process by eliminating the charge cycle strokes. In 1886, Karl Benz and Gottlieb Daimler (with Wilhelm Maybach ) simultaneously (and independently) developed the light, high-speed engine from which most modern gasoline engines would descend. Similar engines would also power airships and airplanes in the years that followed. Rudolf Diesel’s “rational heat engine” 1893–1897 was initially only used stationary; this also applies to its forerunner, the engines of George Bailey Brayton and Herbert Akroyd Stuart. It would take centuries before the diesel engine “would enter street traffic.” The fundamental design of the internal combustion engine was duplicated from the steam engine: the crank drive controls the sequence of the thermodynamic process and converts the vapor pressure first into an oscillating and then into a rotary movement. The high development level of the steam engine at the end of the nineteenth century formed the foundation for the engines. The level of mastery in casting, forging, and precise machining of automotive components also increased as a result of the steam engine. It was the one-piece self-tensioning piston ring from John Ramsbottome (1854) that enabled the high working pressures in the combustion chamber of internal combustion engines to be maintained. The piston ring was, therefore, just as much a precondition for the control of the engine process as the knowledge and experience of engine bearings and their lubrication. One of the initial developmental issues with the internal combustion engine was a question of presenting central engine functions. The most difficult problem of the early engines was the ignition. The flame ignition (Otto) and uncontrolled glow tube ignition (Maybach/Daimler) presented an obstacle to the engine development that was overcome only with the advent of electric ignition methods. These ignition types included snapper ignition (Otto), vibrator ignition (Benz), the Bosch magnetic low-voltage ignition with contact-breaking spark, and, finally, the high-voltage magnetic ignition (Bosch) [1.1]. Next, the quality and quantity of the mixture formation had to be improved. Wick-surface and brush carburetors allowed only the low-boiling fractions of the gasoline (final boiling point approximately 100°C) to be used. The fuel particles that could be used did not vaporize simultaneously, creating another problem. In the Wilhelm Maybach nozzle carburetor, the fuel was atomized and no longer “vaporized.” Now, it was possible to use a higher percentage of the gasoline (final boiling point approximately 200°C) productively. The spectrum of fuels that could be used was significantly extended. In particular, the mixture could be formed in practically any quantity (a precondition for a further increase in performance and power). Carburetors with automatic auxiliary air control from Krebs, Claudel (Zenith), as well as Menesson and Goudard (Solex) improved the operating behavior of the engines and reduced the fuel consumption. With the increase in power, more heat had to be dissipated with the coolant. Now, it was the simple evaporation cooling that proved to be the power-limiting factor. Heat dissipation

was too low with the cooling system of the time. Critical components could not be adequately and reliably cooled with a natural water circulation (thermal siphon), creating another problem. It required a large amount of water to be stored (and transported) on the vehicle to work effectively. The Wilhelm Maybach honeycomb cooler offered the physically “workable” solution that allowed for the intensification of the heat transfer on the side of the weak thermal transition (on the air side). Once these basics had been established on the engine side of vehicle technology, the motor vehicle industry developed rapidly. Advances on the engine side inspired the advances on the vehicle side (and vice versa). More and more companies took up the production of motor vehicles and engines. To increase power and enhance the smoothness of running, the number of cylinders was increased—from one to two and then to four, as in the Mercedes Simplex engine. The splitting of the combustion chamber into several cylinders enabled higher speeds and a better utilization of the combustion chamber, that is, higher specific work (effective mean pressure). The construction of motor vehicles and engines had also started in other countries (France, Italy, England, and, later, the United States). These engines were initially modeled using the German design, but soon other manufacturers began to create their own designs. The engine technology enjoyed an enormous boost from the aircraft development, from which the motor vehicle engines also benefited. Experience was shared so that the errors made in the aircraft engine development (and recognized as such) could be avoided from the outset in the motor vehicle engines. Nevertheless, there was competition among several drive concepts, including the technically mature steam engine. This design had benefits as a power source for road vehicles, including the fact that the engine was self-starting, had an elastic operating curve to match the required tractive power of the vehicle, and was smooth running. The electric drive appeared to offer even greater benefits, but the disadvantages of this drive concept quickly became apparent. As the engine power increased, so did the speed and weight of the vehicles. Now, it was a question of adapting the engine functions such as mixture composition, ignition timing, lubrication, and cooling to the conditions of road operation. The complex technical system engine had to be made controllable even for untrained personnel (namely, the vehicle owner). Fuel and oil consumption had to be reduced, the latter not only for cost reasons but also because the exhaust gases enriched with fully and partially combusted oil were a cause of public annoyance. This mixture of demands, faults, experience, and new findings led to the development of engine concepts with different but also with similar design elements. W-type, radial-type, single-shaft reciprocating piston, and rotary piston engines were only occasionally built for motor vehicles. The standard design was the inline engine with four, six, and eight cylinders. V-engines with eight, twelve, or even sixteen cylinders were also built. The “typical” engine consisted of a low crankcase with mounted single or twin cylinders. The cylinder and cylinder

Internal Combustion Engine Handbook | 3

6606_Book.indb 3

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 1 Historical Review

head were cast in one piece, and the uptight valves were driven by the cam-shaft(s) mounted low in the crankcase. The crankshaft was suspended in bearing brackets with bearings after only every second or even third throw. Although the automatic intake valves had been replaced by driven valves, the valve timing still presented several problems: valves burned through, valve springs broke, and the noise level became high. For this reason, the smooth running Knight slide valve gear appeared to be superior at the time. Knight sleeve valve engines were built in England by Daimler Co., in Belgium by Minerva, in the United States by Willys, and in Germany by Daimler-Motoren-Gesellschaft. But ultimately the valve timing system with its simpler design and operation was preferred. In the United States, the personal vehicle changed from a leisure pastime of the wealthy to an article of daily use before World War I. In 1909, Henry Ford started the production of the Model T (Tin Lizzie). By 1927, more than fifteen million of these vehicles had been manufactured [1-4]. In Europe, the widespread use of motor vehicles (predominantly commercial vehicles) started during World War I. The mass production necessitated a certain unification and standardization of parts. Operation under the extreme conditions at the front mercilessly revealed design errors. The operation, maintenance, and repair of so many vehicles necessitated the training and qualification of the operating personnel. The development of the aircraft engines driven by the war gave powerful impetus to the improvement of motor vehicle engines in the early 1920s, and this applies to both the design (basic construction) and the details of individual parts. Alongside upright valves with L- and T-shaped cylinder heads, engines with suspended valves and compact combustion chambers were built enabling higher compression ratios—a precondition for more power and lower consumption. With the piston competition of 1921 organized by the German Imperial Ministry of Transport, the German engine industry quickly discovered the benefits of the light alloy piston compared with the cast iron piston. As a result, the engines of the 1920s were changed to light alloy pistons. In spite of many setbacks, this resulted in a significant increase in power and efficiency. The controlled piston enabled piston knock to be reduced and ultimately eliminated. In the early 1920s, there had been significant problems with the conrod bearings of the aircraft engines; they had reached the limits of their load bearing capacity. The steel leaded bronze bearing, developed by Norman Gilmann at Allison (USA), provided the remedy. These bearings were first used in diesel engines for commercial vehicles and later in high-performance car engines. The next step in development was the three-material bearing, consisting of a steel supporting shell, a leaded bronze intermediate layer, and a babbitt metal running layer; they had been developed by Clevite in the United States. Higher speeds and increased demands on the reliability of the engines required better engine lubrication. This development advanced from wick and pot lubrication (lubrication from storage vessels) and lubrication with hand pumps [1.9]. Consumers were supplied with lubricant, and by

immersion of engine parts or by special scoop mechanisms, they were able to lubricate various components. This solution was followed by the forced circulation lubrication as was the method commonly used in aircraft engines. Two-stroke engines operated with mixture lubrication, that is, by adding oil to the fuel. Thermal siphon cooling did not allow sufficient heat to be dissipated from the parts subject to high thermal loads. As a result, forced circulation cooling was introduced. Piston knock had become a power-limiting criterion in gasoline engines even during World War I. In 1921, Thomas Midgley, Jr., and T.A. Boyd in the United States discovered the effectiveness of tetraethyl lead (TEL) as an “antiknock additive.” The addition of TEL to the fuel reduced the knock, permitted higher compression ratios, and resulted in higher efficiencies. In the 1920s, many small automobiles were developed whose engines had to be light, simple, and arid cheap. The two-stroke system with its high power density was an obvious choice. There were two mutually exclusive arguments in favor of this solution: high power density and design simplicity. Valveless two-stroke engines with crankcase scavenging were suitable for motorbikes and small automobiles. The development of Schnürle reverse or loop scavenging from DKW was an important advancement compared with the cross-flow scavenging method because it permitted better scavenging of the cylinder. This method also enabled flat pistons to replace stepped pistons (with high thermal load). The “Roaring Twenties” heralded the era of the “great” Mercedes, Horch, Stahr, and Maybach with eight-cylinder inline and twelve-cylinder V-engines. In England, there were Rolls Royce, Bentley, and Annstrong-Siddeley, in France Delage and Bugatti, and in the United States, Pierce Arrow, Duesenberg, Auburn, Cord, Cadillac, and Packard. Influenced by the development in aircraft engine construction, the engine builders started to turbocharge the engines with displacement-type fans (Roots blowers) that could be switched on and off, depending on the power requirements (Mercedes-Benz, Itala and Bentley—radial blower (turbocompressor): Duesenberg). The air cooling of the aircraft engines also appeared to offer benefits, but this proved to be far more difficult with motor vehicle engines because of the low vehicle speed and less favorable operating conditions. A pioneer of air cooling was the Franklin Mfg. Co. from the United States. This company manufactured an air-cooled six-cylinder inline engine even before World War I. General Motors also tried air cooling with a Chevrolet (Chevrolet copper engine), where the cooling fins were made of copper to improve the heat dissipation. Because of technical problems, however, this engine never went into mass production. In Europe, air-cooled motor vehicle engines were also developed and built in the 1920s and 1930s: commercial vehicle engines from Krupp and Phanomen, and car engines from Tatra and Ferdinand Porsche for the new Volkswagens were produced. The aircooled opposed-cylinder (boxer) engine from Volkswagen became a synonym for reliability arid sturdiness (first in the jeep and the amphibian vehicle and later in the “Beetle”).

4 | Internal Combustion Engine Handbook

6606_Book.indb 4

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Historical Review

In the 1920s, a highly efficient accessory industry was built up in symbiosis with the automotive and engine industry. It served as a development center that united not only knowledge and experience in the various areas but also enabled more cost-effective production. This industry produced for several (or even all) of the engine manufacturers and thus was able to offer proven, more or less standardized, and inexpensive accessories such as pistons, bearings, radiators, carburetors, electrical equipment, and diesel injection systems. The motor vehicle development promoted and enhanced the construction and expansion of long-distance highways. Better roads permitted higher speeds and wheel loads. The traffic density increased slowly but surely. The operation of the engines was simplified, particularly by the electric starter introduced by Charles F. Kettering at General Motors that made starting not only easier but also safer. Ignition timing (advance-retard) and mixture composition (lean-rich) no longer had to be adjusted by the driver and were controlled automatically. In the 1930s, cars were increasingly driven during the winter months. Up to this point, many cars had not been used in winter. The year-round operation of vehicles required different oils depending on the outside temperature (i.e., summer oil–winter oil). Consideration had to be given to the outdoor temperatures by controlling the coolant temperature, first by covering the radiator with leather blankets, then by using adjustable radiator shutters, and finally by using a thermostat to control the coolant temperature. In the 1930s, alternative concepts were developed for vehicle engines. In Europe, the steam engine was used in commercial vehicles (Foden, Sentinel, Leyland, and Henschel) to cut fuel costs and to achieve higher power outputs than were possible at the time with vehicle diesel engines. Even the thought of a cost-effective independent operation played a role in the development of these engines. In the United States, Doble automobiles powered by steam engines had become known for their quiet running. Despite the favorable tractive force curve, the steam engine ultimately failed to assert itself against the internal combustion engine. Commercial vehicles were operated with gas from an accumulator or with generator gas. During World War II and in the time period thereafter, automobile engines had to be converted to generator gas because of the shortage of fuel (Figure 1.2). Fuel injection using compressed air (“air injection”) had been an obstacle to the use of diesel engines in the motor vehicle. In the early 1920s, intensive work on a “compressorless (airless) injection” was carried out in various areas. Based on the preliminary work conducted before and during World War I (L’Orange and Leissner) (airless) diesel engines for motor vehicles were developed—in Germany by MAN, Benz (later, Mercedes-Benz), and Junkers. On the basis of Acro patents, Robert Bosch developed complete fuel injection systems for vehicle diesel engines [1-1]. The fuel injection pumps had helix and overflow control. But because direct fuel injection had not been mastered for motor vehicle engines with their wide speed range, indirect injection (prechamber and whirl chamber, air accumulator) was preferred. The diesel engine

proved to be effective in heavy commercial vehicles and was increasingly used in light commercial vehicles and, ultimately, also in the automobile (Mercedes-Benz, Hanomag, Oberhansli, Colt, Cummins, etc.). One of the first automobiles with a diesel engine was a Packard with a Cummins engine. To demonstrate the suitability of the diesel engines for cars, specially modified vehicles were entered in races. In 1930, a Packard Roadster powered by a Cummins diesel engine achieved a speed of 82 miles/h (131 km/h) on the Daytona Beach racetrack in Florida. In Germany, a Hanomag streamlined vehicle with a diesel engine reached 97 miles/h (155.6 km/h); in 1978, it was a Mercedes-Benz C 111 that set the record at 197 miles/h (316.5 km/h). Imbert wood gas generator for car engines Cover

Centrifugal cleaner

Gas radiator

Wood Gas

Air nozzles

Gas

Air Air and ignition hole

Mixture throttle Wood charcoal Buddle valve to engine

Fan

Gas/ air mixer

Cover flap

Gas

Controller throttle valve Air

Air throttle valve

Wood wool Fine filter

Figure 1.2 Wood gas generator for car engines [1-3].

Despite the benefits of diesel engines, large gasoline engines were used to drive commercial vehicles. In the United States and in Germany, the twelve-cylinder engine of the Maybach Zeppelin powered omnibuses, fire engines, and half-track vehicles. The Opel Blitz commercial vehicle (with the six-cylinder inline engine of the Opel Admiral) became the standard vehicle of the German Wehrmacht. Small delivery vehicles (Tempo, Goliath, and Standard) were also driven by gasoline engines. Gradually, the diesel engine also broke into the automobile sector. The most common automobile powered by a diesel engine was the taxi. During World War II, the development of automobile engines stagnated worldwide because other things now had priority. After the war, the production of prewar engines started again.

Internal Combustion Engine Handbook | 5

6606_Book.indb 5

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 1 Historical Review

In the United States, car owners could afford large engines, and six-cylinder inline and eight-cylinder V-type engines were common. In Europe, many compact and subcompact cars were built with air and water-cooled two-stroke and four-stroke engines. German manufacturers included: Gutbrod, Lloyd, Goliath, and DKW. France also had several manufacturers (Dyna-Panhard, Renault 4 CV, and Citroen 2 CV). England had Austin and Morris, and Italy had Fiat. To avoid the high fuel consumption of two-stroke gasoline engines because of the scavenging losses, Gutbrod and Goliath engines had a mechanical fuel injection system. During the “economic boom” in Germany, the demand for small cars fell, preventing the two-stroke engine from establishing itself in the automobile (except of the Wartburg and Trabant cars, which were equipped with this engine type until the end of the 1980s in the German Democratic Republic). In the early 1950s, many four-stroke car engines still had side valves, and the crankshaft rested in bearings only after every second throw. After this time period, engines started to show a more modern design: crankcase drawn down well under the middle of the crankshaft, bearings for the crankshaft after every throw, compact combustion chambers with overhead valves (OHV), bucket tappets with overhead camshafts (OHC) at higher engine speeds, and increased piston displacement. Mercedes-Benz successfully competed in races again; the engines of the Silver Arrows had a gasoline injection system and positive-closing valves (desmodromic (positive) control) derived from aircraft engines. The economic upswing in the western world allowed prosperity to rise in general, and broad segments of the population could afford automobiles. As a result, vehicle production increased. There was a plenty of opportunities for vehicle development. In Japan, a new producer appeared on the world market that revolutionized automobile production with a high standard of quality, a reduction of the manufacturing depth, the splitting of production, improved assembly and development processes, and (just-in-time) delivery. Global competition necessitated even tighter cost control; the engines were produced in much larger quantities and were built with cost-effective production in mind. These engines required only simple maintenance and repair after production. Electronic data processing started to establish itself in research and development in the 1970s. This practice utilized computeraided design to simulate engine processes using a technique called finite element method (FEM). FEM resulted in rationalized, accelerated, and higher precision development. The concept of the reciprocating piston engine was questioned time and again. At the end of the 1940s, Rover in England had developed a vehicle with a gas turbine engine (Figure 1.3). High power density, compact design, some moving parts, no free mass effects (hence smooth running), good pollution control (thanks to smoke-free exhaust), and good cold-starting properties are major points in favor of the gas turbine. However, it was discovered that gas turbines are not suitable for the low powers and operating conditions of automobile engines. The gap losses are too high, resulting in poor efficiency.

In the 1960s, the rotary piston engine of Felix Wankel developed by NSU Motorworks AG (Figure 1.4) appeared to offer an alternative to the reciprocating piston engine. Its kinematics, power density, and compact design are the benefits compared with the reciprocating piston engines. However, the disadvantages outweighed the benefits: limited compression ratio, unfavorable combustion chamber, combustion with high constant pressure ratio, “late” combustion into the expansion phase, and problematical sealing of the combustion chamber led to high fuel consumptions and poor exhaust emission values. Only Mazda managed to build sporty vehicles with rotary piston engines with any degree of success. Rover gas turbine

Figure 1.3 Gas turbine for Rover car [1-2].

KKM 612

Figure 1.4 Rotary piston engine, type NSU KKM 612 [1-11].

The energy crises in the 1970s and the heightened public awareness of environmental problems led to a call for more economical engines with lower exhaust emissions. Starting from mechanical injection, a low-pressure fuel injection system with electronically controlled fuel metering was designed (much of

6 | Internal Combustion Engine Handbook

6606_Book.indb 6

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Historical Review

the work done by Bosch [1-1]). Despite the high development level of carburetor technology (twin carburetors, two-stage carburetors, and constant-pressure carburetor), fuel injection quickly became the established solution. Electronics became more and more involved in the engine control. A common microprocessor-controlled electronic system with map storage controls ignition and mixture formation. As measures inside the engine were no longer sufficient to reduce pollutant emissions to legally specified limits, three-way catalytic converters were employed that demanded precise control of the stoichiometric excess-air factor (lambda). Continuous measurement of the oxygen content in the exhaust gases using the Lambda sensor allows the pollutant emissions to be reduced. An additional improvement is achieved with controlled exhaust gas recirculation. Exhaust gas turbocharging as a means of increasing power and reducing consumption began to be employed in commercial vehicle engines from the 1960s. With increasing development levels, exhaust gas turbochargers could be “miniaturized” to such an extent that automobile gasoline engines could also be equipped. Because the fluid mechanics-based exhaust gas turbocharger and the reciprocating piston-powered internal combustion engine exhibited different operating behaviors, the “air supply” of the turbocharger and the “air demand” of the engine had to be balanced in order for these two machines to work together properly. Bypassing the turbine with part of the exhaust gas stream (waste gate control) and, for diesel engines, using variable turbine geometry were some of the initial advancements. A further improvement was achieved by cooling the charge air in the intercooler. As far as their response behavior for automobiles is concerned, mechanically powered turbochargers are at an advantage. Volkswagen developed a spiral turbocharger (G charger), and MercedesBenz uses Roots blowers for its “sporty” vehicle engines. An outstanding concept for turbocharging is the pressure wave supercharger (Comprex charger) from Brown, Boveri & Ci (BBC), in which the energy from the exhaust gas is transferred dynamically directly to the charge air, that is, without exhaust gas turbine and without a turbo compressor. Despite enormous development efforts, however, this principle has been unable to establish itself in the automotive industry (one of the reasons is because of high cost). Another drawback to this process is the fact that the exhaust gas temperature of gasoline engines is too high for proper operation. The automotive diesel engine was at series-production maturity as early as the 1930s. It found an admittedly limited, but loyal, group of fans in the 1950s among taxi drivers and high-mileage drivers who attached less importance to sporty driving and more to low fuel consumption and long service life. Apart from the Mercedes-Benz and Borgward engines, Peugeot and Fiat were the only other diesel engine manufacturers at the time. In the 1970s, Volkswawgon (VW) introduced an automotive diesel engine shortly followed by other German manufacturers (Opel, BMW, Ford, and Audi). The interest in the diesel cars was low in the United States. The distributor injection pump arrived on the scene in the 1960s/1970s and

proved to be ideal, particularly for the small injection volumes required by automotive diesel engines. Direct fuel injection offered significant consumption benefits for these engines, thus helping them to establish themselves for commercial vehicle engines in the 1960s. By the late 1980s, Ford had already equipped a delivery van with an engine using direct fuel injection. Audi was also in the process of delivering low-pollution car engines with direct injection around this time. Other companies followed suit, making direct injection a standard for diesel engines that still exists today. Turbochargers and intercoolers are becoming more and more common in diesel engines. High injection pressures are achieved with the unit injector system and, more recently, with the accumulator injection system (common rail). To reduce the thermal load on the diesel engine pistons, they are cooled either by spraying the undersides of the pistons or by using cooling channels. In the 1980s and 1990s, the charge cycle became one of the major developmental focal points. Flow coefficients and volumetric efficiency were improved with the multivalve technology. Further improvements came from variable valve timings and valve strokes, as well as variable-configuration intake manifolds. The development trend is now toward electromagnetically actuated and controlled valves. As a result, the intake cycle can be dethrottled reducing one of the major problems with this engine type. The direct injection into the cylinders of gasoline engines results in higher performance, reduced pollutant emissions, and lower consumption. On the engine side, the fuel consumption has been reduced by means of a whole range of measures: smaller dimensions and weights of the engine (downsizing), roller tappets rather than sliding tappets in the control, low-viscosity oils (that demand controlled operation of blowers and pumps), and so on. Increasing engine speeds of the five-cylinder inline and V-6 engine designs demanded measures to improve the machine dynamics of the engines. Differentials ensure the desired smoothness of running, as do rotational oscillation dampers. Limited resources and generally higher pollutant emissions are driving the search for different drive concepts. On the one hand, it is a question of finding a substitute for crude oil, and, on the other, of relieving the environment. A solution strongly favored by politicians for a time was the use of regenerative energies in the form of vegetable oil (rape oil methyl ester). The rape growing areas are not sufficient for an adequate supply of fuel (quite apart from the ecological problems associated with monocultures) nor are they technically expedient to replace mineral oils in motor vehicle engines. Another development is aimed at the use of hydrogen as fuel. Hydrogen, in conventional reciprocating piston engines as well as in fuel cells, can help to alleviate the pollutant situation. On the downside, hydrogen is difficult to produce. It has to be “generated” either by reverse electrolysis that requires a great deal of energy or by converting methanol or gasoline (not a good method for conserving resources). A feasible scenario lies in the increased use of natural-gas-powered engines. This solution would ensure the energy supply with ever-decreasing resources of crude oil while preparing the way for the advent of a gas technology using hydrogen.

Internal Combustion Engine Handbook | 7

6606_Book.indb 7

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 1 Historical Review

Bibliography

1-6. Kirchberg, P. 2000. Plaste, Bleche und Planwirtschaft. Die Geschichte des Automobilbaus in der DDR. Berlin: Nicolasche Verlagsbuchhandlung: Berlin.

1-2. Bussien, R. (Hrsg.). 1965. Automobiltechnisches Handbuch. 18. Aufl. Technik Verlag H. Cram: Berlin.

1-7. Krebs, R. 1994. 5 Jahrtausende Radfahrzeuge. Springer: Berlin.

1-1. Robert Bosch GmbH (Hrsg.). 1952. Bosch und die Zündung. BoschSchriftenreihe Folge 5: Stuttgart.

1-3. Eckermann, E. 1986. Alte Technik mit Zukunft. R. Oldenbourg: (Hrsg. Deutsches Museum) München. 1-4. von Fersen, O. (Hrsg.). 1986. Ein Jahrhundert Automobiltechnik – Personenwagen. VDI-Verlag: Düsseldorf. 1-5. von Frankenberg, R. and Mateucci, M. 1988. Geschichte des Automobils. Siegloch: Künzelsau.

1-8. Sass, F. 1962. Geschichte des deutschen Verbrennungsmotorenbaues. Springer: Berlin. 1-9. Pierburg. 1979. Vom Docht zur Düse. Ausgabe 8/1979. Fa. Pierburg: Neuss. 1-10. Zima, S. 1999. Kurbeltriebe. 2. Aufl. Vieweg: Wiesbaden. 1-11. ATZ 69. 1967. 9, pp. 279–284.

8 | Internal Combustion Engine Handbook

6606_Book.indb 8

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

2 Definition and Classification of Reciprocating Piston Engines 2.1 Definitions Piston machines are machines in which energy is transferred from a fluid (a gas or a liquid) to a moving displacer (e.g., a piston) or from the piston to the fluid [2-1], [2-2], [2-3]. They are thus part of the category of fluid energy machines that, as driven machines, absorb mechanical energy to increase the energy of the conveyed fluid. In drive machines, on the other hand, mechanical energy is released in the form of useful work at the piston or at the crank mechanism. The occurrence of a periodically changing working chamber as a result of the motion of the displacer (piston) is characteristic of the manner of operation of piston engines. One differentiates between reciprocating displacer engines and rotary displacer Type of working process

engines depending on the nature of the displacer’s movement. In reciprocating piston engines, the displacer takes the form of a cylindrical piston that moves between two extreme positions, the “dead centers,” in a cylinder. The term “piston” is also frequently applied to noncylindrical displacers. In rotary piston engines, a rotating displacer is normally responsible for varying the working chamber. Combustion engines are machines in which chemical energy is converted into mechanical energy as a result of the combustion of an ignitable mixture of air and fuel. The best known combustion engines are internal combustion engines and gas turbines. Figure 2.1 provides an overview.

Open process

Closed process

Internal combustion

External combustion

Combustion gas = working fluid

Combustion gas ≠ working fluid Change of phase of the working fluid No

Type of combustion Type of ignition Machine type Mixture type

Cyclical combustion Autoignition

Yes

Continuous combustion Supplied ignition

Engine

Diesel

Hybrid

Gasoline

Rohs [2-5]

Stirling [2-6]

Steam [2-7]

Turbine

—

—

—

Gas

Superheated steam

Steam

Homogeneous (heterogeneous)

Heterogeneous (in the continuous flame)

Heterogeneous (homogeneous)

(in the combustion chamber) Figure 2.1 The classification of combustion engines (after [2-8]).

Internal Combustion Engine Handbook | 9

6606_Book.indb 9

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 2 Definition and Classification of Reciprocating Piston Engines

Cylinders

External rotor

Housing

Housing

Eccentric shaft

Pistons

Connecting rod

Crankshaft a

Pistons b

Pistons c

Internal combustion engines are piston engines. One differentiates between reciprocating piston engines (featuring oscillating piston movement) and rotary piston engines (featuring rotating piston movement) depending on the geometry of the gastight, changing working chamber and on the type of piston motion. Rotary piston engines are, for their part, subclassified again into rotary engines (featuring an internal and an external rotor with purely rotary motion about fixed axes) and planetary rotary engines (that feature an internal rotor, the axis of which describes a circular motion) [2-4]. Figure 2.2 shows the differing working principles. Only the Wankel engine, a planetary piston engine, has achieved any significance. It is also necessary, depending on the type of working process, to differentiate between combustion engines with internal combustion and those with external combustion. In engines featuring internal combustion, the working fluid (air) is simultaneously the source of the oxygen necessary for combustion. Combustion of the fuel fed produces waste gas, which must be replaced in a “gas exchange” cycle prior to every working cycle. Combustion is therefore cyclical, differentiation being made among gasoline, diesel, and hybrid engines, depending on the combustion process. In the case of external combustion engines (such as the Stirling engine, for example), the heat produced outside the working chamber as a result of continuous combustion is transferred to the working fluid. This permits a closed-circuit working process and the use of any fuel. Only reciprocating piston engines featuring internal, cyclical combustion are examined from this point on.

2.2 Potentials for Classification The potentials for the classification of reciprocating engines are extremely diverse because of the complex interrelationships involved. Internal combustion reciprocating engines [2-9] can be differentiated by their

Figure 2.2 The working principles of reciprocating piston engines, rotary engines, and planetary piston engines. Housing with epitrochoid internal contour and power-yielding internal rotor that rotates eccentrically around a pinion and seals simultaneously.

•• Gas exchange control system •• Charging system •• Configuration. Further differentiating features may take the form of the following [2-10], [2-11]: •• Ignition system •• Cooling system •• Load-adjustment system •• Application •• Speed and output graduations. A number of differentiating features are now only of historical significance, however.

2.2.1 Combustion Processes

Among the combustion processes, differentiation is made primarily between the Otto cycle and the diesel cycle. Hybrid engines exhibit characteristics of both the Otto cycle and the diesel cycle. The gasoline engine is a combustion engine in which combustion of the compressed fuel + air mixture is initiated by means of synchronized extraneous ignition. In the diesel engine, on the other hand, the liquid fuel injected into the combustion chamber ignites on the air charge after this has previously been heated, by means of compression, to a temperature sufficiently high to initiate ignition [2-9]. In the case of hybrid engines, one differentiates between engines featuring charge stratification and multifuel engines [2-8] (also see chapters 15.1 and 15.2).

2.2.2 Fuel

•• Combustion process

Gaseous, liquid, and solid fuels can be combusted in combustion engines: •• Gaseous fuels: Methane, propane, butane, natural gas (compressed natural gas (CNG)), generator, blast furnace, biogas (sewage treatment and landfill gas), and hydrogen.

•• Fuel

•• Liquid fuels:

•• Operating principle •• Mixture formation

Light liquid fuels: Gasoline, kerosene, benzene, alcohols (methanol, ethanol), acetone, ether, and liquefied gases

10 | Internal Combustion Engine Handbook

6606_Book.indb 10

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

2.2 Potentials for Classification

(Liquified Natural Gas (LNG) and Liquified Petroleum Gas (LPG)). Heavy liquid fuels: Petroleum, gas oil (diesel fuel), fatty-acid methyl esters, and, primarily in Europe, rape-seed86 methyl esters (RME), also referred to as “biodiesel,” vegetable oils, heavy fuel oils, and marine fuel oil. Hybrid fuels: Diesel–RME, Diesel–water, gasoline–alcohol. •• Solid fuels: Pulverized coal development has long been discontinued.

2.2.3 Working Cycles

In the field of working cycles, differentiation is made between four-stroke and two-stroke processes. Common to both is the compression of the charge (air or a fuel vapor + air mixture) in the first step (stroke) by the reduction of the working chamber and ignition occurring shortly before the reversal of piston motion. Also, combustion associated with an increase in pressure up to the maximum cylinder pressure and the expansion of the working gas in the subsequent stroke, during which work is applied to the piston, is similar in both processes. The four-stroke process requires two further strokes to remove the combustion gas from the working chamber by means of displacement and to fill the working chamber with a fresh charge by means of natural induction (normal aspiration). In the two-stroke process, gas exchange occurs in the vicinity of bottom dead center (BDC) as a result of expulsion of the combustion gases by the fresh charge with only a slight change in the working volume, with the result that the complete stroke is not exploited for compression and expansion. An additional scavenging blower is necessary for the scavenging process.

2.2.4 Mixture Generation

Combustion engines can be differentiated in terms of their type of mixture generation: •• External mixture generation: Formation of the fuel–air mixture in the inlet system •• Internal mixture generation: Formation of the mixture in the working chamber on the basis of the quality of mixture generation: •• Homogeneous mixture generation: Carburetor and intake manifold injection in the case of the gasoline engine, or gasoline direct injection (GDI) during the induction stroke •• Nonhomogeneous mixture generation: Injection at extremely short intervals in the diesel engine and in gasoline engines with GDI during the compression stroke and on the basis of the location of mixture generation: •• Direct injection into the working chamber in the case, for example, of DI diesel engines and GDI engines. Injection may be air-directed, jet-directed, or wall-directed •• Indirect injection into a subsidiary chamber, such as antechamber, swirl-chamber, and air-chamber diesel engines •• Intake manifold injection (in gasoline engines).

2.2.5 Gas Exchange Control

Valve, port, and slide-valve tuning systems are used for control of the gas exchange. In the case of valve timing mechanisms, one differentiates between overhead- and side-actuated engines [2-9]. The overhead-actuated engine has overhead valves, that is, the closing movement of the valves occurs in the same direction as the movement of the piston toward the top dead center. The side-actuated engine, on the other hand, has vertical valves, and closure of the valves occurs in the same direction as the movement of the piston toward BDC. Only the overhead valve engine (OHV) arrangement, with overhead valves located in the cylinder head, is used in modem four-stroke engines. The camshaft may be located in the cylinder head or in the crankcase. Two-stroke engines mainly employ port-based timing systems (slots, or “ports” in the cylinder sleeve, with the piston acting as a slide valve), and bevel slide valves, disk valves, slide valves, and diaphragm timing systems in individual cases. In addition, a valve timing system (an exhaust valve in many cases) is also used in some recent motor-car and large marine engine developments.

2.2.6 Supercharging

In a normally aspirated engine, the fresh charge (air or mixture) is drawn into the cylinder by the working piston (natural aspiration). Supercharging enlarges the quantity of the charge as a result of precompression; a supercharger conveys the fresh charge into the cylinder. The primary aims of supercharging are the enhancement of power and torque output and the reduction of fuel consumption and exhaust gas emissions. Figure 2.3 shows an overview of possible types of supercharging (after [2-12]). The most widely used and effective variant in practice is self- or auto-supercharging, using a compressor: •• Mechanical supercharging: The compressor is driven directly by the engine. •• Exhaust turbo-supercharging: A turbine (exhaust turbine) powered by the engine exhaust drives the compressor. Processes without a compressor, which exploit the gasdynamic processes in the intake and exhaust systems to increase the charge, are also used.

2.2.7 Configuration

Many variants of cylinder arrangement have been suggested in the more than 120-year history of the internal combustion engine. Only a few standard configurations have stood the test of time [2-10], [2-11]. Starting from the single-cylinder engine, the number of cylinders selected can range up to as high as twelve in the case of vehicle engines. Aircraft engines with up to twenty-eight, or even as many as forty-eight cylinders, and high-performance engines with up to fifty-six cylinders have also been constructed.

Internal Combustion Engine Handbook | 11

6606_Book.indb 11

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 2 Definition and Classification of Reciprocating Piston Engines

Supercharging

Self charging

External charging

with compressor

without compressor

with exhaust gas use

without exhaust gas use

with exhaust gas use

Exhaust gas turbocharging

Mechanical supercharging

Pressure wave supercharging

without exhaust gas use

Resonance charging

Combined mechanical and exhaust gas turbocharging

Combined external and exhaust gas turbocharging

Constant pressure turbo charging

Pulse turbo charging

Ram pipe super charging

Figure 2.3 Various supercharging methods (after [2-12]).

There are many possible combinations for the cylinder arrangement, some of which are identified in a self-explanatory manner by letters. Figure 2.4 shows a selection of possible cylinder arrangements and configurations. The following are presently of significance: •• The inline engine (one bank of cylinders and one crankshaft). •• The V-engine (two banks of cylinders arid one crank-shaft): Two connecting rods are coupled to each crank pin. Common V-angles are 45°, 60°, 90°, and 180°. The VR engine [2-14] has a V-angle of 15°, the crankshaft having a separate crank pin for each connecting rod. •• The W-engine (three banks of cylinders and one crankshaft): Three connecting rods are connected in each case to one crank pin. A V-engine consisting of two VR banks is referred to as a V-VR engine, or also as a W-engine [2-14]. •• The boxer (flat-opposed) engine: Unlike the 180° V-engine, each connecting rod is connected to a separate crank pin. The crank mechanism has proven its value in engine design. Trunk piston engines and crosshead engines may be differentiated as variants. Slider crank mechanisms and cam engines are also described in the relevant literature, as are crankshaftless engines (curved-plate, curved-track, and swash-plate engines) [2-10]. Single- and double-acting engines can be differentiated according to their manner of action, depending on whether the combustion gases act on only one side or on both sides of the piston. The double-piston engine has two pistons to each

combustion chamber, the pistons being arranged either opposing (opposed-piston engine) or concurrent (U-piston engine). Vertical, horizontal, and overhead engines are differentiated on the basis of the location of the cylinder axis, and overhead- and side-actuated engines by the location of the timing mechanism.

2.2.8 Ignition

The fuel–air mixture may be ignited by means of supplied ignition or compression ignition: •• Supplied ignition (gasoline engine): An electrical spark ignites the mixture in the cylinder (spark ignition). •• Autoignition (diesel engine): The fuel injected ignites spontaneously in the air heated by compression in the cylinder (compression ignition).

2.2.9 Cooling

In view of the high temperatures that occur, the combustion engine needs to be cooled, to protect its components and the lubricating oil. It is necessary to differentiate between direct and indirect engine cooling. Direct cooling is accomplished using air (air cooling) either with or without the assistance of a fan. In the case of indirect cooling, the engine is cooled with a mixture of water, antifreeze, and corrosion inhibitors, or with oil (liquid cooling). Removal of heat to the environment is accomplished via a heat exchanger arrangement. One

12 | Internal Combustion Engine Handbook

6606_Book.indb 12

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

2.2 Potentials for Classification

Inline engine

V-Engine

W-Engine

Radial series engine

Boxer engine

X-engine

Dual radial engine

Radial engine

Twin bank engine

Quad radial engine

H-Engine

Dual shaft opposed-piston engine

Three shaft opposedpiston engine

Swash-plate engines/ inline engines

differentiates among evaporative, recirculating, once-through, and hybrid cooling.

2.2.10 Load Adjustment Motor output P

P = M ⋅w = M ⋅2⋅p ⋅n

(2.1)

can be matched to the power requirement by modifying both speed n and torque M (load). In the context of load adjustment, it is necessary to differentiate between

Figure 2.4 Cylinder arrangements in reciprocating piston engines [2-10], [2-13].

•• Quantity control and filling control: With an approximately constant air ratio λ , a throttling element (butterfly, rotary disk, slide, or other valve) controls the quantity of mixture that flows into the cylinder (conventional gasoline engine). •• Quality control: In diesel engines, and in GDI gasoline engines in certain operating ranges, the fuel is metered in as required. The injection flow is varied, with a practically constant flow of air (variable air ratio λ ).

Internal Combustion Engine Handbook | 13

6606_Book.indb 13

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 2 Definition and Classification of Reciprocating Piston Engines

2.2.11 Applications

Bibliography

•• Land-based vehicles: Road vehicles (motorcycles, automobiles, buses, and commercial vehicles), off-road vehicles, and rail vehicles

2-2. Kleinert, H.-J. (Hrsg.). 1989. Taschenbuch Maschinenbau – Bd. 5: Kolbenmaschinen, Strömungsmaschinen, 1. Aufl. Verlag Technik: Berlin.

A number of examples of the use of combustion engines are as follows:

•• Marine craft: Boats, inland, coastal, and ocean-going ships •• Aircraft: Airplanes and airships •• Agricultural machines and vehicles •• Commercial and industrial applications: Construction machines, handling, conveying, and lifting equipment, tugs, and tractors •• Stationary engine installations: Engine-powered generating plants, unit-type cogeneration plants, electrical generating sets, emergency-power sets, and supply systems.

2.2.12 Speed and Output Graduations

An extremely broad range of combustion engine speeds and outputs are used. Power ranges extend from model engines of 0.1 kW up to large-scale commercial installations of as much as 50,000 kW. An engine’s speed range also defines its output and size. The following can be differentiated by their speed [2-1]: •• Low-speed engines used, for example, in ships (60–200 rpm, in the case of diesel engines) •• Medium-speed engines (200–1000 rpm in diesel engines, maximum speed <4000 rpm in gasoline engines)

2-1. Grote, K.-H. and Feldhusen, J. (Hrsg.). 2007. Dubbel – Taschenbuch für den Maschinenbau, 22. Aufl. Springer: Berlin; Heidelberg; New York.

2-3. Eifler, W., Schlücker, E., Spicher, U., and Will, G. 2009. Küttner Kolbenmaschinen, 7. Aufl. Vieweg+Teubner: Wiesbaden. 2-4. Bensinger, W.-D. 1973. Rotationskolben-Verbrennungsmotoren. Springer: Berlin; Heidelberg. 2-5. Rohs, U. Kolbenmotor mit kontinuierlicher Verbrennung. Offenlegungsschrift DE 199 09 689 A 1, veröffentlicht: 07.09. 2000. 2-6. Werdich, M. and Kübler, K. 2007. Stirling-Maschinen: Grundlagen – Technik – Anwendung, 11. Aufl. Ökobuch: Staufen. 2-7. Buschmann, G., et al., “Zero Emission Engine – Der Dampfmotor mit isothermer Expansion,” MTZ 61(5): 314–323, 2000. 2-8. Robert Bosch GmbH (Hrsg.). 2007. Kraftfahrtechnisches Handbuch, 26. Aufl. Vieweg: Wiesbaden. 2-9. DIN Deutsches Institut für Normung (Hrsg.). 1976. DIN 1940: Verbrennungsmotoren – Hubkolbenmotoren – Begriffe, Formelzeichen, Einheiten. Beuth: Berlin. 2-10. van Basshuysen, R. and Schäfer, F. (Hrsg.). 2006. Lexikon Motorentechnik, 2. Aufl. Vieweg: Wiesbaden. 2-11. Beier, R., et al. 1983. Verdrängermaschinen, Teil II: Hubkolbenmotoren. TÜV Rheinland: Köln. 2-12. DIN Deutsches Institut für Normung (Hrsg.). 1976. DIN 6262: Verbrennungsmotoren – Arten der Aufladung – Begriffe. Beuth: Berlin. 2-13. Zima, S. 1999. Kurbeltriebe, 2. Aufl. Vieweg: Wiesbaden. 2-14. Braess, H.-H. and Seiffert, U. (Hrsg.). 2007. Vieweg Handbuch Kraftfahrzeugtechnik, 5. Aufl. Vieweg: Wiesbaden.

•• High-speed engines, for use, for example, in motor cars (maximum speed >1000 rpm in diesel engines and >4000 rpm in gasoline engines). Engines for sports and racing vehicles reach speeds of up to 22,000 rpm.

14 | Internal Combustion Engine Handbook

6606_Book.indb 14

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3 Characteristics Engine characteristics serve the developers, designers, and users of internal combustion engines as important aids in designing the fundamental dimensions, assessing engine power, and consumption, and evaluating and comparing different engines. A distinction is made between engine characteristics such as stroke, bore, piston displacement, and compression ratio and operating characteristics such as power, torque, engine speed, mean pressure, volumetric efficiency, and fuel consumption.

3.1 Piston Displacement



l ⋅ sin b = r ⋅ sin a

(3.3)

⎛r ⎞ b = arcsin ⎜ ⋅ sin a ⎟ . ⎝l ⎠

(3.4)



2 cos β = 1 − sin 2 b = 1 − ( r / l ) ⋅ sin 2 a

(3.5)

Allowing for

and inserting the connecting rod ratio ls =



The piston displacement or swept volume V h for an engine cylinder is the distance traveled by the piston during one piston stroke from BDC to TDC:

r l

(3.6)

we obtain the equation for the piston stroke:



l l 2 ⎞ ⎛ sa = r ⋅ ⎜ 1 + − cosa − ⋅ 1 − ( r / l ) ⋅ sin 2 a ⎟ ⎠ ⎝ r r

(3.8)

where



⎡ ⎤ 1 sa = r ⋅ ⎢( 1 − cosa ) + ⋅ 1 − 1 − ls2 ⋅ sin 2 a ⎥ ls ⎣ ⎦

s = Piston stroke

or sa = r ⋅ f (a) ,

(3.9)

VH = Vh ⋅ z =



p ⋅ dK2 ⋅ s ⋅ z , 4

(3.1)

dK = Piston diameter or cylinder bore



V h = Piston displacement for one cylinder z = Number of cylinders.

3.1.1 Calculation of Stroke and Piston Displacement from the Crankshaft Position (Figure 3.1) sa = r + l − x = r + l − r ⋅ cosa − l ⋅ cos b ,

)

with f(α ) = Stroke function. The connecting rod ratio λ s for car engines normally lies in the range from 0.2 to 0.35. It is difficult to work with the equation for the piston travel, particularly when piston speed or piston acceleration is to be calculated. An approximation equation can normally be used for simplicity in which the radical of a power series (MacLaurin series) is developed:

VH = Total swept volume of the engine



(

(3.7)

1 − ls2 ⋅ sin 2 a =

(3.2)

where r = Crankshaft radius l = Conrod length. Between the crank offset α and the connecting rod sweep angle β (connecting rod offset), we have the relationship:



1 1 1 1 − ⋅ ls2 ⋅ sin 2 a − ⋅ ls4 ⋅ sin 4 a − ⋅ ls6 ⋅ sin 6 a −… 8 16 2

(3.10)

Because of the values of λ s ≈ 0.2 to 0.35, the third term is already very small compared with the first term so that

Internal Combustion Engine Handbook | 15

6606_Book.indb 15

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 3 Characteristics

1 1 − ls2 ⋅ sin 2 a ≈ 1 − ⋅ ls2 ⋅ sin 2 a 2



(3.11)

can be assumed. Using the trigonometric function sin 2 a =



1 ⋅ ( 1 − cos 2a ) , 2

Vmin

OT

(3.12)

s

we then obtain for the piston travel sα as



⎡ 1 ⎛ 1 ⎞⎤ sa ≈ r ⋅ ⎢( 1 − cosa ) + ⋅ ⎜ 1 − 1 + ⋅ ls2 ⋅ sin 2 a ⎟ ⎥ ⎠⎦ ls ⎝ 2 ⎣

(3.14)



1 1 sa ≈ r ⋅ ⎡⎢( 1 − cosa ) + ⋅ ls ⋅ ⋅ ( 1 − cos 2a )⎤⎥ 2 2 ⎣ ⎦ l l ⎡ ⎤ sa ≈ r ⋅ ⎢1 − cosa + s − s ⋅ cos 2a ⎥ . 4 4 ⎣ ⎦

(3.15)



(3.13)

UT

Vh

sα

Vmax

b

l

For the momentary combustion chamber volume Vα , we obtain Va = Vc + AK ⋅ sa ,



a

r

(3.16)

where VC = Compression ratio (see Figure 3.2)

Figure 3.1 Swept volume and compression ratio.

AK = Piston surface. We thus obtain



1 Va ≈ Vc + AK ⋅ r ⋅ ⎡⎢1 − cosa + ⋅ ls ⋅ ( 1 − cos 2a )⎤⎥ . 4 ⎣ ⎦

(3.17)

3.2 Compression Ratio The compression ratio is defined as the quotient of the maximum and minimum cylinder volumes: the maximum cylinder volume is when the piston is at BDC. When the piston is at TDC position, the volume is minimal and is referred to as compression or dead volume. The compression volume is made up of the combustion chamber volume of the cylinder head, the valve pockets in the piston, a piston recess, and the top land volume up to the upper compression ring. Compression volume and piston displacement can be determined by gauging in liters. Figure 3.1 shows the swept volume and compression volume schematically. For the compression ratio of a four-stroke engine, we thus obtain

e=

Vmax Vh + Vc = Vmin Vc

(3.18)

where VC = Vmin = Compression or dead volume The compression ratio of a spark ignition (SI) engine is limited upward by the knock and by autoignition.

In SI engines with direct fuel injection, an increase in the compression ratio is possible because of the improved internal cooling by the internal mixture preparation. This gives them a higher efficiency compared with the SI engine with intake manifold injection. For the diesel engine, the compression ratio has to be selected at least so large as to ensure reliable starting when cold. In general, the thermodynamic efficiency increases with increasing compression ratio. An excessively high compression ratio, however, results in a decrease in the effective efficiency at full load because of the sharply increasing friction forces. In part-load operation, a high compression ratio has a positive effect on the efficiency [3-1]. Irrespective of that, the peak pressure that is limited by the material strength limits the compression ratio that can be achieved in practice. Figure 3.2 shows the influence of compression ratio on the effective efficiency and on the mean effective pressure in an SI engine during full-load operation. The ignition timing was set to maximum torque. The increase in efficiency up to a compression ratio of approximately 17:1 is clearly seen. The efficiency then drops in this case because of increasing frictional forces and a less favorable combustion chamber form because of increasing percentages of quench areas. With increasing compression, the NOx and HC emissions initially continue to increase. The nitrous oxides rise because of the increased combustion temperatures in the combustion chamber, and the HC emissions rise because of the greater splitting of the combustion chamber (larger relative proportion of gaps) and the increase in the ratio of combustion chamber surface area to combustion chamber volume (surface-to-volume ratio). To avoid this, combustion chambers must be designed

16 | Internal Combustion Engine Handbook

6606_Book.indb 16

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3.3 Rotational Speed and Piston Speed

as compactly as possible. With increasing compression, the exhaust gas temperature also drops because of the better efficiency so that postreactions of unburned hydrocarbons and carbon monoxide in the exhaust system are prevented. At the same time, however, an increase in compression results in a better lean-off capability and allows the ignition to be retarded because of the faster combustion. This enables the HC and NOx emissions to be reduced further.

9 7

9

11

13

15

17

19

SI engine Turbo

10

Knock, autoignition

14

Knock, autoignition

9

12

Knock, autoignition

Diesel (indirect injection)

18

24

Loss of efficiency at full load, component load, noise

Diesel (direct injection)

17

21

Loss of efficiency at full load, component load, noise

New developments are geared to varying the compression ratio according to the operating point while the engine is running. In the SI engine, the compression ratio is selected for optimum efficiency in part-load operation, whereas in full-load operation, the compression ratio is reduced to prevent knock. In the diesel engine, the compression ratio is limited by the maximum cylinder pressure (because of the component load). For diesel engines, the geometric compression ratio for full load can be optimally selected between a high efficiency and a maximum component load. For reliable cold starting, the compression ratio is set as high as possible.

eff. efficiency

e′ =

Vh′ + Vc , Vc

(3.19)

Vh′ =

p ⋅ dK2 ⋅ s′ , 4

(3.20)

3.3 Rotational Speed and Piston Speed 3.3.1 Rotational Speed

s′ = Residual stroke above the slots.

8 11

0,3 21

where Vh′ = Dead volume above the slots

Knock, autoignition

Figure 3.4 Compression ratios of modern engines.

In two-stroke engines with slot control, a distinction is made between the geometric compression ratio ε and the effective compression ratio ε ′. Figure 3.3 shows the difference. The effective compression begins only after the piston has closed the intake and exhaust slots. The effective compression ratio is calculated as



Autoignition

11

0,32

Figure 3.2 Influence of compression ratio on mean effective pressure and effective efficiency at full load of a gasoline engine [3-2].

where

10

9

Direct injection engine Turbo

Compression ratio e [-] Pme

7.5

Self-injection SI engine

0,4

0,34

upper limit by to

Two-stroke SI engine

0,42

0,36 10

from

Direct injection SI engine

0,38

11

ε

Engine type

eff. efficiency he [-]

eff. mean pressure pme [bar]

12

Figure 3.4 shows the possible ranges of the compression ratios for common engines.

n=

Number of crankshaft revolutions Time

(3.21)

3.3.2 Angular Velocity w = 2⋅p ⋅n



(3.22)

3.3.3 Piston Speed

The piston speed as a function of the crank angle is determined by the temporal derivation from the equation of the movement of the crank drive together with the angular velocity:

s′ s A

E

Figure 3.3 Geometric and effective compression ratio of the two-stroke engine.





sa =

dsa dsa da = ⋅ dt da dt

da = w = 2 ⋅ p ⋅ n . dt

(3.23)

(3.24)

Internal Combustion Engine Handbook | 17

6606_Book.indb 17

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 3 Characteristics

Consequently

1,08

ds 1 sa = w ⋅ a ≈ w ⋅ r ⋅ ⎡⎢ sin a + ⋅ ls ⋅ sin 2a ⎤⎥ . da 2 ⎣ ⎦

(3.25) 1,06

cm = 2 ⋅ s ⋅ n

(3.26)

The mean piston speed is a measure for comparing the drives of various engines. It provides information on the load on the sliding partners and indications of the power density of the engine. Figure 3.5 lists the rotational speeds and piston speeds of modern engines for orientation. Maximum speed Mean piston (rpm) speed (m/s) approximately approximately

Racing engine (Formula 1)

18,000

Small engines (two-stroke)

20,000

19

Motorcycle engines

13,500

19

Car SI engine

7500

25

1,00 0,00

0,10

0,20

0,30

0,40

Push rod ratio λs

The power at any working point of the engine is calculated from the torque and engine revolutions: Pe = Md ⋅ w = Md ⋅ 2 ⋅ p ⋅ n



(3.27)

According to this equation, an increase in power can be achieved by increasing the rotational speed or the torque. Both are subject to certain limits (see Chapter 3.3). As an example, Figure 3.7 shows the motor characteristics of a diesel engine. 180

20

560

160

520

140

480

120

440

100

400

80

360

60

5000

15

Truck diesel engines

2800

14

Larger high-speed diesel engines

2200

12

Medium high-speed engines (diesel)

1200

10

Crosshead engines (two-stroke diesel)

 150

8

Figure 3.5 Maximum rotational speed and mean piston speed at rated revolutions of modern engines.

3.3.5 Maximum Piston Speed

For the evaluation and design of piston rings and the slide systems piston cylinder running surface, not the average but the maximum piston speed is important. Assuming an infinitely long conrod (λ s = 0) simplifies the maximum piston speed: maximum piston speed for cmax = ω ⋅ r.

For the consideration of the conrod of finite length, the maximum of Eq. (3.25) must be determined. A correction according to Figure 3.6 shows the effect clearly: cmax = ω ⋅ r ⋅ kλ s.

3.4 Torque and Power

600

Car diesel engines



1,02

Torque [Nm]

Engine type

1,04

Figure 3.6 Correction of the maximum piston speed by the real λ s.

3.3.4 Mean Piston Speed

Factor kλs

With increasing piston speed, the mass forces, wear, flow resistance during intake, friction, noise also increase. The maximum permissible mass forces, in particular, limit the piston speed and hence the maximum rotational speed. In engines with internal mixture formation, that is, diesel engines and SI engines with direct injection, the rotational speed is additionally limited by the time necessary for the mixture formation. In diesel engines, this is one of the reasons for the significantly lower maximum revolutions compared with an SI engine of a similar size.

320

Output [kW]



40 0

1000

2000 3000 4000 Engine speed [1/min]

Output

5000

Torque

Figure 3.7 Power and torque curves for a turbocharged diesel engine [3-6].

The maximum torque and the maximum power are each plotted against the engine revolutions. The maximum power is not necessarily always achieved at the maximum engine revolutions. Not only the peak values for power and torque but also their curves against the engine revolutions are critical for the assessment of the interplay between engine and vehicle or engine and machine (see also Chapter 3.6: Gas work and mean pressure).

18 | Internal Combustion Engine Handbook

6606_Book.indb 18

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3.5 Fuel Consumption

If the effective power Pe is related to the swept volume VH, we speak of the specific power output Pl or power output per liter displacement: Pl =



Pe VH

(3.28)



where

K m 1 = Pe he ⋅ H u

(3.33)

1 he ⋅ H u

(3.34)

η e = Effective efficiency. The equation

The power-to-weight ratio mG is obtained from the motor mass mM based on the output: mG =

be =

mM Pe

(3.29)

be =

graphically displayed in Figure 3.9 clarifies the relation between effective efficiency and effective specific fuel consumption.

Figure 3.8 shows typical values. Power-toweight ratio (kg/kW)

Engine type

Specific power output (kW/l)

Racing engine (Formula 1)

200

0.4

70

2.0

Turbocharged car SI engine

100

3.0

Car diesel engine (naturally aspirated)

45

5.0

Turbocharged car diesel engine

64

4.0

Commercial vehicle diesel engine

30

3.0

Larger high-speed diesel engines

50

11.0

Medium-speed diesel engine

25

19.0

Car SI engine

he [-] 0,45 0,4

Euro-Super Diesel

0,35 0,3

3.0

0,2

55.0

0,15 200 225 250 275 300 325 350 375 400 425 450 475 500 be [g/kWh]

Figure 3.8 Empirical values for the specific power output and powerto-weight ratio.

150

200

250

300 350 be [g/PSh]

3.5 Fuel Consumption The energy admitted with the fuel per unit of time is calculated as EK = mK ⋅ H u



(3.30)

where mK = Weight of fuel admitted Hu = Net calorific value of the fuel.

Figure 3.9 Efficiency of the fuel consumption (HU, Euro-Super = 42.0 MJ/kg and HU, Diesel = 42.8 MJ/kg).

Figure 3.10 to Figure 3.12 show examples of power and fuel consumption curves for a car SI engine, a car diesel engine, and a commercial vehicle diesel engine. The isolines (bellshaped curves) indicate working points of equivalent fuel

The fuel consumption is measured as a volumetric flow or as a mass flow

mK = rK ⋅ VK t

(3.31)

where

ρ K = Density of the fuel. For better comparability, the fuel consumption can also be referred to the indicated or effective power. Indicated specific fuel consumption bi =

K m 1 = Pi hi ⋅ H u

where

η i = Indicated efficiency or internal effectiveness. Effective specific fuel consumption is

(3.32)

Mean pressure [bar]

K = m

18

180

16

160

14

140

12

120

255

10 260

8

80

290

6

310

4 2 0 1000

100

270

330 360 400 525 g/kWh

2000

3000

4000

5000

60

Output [kW]

Slow large diesel engine (two-stroke)

0,25

40 20 0 6000

Engine speed [1/min]

Figure 3.10 Power and consumption curves for a car SI engine [3-7].

Internal Combustion Engine Handbook | 19

6606_Book.indb 19

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 3 Characteristics

consumption. To assess the fuel consumption of an engine, the fuel consumption not only at the best point but also at all the working points has to be taken into consideration.

Efficiency (%)

minimal

maximal

500

Small engines (two-stroke)

350

24

450

Motorcycle engines

270

31

Car SI engine

250

34

Indirect injection car diesel engines

240

35

Turbocharged DI car diesel engines

200

42

Truck diesel engines with turbocharger

185

45

Larger high-speed diesel engines

190

44

Medium-speed engines

185

45

Crosshead engines (two-stroke diesel)

156

54

205 210

400

Torque [Nm]

Engine type

specific fuel consumption (g/kWh)

350

220

300

230

250

240 250

200 150

260 270 280

290 g/kWh

100 50

Figure 3.13 Empirical values for fuel consumption and efficiency at the best point.

0 400

800 1200 1600 2000 2400 2800 3200 3600 4000 4400 4800

Engine speed [1/min]

Figure 3.11 Consumption curve, car diesel engine V8-TDI (Turbocharged Direct Injection) [3-8].

20

Gas work is the work done by the cylinder pressure at the piston. With the mean pressure, we distinguish between indicated and effective mean pressure and the factional mean pressure.

Output line distance: 20 kW Usage line distance: 5 g/kWh

3.6.1 Indicated Mean Pressure

18

The indicated mean pressure pmi is equivalent to the specific work acting on the piston. The indicated mean pressure is determined from the cylinder pressure curve and the swept volume (Figure 3.14). The indicated mean pressure can be determined from the p-V diagram by planimetry (measurement of the area). If the surface enclosed by the curve is surrounded in a clockwise direction, we have a positive indicated mean pressure; if it is surrounded in a counterclockwise direction, we have a negative indicated mean pressure. Therefore, a distinction can be made between an indicated mean pressure of the high-pressure section and

194 g/kWh

16

195 14

Mean pressure [bar]

3.6 Gas Work and Mean Pressure

200

12

205 10 100

14

8

12

75 6

p [bar]

10

4 29 800

1000

1200

1400

1600

1800

6 4

50 kW

2

8

2000

Engine speed in [1/min]

pAmbient

2 0 0

Figure 3.12 Power and consumption curves for a commercial vehicle engine with VH = 12% [3-1].

Figure 3.13 shows empirical values for the specific fuel consumption.

100

200

300

400

500

V [cm3] Figure 3.14 Cylinder pressure over swept volume (2000 rpm, pmi 2 bar, and Vh = 500 cm3).

20 | Internal Combustion Engine Handbook

6606_Book.indb 20

1/19/16 8:28 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3.6 Gas Work and Mean Pressure

an indicated mean pressure of the gas exchange cycle. The sum of these two portions gives the indicated mean pressure of the engine pmi (Figure 3.15). The indicated mean pressure of the gas exchange cycle pmiGW composes of the intake and exhaust work and can, therefore, be regarded as a measure of the quality of the gas exchange [3-9].

pmi =



pmi

= Vh

pmi ⋅Vh = WKA



For naturally aspirated engines, the pmiGW is generally negative, that is, a work loss. For turbocharged engines, this portion is normally positive. The indicated mean pressure (Figure 3.15) can be derived from the work of the gas force transmitted to the piston during a working cycle: dWKA = p ⋅ AK ⋅dsa



(3.35)

PiZ = i ⋅ n ⋅ pmi ⋅Vh

Pi = i ⋅ n ⋅ pmi ⋅Vh ⋅ z = i ⋅ n ⋅ pmi ⋅VH

AK = Piston or cylinder surface area



WKA =  ∫ p ⋅dVa

(3.37)

The indicated power PiZ of a cylinder is hence calculated as PiZ = nA ⋅WKA



(3.38)

with

(

1 n ∑ pmii − pmi n − 1 i=1

(3.44)

)

2



(3.45)

pmi = Mean value of the indicated mean pressure. By an analogy with the indicated mean pressure (pmi), we also define the effective mean pressure (pme) and the friction mean pressure (pmr).

3.6.2 Effective Mean Pressure

The effective mean pressure can be determined from the torque Md:

i = Working cycles per revolution

pme =

Md ⋅ 2p VH ⋅ i

(3.46)

where Md = Torque of the engine i = Working cycles per revolution (0.5 for four-stroke, 1 for two-stroke engines)

For four-stroke engines:  i = 0.5 For two-stroke engines:  i = 1.

VH = Total swept volume of the engine.

Thus, the following applies for the cylinder output: PiZ = i ⋅ n ⋅WKA



s pmi = Standard deviation of the indicated mean pressure

nA = Working cycles per unit of time = i ⋅ n n = Engine revolutions per unit of time



pmi

COV = Variance (coefficient of variation) (3.36)

where dVα = change in volume = f (crank angle α ), and integration over the whole working cycle gives

s pmi

where

with change in volume, depending on piston travel AK ⋅dsa = dVa ,

s pmi =



WKA = Gas work at the piston per working cycle

(3.43)

The indicated mean pressure of several consecutive cycles is used to assess the regularity of the combustion, for example, by calculation of the variance. Irregular combustion and misfiring can be determined in this way. These are the criteria for hydrocarbon emissions, power, and smooth running of the engine. For well-designed engines, the variance of the indicated mean pressure is less than 1%, whereby the variance increases with increasing engine revolutions. The variance is calculated as follows:



sα = Piston travel = f (crank angle α )

(3.42)

This equation is true for one cylinder. An engine with several cylinders (z=number of cylinders) has the indicated power:

COV =

with p = combustion pressure or cylinder pressure

(3.41)

The indicated cylinder power can be expressed by



Figure 3.15 Determination of the indicated mean pressure from the areas over the swept volume.

(3.40)

or



–

WKA Vh

(3.39)

The gas work WKA referred to the swept volume V h per working cycle is defined as indicated mean pressure pmi:

Figure 3.16 shows examples of the effective mean pressure of modern engines.

Internal Combustion Engine Handbook | 21

6606_Book.indb 21

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 3 Characteristics

The indicated and the effective efficiencies are essentially determined from the energy stored in the fuel [3-3]. The energy admitted with the fuel per unit of time is calculated as

Effective mean pressure (bar)

Engine type

up to Motorcycle engines

12

Racing engine (Formula 1)

16

Car SI engines (without turbocharger)

13



Car SI engines (with turbocharger)

17

where

Car diesel engines (with turbocharger)

20

Truck diesel engines (with turbocharger)

24

Larger high-speed diesel engines

28

 K = Admitted mass of fuel per unit of time m Hu = Net calorific value of the fuel.

Medium-speed diesel engines

25

Crosshead engines (two-stroke diesel)

15

h=

3.6.3 Friction Mean Pressure

The friction mean pressure is the difference between indicated mean pressure and effective mean pressure: pmr = pmi − pme

(3.47)

The friction mean pressure according to Society of Automotive Engineers is the power loss because of mechanical friction in the engine and the pump losses in the crankcase. The friction in the engine is primarily dependent on the engine revs and hence on the piston speed, where the friction increases with increasing engine revs [3-4]. The cylinder pressure, that is, engine load and engine temperature, and the oil viscosity have a lesser effect on the friction. The friction losses according to Deutches Institut für Normung (DIN) (German Industry Standard) also include the drive powers for auxiliary components of the engine such as the alternator, air conditioning compressor, or servo pump.



Power P P = = EK m  K ⋅ Hu Fuel t

hi =



In the internal combustion engine, a distinction is made among the indicated, effective, and mechanical efficiencies.

Pi  K ⋅ Hu m

(3.50)

3.7.2 Effective Efficiency he =



Pe  K ⋅ Hu m

(3.51)

The ratio of effective efficiency to indicated efficiency is described by the mechanical efficiency.

3.7.3 Mechanical Efficiency

3.7 Efficiency

(3.49)

3.7.1 Indicated Efficiency

hm =

100% fuel energy

(3.48)

If we consider the engine power P as the output of the engine process and the admitted fuel energy per unit of time as the input, then the efficiency η can be calculated as

Figure 3.16 Effective mean pressure of modern engines.



EK  K ⋅ Hu =m t

he Pe = hi Pi

(3.52)

Figure 3.17 shows the breakdown of the admitted fuel energy into thermal losses and useful and frictional work. It also shows the breakdown of the frictional work or inertia work into the various portions.

100% drag

Effective output 28,5%

Crankshaft 11,0%

Piston ring 9,0% Piston 7,5%

9,8% Conrod 7,0%

Thermal losses 61,7%

Gas change and engine accessories 65,5%

Figure 3.17 Classification of the efficiency in a four-stroke SI engine [3-4].

22 | Internal Combustion Engine Handbook

6606_Book.indb 22

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3.8 Air Throughput and Cylinder Charge

3.8 Air Throughput and Cylinder Charge

where VG = Volumetric charge input per working cycle of a cylinder

The power of an engine is dependent on the cylinder charge. The air expenditure λ a and the volumetric efficiency λ l are used to assess and characterize the cylinder charge.

The air expenditure is a measure of the fresh charge admitted to the engine. It is assumed that the charge is in gaseous form. For the air expenditure, we have the relationship:



mG ges mG mG = or la = mth Vh ⋅ rth VH ⋅ rth

For the SI engine la =



3.8.1 Air Expenditure

la =

VG ges = Volumetric charge input per working cycle of the motor.

(3.53)

where mG = Total fresh charge mass admitted to a cylinder per working cycle mG ges = Total fresh charge mass admitted to the engine per working cycle mth = Theoretical charge mass per working cycle (cylinder or complete engine)

ρ th = Theoretical charge density.



la =



mG = mL or mG ges = mL ges

The volumetric efficiency is a measure of the fresh charge remaining in the cylinder at the end of the charge cycle. As with the air expenditure, this is referred to as the theoretical charge density: ll =

The theoretical fresh charge mass is calculated from the geometric swept volume and the ambient state of the charge. For turbocharged engines, the thermodynamic state up line of the intake organs is used instead of the ambient state. For engines with internal mixture formation, the charge consists of air, and for engines with external mixture formation, the charge consists of air and fuel. The gas equation gives us

pu ⋅Vh = mth ⋅ R ⋅Tu or pu ⋅VH = mth ges ⋅ R ⋅Tu

(3.56)

with R = RG (Gas constant of the mixture) in the SI engine R = R L (Gas constant of air) in the diesel engine or direct injection SI engine Tu = Ambient temperature pu = Ambient pressure. If the density of the mixture or air taken in is assumed to be equal to the theoretical charge density ρ th, the air expenditure can also be calculated using volumetric parameters

mG = VG ⋅ rG or mG ges = VG ges ⋅ rG

(3.57)

mZ ges mZ mZ = or ll = mth Vh ⋅ rth VH ⋅ rth

(3.60)

The cylinder fresh charge is calculated as mZ or mZ ges gilt. For the SI engine

(3.54)

(3.55)

(3.59)

3.8.2 Volumetric Efficiency

2. Diesel engine

VL ges VL or la = Vh VH

To determine the air expenditure empirically at the engine, the intake air volume or air mass is measured. In addition, the pressure and temperature of the air and the ambient conditions, as well as the fuel consumption with the SI engine, have to be recorded.



mG = mK + mL or mG ges = mK ges + mL ges

(3.58)

For the diesel engine

The total fresh charge mass admitted consists of 1. SI engine

VG ges VG or la = Vh VH

mZ = mZL + mZK or mZ ges = mZL ges + mZK ges



(3.61)

For the diesel engine

mZ = mZL or mZ ges = mZL ges

(3.62)

where mZL = Air mass in one cylinder mZL ges = Air mass in all the engine cylinders mZK = Fuel mass in one cylinder mZK ges = Fuel mass in all cylinders. The charge mass remaining in the cylinder or in all the engine cylinders cannot be calculated or measured directly. The following method is employed as an approximation: (a) Cylinder pressure indication in one or all the engine cylinders ssumption that the cylinder charge temperature at the (b) A moment the “intake valve closes” is roughly the same as the temperature in the intake duct upline of the intake valve (measurement of this temperature using a thermocouple) (c) A pplication of the gas equation at the moment the “intake valve closes” pZEs ⋅ VEs = mZ ⋅ R ⋅ TZEs (d) R G or RL RL is assumed again for the gas constant R.

Internal Combustion Engine Handbook | 23

6606_Book.indb 23

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 3 Characteristics

With four-stroke SI engines, the crank angle range of the valve overlap (the time during which both intake and exhaust valves are open at the same time during the charge cycle) is relatively small. For the case of the small valve overlap, λ a ≈ λ l can be assumed as a good approximation. For engines without a turbocharger, λ a and λ l are always smaller than 1, as flow resistance during the intake and exhaust prevents a complete scavenging of the geometric swept volume. Turbocharged engines and engines with ram-effect supercharging are examples of engines that have operating states in which λ a and λ l are larger than 1. Diesel engines, particularly those with a turbocharger, have large valve overlaps to achieve internal cooling and a better scavenging of the remaining gas out of the combustion chamber. Here, λ a » λ l can become A. With slot-controlled two-stroke engines, a considerable difference exists between air expenditure and volumetric efficiency because of the overflow losses. The quotient of volumetric efficiency and air expenditure gives the retention rate that is a measure of the fresh charge remaining in the cylinder.

During combustion in the engine, the ratio of the air mass actually in the cylinder mL to the stoichiometric air mass mL, ∑ τ is referred to as the air/fuel factor λ . The stoichiometric air requirement LSt is defined as the quotient of the air mass and the fuel mass under stoichiometric conditions







(3.65)

with the stoichiometric components x=

MK M M M ⋅c y = K ⋅h q = K ⋅s z = K ⋅o MC MH MS MO

where c, h, s, o = Percentages by weight of the elements carbon (c), hydrogen (h), sulfur (s), and oxygen (o) contained in the fuel MC, M H, MS, MO = Molar weights of the elements in the fuel MK = Molar weight of the fuel. Allowing for the percentage by weight of oxygen in the air ξ O2,L we obtain for the stoichiometric air requirement LSt =

1 mO2 ,St 1 MO2 nO2 ,St ⋅ = ⋅ ⋅ xO2 ,L mK xO2 ,L MK nK

(3.66)

where MO2 = Molar weight of oxygen nO2; nK = Volumes of oxygen and fuel y z With the relations nO2 ,St = x + + q − and nK = 1 from the 4 2 chemical reaction equations, we obtain

3.9 Air–Fuel Ratio

LSt =

y z⎞ y ⎛ C x H y S q O z + ⎜ x + + q − ⎟ ⋅ O 2 ⇒ x ⋅ CO 2 + ⋅ H 2 O + q ⋅ SO 2 ⎝ 4 2⎠ 2

mL,St mK

m mL l= L = mL, St mK ⋅ LSt

(3.63)

(3.64)

where

LSt =

MO 2 ⎞ 1 ⎛ MO 2 1 MO 2 ⋅ ⋅c + ⋅ ⋅h+ ⋅ s − o⎟ MS xO2 ,L ⎜⎝ MC 4 MH ⎠

(3.67)

LSt =

1 ⋅ ( 2.664 ⋅ c + 7.937 ⋅ h + 0.988 ⋅ s − o ) . 0.232

(3.68)





Figure 3.18 shows exemplary data of a fuel analysis. Unit

Value

Mean molar mass of the fuel

G/mol

99.1

Composition of the fuel specimen:

wt.%

 87.08

Carbon

wt.%

 12.87

Hydrogen

mL,St = Air mass under stoichiometric conditions mK = Fuel mass. The stoichiometric air requirement can be calculated from the percentage by weight of the chemical elements contained in the fuel, whereby the combustion products (exhaust gases) resulting from the combustion also have to be taken into consideration. The combustion process proper covers many intermediate reactions in which numerous, but also predominantly short lived, compounds or “radicals” are involved. The most important combustion products with complete combustion are carbon dioxide (CO2), water (H2O), and sulfur dioxide (SO2), as well as the air nitrogen (N2, inert gas) that is practically unchanged by the combustion. For complete combustion of a fuel with the composition CxHySqOz, we thus obtain the chemical reaction equation:

Theoretical total formula

wt.%

  0.05

Oxygen

—

 7.2

Carbon

—

12.6

Hydrogen

—

 0.0

Oxygen

Gross calorific value (Ho)

MJ/kg

 45.72

Net calorific value (Hu)

MJ/kg

 42.88

Theoretical stoichiometric air demand

kg air kg fuel

 14.47

Figure 3.18 Example of a fuel analysis, Euro Super.

The fuel metering during engine operation is influenced by the stoichiometric air requirement. For this reason, the mixture forming system has to be adapted accordingly when using different fuels (e.g., gasoline- and alcohol-based fuels). During combustion in the engine, the mixture ratio deviates more or less from the stoichiometric ratio.

24 | Internal Combustion Engine Handbook

6606_Book.indb 24

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

3.9 Air–Fuel Ratio

A mixture with excess air (λ > 1)is referred to as a “lean mixture” (lean operation), while a mixture with an air deficiency (λ < 1) is referred to as a “rich mixture”. SI engines with intake manifold injection are operated today in wide program map ranges almost exclusively with a stoichiometric mixture (λ = 1). SI engines with direct injection can be operated homogeneously with λ = 1, homogeneous lean (λ > 1), and stratified lean (on average for the combustion chamber λ ≫ 1, but also partially with λ = 1). Diesel engines are always operated with excess air (λ  > 1), and small two-stroke engines are predominantly operated in the air deficiency range (λ  < 1).

Bibliography

3-1. Mollenhauer (Hrsg.). 2007. Handbuch Dieselmotoren. Springer: Berlin. ISBN 978-3-540-72164-2.

3-3. Spicher and Ulrich. 2004. Umdruck zur Vorlesung Verbrennungsmotoren. Universität: Karlsruhe. 3-4. Mahle (Hrsg.). 1994. Einflussgrößen auf die Reibleistung der Kolbengruppe. Technische Information Nr. 7148. Stuttgart. 3-5. Robert Bosch GmbH (Hrsg.). 2007. Kraftfahrtechnisches Taschenbuch. 26. Aufl. Vieweg: Wiesbaden. ISBN 978-3-8348-0138-8. 3-6. Anisizs, F., Borgmann, K., Kratochwill, H., and Steinparzer, F. “Der erste Achtzylinder-Dieselmotor mit Direkteinspritzung von BMW,” MTZ 60(6):362–371, 1999. 3-7. Fortnagel, M., Heil, B., Giese, J., Mürwald, M., Weining, H.-K., and Lückert, P. “Technischer Fortschritt durch Evolution: Neue Vierzylinder Ottomotoren von Mercedes-Benz auf der Basis des erfolgreichen M111,” MTZ 61(9):582–590, 2000. 3-8. Bach, M., Bauder, R., Endress, H., Pölzl, H.-W., and Wimmer, W. “Der neue TDI-Motor von Audi: Teil 3 Thermodynamik,” MTZ 60:40–46, 1999, Sonderausgabe 10 Jahre TDI-Motor von Audi. 3-9. Kuratle, R. 1995. Engine technology. 1. Aufl. Vogel: Würzburg. ISBN 978-3-8033-1554-4.

3-2. Heywood and John B. 1988. Internal Combustion Engine Fundamentals. Mc Graw-Hill: New York. ISBN 0-07-100499-8.

Internal Combustion Engine Handbook | 25

6606_Book.indb 25

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 26

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

4 Curves The working point of an internal combustion engine is defined by its speed and its torque. The full range of all possible working points in a two-dimensional presentation gives the “engine map”. In this map, the working range of the internal combustion engine is limited by the full-load curve and by the minimum and the maximum engine revs (Figure 4.1). The power output by the engine at any particular working point is calculated from the equation Pe = 2 ⋅ π ⋅ M ⋅ n. Lines of constant power are referred to in the engine map as power hyperbolas.

Pme M

Runningresistance curves

Load

Pe = const.

Full-load

Pe, max

Mountain Plane Valley nmin

Overrunning

Rotational Speed nmax

Figure 4.1 Engine map.

The engine map is used to document certain engine characteristics as a function of the working point. This representation can consist of an indication of discrete values in individual points. When many individual values are available over the whole working range of the engine, lines of equal value for the respective engine characteristic, the “isolines,” can be created from these individual values by interpolation. The most common curve representation is of the specific fuel consumption whose isolines are presented in the map as the “conchoids” (see also Figure 4.3).

Apart from the engine characteristics, the characteristics of the vehicle and its powertrain can also be displayed in the map. This normally takes the form of the running-resistance curve. These curves show the relationship between the engine revs and the torque drawn by the powertrain for each gear during constant travel on a level road. Uphill or downhill travel results in a parallel shifting of the running-resistance curve (see Figure 4.1). If the working point of the engine is above the runningresistance curve, the vehicle accelerates; if it is below the curve, the vehicle brakes. The surplus power available for acceleration results from the current engine revs and the surplus torque corresponding to the distance between the running-resistance curve and the full-load curve. A gear shift results in a different torque for the same travel speed because of the change in engine revs with an approximately equal power requirement, that is, the working point is shifted along the power hyperbola up to the intersection with the running-resistance curve corresponding to the gear shift. In this way, the engine map allows the changes in the operating or emission behavior to be assessed in relation to the boundary conditions of the vehicle and the method of operation. For operating conditions with a low power requirement, such as is the case for large portions of the emissions cycles for type testing of a vehicle or in town traffic, operating points with low to medium engine revs/load combinations are of greater relevance. The typical load collectives for highway driving, on the other hand, lie in the top right-hand area of the engine map. For reasons of comparability of engines with different swept volumes, the specific parameters of the load, the specific mean pressure, or the specific work referred to the swept volume are frequently used instead of the torque. Maps are used both for documentation of operating parameters, such as ignition timing, injection timing, or excess air factor to illustrate the operating strategy, and for evaluation of the resulting measured and calculated parameters, such

Internal Combustion Engine Handbook | 27

6606_Book.indb 27

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 4 Curves

Operating strategy: Engine with gasoline direct injection 12

Range of external EGR

Effective mean pressure [bar]

10 l<1

8

l=1

6 4 2 0

l>1 homogeneous

l>1 stratified 1000

2000

3000

4000

Engine speed [1/min]

5000

6000

Figure 4.2 Engine map.

To clarify the operating strategy, characteristic areas of the map are marked differently. In this example, the engine is operated below a load of Pme = 4 bar and up to engine speeds of 3500 rpm by injection during the compression stroke with a stratified fuel–air mixture with a large excess of air. In the rest of the map, the fuel is injected during the intake phase with the consequence that the longer mixture preparation time results in a homogeneous mixture. Even during homogeneous operation, there is a phase with an excess of air in the lower load range for engine speeds between 3500 and 4500 rpm. In the remaining load and engine speed range above the stratified and homogeneous lean operation, this engine is operated like a conventional SI engine with a stoichiometric mixture. As full load is approached (particularly at higher engine revs), the mixture is enriched to protect the catalytic converter against any excessive exhaust gas temperatures and to achieve higher power. This operating map also shows by appropriate marks that an external exhaust gas recirculation takes place in the whole stratified range and in part of the stoichiometric range. Further features characteristic of the engine operating strategy can be illustrated in the operating map in the same way. These include, for example, the balance of a camshaft adjustment or of a controlled intake manifold. For the experienced engineer, the engine map represents a source of highly compacted information from which he/she can derive an assessment of the engine in question. When comparing and evaluating maps on the basis of specific engine parameters, it must be remembered that, in practice, design criteria such as that for swept volume or stroke-to-bore ratio, the compression ratio or the design and arrangement of the injection valves are reflected with only minor differences.

On the other hand, operative measures such as the setting of variable systems (controlled intake manifolds, camshaft adjusters) and of the engine control, as well as measures for exhaust gas posttreatment (e.g., catalytic converter systems and thermal insulation of the exhaust system up to the catalytic converter) result in very significant differences in the operating behavior, even for similar engines. One such example is SI engines with direct injection. In the Japanese market, these engines exhibit a similar behavior to that of the engine described above, whereas the adjustment of the same basic engine for the European market exhibits no stratified or homogeneous lean range in the whole map. This shows clearly that the measures for exhaust gas posttreatment for different markets or even just for more stringent emission certification levels result in more significant differences in the engine map than, for example, manufacturer-specific or design differences would suggest.

4.1 Consumption Curves Figure 4.3 shows a typical consumption map for conventional SI engines with intake manifold injection. As already mentioned, the lines of constant specific fuel consumption are also called conchoids due to their form. The minimum specific fuel consumption is found in the lower engine rev range in the range of high load. Only a flat gradient of the consumption increase is seen in a wider range around the minimum consumption. The gradient rises sharply toward the low load range. One of the main reasons for this is because of increasing throttle losses in the SI engine and the increasing proportion of friction in relation to the useful torque output. These two factors also lead to the visible increase in consumption at constant load and increasing engine revs. Toward the full-load range, the mixture has to be enriched, on the one hand, to counter the knock tendency of the engine and, on the other, to keep the exhaust gas temperature below a critical limit temperature for catalytic converter aging. This leads to a sharper gradient of the consumption increase. Specific fuel consumption [g/kWh] 12

Effective mean pressure [bar]

as emissions, fuel consumption, or temperatures. Figure 4.2 shows how an engine map is used to illustrate the operating strategy of the engine, taking as an example an SI engine with direct injection.

10

300

8

280 260

6

280 300

4

400

2 0

1000

2000

500 3000

4000

5000

6000

Engine speed [1/min]

Figure 4.3 Consumption curve (MPI SI engine).

28 | Internal Combustion Engine Handbook

6606_Book.indb 28

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

4.2 Emission Maps

Specific fuel consumption [g/kWh] 18

Effective mean pressure [bar]

16 14

205

12

220

10 8

250

6 4

300

2 0

400 1000

2000

3000

4000

Engine speed [1/min]

Figure 4.4 Particulate emissions (DI-TCI diesel engine).

If the main vehicle-specific data such as the running resistances and gear ratios are known, the consumption map of an engine can also be used to calculate the fuel consumption of the vehicle. To calculate the consumption in the nonstatic test cycle, the operating curve is broken down into a sequence of static working points as a function of the vehicle-specific parameters, each characterized by the engine revs and the torque. The load points are then entered into the calculation of the cycle consumption and weighted temporally according to their relevance for the driving schedule. The models necessary for the exact calculation of the consumption take into account not only the vehicle-specific data, but also consumption-influencing processes such as the engine warm-up, gear shifting, and other nonstatic effects. These models allow vehicle-related effects on the consumption and emission behavior of the engine in the vehicle to be assessed. Examples of the application of such methods are the transmission setting or the control strategy of a continuously variable transmission (CVT).

4.2 Emission Maps The subjects of the emission maps are generally the raw emissions of the legally limited pollutant components, that is, hydrocarbons, nitrous oxides, and carbon monoxide. Normally, these maps show the work-related specific values (in g/kWh)

or mass flows (in g/h). For diesel engines and for SI engines with direct injection, the maps of the particulate emissions are also of significance. Apart from the raw emissions maps, the emission values down line of the catalytic converters are also often shown. These values permit an evaluation of the conversion in the catalytic converter as well as enable an estimation in the volumes of pollutants emitted by the vehicle in a driving cycle. Figure 4.5 to Figure 4.9 show characteristic maps for conventional SI engines and selected maps of operative parameters relevant to the emission behavior. The engines on which the illustrated maps are based are all equipped with three-way catalytic converters for efficient exhaust gas treatment. The maps relate to the running of the engine at the operating temperature. This and the λ control employed with the selected engines to obtain an exact stoichiometric mixture guarantee a high conversion rate of all the pollutant components in the three-way catalytic converter. The air–fuel ratio shown in Figure 4.5 clearly illustrates the large map area of active Lambda control. As in the example of the SI engine with direct injection shown above, a mixture enrichment is employed here again, both in the full-load range and at high engine revs. In the area of the rated power, the minimum air–fuel ratios are calibrated with values of around λ = 0.80. 12

l [-]

10

Effective mean pressure [bar]

Figure 4.4 shows the typical consumption curve for a diesel engine with direct injection and turbocharger. Of particular note is the slighter increase in consumption with decreasing load, as the quality control of the diesel engine is not related to throttle losses. Despite the far more favorable part-load consumption values compared with the SI engine, the consumptions achieved with the vehicle calibration lie above those for a consumption-optimized setting, particularly in the relevant map area for the European driving cycle. One reason for this is the retarded setting of the injection timing necessary to comply with the permissible NOx and particulate emissions.

0.82

8

0.90

6

0.99 l=1

4 2 0

1000

2000

3000

4000

5000

6000

Engine speed [1/min]

Figure 4.5 Air–fuel ratio (MPI SI engine).

The CO concentration is predominantly a function of the excess air factor, as the maps in Figure 4.5 and Figure 4.6 show. In the map area with active Lambda control, the concentrations generally lie in an uncritical order of between 0.5 and 0.8vol.%. At full load, the combustion takes place with an air deficiency due to the mixture enrichment. The maximum CO concentration of 7.5vol.% occurs at the maximum enrichment rates in the area of the rated power output. This relationship between the CO concentration and the air–fuel ratio illustrated in Figure 4.6 can be regarded as typical for modern SI engines with high specific powers.

Internal Combustion Engine Handbook | 29

6606_Book.indb 29

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 4 Curves

with the maximum EGR rate. Outside the map area of external EGR, we obtain a typical behavior of the NOx emissions. The sharp reduction in NOx emissions recognizable at full load and at high engine revs is a result of the mixture enrichment.

CO [Vol.-%]

Effective mean pressure [bar]

10 6.0

8

12

2.5

6 4 2 0

1000

2000

3000

4000

5000

6000

Engine speed [1/min]

Figure 4.6 CO concentration up line of the catalytic converter (MPI SI engine).

NOx [g/kWh]

Effective mean pressure [bar]

10

10

8

20

15

6 4

5

10

2 0

1000

2000

3000

15

20

4000

25 30

5000

Engine speed [1/min]

Figure 4.7 Specific NOx emissions up line of the catalytic converter (MPI SI engine).

8

6000

4 8

6 15

4

12

2 0

The level of the NOx raw emissions can also be influenced in stoichiometric operation by the adjustment of the operative parameters. When calibrating the engine in the map, a retarded calibration of the ignition angle—within limits—is selected. However, this measure does result in a reduced efficiency so that it also has to be considered when evaluating the consumption map. On the other hand, the exhaust gas recirculation (EGR) in part-load operation offers a significant potential for reducing the NOx raw emissions while at the same time improving efficiency due to the related dethrottling of the engine. Exhaust gas recirculation can be performed either externally via a valve or internally as exhaust gas recirculation by modifying the ignition timing. The map of the specific NOx emissions of an SI engine with external exhaust gas recirculation in Figure 4.7 and the corresponding map of the EGR rates calibrated with the EGR valve in Figure 4.8 show an example of the practical use of exhaust gas recirculation. The minimum NOx emissions are achieved at the working point 12

External EGR [%]

10

0.8

Effective mean pressure [bar]

12

1000

2000

3000

4000

5000

6000

Engine speed [1/min]

Figure 4.8 Exhaust gas recirculation rate (MPI SI engine).

Continuously operating systems for camshaft adjustment are frequently used in mass-produced engines not only to achieve an internal exhaust gas recirculation but also to improve the torque behavior at full load. Optimization of the ignition timing as a function of the engine revs allows air expenditure benefits to be achieved as a result of the improved torque curve. In contrast to the NOx and CO emissions, the level of the HC raw emissions is influenced far more strongly by design parameters. The first aspect here is the form of the combustion chamber, with the surface-to-volume ratio representing a characteristic parameter. Although the HC emissions are sensitive to operative parameters with the engine at operating temperature, it is of less significance in the normal range of variations. An internal EGR using variable valve timing can have a positive effect on HC emissions as the typical HC peak observed toward the end of the exhaust cycle is returned to the combustion. A typical HC emissions map of an SI engine with single-stage intake camshaft adjustment is shown in Figure 4.9. The maps of the emissions down line of the catalytic converter for modern SI engines with a three-way catalytic converter not illustrated here are characterized by the practically complete conversion of the pollutants. Deviations from the extremely low emission levels occur in the map areas with substoichiometric operation where the catalytic oxidation of the HC and CO contents remains limited due to oxygen insufficiency. Because of the combustion with excess air typical for the diesel engine, the carbon monoxide and HC emissions are significantly lower compared with the SI engine (Figure 4.10 and Figure 4.11). The residual oxygen that always exists in the exhaust gases from diesel engines permits a further reduction of these pollutant components in oxidation catalytic converters.

30 | Internal Combustion Engine Handbook

6606_Book.indb 30

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

4.2 Emission Maps

HC [g/kWh]

10 8

4.0

6

3.0

4 2 0

7.0

5.0

1000

2.5

5.0

4.0

3.0

18

16

NOx [g/kWh]

16

4.0

10.0

14

3.0

2000 3000 4000 Engine speed [1/min]

5000

6000

Figure 4.9 Specific HC emissions (MPI SI engine).

18

to limit the occurrence of NOx by influencing the combustion process. The measures employed here are the same as with the SI engine, exhaust gas recirculation, and the retarding of the injection process that is more or less the equivalent of the retarded ignition of the SI engine.

Effective mean pressure [bar]

Effective mean pressure [bar]

12

12 10 8

4.0

8.0

6

2.0

4

1000

2000

2.0

0.5

5.0 20.0

2 1000

2000

10.0 50.0 3000

4000

Engine speed [1/min]

Figure 4.10 Specific CO emissions (DI-TCI diesel engine).

Effective mean pressure [bar]

16

HC [g/kWh] 18

0.1

14

AGR [%]

16

12

14

10 8

12

0.2

6 4 2 0

To increase the reduction effect of the EGR for the emitted nitrous oxides, the recirculated exhaust gases of the diesel engine are cooled. The map of the EGR rates in Figure 4.13 shows that in this example, the exhaust gas recirculation is essentially calibrated for the emission-relevant map area. The exhaust gas recirculation rates can be as high as 50%, and thus lie far higher than those for an SI engine. In contrast with the SI engine, the possibilities of exhaust gas recirculation are not limited here by the occurrence of combustion misses. It must be remembered here that combustion takes place with a large air surplus and that the oxygen concentration in the exhaust gas is still as high as 15 vol.%.

5.0 1000

2.0

1.0 2000

Effective mean pressure [bar]

Effective mean pressure [bar]

0.5

8

18

4000

Figure 4.12 Specific NOx emissions (DI-TCI diesel engine).

.5

0.5

10

0

3000

10.0

Engine speed [1/min]

12

4

6.0 8.0

14

6

4.0

2 0

CO [g/kWh]

6.0

8.0

10

0.5

3000

4000

Engine speed [1/min]

Figure 4.11 Specific HC emissions (DI-TCI diesel engine).

More critical for diesel engines, however, are the NOx raw emissions (Figure 4.12). Because catalytic posttreatment with excess air is not effective here, the primary solution pursued is

8

5 10

6 4 2 0

50 1000

40

20 30

2000

3000

4000

Engine speed [1/min]

Figure 4.13 Exhaust gas recirculation rate (DI-TCI diesel engine).

Internal Combustion Engine Handbook | 31

6606_Book.indb 31

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 4 Curves

In addition, for diesel vehicles, the law regulates the amount of particulate emissions that can come out of the exhaust. A common method for assessing the particulate emissions from diesel engines is the filter smoke number (FSN). The increased black smoke values in the emission-relevant map area (Figure 4.14) are indicative of the relationship between particulate formation and EGR. Filter Smoke Number FSN [-]

16

Effective mean pressure [bar]

14 12

2.0

10

1.0

8 6 4 2 0

0.5 0.5

1.0

2.0

2.0 3.0 0.5 1000

2000

3000

12 4000

Engine speed [1/min]

Figure 4.14 Particulate emissions (DI-TCI diesel engine).

This relationship also draws attention to the known conflict of goals between NOx and particulate emissions. Outside the area of the map geared to EGR, the level of the smoke numbers is relatively low and increases significantly only as full load approaches, particularly at low engine revs because of the lower air–fuel ratio prevailing here. Particulate formation has to be countered by good preparation of the injected diesel fuel. That is why the high-pressure injection with high-quality atomization represents one of the major development directions of modern diesel engines. With a further tightening of the particulate emission limits, the use of particulate filter systems will permit major developmental steps to be taken in addition to the internal motor measures. Despite the relatively high cost, this technology is already in use in mass-produced vehicles. An almost nationwide use can be assumed in sensitive markets. In contrast with the stationary calibration documented in the engine maps, the intermittent regeneration of the particulate filter necessitates intervention in the calibration of the engine that serves to temporarily increase the exhaust gas temperatures in certain map areas to promote the burnoff of the particles collected on the filter surface.

4.3 Ignition and Injection Maps The typical calibration of the ignition angle in conventional engines with λ control exhibits a strong dependence on the operating point. In the middle of the part load, the ignition

Ignition timing [°KW v. OT]

10

Effective mean pressure [bar]

18

angle is generally calibrated in the area of optimum efficiency. Figure 4.15 shows a fundamental trend to an increasing need for advanced ignition with increasing engine revs and decreasing load. This behavior is superimposed by further effects. In the lower load range, a significant advance adjustment of the ignition is seen (even at low engine revs). For the engine shown, an external exhaust-gas recirculation is calibrated in this area. The recirculated exhaust gas that acts as an inert gas delays the combustion process that therefore must be initiated earlier. Furthermore, a retarding of the ignition is seen close to full load at engine revs of approximately 4500 rpm. This behavior is attributable to the frequently observed tendency to knock in the area of the highest air expenditure. In line with the latest state of the art, the disadvantages resulting from this measure can be minimized with the use of dynamic knock control systems. These permit a torque-optimized preignition angle without the risk of engine damage caused by knocking combustion.

8

5 10

15

20 25

6 30

4

35

2 0

1000

2000

3000

4000

5000

6000

Engine speed [1/min]

Figure 4.15 Ignition timing (MPI SI engine).

In diesel engines, the combustion is primarily controlled by the fuel injection process. The start of injection therefore has a significant effect comparable with that of the ignition angle in the SI engine. With the transition to direct injection that predominates today, the fast combustion with sharp gradients of the cylinder pressure curve leads to acoustic problems. An effective measure for reducing the cylinder pressure gradients with modern electronic diesel control (EDC) engine control systems is the preinjection. With preinjection, the combustion is first triggered by a smaller injection of fuel. Then the remaining volume of fuel is admitted to the process during the main injection. Figure 4.16 and Figure 4.17 show the maps for the start of injection for the preinjection and main injection in a modern car diesel engine. It can clearly be seen how the preinjection is limited to a specific engine speed range by the EDC engine control system. Furthermore, the relatively late positions of the start of injection of the main injection indicate the use of the measures to

32 | Internal Combustion Engine Handbook

6606_Book.indb 32

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

4.4 Exhaust Gas Temperature Maps

reduce the NOx emissions described above. In the map area without preinjection, on the other hand, the main injection is shifted earlier.

18 16

Start of preinjection control [°KW v. OT]

4.4 Exhaust Gas Temperature Maps The behavior of the exhaust gas temperature of an SI engine is shown in Figure 4.18. The sharp increase in exhaust gas temperature to high loads necessitates specific measures to protect the exhaust gas catalytic converter from thermal aging or even destruction.

Effective mean pressure [bar]

14

12

12

6 4 2 0

10

40

8 30 15 10

No preinjection

35

20

1000

2000

3000

4000

Engine speed [1/min]

Start of main injection control [°KW v. OT]

16

Effective mean pressure [bar]

14 12 10 8 6 4 2 0

0 1

2

16

12 8 10

4

–2 1000

2000

3000

4000

Engine speed [1/min]

Figure 4.17 Start of main injection control (DI-TCI diesel engine).

850

8

900

800 6

700 600

4

500 400

2

300

0 1000 2000 3000 4000 5000 6000 Figure 4.18 Exhaust gas temperature map at the entry to the catalytic converter (MPI SI engine). Engine speed [1/min]

Figure 4.16 Start of preinjection control (DI-TCI diesel engine).

18

Effective mean pressure [bar]

10

Temperature upstream of cat. [°C]

Both design measures and the calibration of the engine operating parameters are employed here. For engines with exhaust gas turbocharging, the gas temperature at the turbine inlet is also critical for component protection. For SI engines, an enrichment of the fuel–air mixture is therefore employed as an effective component protection measure in the map area with critical exhaust gas temperatures as described above. For operation with low load points, on the other hand, an excessively low exhaust gas temperature must be avoided so that the catalytic converter does not cool down. For this reason, a relatively retarded ignition timing would be necessary. In addition to these measures recognizable in the static maps, deviating control parameters for the ignition angle and EGR rates are normally calibrated after the engine cold start so that the catalytic converter quickly reaches the light-off temperature necessary for conversion of the raw emissions into harmless components.

Internal Combustion Engine Handbook | 33

6606_Book.indb 33

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 34

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5 Fundamentals of Thermodynamics Hence, to obtain qualitative information on the relationship of certain process variables to predetermined parameters, more or less simple model calculations are used. This allows basic conclusions to be made regarding the effect of the conversion of energy based on engine-related parameters that are much less complex. A series of methodologies were produced in the past that extended from a simple closed process control to more or less complicated open multizone models [5-9], [5-11], [5-12], [5-13].

5.1 Cyclical Processes

qin

qout

Temperature T

To obtain basic information, simplified models are created and described as cyclical processes. Cyclical processes are sequential state changes of a fuel in which the fuel is returned to its initial state. They are described as closed cyclical processes with the supply and removal of heat (Figure 5.1) [5-14]. Pressure p

Internal combustion engines are heat engines in which chemically bound energy is converted into mechanical energy [5-1], [5-2], [5-3]. This is done by means of a reaction, the combustion process, in which energy is released. Part of this heat released into the combustion chamber of the cylinder is converted into mechanical energy by the crankshaft drive, and the remaining energy is carried away with the exhaust and released to a coolant via the walls neighboring the combustion chamber as well as directly into the environment. The goal of the process of converting chemical energy into mechanical energy is to attain the greatest possible process efficiency (strongly dependent on the thermodynamics). These conversion processes are very complex, especially the combustion process with its energy-substance exchange processes and the chemical processes of the gas in the cylinder [5-4]. In addition, the process of transfer of heat from the gas to the wall directly surrounding the combustion chamber, the neighboring engine components, and the coolant or oil can be approximated only with great effort [5-5], [5-6], [5-7], [5-8]. Because the fuels for spark-ignition and diesel engines are mixtures consisting of various hydrocarbons, it is practically impossible to describe the reaction kinetics of the many reactions. Frequently, pure substances such as methanol, methane, and hydrogen are used that have a sufficiently precise reaction mechanism with all the associated substance data. Depending on the methodology, it is sufficient to use specific reaction processes such as formation of NO [5-9] or the simplifying assumption of an O–H–C equilibrium at the flame front [5-10]. If the process is considered from a locally multidimensional, nonstationary perspective with all the transport mechanisms that actually exist in the gas, complex mathematical models result that yield somewhat imprecise substance data (if any at all) for the physical–chemical description.

qin

qout

Volumen v

Entropy s

Figure 5.1 State changes and work in a cyclical process [5-15].

Internal Combustion Engine Handbook | 35

6606_Book.indb 35

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

This type of model does not treat the conversion of the initial products of combustion such as air and fuels into exhausts (CO, HC, NOx, CO2, HCO, H2, N2, etc.). The four cycles of the combustion engines are compression, supply of heat as a “replacement” for the combustion process, expansion, and heat removal as a replacement of the charge cycle. The state of the medium, for example, at the beginning of compression and at the end of heat removal is identical. State diagrams for internal combustion engines are as follows: •• Pressure–volume diagram (p–v diagram): The contained area represents work that is termed the indicated work. •• Temperature–entropy diagram (T–s diagram): The areas represent heat. The cyclical process work is the difference between the supplied and the removed heat. The area enclosed by the lines of the state changes is a measure of the useful work of the cyclical process. Essential information on the engine process attainable with such cyclical processes relates to the process efficiency. A definition of one such type of efficiency, thermal efficiency, is hth =



qzu − qab q = 1 − ab qzu qzu

(5.1)

where qzu = supplied quantity of heat and qab = removed quantity of heat. The theory of cyclical processes originates from the French officer Sadi Carnot (1796–1832), who recognized that to convert heat into work there must be a temperature gradient. He also noticed that the thermal efficiency of a heat engine increases with the increase in temperature where the heat is supplied and with the decrease in temperature where it is removed. This becomes particularly clear with the optimum cyclical process that he described, the Carnot process (Figure 5.2). The state changes of the Carnot process are

•• Isothermic expansion •• Isentropic expansion. In the T–7s diagram, the Carnot process is portrayed as a rectangle. The thermal efficiency results as a ratio of useful work to supplied heat: hth =



hth c = 1 −



qzu − qab q = 1 − ab qzu qzu

Tmin ⋅ ( s1 − s2 ) T = 1 − min . Tmax ⋅ ( s4 − s3 ) Tmax

(5.2)

(5.3)

The thermal efficiency assumes the highest attainable value at a given temperature ratio in the Carnot process. In the p–v diagram, the diagram area of the Carnot process is so small that the temperatures and pressures would have to be raised to an unacceptable level to obtain acceptable useful work (corresponding to the area in the p–v diagram). This was realized by Rudolf Diesel when he wanted to implement the Carnot process with his rational heat engine. A rectangular process in the p–v diagram yields the greatest amount of work, but is much less efficient because of the small area in the T–s diagram. A rectangular process is therefore not suitable in practice. The cyclical processes that are technically feasible with a heat engine are subject to the restrictions of the geometry and kinematics of the respective machine type, the conditions of energy conversion, and the state of the art. The evaluative criteria for comparative processes that are described in the following are •• Efficiency •• Work yield •• Technical feasibility.

•• Isothermic compression •• Isentropic compression

pmax qin

qin Pressure p

Temperature T

pmax Tmax

Tmax pmin

Tmin

pmin

qout Entropy s

qout Tmin Volume v

Figure 5.2 State changes in the Carnot process [5-15].

36 | Internal Combustion Engine Handbook

6606_Book.indb 36

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.2 Comparative Processes

5.2 Comparative Processes

•• Isentropic compression

5.2.1 Simple Model Processes

•• Isentropic expansion

•• Isochoric supply of heat •• Isochoric heat removal.

The cyclical processes of an engine describe the energy conversion where the individual state changes of the fuel most closely approximate the actual behavior in the engine. With this in mind, internal combustion engines represent closed systems in which the energy conversion is discontinuous. A characteristic of the cyclical processes of internal combustion engines is that state changes occur in a work area whose size changes as a result of the movement of the crankshaft drive over the course of the combustion cycle. Compression and expansion can be described by simple state changes. The combustion and the charge cycle are replaced by heat addition and heat removal. Ideal cyclical processes for internal combustion engines are differentiated according to the type of heat supply. A general process can be represented by the heat supply at a constant volume (isochoric) and at constant pressure (isobaric), as described by Myron Seiliger (1874–1952) as the Seiliger process. Borderline cases can be derived from this such as pure constant volume (only an isochoric supply of heat) and pure constant pressure (only isobaric heat supply) cycles.

It exhibits the thermodynamically most favorable process that can be realized—with acceptable technical effort—in a machine with a cyclically changeable work area [5-1]. The resulting thermal efficiency is larger than the Seiliger and constant pressure cycle at the same compression ratio. The efficiency depends on the type of gas (isentropic exponent) and the compression ratio. It increases with the rising compression ratio and is calculated as follows: hth =



qzu v − qab qzu v



(5.4)

5.2.1.2 Constant Pressure Cycle The state changes of the constant pressure cycle are shown in Figure 5.4. The sequence of the state changes in this process is •• Isentropic compression •• Isobaric supply of heat •• Isentropic expansion

5.2.1.1 Constant Volume Cycle Figure 5.3 presents the state change during the constant volume cycle. The sequence of the state changes in this process is

•• Isochoric heat removal. It can then be used as a comparative process when, for reasons of component load, the maximum pressure must

Temperature T

Pressure p

Constant volume cycle 3

qin

3

qin

2

2 4 qout 1

4 1

qout

2

Entropy s

qin

qin

Temperature T

Pressure p

Volume v

3

4 1 qout Volume v

Figure 5.3 State changes in the constant volume cycle [5-15].

3

2 4

1

qout Entropy s

Figure 5.4 State changes in the constant pressure cycle [5-15].

Internal Combustion Engine Handbook | 37

6606_Book.indb 37

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

be limited. The thermal efficiency is calculated as follows: hth =



qzu p − qab qzu p



At a given compression ratio, a maximum pressure limit must be specified. The heat supply is partly isochoric and partly isobaric. The thermal efficiency from this process control is

(5.5)

The efficiency of this process depends on the gas type (isentropic exponent), the compression ratio, and the supplied quantity of heat at a constant pressure. It rises as the compression ratio increases and falls as the supply of heat increases. Of the three process controls considered, the constant pressure cycle has the least efficiency.

hth =



qzu v + qzu p − qab qzu v + qzu p

.

(5.6)

It must be noted that the quantity of heat qzu v is supplied at a constant volume, and therefore, it is measured from the temperature difference with reference to the specific heat at a constant volume (cv). The supply of heat at a constant pressure qzu p is measured from the temperature difference with reference to the specific heat at a constant pressure (cp). Depending on the distribution of the supplied quantity of heat between the isochoric and isobaric state changes, the thermal efficiency results as a limit curve that would exist with constant volume and constant pressure. Applying this process to a supercharged engine yields the relationships shown in Figure 5.6.

5.2.1.3 Seiliger Process The state changes of the Seiliger process are shown in Figure 5.5. Taken individually, these are •• Isentropic compression •• Isochoric supply of heat •• Isobaric supply of heat •• Isentropic (adiabatic reversible) expansion •• Isochoric heat removal.

3

Temperature T

Pressure p

qin 4

qin 2

qin qin

4

3

2 5

5 qout 1

1

qout Entropy s

Pressure

Volume v

q23

q34 3

4

p01

p67 p5

Compressor

Exhaust collecting pipe

2

Turbine

pL Charge pressure pA Exhaust back-pressure p0 Atmosph. pressure 5 1

pL 11 pA 10

q51 7¢

7

6

p0 9 Surface Surface Surface Surface Surface

Figure 5.5 State changes in the Seiliger process [5-15].

0 1-2-3-4-5-1 0-1-11-9-0 7-8-9-10-7 5-7¢-6-5 7¢-7-8-8¢-7¢

Engine work Compressor work Turbine work Loss of kinetic energy Conversion of kinetic energy to heat and utilization in turbine

8¢

q80

8 Volume

Figure 5.6 State changes in the Seiliger process of an exhaust gas turbocharging engine [5-15].

38 | Internal Combustion Engine Handbook

6606_Book.indb 38

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.2 Comparative Processes

Comparison of processes

Constant volume cycle Mixed cycle Constant pressure cycle

Volume V

Temperature T

Pressure p

Equal compression ratio e Same amount of supplied heat q

Constant volume cycle Mixed cycle

Constant pressure cycle

Entropy s Additional removed heat in mixed cycle Additional removed heat in constant pressure cycle

Figure 5.7 Comparative engine processes [5-15].

In principle, supercharging does not change the process in the engine; only the pressure level rises. The compression in the engine is upstream from the compression in the compressor, and the expansion in the engine follows the expansion in the engine and expansion in the exhaust pipe:

An illustration of this in an example of the constant volume process is provided in Figure 5.8. In the p–V diagram, two types of process loss can be illustrated:

•• Isentropic compression in the compressor

•• The area 5–6–1–5 would be useful if the medium is expanded to the initial pressure as well as to the initial temperature. This must be followed by isothermic compression to the initial pressure.

•• Isentropic compression in the engine •• Isochoric supply of heat in the engine •• Isobaric supply of heat in the engine

•• If the medium is expanded from point 4 to point 5, that is, to the initial pressure, the work (area 4–5–1–4) would be useful.

•• Isochoric heat removal from the engine •• Isobaric supply of heat to the turbine •• Isentropic expansion in the turbine

Pressure

•• Isentropic expansion in the engine 3

•• Isobaric heat removal from the turbine. The work of the exhaust turbine and the compressor are correspondingly represented as the areas in the p–V diagram (Figure 5.6).

5.2.2 Exergy Losses

The exergetical perspective of the discussed process controls shows that the exergy of the supplied energy can be only partially converted into mechanical work. Exergy is the energy that can be converted into any other form of energy in a predetermined environment. Anergy is the part of the energy that cannot be converted into exergy [5-1].

4

Pu

1

I

5

6

II

v 3

Temperature

5.2.1.4 Comparison of the Cyclical Processes Figure 5.7 compares the three processes considered in the p–V and T–s diagrams. The efficiency of the constant volume cycle is the maximum attainable, given an equivalent compression ratio. This is because of the lower quantity of heat that is removed, given an equivalent compression ratio and the same amount of supplied heat in comparison with the two other process controls.

2

v2 = const. v1 = const.

2

Tu 0

4

I

5 p1 = const.

II

1

6

III

a

b

Figure 5.8 Thermodynamic loss with the example of the constant volume cycle.

Internal Combustion Engine Handbook | 39

6606_Book.indb 39

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

However, in a real engine, this would require a substantial amount of additional engineering that would be out of proportion to the gain. The third loss arises from the anergy of the supplied energy. It is not directly attributable to the process control. If a medium reaches the environmental temperature and environmental pressure, it is in a thermal and mechanical equilibrium with the environment. The second law of thermodynamics prevents the conversion of internal energy into exergy or useful work [5-1].

and the composition of the fresh gas characterized by the air–fuel ratio λ . This is defined as l=



 Air m  Fuel ⋅ mAir stoich m

(5.7)



where ṁAir refers to the air mass, ṁFuel the fuel mass, and ṁAir stoich the stoichiometric air mass of the corresponding fuel. The compression ratio can be taken from the corresponding test engine. As a representative of gasoline, isooctane (C8H18) can be used because it more or less approximates the physical and chemical properties of commercially available fuels.

5.3 Open Comparative Processes

It is assumed that the gas composition during compression remains constant. The final compression state can be calculated with the aid of isentropic relationship S1,T1 = S2,T2 and the thermal state equation for ideal gases p ⋅ v = ∑ s i ⋅ Rm ⋅T , with p = pressure, v = specific volume,

5.3.1 Work Cycle of the Perfect Engine

The ideal cyclical processes are only a crude approximation that can be used to arrive at a few basic conclusions. In regard to efficiency, they yield overly “good” values in comparison with reality: the work yield is greater and the efficiency is better than in real engines because the properties of the working gas, air, are treated as those of a real gas. Further, the heat loss, charge cycle loss, friction loss, and chemical reactions are not included. To obtain more detailed information on the process cycle and answers regarding the optimum process control, further processes have been defined that allow a better approximation of real engines. This is possible with open comparative processes. A helpful and frequently used comparative process is the “perfect” engine process. The parameters under which this process occurs are as follows:



•• The charge in the combustion chamber has no residual exhaust gas.

CO, CO2, N2, NO, NO2, NH3, O2, O, H, N, H2, H2O, and OH.

•• The air–fuel ratio is the same as that in the actual engine. •• There is a loss-free charge cycle (no flow and leakage loss). •• Combustion occurs according to set laws. •• Heat-insulating walls are present. •• Isentropic compression and expansion occur with specific heats cp and cv depending on the temperature. •• The combustion products are in chemical equilibrium. With the process defined in this manner, we can determine the influences of the parameters of compression and air–fuel ratio on average pressure, process effectiveness, and a few concentrations of substance components (Figure 5.9). Depending on the methodology, a process control can be selected that uses simple cyclical processes. This can be an isochoric (constant volume combustion), an isobaric (constant pressure compression), or a mixed isochoric/isobaric cycle. 5.3.1.1 Elements of Calculation The calculation of the cycle of the perfect engine can be divided into the following steps. (a) Isentropic compression of the fresh mixture. The initial state is described by the pressure p0, the temperature T0,

i

T = temperature, Rm = general gas constant, and σ i = specific moles of component i:

∑s i

i,1

⎛ 0 ⎛ 0 p1 ⎞ p2 ⎞ ⋅ ⎜ si,T 1 − Rm ⋅ ln 0 ⎟ = ∑ s i,2 ⋅ ⎜ si,T 2 − Rm ⋅ ln 0 ⎟ , (5.8) p ⎠ p ⎝ ⎝ ⎠ i

0 where si,T 1 is the entropy of the component i at standard pressure p0 and temperature T.

The solution to the equation can, for example, be found using an iterative process. (b) Isochoric adiabatic combustion. It is assumed that there is total chemical equilibrium. The combustion products consist, for example, of the components: The state of the gas mixture in the cylinder after combustion is characterized by the pressure p3, the temperature T3, and the specific moles of the participating (in this case, thirteen) components. To determine these quantities, fifteen independent equations are necessary. These equations are as follows. 1. First law of thermodynamics for closed systems If it is presupposed that heat is neither supplied nor removed during the combustion and no work is carried out, it follows that du = 0, that is, no change in the internal energy. We thus obtain:

∑s i

i,2

⋅ ui,T 2 = ∑ s i,3 ⋅ ui,T 3

(5.9)

i

2. Thermal state equation This is expressed by the following:

p3 ⋅ v3 = ∑ s i ⋅ Rm ⋅T3

(5.10)

i

3. Chemical equilibrium The thirteen gas components that chemically react with each other consist of the basic elements, namely, oxygen, nitrogen, hydrogen, and carbon. To describe the chemical equilibrium,

40 | Internal Combustion Engine Handbook

6606_Book.indb 40

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.3 Open Comparative Processes

12 500

45

ppm

hi

10 000

40

5 hv Vol%

4

n = 2 000

min–1

e = 7,8

35

3 hi measured measured measured

{ {

CO concentration

NO concentration

Process efficiency hi, hv

7 500

Frozen concentrations Chemical equilibrium

5 000

30

2

2 500

25

1

CO NO 0

20

0,8

1,2

1,0

1,4

0

Figure 5.9 Quantities calculated with the working cycle of the perfect engine, and quantities measured on a test bench [5-16].

Air-fuel ratio

nine independent reaction equations are therefore required with the stoichiometric coefficients τ j,i (j = 1–9): ∑ mi ⋅ t j,i = 0. μ is the chemical potential of the component i

i

combustion, the amounts of the four basic materials j = 1–4 O, H, N, and C do not correspondingly change so that the material balances are expressed as follows:

i and is defined as: 0 mi = g i,T + Rm ⋅T ⋅ ln



pi p0

(5.11)

0 i,T

where g represents the molar free enthalpy of component i in a standard state. 4. Material balances The remaining equations for determining the state after combustion are provided by the material balances. During



s B, j = ∑ a j,i ⋅ s i

(5.12)

i

where α j,i is the number of atoms of the basic material j in component i. The nonlinearity equation system that thereby arises consisting of fifteen equations can, for example, be solved using a Newtonian method.

Internal Combustion Engine Handbook | 41

6606_Book.indb 41

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

c)  Expansion. The parameters to represent the expansion state are the chemical equilibrium and constant gas composition. The state change is isentropic. We thus obtain:



∑s i

i,3

⎛ 0 ⎛ 0 p3 ⎞ p4 ⎞ (5.13) ⋅ ⎜ si,T 3 − Rm ⋅ ln 0 ⎟ = ∑ s i,3 ⋅ ⎜ si,T 4 − Rm ⋅ ln 0 ⎟ p ⎠ p ⎠ ⎝ ⎝ i

5.3.1.2 Work of the Perfect Engine The work WVM of the perfect engine results from the difference in the internal energy expressed as follows: WVM = U 4 − U 1 or by





WVM

⎛ ⎞ = m ⋅ ⎜ ∑ s i,1 ⋅ ui,T 1 − ∑ s i,4 ⋅ ui,T 4 ⎟ ⎝ i ⎠ i

where uT and sT,P mean the specific internal energy or entropy at temperature T and pressure p and u0,T0 and s0,T0,P0 are quantities that result when the combustion gases are in a thermodynamic equilibrium with the environment. The relative exergy loss EV of the combustion can be defined as follows: EV =



(5.19)

whereas the relative exhaust gas loss is

(5.14)

EA =

(5.15)

E2 − E3 E1

E4 E1

(5.20)

Figure 5.10 shows the characteristic of the relative exergy loss in a perfect spark-injection cycle.

where U and ui represent internal work. 60



hVM =

WVM mKr ⋅ H u

(5.16)

with Hu as the bottom calorific value of the fuel and mFuel as the fuel mass. If the effectiveness is defined as the ratio of the obtained process work WVM to the maximum theoretically obtainable work mFuel × Hu must be replaced by the term Wtheoretical. The quantity Wtheoretical then can be defined as the maximum obtainable work in a reversible process control or as the reversible reaction work. This results from the difference in the free enthalpy from the state of the fresh mixture and the exhaust gas with Wtheoretical:

Wtheoretical

⎛ ⎞ n nn HTn 0 − HTnn0 − T0 ⋅ ⎜ ∑ Si,p0,T 0 − ∑ Si,p0,T 0 ⎟ ⎝ i ⎠ i = ≈ Hu mKr

(5.17)

where HTn 0 and HTnn0 are the enthalpy of the material flows of the combusted and noncombusted materials in reference n nn to the environmental state, respectively; Si,p0,T 0 and Si,p0,T 0 represent the entropy of the component i in the combusted and noncombusted materials in reference to the environmental state, respectively. The differences between the reversible reaction work and bottom caloric value are very low for a few substances defined as substitute fuels such as C7H14, C8H18, or methanol so that Wtheoretical is approximately the same as Hu. For hydrogen, the difference is approximately 6% [5-16]. 5.3.1.4 Exergy Loss in the Perfect Cycle From the basic characteristic of the efficiency in a perfect engine, we can see that the efficiency rises with the air–fuel ratio. To further discuss these results, we need to look at exergy loss. The specific exergy for a closed system is defined as follows:

(

)

eT ,p = uT − u0,T 0 − T0 ⋅ sT ,P − s0,T 0,p0 + p0 ⋅ ( v − v0 )

Rel. energy loss [%]

5.3.1.3 Efficiency of the Perfect Engine The efficiency η VM of the perfect engine is basically defined as follows:

Combustion

40

20 Exhaust 0

0

1

2

4

6

8

Air-fuel ratio

Figure 5.10 Exergy loss from combustion and exhaust (according to [5-16]).

The relative exergy loss of the exhaust falls as the air–fuel ratio increases, whereas the relative exergy loss of combustion rises as the air–fuel ratio increases. The overall result is an increase in efficiency with the air–fuel ratio.

5.3.2 Approximation of the Real Working Cycle

The simple cyclical processes, as well as the process of the perfect engine, provide only limited information as to the real processes occurring in the engine. Models are therefore necessary to further approximate the real process. In particular, information on the indicated average pressure, internal effectiveness, combustion processes (combustion functions), combustion temperatures, pollutant formation, etc., is desirable. Such information is obtained from models that, for example, can be described as two-zone models. Additional model calculations are possible that are based on the specified injection rate that can be used to gain information on the combustion and NO emissions [5-8], [5-17], [5-18] that use single-zone models with a set substitute combustion characteristic [5-19].

(5.18)

42 | Internal Combustion Engine Handbook

6606_Book.indb 42

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.3 Open Comparative Processes

Many of these models do not include the reaction process but instead use suitable functions that describe the energy released from combustion [5-20], such as the Vibe function [5-21]. More extensive thermodynamic analyses can yield models that use local coordinates in addition to the progress over time of parameters. However, because of their multidimensionality, these require a large amount of computing.

(oxygen–hydrogen–carbon) equilibrium. The imaginary, spatially nonexisting flame front, which is also considered to be without mass, separates both zones. These models, which take reaction-kinetic effects into account, permit conclusions about the NOx formation in a defined combustion process [5-9]. Figure 5.11 shows a schematic representation of the two-zone model.

5.3.2.1 Zero-Dimensional Models Simple models for the work cycle calculation of combustion engines are “zero-dimensional” models also known as the filling and emptying method [5-22]. The process variables are dependent only on time, but not on location. Two- or three-dimensional flow fields are ignored. The combustion chamber and the adjacent gas-conducting assemblies such as intake and exhaust pipes or containers, valves, etc., are physically/mathematically described relative to the inflow and outflow processes. The simulation system is created from containers with corresponding volumes, flow resistors (throttles and baffles), and piping. For example, the throttles and baffles simulate throttle valves, EGR valves, and significant diameter changes [5-23]. For charged engines, the compressor and turbine are considered by using corresponding maps. The result of the differential equation system of the balance equations for mass and energy returns the mass, temperature, and pressure variables in the corresponding model element, taking the thermal state equation into account. In the single-zone model as a “zero-dimensional” model, a substitute combustion characteristic is frequently used to specify the release of the heat in the chemical energy of the fuel. The system is described in the mass and energy balance within defined system limits. The necessary balance equations take into account the various options for fuel supply into the combustion chamber; the resulting differing parameters are important for the treatment (vaporization) of the fuel. In addition to the release of energy, assumptions regarding the thermal transition [5-6], [5-7], [5-24], [5-25] and charge cycle are required. Another variant of the “zero-dimensional” models are two-zone models. The combustion chamber is divided into two zones separated by a so-called flame front. Zone I represents the uncombusted air–fuel mixture and Zone II the combustion products in the OHC

5.3.2.2 Models to Determine Combustion Behavior Because it is practically impossible to directly determine the conversion of material over time during combustion in the engine, model calculations are used. Despite the simplification, experience shows that they can yield at least very good qualitative information. We now discuss a model based on thermodynamics that is defined as follows:

Volume change work Zone 1

Zone 1 Air and fuels

•• The use of the pressure characteristic is measured in the engine for calculating the cycle. •• At the time of ignition, the contents within the cylinder consist of residual exhaust gas and a fresh mixture. •• The mass flowing into the cylinder remains completely in the cylinder (no mass loss). •• During compression, no chemical reactions occur. •• The charge in the cylinder during combustion consists of two homogeneous areas in reference to pressure, temperature, and composition (area I = noncombusted material and area II = combusted material). •• The two homogeneous areas are separated by an infinitesimally thin flame front and exchange mass but no heat. •• The state change of area I occurs at constant enthalpy. •• The gas leaving the “flame front” is conveyed into area II and mixes with it to form a new state of equilibrium. •• The transfer of heat among the respective areas (“combusted”, “noncombusted”) to the combustion chamber wall occurs according to fixed laws. •• The composition in area I does not change during combustion. The goals are to determine temperature as a function of time in the combusted and uncombusted materials, the specific moles in the combusted material, and the so-called combustion function that expresses the ratio of combusted fuel mass to the overall fuel mass. In the uncombusted material, the specific moles do not change by definition. These quantities

Zone 2 Combustion products

Volume change work Zone 2

OHC equilibrium Wall heat losses Zone 1

Wall heat losses Zone 2 Infinitesimally thin flame front

System limit

Figure 5.11 Schematic representation of the two-zone model.

Internal Combustion Engine Handbook | 43

6606_Book.indb 43

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

can yield information on the combustion speed, the duration of combustion, and the combustion delay. The process is then calculated with the following steps as a function of time or the crankshaft angle α . 1. Cylinder charge and the beginning of the reaction The temperature can be determined with the thermal state equation p ⋅ v = ∑ s i ⋅ Rm ⋅T



kII

dTII = da



1 kI

∑s i=1



iI

⋅ c p mi (TI )

xB =



mII mtotal

Accordingly, r equations result in the following form: ⎛ ⎞ k ⎜ ⎟ s p dp R ⋅T II dT ∑ vi, j ⋅ ⎜⎜ Smi0 (TII ) − Rm ⋅ ln kII iII ⋅ p0 ⎟⎟ ⋅ daII − da ⋅ mp II ⋅ ∑ vi, j i=1 i=1 ⎜ ∑ s iII ⎟⎠ ⎝ i=1 kII

kII ⎛ ⎞ vi, j ⎟ ⎜v ∑ ds i, j = Rm ⋅TII ⋅ ∑ vi, j ⋅ ⎜ − i=1 ⎟ ⋅ iII ⎜ s iII kII ⎟ da i=1 ⎜ ∑ s iII ⎟⎠ ⎝ i=1 kII



and b equations from the basic material balance:

(5.24)



(5.25)

1 kII

xB ⋅ ∑ s i, j ⋅ cpmi (TII )

kII



−xB ⋅ ∑ H mi (TII ) ⋅ i=1

(5.26)

ds i,II da

The equation for the percent fuel conversion is 1 dxB = kII kI ⎞ da R ⎛ m ⋅ ⎜ TII ⋅ ∑ s i,II − T ⋅ ∑ s i,I ⎟ p ⎝ i=1 ⎠ I i=1

⎛ dq R ⋅T dp kI ⎞ ⋅⎜ I + m I ⋅ ⋅ ∑ s iI ⎟ (5.23) p da i=1 ⎠ ⎝ da

Zone II (combusted material): The unknown quantities are the k II specific moles in the combusted material σ i II, the temperature TII, and the converted mixture mass. It is useful to use the components CO2, CO, OH, H, O, O2, H2O, H2, and N2 as inert components for the gas composition in zone II. r-independent equations for the chemical equilibrium and b equations from the basic material balances (k II = r + b) serve to determine the k II specific moles. The equation system is completed with an independent equation for the temperature in the combusted material and the material conversion. This is characterized by the combustion function that is defined as follows:

ds i,II = 0 with l = 1…b da

i=1

i

dTI = da

⋅

kII ⎛ ⎛ kII ⎞ dx ⎞ ⎜ ⎜ ∑ hi,Flame − ∑ s i, j ⋅ H mi (TII )⎟ B ⎟ i=1 ⎜ ⎝ i=1 ⎠ da ⎟ ×⎜ ⎟ k ⎜ ⎟ dqII xB dp II + ⋅ Rm ⋅TII ⋅ ⋅ ∑ s i, j ⎟ ⎜ +xB ⋅ da p da i=1 ⎠ ⎝

and the empirically determined quantities of combustion chamber pressure, the volume above the piston, and fresh gas composition.

(5.22)

i,l

The equations for the temperature of the noncombusted material include

i

p ⋅ vI = ∑ s iI ⋅ Rm ⋅TI and

i=1



(5.21)

2. Combustion process Zone I of noncombusted material: The thermal state equation and the first law of thermodynamics for open systems yield

∑a

⎛ dV xB ⋅ Rm ( 1 − xB ) ⋅ Rm ⎞ ⋅ ⎟ ⎜ m ⋅da − p − p ⎜ ⎟ ⋅⎜ kI kI ⎛ ⎞⎟ ⎜ ⎜ TII ⋅ ∑ ds − TI ⋅ dp ∑ ds ⎟ ⎟ ⎜⎝ ⎝ da i=1 i,I p da i=1 i,I ⎠ ⎟⎠



kII ⎛ dT kII ds T dp × ⎜ II ⋅ ∑ s i,II + TII ⋅ ∑ i,II − II ⋅ da da p da i=1 i=1 ⎝

(5.27)

kII

∑ ds i=1

i,II

⎞ ⎟ ⎠

There are accordingly k II + 3 equations for determining the combustion function xB, the temperature in the uncombusted material TI, the temperature of the combusted material TII, and the composition of the combustion materials σ 1,II … σ 8,II. Typical definitions that can be represented with such models are shown in Figure 5.12 and Figure 5.13. 5.3.2.3 Multidimensional Models Multidimensional models describe the processes during process simulation of the engine behavior as a function of time and place, taking three local coordinates into account. The inflow behavior and the flow behavior in the cylinder can be well represented. This is important for the calculation and consideration of process-relevant parameters such as swirl and tumble formation. Discharge processes, charge cycle, EGR, and more can also be mapped. These CFD simulations are time consuming because of the mesh generation required for the calculation and determination of the corresponding initial and boundary conditions [5-26], [5-27]. Furthermore, the degree of complexity is significantly increased when reaction-kinetic processes and energy and material transport processes are included in the calculations.

44 | Internal Combustion Engine Handbook

6606_Book.indb 44

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.4 Efficiency

100

l = 0,92

%

Base: Pressure progression in the combustion chamber at n = 4000 1/min Optimal ignition timing

3000 K

2500

50 OT

Flame temperature

Temperature

Converted mixture mass

l = 1,6

l = 2,97

2000

OT OT 0

0

50

100

Angle since ignition

°kW

1500 150

Figure 5.12 Calculated combustion functions and flame temperatures in a two-zone model (Methanol–H2).

known only in part or fragmentarily—for example, gasoline being a mixture of a multitude of individual components. For this reason, model calculations are frequently possible only with individual components of the fuel hydrocarbon mixture. A closer approximation to the real combustion process in the engine compared with, for example, zero-dimensional two-zone models can be achieved when transport processes such as diffusion and thermal transmission in the gas are included in the model calculation. This requires the description of both the temporal and local behavior of important process quantities. The necessary formulation of the balance equations is based on the thermodynamics of irreversible processes. Continuous systems are considered, that is, the intensive state variables such as the temperature, pressure, and density are always functions of time and place. The balance equations describe the local changes in each volume element. In addition to the source therm for production or decomposition of the permitted components, there is an exchange of energy and material with the neighboring element [5-4]. If friction influences and the temporal and local pressure gradients are not included, the essential equations to describe such systems are the quantity balance and energy balance. (a) Quantity balance: Taking into account chemical reactions and diffusion, the following results for the change of specific moles σ i:



6

Base: Pressure progression in the combustion chamber at n = 4000 1/min Optimal ignition timing

(

)

r ∂s i ∂s ∂I = −v ⋅ r i − i + ∑ v nj,i − v nn j,i ⋅ J j . ∂t ∂x ∂x j=1

(5.28)

(b) Energy balance: Not included are external force fields, friction effects, and local and temporal pressure gradients: k

∑s ⋅ r

l = 0,99

Combustion speed

r

i

i=1

4

k ∂H m,i = − ∑ H m,i ∂t i=1

∑ ⋅ ( vnj,i − vnn j,i ) ⋅ J r

j

j=1

 − div IQ ∑ s i ⋅ r ⋅ v ⋅degree H m,i (5.29) k

l = 0,7

i=1

 − ∑ I i ⋅degree H m,i k

l = 1,7

2

0



l = 2,2

0

50

100

Angle since ignition

°kW

150

Figure 5.13 Calculated combustion speeds in a two-zone model (Methanol–H2).

5.3.2.4 Multidimensional Model for the Combustion Simulation Without Fluid Mechanical Overlapping Even without an overlap of reaction-kinetic processes by fluid mechanical givens, the calculation effort is extremely high. The principal obstacle is the specific fuel reaction, which may be

i=1

(c) w here i means the number of permitted components in the gas, j means the number of permitted chemical reactions, nn characterizes the combusted materials, n is the noncombusted materials, Ij is the diffusion flow density, Jj is the reaction speed of reaction j, IO characterizes the heat flow, and Hm,i is the partial molar enthalpy of component i.

5.4 Efficiency The consideration of the simple cyclical processes (Section 5.2) yields efficiency defined as thermal efficiency η th which can be evaluated as the maximum possible efficiency depending on the selected process. Given the previously cited prerequisites, the “perfect engine” yields efficiency η v, which is less efficient than η th with the same process control.

Internal Combustion Engine Handbook | 45

6606_Book.indb 45

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 5 Fundamentals of Thermodynamics

As the computational models grow closer in their approximation of the real process, we grow further away from the ideal. The obtained efficiency continually falls and more closely approximates reality. The deviations in efficiency of the “perfect engine” from the internal efficiency η i of a real engine are determined by the following: •• Incomplete combustion and combustion process. The exhaust still contains components that can be further oxidized and hence represents a calorific value that is not exploited in the process. In addition, the real combustion process deviates from the comparative process. •• Leaks, heat losses, and charge cycle loss. The internal efficiency η i of a real engine can be determined by the indication of the high-pressure and low-pressure loops. The next step to obtain the effective efficiency η e is to consider additional loss such as friction loss (power train friction, accessories, auxiliary drives, etc.).

5.5.1 Balance Equation

If the material and energy flows that pass through the control chamber are balanced, the following results: H Fuel + H Air + H KWE = H KWA + Pc + Qres + H ExhaustT2 (5.30)



The energy difference from different gas flow speeds between entering and leaving the engine is not considered. The air and fuel are converted by means of a chemical process into exhaust. For calculation, the definition of the calorific value is used: Hu =



If an engine is operated while stationary, that is, with a fixed operating point, the process is a stationary flow process in which technical work is accomplished. To portray an energy balance, a system limit is defined, and the material and energy flows that go beyond this limit are considered (Figure 5.14). Fuel

Air

System limit

Effective power

Engine

Cooling power

(5.31)

where H 1′ is the enthalpy flow of the noncombusted materials at temperature T1 and H 1′′ of the enthalpy flow of the combusted materials (exhaust) at temperature T1. The temperature T1 of the combusted materials is attained by cooling the combusted materials to the initial temperature. The enthalpy flows are defined as H 1′ = H Air + H Fuel and H 1′′ = H Exhaust T1



5.5 Energy Balance in the Engine

H 1′ − H 1′′  Fuel m

(5.32)

It accordingly follows that  Fuel + H Exhaust T1 (5.33) H − H KWE + Pe + Q Res + H Exhaust T2 = H um KWA or  Fuel = ∆H KW + Pe + Q Res + ∆H Exhaust H um



(5.34)

It must be remembered that ∆ḢExhaust is the enthalpy difference between the exhaust at the respective exhaust temperature T2 and temperature T1. From the preceding equation, we can clearly see the distribution of the energy supplied by the fuel or the calorific value. It is divided into effective power, residual heat, the enthalpy difference of the cooling water, and the enthalpy difference of the exhaust gas. The enthalpy of the cooling water is calculated with the following equation:  KW ⋅ c W ⋅ (TKWA − TKWE ) ∆H KW = m



(5.35)

with Residual heat Exhaust

Figure 5.14 Material and energy flows in the engine.

In particular, the following flows go beyond the system limit:

ṁKW = Flow rate of the cooling water, cW = Specific heat of water (4.185 kJ/kg K), TKWA = Cooling water temperature at outlet TKWE = Cooling water temperature at inlet. The enthalpy difference of the exhaust is found with the equation:

(

Pe

= Effective performance

QRes

= heat (flow of heat into the environment because of heat radiation, heat conduction, and convection)



ḢAir

= Enthalpy flow of the air

ḢFuel

= Enthalpy flow of the fuel

ṁExhaust = Mass flow of the exhaust

ḢKWE

= Enthalpy flow of the cooling water (inlet)

c p Exhaust = Average specific heat of the exhaust. 0

ḢKWA

= Enthalpy flow of the cooling water (outlet)

 Exhaust = m L +m  Kr . The exhaust mass flow is m

T2

T1

0

0

)

 Exhaust ⋅ c p Exhaust T2 − c p Exhaust T1 ∆H Exhaust = m

(5.36)

with T

Ḣexhaust = Enthalpy flow of the exhaust gas.

46 | Internal Combustion Engine Handbook

6606_Book.indb 46

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

5.5 Energy Balance in the Engine

The residual heat that essentially consists of the radiated heat, conducted heat, and convection can accordingly be calculated because all other quantities can be calculated from the measured data:

 Kr − Pe − ∆H KWA − ∆H Exhaust Q Res = H u ⋅ m

(5.37)

Bibliography

5-1. Behr, H.D. 1989. Thermodynamik. Springer: Berlin; Heidelberg; New York. 5-2. Pischinger, R., Kraßnig, G., Taucar, G., and Sams, T. 1989. Thermodynamik der Verbrennungskraftmaschine. Die Verbrennungskraftmaschine Neue Folge Band 5. Springer: Wien. 5-3. Heywood, J.B. 1988. Internal Combustion Engine Fundamentals. McGraw Hill International Editions: New York. 5-4. Schäfer, F. 1983. Thermodynamische Untersuchung der Reaktion von Methanol-Luft-Gemischen unter der Wirkung von Wasserstoffzusatz, VDI Fortschrittberichte, Reihe 6, Energietechnik/Wärmetechnik Nr. 120. VDI Verlag: Düsseldorf. 5-5. Eiglmeier, C. and Merker, G.P. “neue Ansätze zur phänomenologischen Modellierung des gasseitigen Wandwärmeübergangs im Dieselmotor,” MTZ 61:5, 2000.

5-13. Ohyama, Y. and Yoshishige, O. Engine Control Using a Real Time Combustion Model, SAE 2001-01-0256. 5-14. van Basshuysen, R. and Schäfer, F (Hrsg.). 2006. Lexikon Motorentechnik. Der Verbrennungsmotor von A–Z. Vieweg Verlag: Wiesbaden. 5-15. Zima, S. Unveröffentlichte Darstellungen 5-16. Jordan, W. 1977. Erweiterung des ottomotorischen Betriebsbereiches durch Verwendung extrem magerer Gemische unter Einsatz von Wasserstoff als Zusatzkraftstoff, Dissertation, Universität Kaiserslautern. 5-17. Chmela, F., Orthaber, G., and Schuster, W. “Die Vorausberechnung des Brennverlaufs von Dieselmotoren mit direkter Einspritzung auf der Basis des Einspritzverlaufs,” MTZ 59:7, 1998. 5-18. Sams, T., Regner, G., and Chmela, F. “Integration von Simulationswerkzeugen zur Optimierung von Motorkonzepten,” MTZ 61:91, 2000. 5-19. Barba, C., Burkhard, C., Boulouchos, K., and Bargende, M. “Empirisches Modell zur Vorausberechnung des Brennverlaufs bei Common-Rail-Dieselmotoren,” MTZ 60:4, 1999. 5-20. Codan, E. “Ein Programm zur Simulation des thermodynamischen Arbeitsprozesses des Dieselmotors,” MTZ 57:5, 1996. 5-21. Vibe, I. 1970. Brennverlauf und Kreisprozess von Verbrennungsmotoren. VEB Verlag Technik: Berlin. 5-22. Ramos, J.I. 1998. Internal Combustion Engine Modelling, Hemisphere Publishing Corporation: New York.

5-6. Bargende, M. 1990. Ein Gleichungsansatz zur Berechnung der instationären Wandwärmeverluste im Hochdruckteil von Ottomotoren, Dissertation, TH Darmstadt.

5-23. Seiffert, H. 1962. Instationäre Strömungsvorgänge in Rohrleitungen an Verbrennungskraftmaschinen. Springer Verlag: Berlin; Göttingen; Heidelberg.

5-7. Woschni, G. “Die Berechnung der Wandwärmeverluste und der thermischen Belastung der Bauteile von Dieselmotoren,” MTZ 31, 1970.

5-24. Kleinschmidt, W. 1993. Der Wärmeübergang in aufgeladenen Dieselmotoren aus neuerer Sicht, 5. Aufladetechnische Konferenz, Augsburg.

5-8. Mollenhauer, K. 1997. Handbuch Dieselmotoren, Springer.

5-25. Merker, G.P. and Kessen, U. 1999. Technische Verbrennung: Verbrennungsmotoren, Teubner Verlag: Stuttgart.

5-9. Heider, G., Woschni, G., and Zeilinger, K. “2-Zonen Rechenmodell zur Vorausberechnung der NO-Emission von Dieselmotoren,” MTZ 59:11, 1998. 5-10. Torkzadeh, D.D., Längst, W., and Kiencke, U. Combustion and Exhaust Gas Modeling of a Common Rail Diesel Engine – an Approach. SAE 2001-01-1243. 5-11. Jungbluth, G. and Noske, G. “Ein quasidimensionales Modell zur Beschreibung des ottomotorischen Verbrennungsablaufs,” Teil 1 und Teil 2. MTZ 52, 1991. 5-12. Stiech, G. 1999. Phänomenologisches Multizonen-Modell der Verbrennung und Schadstoffbildung im Dieselmotor, VDI Fortschrittberichte, Reihe 12, Verkehrstechnik/Fahrzeugtechnik Nr. 399. VDI Verlag: Düsseldorf.

5-26. Ferziger, J.H. and Peric, M. 1996. Computational Methods for Fluid Dynamics. Springer: Berlin; Heidelberg; New York. 5-27. Merker, G., Schwarz, C., Stiesch, G., and Otto, F. 2004. Verbrennungsmotoren Simulation der Verbrennung und Schadstoffbildung. Teubner Verlag. 5-28. Schwaderlapp, M., Bick, W., Duesemann, M., and Kauth, J. 200 bar Spitzendruck, Leichtbaulösungen für zukünftige Dieselmotor-blöcke. MTZ Motortechnische Zeitschrift 65, 2004. Vieweg Verlag: Wiesbaden.

Internal Combustion Engine Handbook | 47

6606_Book.indb 47

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 48

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6 Crank Gears 6.1 Crankshaft Drive 6.1.1 Design and Function

The crank gear, a colloquial term for the crankshaft drive, is a functional group in reciprocating piston engines that efficiently transforms oscillating movement (back-and-forth movement) into rotary movement (and vice versa), but is also excellent at converting thermodynamic processes to yield the maximum work, efficiency, and technical feasibility. These advantages are gained at the cost of serious disadvantages, however:

•• is articulated to the small conrod eye by the piston pin and also moves back and forth. •• with the large conrod eye—articulated to the crank pin—also rotates. •• with the conrod shaft swings within the plane of the crank circle. •• The crankshaft rotates (Figure 6.2).

•• limitation of speed—and, hence, the development of power— due to free inertia •• uneven force transmission that requires special measures in the form of multiple cylinder crank gears, a suitable throw and firing sequence, mass balancing, and mass balancing gears •• excitation of rotational oscillations that place a great deal of stress on the crankshaft and the drivetrain •• high fluctuations in the force characteristics in comparison to the nominal values for these forces •• problematic component geometry in regard to the flow of force with high stress peaks •• tribological problem. The crankshaft drive consists of pistons with rings, piston pins, conrods (connecting rods), a crankshaft with countermass(es) (counterweights), bearings (connecting rod bushing, connecting rod bearing, and crankshaft main bearing), and the lubricant (Figure 6.1). In the following discussion, we refer to the kinematically relevant parts of the crankshaft drive. The individual parts of the crankshaft drive execute various movements: •• The piston oscillates in the cylinder. •• The conrod

Figure 6.1 Crank gear of a V-8 passenger car SI engine.

During a single rotation of the crankshaft, the piston moves from top to bottom and returns to top dead center (TDC); it, thereby, executes two strokes. It accelerates and decelerates while executing this movement. The crank gear movement, that is, the respective position of the piston, is described by the crankshaft angle φ —the angle between the cylinder axis and the crankshaft throw.

Internal Combustion Engine Handbook | 49

6606_Book.indb 49

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

s

l · cosj

l

Oscillating movement

r

Rotation

f

r · cosf

Rocking motion

l j

sx

s0

Oscillating movement

r Rotation

l · sinj = r · sinf Figure 6.2 Movements of the crank gear parts.

Figure 6.3 Geometric relationships in the crankshaft drive.

The crankshaft angle is a measure of both path and time, since it indicates the time in which the crank gear has reached a certain position independent of the respective speed. The following numerical value equation applies:

The relationship between the crankshaft anglej φ and the conrod angular travely ψ can be represented as follows:

j [°KW] = 6 ⋅ n [min ]⋅t [s] −1



(6.1)

The piston movement is calculated with the piston travel equation, that is, by the relationship of the piston travel to the crankshaft angle, s = f(φ ); it results from the geometric relationships (Figure 6.3). r = Crankshaft radius s = Piston travel l = Conrod length v = Piston speed

l=

r Conrod ratio l

y = arctan

l ⋅ sin j

1 − l 2 ⋅ sin 2j



(

(6.6)

)

1 s = r ⋅ ⎡⎢1 − cosj + ⋅ 1 − 1 − λ 2 ⋅ sin 2 j ⎤⎥ . λ ⎣ ⎦

(6.7)

Since it is difficult to use the radical in the piston travel equation, it is replaced by a quickly converging series that can be terminated after the second element, because both λ and sin φ are <1, and their exponents or products are much less (6.8)



1 1 1 1 + x = 1 + x − x 2 + x 3 −… 2 8 16



x = − l 2 ⋅ sin 2 j

(6.9)

The simplified piston travel equation is accordingly (Figure 6.4):

a = Piston acceleration

s0 = l + r

(6.2)



sx = r ⋅ cosj + l ⋅ cosy

(6.3)



s = s0 − sx

(6.4)

By including time, we gain the piston speed (Figure 6.5)



s = l + r − ( r ⋅ cosj + l ⋅ cosy )

(6.5)

1 ⎛ ⎞ v = r ⋅ w ⋅ ⎜ sin j + ⋅ l ⋅ sin 2j ⎟ . ⎝ ⎠ 2





1 ⎛ ⎞ s = r ⋅ ⎜ 1 − cosj + ⋅ l ⋅ sin 2 j ⎟ . ⎝ ⎠ 2

(6.10)

(6.11)

50 | Internal Combustion Engine Handbook

6606_Book.indb 50

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

The mean piston speed is the travel of two shafts completed during one rotation and relative to the corresponding time t = 1/n vm = 2 ⋅ s ⋅ n



(6.12)

harmonic movement. Large conrod ratios, that is, relative to the stroke of short connecting rods, reduce the engine height, yet they produce greater friction because of the stronger angle of the connecting rods. Currently used λ values for vehicle engines are mostly within 0.2 and 0.35.

By double differentiating the piston travel equation according to time, we gain the piston acceleration (Figure 6.6) a = r ⋅ w 2 ⋅ ( cosj + l ⋅ cos 2j )





15000

l = 0,35

(6.13)

l = 0,25

Piston acceleration [m/s 2]

10000

100 l=0

Piston travel [mm]

80

l = 0,15

60

l = 0,25

l = 0,15 0

0

45

90

135

180

–5000

40 l = 0,35

20

0

–10000 Crankshaft angle [°]

0

45 90 Crankshaft angle [°]

135

180

Figure 6.4 Piston travel as a function of the crankshaft angle for different conrod ratios.

Different approaches are used to try and reduce the oscillating masses or keep them from increasing despite an increase in output:

•• load-optimized wristpin geometry (for example, conic inner bores or so-called profiled piston pins with optimized outer contour)

20 l=0 l = 0,15

15

•• mass reduction in the area of the small conrod eye due to so-called trapezoid or stepped conrods

l = 0,25

•• clamp-type conrods (piston pin shrunk into the small conrod eye results in the elimination of piston pin securing ring and bearing bush)

l = 0,35

10

•• small conrod ratios to reduce the second-order inertial forces (portion superimposed on the harmonic vibration).

5

0

Figure 6.6 Piston acceleration as a function of crankshaft angle for different conrod ratios.

•• pistons with reduced compression heights, reduced piston ring height, and, in some cases, reduced number of piston rings—optimized internal geometry (drawn-in bolt eyes and reduced eye spacing)

25

Piston speed [m/s]

l=0

5000

0

45 90 Crankshaft angle [°]

135

180

Figure 6.5 Piston speed as a function of crankshaft angle for different conrod ratios.

The piston travel, speed, and acceleration are influenced by the conrod ratio λ . Assuming an infinitely long conrod (λ = 0), 1 2 the term 2 ⋅ l ⋅ sin j. is superimposed to the purely harmonic cosine vibration in the piston travel equation. In general, the bigger the conrod ratio λ , the larger the deviation from the

The masses increase substantially with the engine size and load; for example, the mass of the “naked” piston of the V8 Audi spark-ignition (SI) engine is 355 g [6-1] and that of the complete piston of the Porsche Carrera is 650 g [6-2]. By transposing or “deaxising” the crankshaft drive, the movement of the crankshaft drive can be changed as desired (Figure 6.7). There are •• axially offset crankshaft drives in which the piston pin is moved from the middle of the cylinder •• shifted crankshaft drives in which the crankshaft center is moved from the middle of the cylinder.

Internal Combustion Engine Handbook | 51

6606_Book.indb 51

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

Pressure side

Counterpressure side

Shifting

Counterpressure side

Pressure side

Deaxising

Figure 6.7 Shifting and deaxising the crankshaft drive.

It is also possible to combine shifting and deaxising. By shifting, the movement is changed so that the elongated positions of the crank gear no longer lie in the cylinder axis, the piston travel is no longer symmetrical with BDC, and the piston speeds in the advance and return strokes assume different values. Depending on whether the shifting is dome at the pressure or counter-pressure side, we have different signs for e. The piston travel, speed, and acceleration of the shifted crankshaft drive can be determined with the shift y that refers to the conrod length:

e=

y to [3; 4]: l

(6.14)

(6.15)



1 2 s = r ⋅ ⎡⎢ cosj + ⋅ 1 − ( l ⋅ sin j + e ) ⎤⎥ l ⎣ ⎦

(6.16)



⎡ cosj ⋅ ( l ⋅ sin j + e ) ⎤⎥ v = −r ⋅ w ⋅ ⎢ sin j + 2 ⎢ 1 − ( l ⋅ sin j + e ) ⎥⎦ ⎣

The forces in the crankshaft drive of an internal combustion engine arise from the gas pressure in the combustion chamber and from inertial forces (Figure 6.9). The shares of gas and inertial forces within the crank gear forces depend on the •• thermodynamic process: SI engine/diesel engine •• engine design: naturally aspirated engine/EGT engine •• load point in the map, for example: •• high gas force and low inertial force •• low gas force and high inertial force. Because of the nonuniform processes of work and movement in the reciprocating-piston engine, the size and direction of the forces in the crank gear change during a work cycle. The following act on the crank gear: •• oscillating inertial force of the piston •• oscillating inertial force of the conrod •• rotating inertial force of the conrod •• rotating inertial force of the crankshaft throw •• rotating inertial force of the countermass.

l ⋅ cos 2 j − sin j ⋅ ( l ⋅ sin j + e ) ⎤ ⎥ + 2 ⎥ 1 − ( l ⋅ sin j + e ) ⎦



6.1.2 Forces Acting on the Crankshaft Drive

•• gas force

⎡ l ⋅ cos 2 j ⋅ ( l ⋅ sin j + e ) a = −r ⋅ w 2 ⋅ ⎢ cosj + 2/3 ⎢ ⎡1 − ( l ⋅ sin j + e )2 ⎤ ⎢⎣ ⎣ ⎦

⎡ cosj ⋅ ( l ⋅ sin j + e ) ⎤⎥ − r ⋅ w 2 ⋅ ⎢ sin j + 2 ⎢ 1 − ( l ⋅ sin j + e ) ⎥⎦ ⎣

There are different reasons for shifting and deaxising. In the early period of engine construction, the crankshaft drive was limited to 1/10 of the stroke [6-5]. This was to align the connecting rod with the cylinder axis when it passed through TDC to reduce the normal force (piston-side force) around the ignition and, hence, reduce the load and wear. Today, shifting is used with VR engines (V-engines with V-angles between 10° and 20°) to allow for the necessary free travel of the opposing cylinder [6-6], [6-7]. Deaxising in the direction of pressure (direction in which the piston contacts the cylinder barrel in the expansion stroke) causes an earlier contact change for the piston when the normal force on the piston is weaker. The tilting movement of the piston causes it to first contact the cylinder with the “soft” bottom part (piston skirt), which reduces impact. One, therefore, speaks of deaxising to reduce noise. The optimum amount of axially offsetting has been determined experimentally; for a boxer engine, it is 0.9 mm, for example. For an opposed cylinder engine, this is, for example, 0.9 mm. In automotive diesel engines, thermal deaxising is used—axially offsetting to the counter-pressure side. This allows the piston (within the piston play) to stay more in the middle of the cylinder, which has a positive effect on the seal of the piston rings and counteracts the collection of carbon deposits on the fire land (Figure 6.8).

.

(6.17)

The inertial force from the rotating motion of the conrod is not included. In the following discussion, the cited forces are those that occur briefly after ignition TDC with a crankshaft angle of 30° after TDC (Figure 6.10).

52 | Internal Combustion Engine Handbook

6606_Book.indb 52

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

Influence of deaxising the piston pin on the change in contact after ignition TDC Pin not deaxised

Pin deaxised toward the minor thrust face Major thrust face

Minor Major thrust face thrust face Strong pulse, rigid liner

Minor thrust face Weak pulse, rigid liner

Weak pulse, yielding liner

Strong pulse, yielding liner

Slight cavitation Crankshaft direction of rotation

Strong cavitation Minor thrust face Pressure side Major thrust face Counterpressure side

Figure 6.8 Piston pin deaxising.

FK = FGas + FPiston osc + FConrod osc p FGas = p(j)⋅ AK AK = ⋅ d 2 4 Fosc = −mosc ⋅ r ⋅ w 2 ⋅ ( cosj + l ⋅ cos 2j )

Gas pressure in the cylinder

(6.18)

mosc = mPiston osc + mConrod osc Uneven movement of the valve gear parts



FK = p(j)⋅ APiston − r ⋅ w 2 ⋅ mosc ⋅ ( cosj + l ⋅ cos 2j )

FGas

Inertial forces

FPistons osz Valve gear forces, bearing forces

F

Conrod osz

Figure 6.9 Diagram of crank gear forces.

The gas pressure that arises from the combustion of the mixture depends on the amount and change of different influences, such as the

FConrod rot

•• thermodynamic process •• combustion process •• power •• operating point in the program map at which the engine is driven. The gas pressure is determined with a process calculation or by measurement (indication) (Figure 6.11). To express it simply as is often done, the oscillating inertial forces are summarized as a single force Fosc. This counters the gas force applied to the piston. The gas and inertial forces together yield the piston force FK.

FCrankshaft rot.

FCounterweight Figure 6.10 Forces acting on the crank gear.

Internal Combustion Engine Handbook | 53

6606_Book.indb 53

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

100 % Output

Gas pressure in the cylinder [bar]

160

120 86 %

80 71 % 40

Four-stroke diesel engine, exhaust gas turbocharging Performance points corresponding to a running resistance curve

48 %

39 %

0 300

360

420

480

Figure 6.11 Gas pressure characteristics of a supercharged diesel engine with direct fuel injection.

540

Crankshaft position [°KW]

FST =



FK cosy

(6.19)

FN = −FK ⋅ tan y



FK

Piston force FK

Since the connecting rod, apart from the dead centers, assumes a position that deviates from the direction of the cylinder axis, the piston force FK must be correspondingly diverted. This results in the rod force FST and the normal force, that is, perpendicular to the cylinder wall FN (also termed piston-side force) (Figure 6.12 and Figure 6.13).

180

360

(6.20)

Results become clearer when we look separately at the paths of the individual forces from the combustion chamber to the crankshaft bearing or engine suspension [6-11].

540

720

Crankshaft angle [°}

Figure 6.13 Characteristic of the piston force of a fast-running four-stroke diesel engine over a power cycle.

•• Forces acting on the piston The gas pressure acting on the piston produces the gas force; it is counteracted by the (oscillating) inertial force of the pistons. The sum of these two forces produces the piston force FK′ . The piston force alternates between positive and negative several times over the course of the power cycle of a four-stroke engine and subjects the piston to a dynamic load.

FN

FST FK



FK′ = FGas + FPiston osc

(6.21)



FPiston osc = −mPiston osc ⋅ r ⋅ w 2 ⋅ ( cosj + l + cos 2j )

(6.22)

•• Forces acting on the piston pin The piston force acting through the top of the piston and the bolt eyes on the piston pin FK′ . is diverted toward the connecting rod. A force parallelogram results from the ′ in the direction of the piston force FK′ , the rod force FST conrod, and a normal force FN′ that is normal (perpendicular) to the cylinder barrel. The wristpin is burdened by the rod force FST ′. FST ′ =

Figure 6.12 Division of the piston force.



FK′ FK′ = cosy 1 − l 2 ⋅ sin 2 j

(6.23)

54 | Internal Combustion Engine Handbook

6606_Book.indb 54

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive



l ⋅ sin j

1 − l 2 ⋅ sin 2 j



(6.24)

•• Forces acting on the connecting rod The oscillating conrod force Fconrod osc acts in the direction of the cylinder axis FConrod osc = −mConrod osc ⋅ r ⋅ w 2 ⋅ ( cosj + l + cos 2j ) . (6.25)



It divides into a component in the direction of the conrod and into a normal component. The first component reduces ′ to FST, and the latter reduces the the conrod force from FST normal force from FN′ to FN. (6.26)



1 F FST = FST = K ′ − FConrod osc ⋅ cosy cosy



FN = − FN′ + FConrod osc ⋅ tan y = −FK ⋅ tan y

(6.27)

FST Crankshaft angle FST

FN′ = − FK′ ⋅ tan y = − FK′ ⋅

180

360

540

720

Crankshaft angle [°}

Figure 6.15 Characteristic of the rod force of a fast-running four-stroke diesel engine over a power cycle.

Normal force characteristic FN

The sign change of the normal force FN means that this occurs several times during a power cycle (Figure 6.14).

FN

180

360

540

720

Crankshaft angle [°}

FT

FR

FST

Figure 6.14 Characteristic of the normal force of a fast-running diesel engine over the power cycle.

Figure 6.16 Division of the rod force.

•• In a cold engine, it becomes manifested in light metal pistons by an annoying noise and piston rattle, which can be reduced by so-called control pistons and/or deaxising. •• The wet cylinder bushings are excited to execute vibrations that the coolant cannot follow, and cavitation may occur. The rod force FST is directed to the crank pin (Figure 6.15). The crank pin rotates under the effect of the rod force along the circle of rotation of the crankshaft radius. Combined with the tangential component of the rod force, the tangential force (FT), the crankshaft radius yields the torque M (Figure 6.16 and Figure 6.17).



FT = FST ⋅ sin ( j + y ) = FK ⋅

sin ( j + y ) cosy

(6.28)

Tangential force FT

The piston is pushed from one side of the cylinder barrel to the other with undesirable consequences: FT

180

360

540

720

Crankshaft angle [°}

Figure 6.17 Tangential force characteristic of a fast-running diesel engine.

The radial component, the radial force FR, does not contribute to the engine torque; it is applied only to the crankshaft throw upon bending (Figure 6.18), and it is a powerless or blind force.

Internal Combustion Engine Handbook | 55

6606_Book.indb 55

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

The crank pin is subject to the rod force FST and the rotating inertial force of the conrod FPL rot. Added geometrically, these forces yield the crank pin force FHZ.

Radial force FR

FR



2 (6.33) FHZ = FST2 + FConrod rot − 2 ⋅ FST ⋅ FConrod rot ⋅ cos ( j + y )

As a reaction to the crank pin force FHZ, the conrod bearing force FPL acts on the conrod bearing (Figure 6.20). 180

360

540

720

FPL = −FHZ



(6.34)

Crankshaft angle [°}

Figure 6.18 Radial force characteristic of a fast-running diesel engine.

FR = FST ⋅ cos ( j + y ) = FK ⋅



cos ( j + y ) cosy

(6.29)

Since every action yields an equal and opposite reaction; a torque opposite—the useful torque on the engine block necessarily arises the reaction torque. This results from the normal force FN and the distance of the normal force b that changes with the piston position. M = FT ⋅ r

MReaction = FN ⋅b

b = r ⋅ cosj + l ⋅ cosy





FConrod bearing

FConrod rot

(6.30)

Hence, the supporting forces FA and FB result from the reaction torque and their distance a to (Figure 6.19) FA = −



FB =



MReaction a

FHZ

(6.31)

MReaction . a

(6.32)

FST Figure 6.20 Crank pin force.

When the size and direction of forces change during a power cycle, as is the case with the conrod bearing force, for example, these forces are represented in polar diagrams by plotting them under the respective angle of their force-transfer direction in the sequence of the crankshaft angle (Figure 6.21). The crankshaft angle and the angle of force are not identical. In order to follow the change of the force over time, the crankshaft angle must be given for the individual points of the force characteristic. It is frequently useful to refer the forces to the following different coordinate systems (Figure 6.22):

MR

FN

r

b

FB

FA

FT

M

•• fixed spatial (or housing) system (main bearing forces, for example) •• fixed pin system (such as the effect of the forces on rotating pins) •• fixed rod system (such as the effect of forces on the conrod bearing).

a

Figure 6.19 Action torque, reaction torque, and supporting forces [6-30].

The crank gear forces are transmitted through the main bearing pin and the main bearing to the crankcase. The rotating inertial force of the crankshaft throws FKR rot, the crank pin force FHZ or its components FT, FR, and Fconrod rot, and the forces

56 | Internal Combustion Engine Handbook

6606_Book.indb 56

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

Polar diagram of the conrod bearing force (fixed shell) Four-stroke diesel engine 0° 360° 720°

15° 10° 5°

20°

330° 690°

25°

30° 300° 0° 20°

30°

The force engagement angle is not identical to the crankshaft angle

35° 40°

710°

300° 660°

90° 420°

200° 500° 485°

610° 260°

455°

270° 630°

225°

540°

100°

80°

680°

430°

675°

60° 450°

290°

415° 240° 600°

385°

The numbers in the locus diagram of the conrod bearing force stand for the 210° corresponding 570° crankshaft angle.

380°

120° 480°

335°

150° 510°

180° 540°

Figure 6.21 Polar diagram of the conrod bearing force of a fastrunning diesel engine.

x

x

x y

y x

y

x y Fixed shell diagram Forces refer to a coordinate system defined in the bore (shell)

x Fixed pin diagram Forces refer to a coordinate system defined in the pin

Figure 6.22 Coordinate systems.

Internal Combustion Engine Handbook | 57

6606_Book.indb 57

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

of the countermass Fcountermass (counterweights) together to form the main bearing force FGL (Figure 6.23).

FGL =

( FKW rot + FR + FConrod rot − FCountermass ) + FT2

(6.35)

The rotating masses of the throw are related to the crank pin axis. mThrow = mCrank pin + 2 ⋅ mCrank web



mCrank web = mCrank ⋅



rCenter of gravity r

(6.36)



(6.37)

supporting forces resulting from each throw are added to yield the overall bearing force. It is useful to calculate the crank gear force by dividing the forces into their X and Y components, totaling the X and Y components—taking into account whether they are positive or negative—and geometrically adding these sums. The direction of the resultant force is obtained from the quotients of the X and Y components—taking into account that the tangent is periodic with π . The quadrants in which the angles lie are obtained from the sign of the individual components

Z=

(∑ X ) + (∑Y ) 2

g = arctan

2



∑ X . ∑Y

(6.38)

(6.39)

The gas force that presses the piston downward also attempts to lift the cylinder head. This is prevented by the cylinder head screws that hold the cylinder head on the cylinder crankcase. On the other hand, the gas force acts through the piston, conrod, and crankshaft on the crankshaft main bearings. These are held by the main bearing bridges (main bearing cap) and the main bearing screws. This closes the flow of force, and the crankcase intermediate wall is (dynamically) stressed from tension (Figure 6.24).

FKR rot FKR rot FHZ

FCountermass FGL Figure 6.23 Main bearing force.

As a reaction to the main bearing force FGL, the equal and opposite main bearing pin force FGZ arises. The main bearing force FGL is divided into the two main bearings, neighboring the crankshaft throw. Apart from single-cylinder engines, the crankshaft has more than two bearings and represents a statically indeterminate system. In view of the fluctuating gas pressure from work cycle to work cycle, the tolerances of the masses, the deformation of the crankshaft and the oil film, and the flexibility of the bearing, the supporting forces are not frequently determined with (apparent) precision. The crankshaft is viewed as consisting of individual throws that are articulated to each other. The difference between the results of the statically indeterminate system and the statically determinate system is slight and can be ignored for fundamental design in particular. The partial

Figure 6.24 Flow of force in the crankcase.

6.1.3 Tangential Force Characteristic and Average Tangential Force

The tangential force (torsional force) also fluctuates with the periodically changing gas and inertial forces. The average tangential force is calculated from the tangential force characteristic over a power cycle. The area enclosed by the tangential force and the diagram axes is a measure of the (indicated or internal) work Wi. If this work is related to the length of the

58 | Internal Combustion Engine Handbook

6606_Book.indb 58

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

Tangential force FT

power cycle, we get the average tangential force FTm. This is only a fraction of the maximum tangential force (Figure 6.25). jp

FTm =

1 ⋅ FT (j)⋅dj j p ∫0

(6.40)

( Sum of positive areas )

FTm

+ ( Sum of negative areas ) = ⋅ mF ⋅ mj jp

10 5 0 –5 360 540 720 Crankshaft angle [°]

Tangential force FT in [kN]

15

20 15 10 5 0 –5 0

25 20 15 10 5 0 –5 –10

180

360 540 720 Crankshaft angle [°]

Cylinders 1 + 2 + 3 + 4 + 5

0

180

450

540

630

720

Cylinders 1 + 2

20 15 10 5 0 –5 –10

0

180

Cylinders 1 + 2 + 3

360

540

720

Crankshaft angle [°]

360 540 720 Crankshaft angle [°] Cylinders 1 + 2 + 3 + 4

Tangential force FT in [kN]

Tangential force FT in [kN] Tangential force FT in [kN]

25

Cylinders 1

20

Tangential force FT in [kN]

Tangential force FT in [kN]

25

–10

360

This evens out the tangential force so that the fluctuations in the tangential force drop to a fraction of that of a singlecylinder crank gear even in a six-cylinder inline crank gear (Figure 6.26). The irregular torsional force characteristic results in fluctuations in the speed, because torsional force FT(φ ) above the average FTm accelerates the crank gear, and decelerates it when the force falls below the average. The fluctuation of the

To even out the tangential force characteristic and increase the power, engines are built with multiple cylinders, with certain exceptions. The tangential forces (torsional forces) of the individual cylinders add up in displaced phases corresponding to the angular ignition spacing over the crankshaft to form the overall torsional force on the clutch side of the engine.

25

270

Figure 6.25 Tangential force characteristic and average tangential force.

φ p = Length of the power cycle crankshaft degrees.

180

180

Crankshaft angle °KW

mφ = Measure of the angle

0

90

(6.41)

mF = Measure of force

–10

Mean tangential force FTm

25 20 15 10 5 0 –5 –10 0 25

180

360 540 720 Crankshaft angle [°]

Cylinders 1 + 2 + 3 + 4 + 5 + 6

20 15 10 5 0 –5 –10

0

180

360

540

720

Crankshaft angle [°]

Figure 6.26 Overlapping tangential forces of a four-stroke six-cylinder inline engine.

Internal Combustion Engine Handbook | 59

6606_Book.indb 59

1/19/16 8:29 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

energy supplied to the crank gear is termed work fluctuation WS. Given the moment of inertia I of the crank gear, we obtain

the following:

1 2 2 ⋅ I ⋅ ( wmax − wmin ) 2 1 = ⋅ I⋅ ( wmax − wmin) ⋅ ( wmax + wmin) 2

WS =



wm = 2 ⋅ p ⋅ n ≈

1 ⋅ ( wmax + wmin ) . 2

(6.42)

(6.43)

The speed fluctuation can be reduced with a flywheel. The flywheel acts as an energy accumulator that stores excess energy if a surplus of tangential force is present and releases this energy when tangential force is lacking. Depending on the type of machine driven by the engine, different requirements are placed on the constant velocity. The speed fluctuation is indicated by the cyclic irregularity δ . The smoother the engine is to run, the lower the cyclic irregularity δ has to be; in particular, when revving the engines under a load, the cyclic irregularity is unpleasant, since it causes the engine accessories to vibrate. d=

wmax − wmin wm

d=



WS WS or I = I ⋅ wm2 d ⋅ wm2

Pi = AK ⋅ s ⋅ z ⋅ wi ⋅ n⋅ i



Pi = Mi ⋅ w

w = 2⋅p ⋅n

Mi = FTm ⋅ r FTm =



s 2

AK ⋅ z ⋅ wi ⋅ ii p

wi = we ⋅

r=

•• The rotational oscillations that cause additional torque in the crankshaft. This vibrational torque can be a multiple of the other types of torque. Figure 6.27 demonstrates the overlapping of the individual stresses of a six-cylinder crankshaft over its individual throws. Stress on the throws of a crankshaft (diagram)

Stress from rotational oscillations

(6.45) (6.46)

The average tangential force can be derived from the internal power of the engine:

•• The pulsating torque results from the strongly fluctuating characteristic of the tangential force. The torsional forces of the individual cylinder add up corresponding to their phase shift (angular ignition spacing). On the clutch side, the pulsating torque evens out; however, the primary factor of the load on the crankshaft is the range of fluctuation in the individual throws.

(6.44)

WS = I ⋅ d ⋅ wm2



•• The useful torque or working torque from the average tangential force that adds up from throw to throw.

1 hm

Pe = FTm ⋅ r ⋅ 2 ⋅ p ⋅ n⋅ hm

AK = Piston surface r = Crankshaft radius s = Stroke z = Number of cylinders Pe = Effective power wi = Indicated specific work we = Effective specific work i = Cycles

η m = Mechanical efficiency.

Stress



The crankshaft is subject to a load by the following:

Stress from fluctuations on torsional force Stress from average torsional force

(6.47) (6.48)

Figure 6.27 Stresses on a six-cylinder crankshaft. (6.49)

6.1.4 Inertial Forces (6.50)

(6.51) (6.52)

In reciprocating-piston engines, inertial effects arise that originate from the movement of the crank gear parts. The inertial forces have both positive and negative effects: •• On the one hand, they are undesirable, since they generate additional loads and impair the development of power of the reciprocating-piston engine. •• On the other hand, they even out the release of force of the crank gear by compensating for the force arising from gas pressure peaks and, hence, reduce force and load. The crank gear executes rotating, oscillating, and swinging motions. To simplify the calculation, the crank gear is reduced to two mass points (Figure 6.28), in which the oscillating and rotating masses are viewed as concentrated: •• on the articulation point of the conrod on the piston (wristpin axis) •• on the articulation point of the conrod on the crankshaft (crank pin axis).

60 | Internal Combustion Engine Handbook

6606_Book.indb 60

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

Oscillating masses of piston and conrod l mosz

f

of the centers of gravity (a, b), so that the center of gravity of the conrods is retained. In connecting rods for automotive engines, this corresponds approximately to a ratio of 1/3 (oscillating mass) to 2/3 (rotating mass). (6.53)



a mConrod osc = ⋅ mConrod l b mConrod rot = ⋅ mConrod l

(6.54)



l j

mrot

f

r Rotating masses of conrod and crankshaft Figure 6.28 Reduction of the crankshaft drive to two mass points.

The conrod also executes a swinging motion (Figure 6.29) that results in inertial forces. In fast-running engines, this cannot be ignored.

These inertial forces and the inertial torque that they generate proceed in an outward direction as free forces and free torques that try and move the crankcase back and forth in a horizontal and perpendicular direction. In addition, they cause the engine axes to tip. These free forces and torques can be more or less compensated (even completely compensated with a corresponding effort) by countermasses (counterweights) and/ or by a corresponding number and arrangement of throws to make the engine externally stable. 6.1.4.1 Inertial Forces in Single-Cylinder Crank Gears In crank gears, rotating inertial force arises, as well as oscillating inertial forces of the first order and higher. If the demands of precision are not particularly high, only the oscillating inertial forces up to and including those of the second order are taken into account. •• Rotating inertial force The rotating inertial force is a centrifugal force; it stays the same at a constant engine speed, but its direction changes with the crankshaft angle. The rotating inertial force rotates at the crankshaft frequency. Its locus diagram is a circle. Frot = mrot ⋅ w 2 ⋅ r



(6.55)

•• Oscillating inertial forces The oscillating inertial forces act in the direction of the cylinder axis, and their size and sign (direction) change over the course of the piston stroke:

Fosc = mosc ⋅ w ⋅ r ⋅ ( cosj + l ⋅ cos 2j )

(6.56)



Fosc = mosc ⋅ w ⋅ r ⋅ cosj + mosc ⋅ w ⋅ r ⋅ l ⋅ cos 2j

(6.57)

2

2

•• First-order inertial force To be understood as an “order” in this context is “the frequency at which an event occurs in relationship to the crankshaft speed.” The amount of the first-order inertial force changes with the crankshaft frequency—hence “first order”—and changes direction twice per rotation. Figure 6.29 Conrod pattern: envelope curve of all conrod positions during a crankshaft rotation.

The mass of the conrod is divided into an oscillating and a rotating part inversely proportional to the respective spacing

FI osc = mosc ⋅ w 2 ⋅ r ⋅ cosj

(6.58)

•• Second-order inertial force The maximum is only the λ part of the oscillating first-order inertial force (Figure 6.32); its amount changes at twice the crankshaft frequency, and it changes direction four times per rotation.

Internal Combustion Engine Handbook | 61

6606_Book.indb 61

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

Resulting oscillating inertial force

Force

Oscillating inertial force 1st order

Oscillating inertial force 2nd order

90

270 180

0

360 Crankshaft angle [°]

Figure 6.30 Resulting oscillating inertial force at a singlecylinder crank gear.



FII osc = mosc ⋅ w 2 ⋅ r ⋅ l ⋅ cos 2j

(6.59)

The characteristics of the oscillating inertial forces of the first order and of the second order add to form the resulting oscillating inertial force (Figure 6.30). This overall inertial force for a cylinder results from the vectoral addition of the rotating and oscillating inertial forces of the first and second orders, and possibly the forces of a higher order (Figure 6.31).

6.1.4.2 Inertial Forces in a Two-Cylinder V-Crank Gear If two cylinders at an angle δ to each other act together on a crankshaft throw (V-engine), the inertial forces of both cylinders are added as vectors (Figure 6.32).

V angle d

B

A

Oscillating inertial force 2nd. order Oscillating inertial force 1st order

f

fB

fA

Rotating inertial force

Figure 6.32 Crankshaft angle designation in a V-crank gear.

The locus diagram of the rotating inertial forces of both cylinders is a circle, and the locus diagrams of the oscillating inertial forces (depending on the V-angle δ and the order of the force under consideration) can be circles, ellipses, and straight lines (Figure 6.33). •• Rotating inertial force As is the case with a single-cylinder crank gear, the resulting rotating inertial force is constant with a vector that revolves at the crankshaft speed. The rotating mass is composed of the rotating masses of the two conrods and the rotating mass of the crankshaft throw; its locus diagram is a circle.

Figure 6.31 Locus diagram of the resulting inertial force in a singlecylinder crank gear.



FV2 rot = mV2 ⋅ w 2 ⋅ r

(6.60)

62 | Internal Combustion Engine Handbook

6606_Book.indb 62

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

cylinder axes A and B. The momentary values of the inertial forces of the two cylinders determined in this manner are added vectorally, and they yield the resulting inertial force vector of the first order (Fosc I res) (Figure 6.34).

V angle d

•• Oscillating second-order oscillating inertial force

A

The resulting second-order inertial force also results from the inertial forces of the cylinders A and B.

B

2.0

Since the oscillating second-order inertial force changes at twice the crankshaft frequency, the vectoral rotary angle is twice that of the first order. This amount is the λ th portion of the first-order inertial force.

1.0



j A = 2 ⋅j + d

(6.66)



jB = 2 ⋅j − d

(6.67)



FII osc A = FII ⋅ cos ( 2j + d )

(6.68)



FII osc B = FII ⋅ cos ( 2j − d )

(6.69)



FII = l ⋅ mosc ⋅ w 2 ⋅ r

0°

0 30° 45°

1st order

80° 90°

V angle d

120° 135° 150°

1.0



0.5 2nd order

Figure 6.33 Locus diagrams of the free inertial forces of V-crank gears depending on the V-angle.



mV2 = 2 ⋅ mConrod rot + ( mKW rot − mCountermass )

(6.61)

•• Oscillating first-order oscillating inertial force The resulting oscillating first-order inertial force results from the vectoral addition of the inertial forces of the two cylinders A and B. If the crankshaft angle φ of the crankshaft throw is measured from the bisector of the V-angle, then the crankshaft angle of the cylinder A is φ A = φ + (δ /2) and the angle of the cylinder B is φ B = φ − (δ /2). Between the oscillating inertial forces of the cylinders A and B, there is an operating time difference equal to the V-angle δ . (6.62)



d⎞ ⎛ FI osc A = FI ⋅ cos ⎜ j + ⎟ ⎝ 2⎠

(6.63)



d⎞ ⎛ FI osc B = FI ⋅ cos ⎜ j − ⎟ ⎝ 2⎠



FI = mosc ⋅ r ⋅ w 2

(6.64)



FI osc res

d = 2 ⋅ FI ⋅ cos d ⋅ cos 2 j + sin 4 2

(6.70)

FII osc res = 2 ⋅ FII

180°

0



(6.65)

The resulting force can be graphically determined by representing the crankshaft throw in its respective position by a vector with the quantity FI. These vectors are projected onto

× cos 2 2j ⋅ ( cos 2d + cos d ) + sin 2 d ⋅ ( 1 − cos d ) .

(6.71)

The resulting force can be graphically determined by the momentary values of the oscillating second-order inertial forces for the cylinders A and B and adding them vectorally. The momentary value of the cylinder A is determined by plotting, from the cylinder axis A, the inertial force vector F2 at the angle φ A = 2φ + δ and projecting it onto the cylinder axis A. The momentary value of the cylinder B is obtained by plotting the vector F2 at the angle φ B = 2φ − δ , but counting from the cylinder axis B, and projecting it on the axis of the cylinder B. 6.1.4.3 Inertial Forces and Inertial Torque in Multicylinder Crank Gears The inertial forces in the individual throws produce torque corresponding to their distance from the engine’s center of gravity—inertial torque. The forces and torques are vectoral quantities, so that the force and torque vectors of the individual throws are shifted in the plane of gravity of the engine, and can be added to form resulting forces and torques. V-engines are two inline crank gears separated by the V-angle. Therefore, the mass effect of one line of crank gears can be determined and added to the other, phase shifted by the V-angle. Or, the resultant force of the crank gears opposing each other across the V-angle can be added like the inline crank gear. The inertial effects are determined by the position of the respective throws (Figure 6.35). •• Inertial forces The rotating forces act in the direction of throw, while vectors rotating in the opposite direction represent the oscillating forces. By projecting the crankshaft throws on the plane of gravity of the engine, that is, the throw or crank diagram (also termed phase direction diagram), the directions of the inertial force vectors are represented.

Internal Combustion Engine Handbook | 63

6606_Book.indb 63

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

A

B

V angle d

fA

f fB d = 60°

de

30°

rA

Cy

de

lin

330°

lin

Cy

rB

360°/0°

300°

mosz · r · v 2 · cosfB

fA

60° F1res

f fB

270°

90°

mosz · r · v 2 · cos fA 240°

120°

mosz · r · v 2

210°

150°

Figure 6.34 Locus diagram of the oscillating first-order inertial force of a 2-V 60° crank gear.

180°

As a reference, the first throw (depending on whether you are counting from the force transmission side or the counterforce transmission) is in the TDC position. The position of the following throws is determined by the respective throw spacing (throw angle).

2a 2a

a a 5

4

1

1

3

2

3

1

5

4

72° 288° 144° 216°

Throw 1st order 2

5

3

2

72° 288° 144° 216°

For the oscillating second-order inertial force, the throw diagram of the second order (phase direction diagram of the second order) is used that is obtained by placing the throws under twice the throw angle. •• Inertial torques

4 Throw 2nd order

Figure 6.35 Crank diagram of a five-cylinder inline engine (ZF 12453).

The torque vector is perpendicular to its plane of action. The sign depends on the position of the relevant throw in reference to the engine’s center of gravity; it, therefore, must be correspondingly taken into consideration. If the throw is to the left of the center of gravity, the vector is positive; if it is to the right, the counting proceeds in a negative direction. From the perspective of the torque diagram, the vectors illustrate the torque that originates from the

64 | Internal Combustion Engine Handbook

6606_Book.indb 64

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

forces to the left of the engine’s center of gravity proceeding from the midpoint of the crankshaft, and proceeding toward the midpoint for torque to the right of the center of gravity. Because the torque vector is perpendicular to its plane of action, that is, perpendicular to its throw, the torque diagram follows the crank diagram by 90°. The torque vectors can, therefore, be drawn in the direction of throw, and the vector of the resulting torque can be set back 90° counterclockwise. In V-engines, the inertial forces of the two cylinders acting on a throw are combined and used to determine the inertial torque.

∑M



X

= a ⋅ Frot ⋅ 0.363





∑ MY = a ⋅ Frot ⋅ ( 2 ⋅ cos 0° + cos 216° + cos 252° + 2 ⋅ cos108°)



∑ MY = a ⋅ Frot ⋅ 0.264



Mrot res = a ⋅ Frot ⋅ 0.3632 + 0.2642 = a ⋅ Frot ⋅ 0.4488 tan d =







0.363 = 1.375 0.264

d = 54°





•• Rotating inertial torques Throw 1st order

The inertial torques result from the rotating inertial force and the respective distance from the plane of gravity. It is correspondingly geometrically added to the throw diagram.

MZyl 1

1

•• Oscillating inertial torques •• Oscillating first-order oscillating inertial torque The vectors of oscillating first-order inertial torque are plotted in the direction of the throw diagram of the first order. After adding the vectors, the resulting torque vector is projected onto the cylinder axis, because the oscillating forces act only in the direction of the cylinder axis. The projection is rotated 90° counterclockwise; this is then the resulting oscillating first-order inertial torque.

MZyl 5

5

MZyl 1 4 108°

MZyl 4

MZyl 4

MZyl 2

MZyl 2

252°

MZyl 5

216°

3

2

MZyl 2

•• Oscillating second-order oscillating inertial torque The same procedure is used for the oscillating secondorder inertial torque, except the throw diagram of the second order is used as a basis.

MZyl 1

MZyl 4 MZyl 5

d 1

Mrot res

6.1.4.4 Example To illustrate these relationships, the functions of a five-stroke shaft are graphed and analyzed. We assume the following:

Mres

2

•• equivalent masses of the crank gears in all the throws

3

4

5 5

4

2

3

1

•• equivalent cylinder spacing a •• the engine’s center of gravity is in the middle of the engine in the crankshaft axis •• the first crankshaft throw is in the TDC position.

Figure 6.36 Determining the resulting rotating inertial torque.

6.1.4.4.1 Rotating Inertial Torque The throw spacing in the throw diagram of the first order is

6.1.4.4.2 Oscillating First-order Inertial Force The effective directions of the vectors are the same as for the rotating inertial torque (Figure 6.37) as is the calculation:

α 1 = 0°



FI = mosc ⋅ r ⋅ ω 2



Mosc I max = a ⋅ FI ⋅ 0.3632 + 0.2642 = a ⋅ FI ⋅ 0.4488

α 2 = 216°

α 3 = 144° (not used since the throw is in the center of gravity) α 4 = 72°

α 5 = 288°.

0.363 tan d = = 1.375 0.264



Taking into consideration the sign of the torque of the individual throws, we get the effective directions of the torque (Figure 6.36):



φ 1 = 0°

φ 2 = 216°

φ 3 not used

72° (+ 180°) = 252°; φ 4 = 252°

6.1.4.4.3 Oscillating Second-order Inertial Force The throw spacing in the throw diagram of the second order is

288° (+ 180°) − (360°) = 108°; φ 5 = 108°

α 1 = 0°

α 2 = 72°

Frot = mrot ⋅ r ⋅ ω 2

α 3 = 288° not used

α 4 = 144°

α 5 = 216°.



∑M

X

= a ⋅ Frot ⋅ ( 2 ⋅ sin 0° + sin 216° + sin 252° + 2 ⋅ sin 108°)

d = 54°





Internal Combustion Engine Handbook | 65

6606_Book.indb 65

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

Throw 1st order

Throw 2nd order MZyl 1

1

MZyl 1

1

MZyl 5

108°

MZyl 4

MZyl 2

MZyl 2

252°

MZyl 5

216°

MZyl 2

3

2

max

3

3

2

Zy l5

M

4

2

4

1

yl

Mosz 1 res

4

3

MZ

5

Mosz 1 max

2

5

1

d 4

5

4 d

M*osz 1 res

MZyl 2

1

MZyl 4

Vectors are enlarged

MZyl 4

MZyl 1

MZyl 5

38° 72°

4

5

MZyl 2

324°

2

MZyl 4

MZyl 5

3

2

MZyl 1

3

sz 2

4

MZyl 4

M*osz 1 res

MZyl 5

5

Mo

MZyl 1

M Zyl MZyl 1

5

1

2

Mosz 2 res

Figure 6.37 Determining the resulting oscillating first-order inertial torque.

Taking into account the sign of the torque of the individual throws, we get the effective directions (Figure 6.38).

φ 1 = 0°

φ 2 = 216°

φ 3 not used

φ 4 = 144° (+180°) = 324°

φ 5 = 216° (+180°) − (360°) = 36°

Figure 6.38 Determining the resulting second-order inertial torque.

(product of the mass and distance from the rotary axis) of the rotating masses and the balancing mass(es) must correspond.

FBalance = Frot



mBalance ⋅ rBalance = mrot ⋅ r r mBalance = mrot ⋅ rBalance

F2 = λ ⋅ mosc ⋅ r ⋅ ω 2

∑M



∑M

X

Y

= a ⋅ F2 ⋅ ( 2 ⋅ sin 0° + sin 72° + sin 324° + 2 ⋅ sin 36°) = a ⋅ F2 ⋅ ( 2 ⋅ cos 0° + cos72° + cos 324° + 2 ⋅ cos 36° )



∑M



∑M



X

= a ⋅ F2 ⋅1.539

Y

= a ⋅ F2 ⋅ 4.736





Mosc 2max = a ⋅ F2 ⋅ 1.5392 + 4.7362 = arot ⋅ 4.98 1.539 tan d = = 0.325 4.736

d = 18°





6.1.5 Mass Balancing

Mass balancing is defined as the compensation of imbalances due to construction. The balancing of manufacturing-related imbalances is merely termed balancing. 6.1.5.1 Balancing Single-Cylinder Crank Gears The rotating inertial force can be balanced by countermass(es), where the condition must be fulfilled that the static torque

(6.72) (6.73)

By dividing the balancing mass into two counterweights, we obtain the following:





mBalance =

1 r ⋅ mrot ⋅ 2 rBalance

(6.74)

To keep the balancing mass small, it must be affixed at the greatest possible distance from the rotary axis (crankshaft axis); this is greatly limited by the constructive conditions. Basically, mass balancing should include a large static torque and a small moment of inertia. Oscillating inertial forces can also be compensated by revolving countermasses, since their force vector is composed of components in the direction of the cylinder axis (Y direction) and perpendicular to the cylinder axis (X direction). The balancing mass is selected so that the component in the direction of the cylinder axis corresponds to the oscillating inertial force; this is balanced, but at the price of a free component perpendicular to the cylinder axis (Figure 6.39).

FBalance = mBalance ⋅ r ⋅ w 2

(6.75)



X Balance = mBalance ⋅ r ⋅ w 2 ⋅ sin j

(6.76)

66 | Internal Combustion Engine Handbook

6606_Book.indb 66

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

YBalance = mBalance ⋅ r ⋅ w 2 ⋅ cosj



(6.77)

components perpendicular to the cylinder axis cancel each other out (Figure 6.40).

F1 osz = mosz · r · v 2 · cosf F1 osz = mosz · r · v 2 · cos f

f

x = mosz · r · v 2 · sinf

y = mosz · r · v 2 · cos f Fbalance

0,5 · mosz · r · v 2 · sinf

0,5 · mosz · r · v 2 · sin f

0,5 · mosz ·r·v2

0,5 · mosz ·r·v2

0,5 · mosz · r · v 2 · cosf

0,5 · mosz · r · v 2 · cosf

= mosz · r · v 2 Figure 6.40 Complete balancing of the inertial forces of the first order.

Figure 6.39 Balancing of oscillating forces using a revolving mass.

Better conditions result when the oscillating first-order inertial force is not completely balanced. Since the crankcase along its height (Y direction) is more rigid than in the transverse direction (X direction), the oscillating first-order inertial force is not completely compensated, so that the free X component does not become too large, and it is only 50% balanced. Completely balancing the rotating inertial force Frot and the 50% balance of the oscillating first-order inertial force is termed a normal balance—it was even used in the nineteenth century for drivetrains of steam locomotives. The mass balancing of designed passenger car engines is 50%–60% of the oscillating inertial force and 80%–100% of the rotating inertial force.

mBalance ⋅ rBalance = ( a1 ⋅ mrot + a 2 ⋅ mosc ) ⋅ r mNormal balance = ( 1⋅ mrot + 0.5 ⋅ mosc ) ⋅

r rBalance

To obtain a balance of the second order, the countermass must rotate at twice the crankshaft speed (Figure 6.41).

(6.78) (6.79)

Then, the two components in the direction of the cylinder axis compensate the oscillating inertial force. The oscillating first-order inertial force is completely balanced when two balancing masses revolving in the opposite direction that are half the oscillating crank gear masses are symmetrically arranged in relation to the vertical engine axis; the two

Figure 6.41 Diagram of mass balancing of the second order in a fourstroke crank gear.

6.1.5.2 Balancing Multicylinder Crank Gears Automobile engines are built with multiple cylinders, that is, with 3–12 (16) cylinders, as three-, four-, five-, and six-cylinder inline engines; V6, V8, and V12 (V16) engines; and VR5 and VR6 engines.

Internal Combustion Engine Handbook | 67

6606_Book.indb 67

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

These engines have three-, four-, five-, and six (eight)-stroke crankshafts so that with a corresponding arrangement, the mass effects of the individual throws cancel each other out (self-balance). For this purpose, the throws are to be evenly distributed in the peripheral direction and the lengthwise direction: •• With centrally symmetrical shafts (equal to the throw spacing across the perimeter), the free forces cancel each other out. •• Centrally and longitudinally symmetrical arrangements of the throws of a four-stroke engine shaft have no free forces and torques of the first order; starting with six strokes, the shafts are completely force-free and torque-free. The criteria for the throw sequence are as follows: •• No or very low free mass effects: A simple rule of thumb for throw sequences with favorable mass balances is presented in [6-8] and [6-9]. •• Additional torques may not arise from mass balancing, and no additional inertial forces may arise from torque balancing. •• Even angular ignition spacing.

The following holds true for crankshafts of fourstroke engines: •• Two-stroke shaft: For two-cylinder four-stroke inline engines, all the three criteria listed above can be met simultaneously only with the balancing mechanisms of a complex design, First-order inertial forces and second torques do not incur in shafts with throws offset by 180°; however, even angular ignition spacing can be achieved only in two-stroke cycles. Even angular ignition spacing in four-stroke cycles can be realized—even if intended for balanced load cycles—only by crankshafts without crank pin offset (throws offset by 0° or 360°), which will eliminate the first- and second-order torques at the same time. However, this shaft design results in first-order forces. An innovative balancing mechanism uses a balancing conrod with attached balancing rocker (Figure 6.43) [6-10]. The rocker system is arranged in the center of the crankshaft to avoid the generation of new torques. Depending on the design, the first-order forces can be fully balanced and the second-order forces are balanced at high portions.

Free first-order inertial torque can be balanced by a shaft rotating in the opposite direction at the crankshaft speed with two countermasses of a corresponding size and lengthwise spacing (torque differential). The arrangement in the engine can be freely selected. Gears or chains provide the drive; frequently, the oil pump drive is connected. To balance the torque of the second order, the differential drive rotates at twice the crankshaft speed (Figure 6.42).

Figure 6.43 Crankshaft drive of a BMW two-cylinder inline engine with balancing mechanism [6-10].

Oil pump drive

•• Three-stroke shaft: Free torque of the first and second orders occurs. The torque of the first order is compensated—especially in V-engines—with a torque differential. •• Four-stroke shaft: In four-cylinder four-stroke inline engines, the inertial forces of the second order are additive. Balancing

Figure 6.42 Torque differential of the Audi V6.

68 | Internal Combustion Engine Handbook

6606_Book.indb 68

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

Setting the tooth face play by means of disk thickness

Spacer plates

Gear drive of the mass balancing gearbox (BMW 318i)

these forces by two shafts with countermasses (differential) oppositely rotating at twice the crankshaft speed becomes increasingly important for engines at nominal speeds >4000 min–1 due to increased demands in terms of comfort. Because of the high peripheral speeds of the bearing pin of this differential—up to 14 m/s—the bearing and drive must be carefully designed. The balance shafts are driven by a gear on a crankshaft web, where the tooth face play of the drive must be harmonized to the shifts and rotational oscillations of the crankshaft (Figure 6.44). By offsetting the height of the balance shaft (Figure 6.45), an additional oscillating torque of the second order can be generated that can also balance with the gas force components of the oscillating torque.

Figure 6.44 Differential for the inertial forces of the second order.

For this reason, the effect of the height offset optimized for speed and load (Figure 6.46) by, for example, the variation of the height offset. •• Five-stroke shaft: Large free inertial torques arise, depending on the selected firing sequence, which are significant either in the first order (for example, ZF 15234) or in the second order (for example, ZF 12453)—or represent a compromise for both the orders. Passenger car and truck engines are built both with and without separate torque balance. •• Six-stroke shaft: Centrally symmetrical and longitudinally symmetrical shafts (starting at six shafts) are balanced by themselves; they do not have any free mass effects. The most important considerations in designing the mass balancing system are as follows: •• complexity of design (differential) •• operating behavior at high speed (second order): bearing, lubrication, and so on •• decreasing the load on the crank gear bearing •• balance of the gas force •• rotational oscillation behavior •• inertia •• friction behavior. The free forces and torques of the different cylinder configurations are summarized in tables in the relevant literature. Mass balancing is used not only on the crankshaft drive but also on the valve gear, that is, on camshafts: •• The body is drilled eccentrically, so that the manufactured imbalance can largely compensate for the free valve mass forces. •• Balancing masses are placed directly on the camshaft (Figure 6.47).

Figure 6.45 Mass balancing of the second order with height-offset balance shafts [6-11].

Internal Combustion Engine Handbook | 69

6606_Book.indb 69

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Free torque 2. Order Engine Long Axis

Force 2. Regulation on the motor bearing

Chapter 6 Crank Gears

Full load without AGW Full load with AGW without height offset Full Load with high offset AGW

Eradication Point

Thrust without AGW Thrust with AGW without height offset Thrust with height offset AGW

Eradication Point

Thrust without height offset Thrust with height offset

Full load without height offset Full load with height offset

Eradication Point

Eradication Point

Engine Speed [min–1]

Engine Speed [min–1]

Figure 6.46 System behavior of balance shafts with and without height offset [6-12]. See color section page 1065.

Internal torque

a

a

a

F

F

F

F

Figure 6.47 Camshaft with balancing mass.

6.1.6 Internal Torques

In addition to imbalanced inertial forces and inertial torques that are perceptible free inertial effects, internal torques also exist in engines. This includes the bending moments that arise in the (freely floating) crankshaft (Figure 6.48). These internal torques provide an additional load on the crankshaft main bearing and subject the crankcase to flexural stress. As the engine speed increases, the internal torques place higher demands on the construction of the engine, predominantly on V12 and V16 engines. The internal moment increases from the crankshaft ends to the middle of the engine. With longitudinally symmetrical shafts, the average bearing is

F·a

Bending moment characteristic in the crankshaft

Bending moment characteristic in the crankcase

Figure 6.48 Diagram of internal torque.

70 | Internal Combustion Engine Handbook

6606_Book.indb 70

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.1 Crankshaft Drive

subject to high loads from the inertial forces of the neighboring throws in the same direction, which can be prevented by internal mass balancing, that is, balancing the inertial forces at their origin, that is, at every throw (Figure 6.49).

In two-stroke engines with a power cycle length corresponding to a 360° crankshaft angle, the throw sequence corresponds to the firing sequence; four-stroke engines have two dead centers with a 720° crankshaft angle power cycle: •• firing TDC •• charge cycle TDC. Hence, for each throw sequence, there are several firing sequences because of the following: •• short angular ignition spacing (V-angle δ ) •• long angular ignition spacing (depending on the direction of rotation: V-angle δ + 360° or V-angle δ − 360°). The number of possible firing sequences for inline engines with k = throw number [6-13] is •• fully symmetrical shaft (four-stroke engines): ⎛k ⎞ ⎜ −1⎟⎠

2⎝ 2

•• partially symmetrical shafts (four-stroke engines with an uneven number of throws; two-stroke engines): Figure 6.49 Four-cylinder SI engine (Opel-Ecotec) balanced on all cheeks.

The advantages of complete internal mass balancing need to be weighed against the disadvantages of increasing the mass, moment of inertia, and cost.

6.1.7 Throw and Firing Sequences

To obtain a very even torque characteristic, the ignition of the individual cylinders must be evenly distributed over the power cycle. The requirement is that the throws must be evenly distributed over the perimeter. Hence, the following throw spacing results: Four-stroke engines 720° crankshaft angle/cylinder number Two-stroke engines 360° crankshaft angle/cylinder number The firing sequence is also determined by the direction of rotation of the crankshaft. For automotive engines, the direction of rotation is established in DIN 73021. •• Clockwise rotation: when viewed from the counterforce transmission side; the cylinders are counted from the counterforce transmission side. •• Counterclockwise rotation: counterclockwise looking at the counterforce transmission side; the cylinders are counted from the counterforce transmission side. The cylinders in V-engines are (viewed from the counterforce transmission side) counted from the right row starting from the left engine row that starts with z/2 + 1, counted from the right row. In V-engines, the same angular ignition spacing can only be achieved when the V-angle corresponds to the power cycle (720° or 360° crankshaft angle) divided by the cylinder number. Other factors for the firing sequence are •• no or very small free inertial effects •• favorable rotational oscillation behavior •• good supercharging conditions.

k!/2 ⋅ k. V-engines represent a good compromise between high power density and a compact basic design. The V-engine is, therefore, a preferred design in passenger car engines as well. A small V-angle requires a longer conrod (smaller conrod ratios l = r/l) and possibly a shifting of the crankshaft drive to provide the necessary free travel of the cylinder. This yields a higher crankcase with reduced piston-side forces, since the angular travel of the connecting rod is shorter. For vehicle engines, the 90° V-angle is preferred, since it allows the first-order inertial forces to be completely balanced with rotating counterweights; in addition, in eight-cylinder V-90° four-stroke engines, the V-angle corresponds to the even angular ignition spacing, the so-called “natural V-angle.” If the number of cylinders and the V-angle do not correspond, an even ignition spacing is still attained by “spreading” the crank pins by the difference between the V-angle and the angular ignition spacing (thrown crank pin, stroke offset, and split-pin crankshaft). Accordingly, six-cylinder passenger car and truck engines are being built today with the V-angles of 90° (such as Audi, Deutz, and DaimlerChrysler), 60° (Ford), and even 54° (Opel), and eight-cylinder engines with a 75° angle (DaimlerChrysler), which requires a total crank offset of 30°, 60°, 66°, and 15°. To select the V-angle, the clearance space of the engine and the harmonization of the engine program must be considered in addition to the crank gear mechanics. 6.1.7.1 Determining the Firing Sequences In two-stroke engines, the firing sequence corresponds to the throw sequence. In four-stroke engines, the two crankshaft rotations of a power cycle are reduced to one rotation. This yields a 0.5thorder phase diagram. The ignitions are evenly distributed over the perimeter and in the lengthwise direction. Viewed in terms of crank gear mechanics, V-engines are two inline engines offset from each other by the V-angle δ with half the

Internal Combustion Engine Handbook | 71

6606_Book.indb 71

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

number of cylinders. The ignition spacing of the cylinders that act together on a throw is as follows: •• δ ° (short angular ignition spacing) •• (δ + 360)° (long ignition spacing). The phase diagrams of the two partial inline engines are superimposed for the short angular ignition spacing by (δ /2)°, and for the long angular ignition spacing by ([δ + 360]/2)° from which the ignition intervals can be determined.

6.2 Rotational Oscillations 6.2.1 Fundamentals

The crank gear is a spring-mass system that is excited to vibrate (oscillating rotational movement of the sequential individual masses on the shaft) by the periodic torsional forces (tangential forces) that overlap the actual rotational movement of the crankshaft. The rotational movement of the crankshaft, therefore, comprises the following three components: •• even rotation corresponding to the speed •• speed fluctuation as a result of the uneven torsional force characteristic (tangential force characteristic) over a power cycle (static speed fluctuation) •• vibration over the displacement angle caused by the torsional force (dynamic speed fluctuation). The movement of the system is described by the angle of twist of the moments of inertia in comparison to the initial position. The kinetic energy stored in the moments of inertia is released to the coil springs and converted into potential energy in order to be reconverted back into kinetic energy. Given loss-free energy conversion, the free vibrations would last forever; the natural frequency depends exclusively on the system properties of spring rigidity and mass. Because of the resistance to the movement, energy is withdrawn from the

system and converted into heat: the vibration is suppressed and slows at a greater or lesser rate depending on the damping. If a periodic force acts on the system from the outside, then it forces the system to assume different vibration behavior; the system vibrates—after a transient phase—at the frequency of the exciting force. If the natural and exciting forces correspond, resonance occurs. Without damping, the vibration amplitude would assume an infinite value. However, the always-present damping limits the amplitude, and the size of the amplitude depends on the strength of the damping. This situation is illustrated by the magnification function V as a function of frequency ratio Ω/ω [The magnification function is the ratio of the (maximum) vibration amplitude of the system to the amplitude that would result if the spring of the system was under a static load from the exciting force.]. If the path of the vibration amplitudes of the individual masses is represented over the length of the shaft as a curve trace, we get the mode of vibrations with the zero transition points of this curve as vibration nodes, in which two neighboring masses vibrate in the opposite direction. No rotational oscillation movement occurs at these points (certainly rotational oscillation stress, however) (Figure 6.50). For each possible form of vibration, there is a natural frequency that the system can use to execute free vibrations in the relevant mode of vibration. The mode of vibrations and the natural frequencies depend on the size and distribution of the torsional rigidities and the moments of inertia in the system. Since the resonance can lead to vibration amplitudes that can destroy the crankshaft (Figure 6.51), it is important to identify such dangerous conditions beforehand and undertake appropriate measures to eliminate them. The properties of the crank gear are, therefore, calculated in this regard. Since it is a complex system, the crank gear must be conceptually simplified (reduced), so that it can be computed with a reasonable amount of effort. The basis of such a simplification (reduction) is the harmonization of the dynamic properties of the reduced system with those of the

Vibration line

cn – 1 c4

fn

f4 f3 f2 f1

c3

ln

c2

l5

c1

l4

Vibration nodes

l3 Vibration amplitude

l2 l1

Figure 6.50 Diagram of a rotational oscillation system.

72 | Internal Combustion Engine Handbook

6606_Book.indb 72

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.2 Rotational Oscillations

actual system. The calculation of the rotational oscillation consists of the following: •• reducing the machine system •• calculating the natural frequencies and modes of natural vibration

For passenger car engines, the BICERA formula is used. Since the shape of the crankshaft throw impairs its rotation, its reduced length is generally greater than the length of the throw (Figure 6.52). c = cred

•• calculating the exciting forces and the exciter works and amplitudes •• calculating the crankshaft excursions in the case of resonance •• calculating the crankshaft stress from the vibration excursions in the case of resonance •• calculating the critical speeds. IThrow

Ireduced

Figure 6.52 Length reduction of a crankshaft throw.

6.2.3 Natural Frequencies and Modes of Natural Vibration

The crank gear consists of coupled moments of inertia and torsional rigidities with mutually influential vibration behavior. Movement equations are created for the individual moments of inertia according to Figure 6.53. I k ⋅ j + c k−1 ⋅ ( j k − jk−1 ) + c k ⋅ ( j k − jk+1 ) = 0



(6.80)

I = Mass moment of inertia

φ = Angle of twist of the moment of inertia c = Torsional rigidity of the shaft piece k = Counter for the moments of inertia. Figure 6.51 Torsion break of a passenger car crankshaft made of GGG 70.

fk + 1

fk

6.2.2 Reduction of the Machine System

The crank gear with the coupled masses (flywheel, crank wheel mechanism, valve gear, belt drive, and so on) is reduced to a simple geometrical model, so that potential and kinetic energies of the actual and reduced systems correspond. •• Mass reduction: The crankshaft with the conrod, piston, and masses that it drives (crank wheel, flywheel, damper, and so on) is replaced by regular cylindrical disks with a constant moment of inertia. Although the moments of inertia of the crankshaft drives change from the piston and conrod movement, for the calculation, constant moments of inertia are assumed. •• Length reduction: The crankshaft throw is replaced with a straight, inertia-less shaft piece with the same diameter as the crankshaft main bearing (or the crank pin) whose length is such that the throw and shaft pieces have the same torsional rigidity (spring constant). There is a series of reduction formulas to accomplish this.

fk – 1

ck lk + 1

ck – 1 lk lk – 1

Figure 6.53 Opposite rotation of the moments of inertia of the reduced crank gears.

A system is obtained for homogeneously coupled linear differential equations with constant coefficients that describe the equilibrium between •• moments of acceleration from the moment of inertia arising from the inertial torque and the angular acceleration •• returning torque from the spring rigidity and difference between the angles of twist on both sides of the examined mass.

Internal Combustion Engine Handbook | 73

6606_Book.indb 73

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

The damping moment can be ignored when determining the natural frequencies, since the natural frequencies are only slightly influenced by the damping when it is weak. The integration of these equations yields the natural frequencies of the system. To solve these differential equations, a model in the form of harmonic movement is created. Systems with more than three moments of inertia make the equation systems overly complex and difficult to deal with; for this reason, different experimental procedures have been developed. Of these, the procedure by Gümbel–Holzer–Tolle has gained broad acceptance. It provides insight into the physical behavior of the vibration processes, and can be carried out using a simple and clear computational approach in which the results of one calculation step are used in the other as a pattern. The basic concept is as follows. An oscillating torque is imagined that acts on the end of a system capable of vibration, so that the system executes forced (undampened) vibrations; the amplitude of this oscillating torque (exciter moment amplitude) is set, so that the vibration excursion of the first mass assumes the value 1. If the exciter frequency is then changed, the exciter moment M (residual exciter moment) also changes, which is necessary to maintain the vibration excursion 1 of the first mass. If the exciter frequency corresponds to one of the natural frequencies of the system, the amplitude Mk of the necessary exciter moment M is zero.



u1 = 1

ve2

ve4 ve5

ve3

Figure 6.54 Residual exciter moment curve.

1.0 1. Natural frequency ve1

(6.81) 0.5

Mk+1 ck+1

(6.82)

M1 = 0

(6.83)

uk+1 = uk −

1. Natural frequency ve1

Exciter frequency V

M = Exciter moment uk = Relative excursion I = Mass moment of inertia c = Spring rigidity Ω = Exciter frequency k = Counter for the moments of inertia. When doing the calculation, the residual exciter moment is calculated for the different exciter frequencies that are necessary to maintain vibration excursion 1 of the first mass, and the residual exciter moment is plotted over the exciter frequency. The intersections of the residual exciter moment curve with the abscissa yield the desired natural frequencies (Figure 6.54). If the calculation is repeated with the natural frequencies found in this manner, we obtain the respective modes of natural oscillation (sum of the amplitudes of all the moments of inertia that define the deformational state of the oscillating system for each frequency). However, only the relative excursions, that is, the excursions of the individual moments of inertia in reference to the excursion of the first moment of inertia (Figure 6.55), are determined.

Relative crankshaft excursions ux



Mk+1 = −Mk + I k ⋅Ω2 ⋅uk

Residual exciter moment MRes

Chapter 6 Crank Gears

0 1.0

0.5

2. Natural frequency ve2

0 1.0

0.5 3. Natural frequency ve3 0

Reduced system

c1

l1

c2

l2

c3

l3

c4

l4

c5

l5

c6

l6

c7

l7

l8

Figure 6.55 Modes of natural oscillation for the three initial natural frequencies of a six-stroke crank gear with crank wheel and clutch.

74 | Internal Combustion Engine Handbook

6606_Book.indb 74

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.2 Rotational Oscillations

We are, therefore, dealing with a problem of intrinsic value whose solution is for only one common factor. To determine the absolute amplitudes, we need the exciting forces. Another solution corresponding to the Gümbel–Holzer–Tolle method is a matrix calculation, which increasingly succeeds thanks to computerized analysis. The relationships derived from the motion equations between the amplitudes of the rotational oscillation excursions and the return torques provide an equation system that can be represented with matrices and solved using a computer. I ⋅ j + D ⋅ j + c ⋅j = M(t) .



of work cycles per unit time) and their integral multiples. They are proportional to the crankshaft speed. All of these exciting frequencies can resonate with one of the natural frequencies (Figure 6.58).

40 20 0 –20

(6.84) 40

The vibration-exciting torsional force (tangential force) is composed of the following: •• gas torsional force (Figure 6.56)

Amplitude in kN

•• torsional force of the oscillating inertial forces (the rotating inertial forces do not participate in the excitation) (Figure 6.57).

Gas torsional force FTG in kN

60 40 20 0 –20 0

100

200

300

400

500

600

700

Crankshaft angle f [°]

Inertial torsional force FTm in kN

Figure 6.56 Gas torsional force characteristic of a four-stroke diesel engine.

60 40

–40 40 20 0 –20

2. primary harmonic k = 2

–40 40 20 0 –20

3. primary harmonic k = 3

–40 40 20 0 –20

4. primary harmonic k = 4

–40 20 0 –20 20 0 –20

5. primary harmonic k = 5

6. primary harmonic k = 6

0

90

180

270

380

450

540

630

720

Figure 6.58 Fourier analysis of a tangential force diagram: the tangential force curve is composed of the first six harmonics.

20 0 –20

1. primary harmonic k = 1

20 0 –20

6.2.4 Exciter Forces, Work, and Amplitudes

0

100

200

300

400

500

600

700

Crankshaft angle f [°]

Figure 6.57 Inertial torsional force characteristic of a four-stroke diesel engine.

Since the gas torsional force is a function of the load (specific work) and the inertial torsional force is a function of the square of the rpm, their influence is investigated separately. The gas torsional force cannot be described by a closed function, and is, therefore, subject to a Fourier analysis; this is composed of a static component (nominal load torque) and a dynamic component (a basic vibration and overlapping harmonics). The exciting frequencies are, hence, the basic frequency (number

The exciter work is the essential determinant in exciting vibration. An exciter force (resulting exciter force amplitude from the amplitudes of the gas and inertial torsional forces for the individual exciter frequencies) generates a greater excursion the farther it acts from the oscillating nodes (exciter work = exciter force × vibration amplitude). The phase angle of the exciting forces, that is, their sequence over time, is represented in phase direction diagrams. The phase direction diagrams of the individual orders result from the order throw diagram of the 0.5th order (four stroke) and of the first order (two stroke) (Figure 6.59). Taking into consideration the vibration amplitude of the individual throws and the phase shift (firing sequence), we get the effective exciter force of the engine.

Internal Combustion Engine Handbook | 75

6606_Book.indb 75

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

k = 0.5

k=1

k = 1.5

1

1,6

1,2,3

4

2

t=

∆j = ( ux − ux+1 ) ⋅ A1

(6.87)

Md cx ⋅ A1 ⋅ ( ux − ux+1 ) = Wp Wp

(6.88)

5



3

In particular, the gas forces excite vibrations of an order that are an integral multiple of the number i of ignitions within a crankshaft rotation.

2,5

3,4

6 1,6

4,5,6 1,2,3,4,5,6

1

3,4

2,5



5

4

3

2

•• Four-stroke engine: i = z/2 ignitions per crankshaft rotation •• Two-stroke engine: i = z ignitions per crankshaft rotation.

6

k=2

k = 2.5

k=3

Figure 6.59  Phase direction diagrams up to the sixth order for an inline six-cylinder four-stroke-crank gear.

The relative crankshaft excursions of the individual cylinders are added geometrically in the direction of the rays of the phase direction diagrams. This shows us that certain orders are particularly dangerous, because their geometric sum becomes very large. The geometric sum is described as the specific exciter work, that is, the exciter work of the engine in reference to force 1. Depending on the order and phase angle, the specific exciter work assumes different values. The amplitude—the absolute excursion—of mass 1 is calculated from the equilibrium of the excitation work and damping work (per vibration). This allows us to determine the absolute excursions A of the individual masses of the substitute system:

A1 =

z  FTk ⋅ ∑ ux z

1

we ⋅ ∑ bx ⋅ ( ux )

2



(6.85)

1

Ax = ux ⋅ A1

(6.86)

FTk = Resulting exciter force amplitude from the amplitudes of the gas and inertial torsional forces (assumed to be the same for all cylinders) ux = Relative crankshaft excursions

ω e = Natural frequency β x = Damping coefficient of the xth cylinder; usually the same damping coefficient is assumed for all cylinders A1 = A mplitude (absolute excursion) of the first mass of the system  ux = Geometric sum of the relative crankshaft excursions Index x = Number of cylinders Index k = Order. The relative twist ∆φ of the masses x and x + 1 from the rotational oscillation stresses the crankshaft in addition to static torsional force.

All integral multiples of z/2 (four-stroke engine) or z (twostroke engine) are dangerous, since the exciters of all the cylinders are aligned for these orders. The critical speeds result from the intersections of the main harmonics with the exciter frequencies. The extent of the danger to the engine at the individual critical speeds can be found by calculating the resonance excursions of the crankshaft.

6.2.5 Measures to Reduce Crankshaft Excursions

Without damping, the excursions of the crankshaft would become increasingly large until the shaft breaks. In practice, however, damping always exists—material damping, friction damping, and damping from the lubrication film. However, these are usually insufficient in today’s highly stressed crank gears, so that additional measures must be taken. To avoid hazardous rotational oscillation states, one can •• influence the exciter work by varying the firing sequence and/or •• shift the natural vibration frequency by changing the mass and spring rigidity. The feasibility and effectiveness of these measures is limited, however. An apparently simple measure is to increase the moment of inertia of the flywheel. This lowers the natural frequency, but at the same time, the oscillating nodes are displaced toward the flywheel, and the shaft load is increased. For these reasons, the only possibility is to reduce the rotational oscillations to a safe level. There are basically two options for this. •• Damping: Converts the vibration energy into heat. In the case of stationary forced vibrations and speed-proportional damping, there is an equilibrium among the moments of mass inertia, damping, return force, and excitation. The greater the damping moment, the smaller the vibration amplitude. •• Absorption: That is, “extinguishing” resonances by detuning the system or, more precisely, shifting the natural frequencies into other speed ranges by counteracting with a mass: by coupling an additional mass, the absorber, the system is given one more degree of freedom. The original natural frequency splits into two natural frequencies that lie closely above and below the original. If the system is excited in the original natural frequency, then it remains unexcited while the absorber vibrates. Such absorbers, however, are effective only for a single frequency. A pendulum attuned to a specific vibration frequency and articulated to the oscillating system enters a reverse phase when this vibration arises and, hence,

76 | Internal Combustion Engine Handbook

6606_Book.indb 76

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.2 Rotational Oscillations

counteracts the exciting moment. The resonance speed is split and shifted upward or downward. Centrifugal force absorbers are speed-dependent.

Rotation angle

Without oscillation damper Sum of 3rd to 6th order

2000

4000

6000 Engine speed [min –1]

Rotation angle

The effect of vibration dampers in passenger vehicle engines is based on both damping and absorbing. With regards to spring rigidity, damping behavior, and mass inertia, they are designed to continuously reduce the rotational oscillation excursions of the system. For passenger car engines, rubber vibration dampers are used: an annular damper mass (secondary part) connected to the primary-side L-shaped driving disk is elastically coupled through a vulcanized rubber layer. The vibration energy is converted by the material damping (hysteresis) of the rubber into heat. The resonance peak is divided into two resonances whose peaks are reduced by the damping. Depending on the design, the damper mass is affixed radially and/or axially to the primary part. Two-stage dampers are also used in which two damper masses are tuned to two different frequencies [6-14] (Figure 6.60). An example of this is seen with the two-mass rubber vibration damper for a five-cylinder diesel engine (2.5L) in which both masses are harmonized to the torsion.

Influence of the rotational oscillation damper on the rotational oscillations of the crankshaft Example of a six-cylinder boxer engine

With oscillation damper Sum of 3rd to 6th order

2000

4000

6000 Engine speed [min –1]

Figure 6.61 Effect of a rotational oscillation damper.

Figure 6.60 Two mass rubber vibration damper (by Palsis) with vulcanized strips of rubber, V-belt strip on the primary side, primary side with shaft sealing flange made of St24W, secondary side made of EN-GJS-400-15, primary mass moment of inertia Θ = 0.008 kg m2, secondary side 0.012 kg m2/220 Hz and 0.006 kg m2/360 Hz, and rubber AEM (Vamac) (source: Palsis).

By reducing the rotational oscillation amplitude (Figure 6.61), not only are the crankshaft and camshaft mechanically relieved, the play-induced noise of the engine and the excitation of the accessories to vibrate are reduced [6-15]. Passenger car engines increasingly require vibration dampers to deal with large engine dimensions (stroke volume) and greater specific work (effective average pressure) because of the stronger excitation. These are also used to lower natural frequencies as a result of greater crank gear masses.

6.2.6 Two-Mass Flywheels

The drivetrain of a vehicle consists of an engine, a transmission, and the vehicle itself. The vibrations excited by the engine are also transmitted to the other components of the drivetrain. Engine-induced vibrations of the transmission are manifested as •• Bucking: The engine excites the system with 0.5th-order vibrations, and it vibrates against the vehicle.

Figure 6.62 Viscose vibration damper with a decoupled belt pulley (torsionally elastic rubber coupling) for inline six-cylinder diesel engines (Palsis). The natural frequencies of passenger car crank gears range from 300 to 700 Hz for which viscous dampers are increasingly used, like the ones that have been used for larger engines.

•• Chatter: The engine excites the transmission primarily with fourth-to-sixth-order vibrations, so that gears and synchronizer rings that do not lie in the flow of force vibrate against each other at comparatively large amplitudes. In addition, the drivetrain is twisted during load changes and swings, which is only slightly dampened. These vibrations are noticeable; they impair driving comfort, and additionally put stress on the components. To improve the vibration and noise behavior of the drive, two-mass flywheels are used:

Internal Combustion Engine Handbook | 77

6606_Book.indb 77

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

Two-mass fly wheel (mechanical torsion damper)

Starter ring gear Primary Grease plate Cover plate

Key

Splash plate Cover plate Secondary flywheel

Pressure spring

Secondary flywheel

Reinforcement ring Rivet Plain bearing

Eccentric plate Shear key

Figure 6.63 Two-mass flywheel (GAT).

the mass of the engine flywheel is divided into a primary part rigidly fixed to the crankshaft and a secondary part articulated to the primary part. The primary and secondary parts are connected by torsionally elastic springs. This isolates the vibration; that is, the operating range is shifted to the supercritical range of the enlargement function. Since different rigidities and damping properties are required to suppress the transmission chatter in the different operating ranges (traction, thrust, and idling), the characteristics of the springs must be correspondingly engineered. This is accomplished, for example, by a series of springs with different rigidities. With correspondingly adjusted feather key systems, friction provides the desired damping [6-16] (Figure 6.63). The rotational oscillation behavior of the engine drivetrain changes because of the lower moment of inertia of the primary part of the flywheel (Figure 6.64). With two-mass flywheels, not only is the driving comfort improved, but the transmission is freed from additional oscillating torque. They are primarily used in passenger car engines with piston displacement ≥2l, particularly in diesel engines [6-16]. Three-mass flywheels are now also being used.

6.3.1 Variable Swept Volume

Swept volume is the product from the stroke and the surface determined by the bore diameter of the cylinder. It is the central value determining the torque and, in conjunction with the speed, the output of an engine. In diesel or SI engines for passenger cars, the swept volume is mostly between one and three liters, depending on the number of cylinders. In most cases, the swept volume per cylinder is somewhere between 350 and 600 cm3. Unlike a diesel engine, the drawn air or mixture mass must be reduced by throttling (quantity control) in a conventional SI engine under partial load. This throttling process generates losses, so that the filling does not match the one that the piston displacement would theoretically support. Of the several methods used to reduce throttling losses in engine operation, the two most important are as follows: System with two-mass flywheel

Conventional system

AAmplitude of angle acceleration inm s–2

6.3 Variability of Compression and Swept Volume

2 · 104

Engine Engine

1 · 104

0

Transmission

1000

2000

3000

Transmission 1000

Engine speed in min –1

2000

3000

Figure 6.64 Effect of a two-mass flywheel.

78 | Internal Combustion Engine Handbook

6606_Book.indb 78

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.3 Variability of Compression and Swept Volume

•• cylinder shutdown. Both the methods enable the reduction in load cycle losses, because throttling is decisively reduced for a specified load point at reduced swept volume, regardless of the method applied. When the displacement is decreased, the motor must be operated at a higher load range to obtain the same specific work, thus reducing the manifold vacuum pressure and throttle losses. The effect of reducing the gas exchange cycle results in less fuel consumption [6-22], [6-23], [6-27]. Because functional reasons prevent the realization of a bore variability, engineering efforts focus on piston stroke variability. At a given bore and stroke reduction, this also changes the stroke-bore ratio that becomes less. This then results in a change in the surface-volume ratio of the combustion chamber with the familiar influence on HC emissions [6-24], [6-25], [6-26]. The efficiency and the NO emissions are also affected [6-25]. Technical solutions for a constant variation of the piston stroke are well known, and represent an optimal solution in view of a reduction in the gas exchange cycle, because, in extreme cases, the throttle valve can be fully dispensed with. The design considerations in conjunction with variable swept volumes always start with a modification of the kinematics of the crankshaft drive. For example, by a lateral shifting of the crankshaft, the stroke is reduced, as is the swept volume. Although experiments with variable swept volumes have been made, the technical solutions derived, however, proved to be very expensive.

6.3.2 Variable Compression

The compression process is one of the four work cycles of a combustion engine. It ensures that temperature and pressure are increased in the work medium and the combustion occurs with higher efficiency. The dependency of the thermal efficiency from the compression ratio is demonstrated in a model process (Figure 6.65). High thermal efficiency of the model process suggests a high degree of efficiency of the engine process with the subsequent minimization of fuel consumption. Figure 6.65, however, shows that the thermal efficiency increases at a continuously decreasing rate, the higher the compression ratio becomes. The consequences thereof, in respect to the realization of a variable compression in the engine and the constructive efforts involved, suggests to increase the compression ratio only to a certain extent. In the engines shown, a compression ratio between 8:1 and 16:1 has been realized [25]. Figure 6.66 shows the possible ranges of the compression ratios for common engines. In an SI engine, increased throttling causes the effective compression ratio to drop at constant geometrical compression. The result is a lower degree of efficiency. This fact becomes even more obvious when a charged SI engine is examined. In consideration of the increased knocking sensitivity when

close to full load, the geometrical compression in a charged SI engine must be reduced in comparison to a naturally aspirated engine, causing a further decrease in efficiency at partial load. Figure 6.67 demonstrates the effective compression ratio in an SI engine in the map.

Figure 6.65 Compression ratio and thermal efficiency. Engine Type

ε

Limited by

SI engine (two stroke)

 7.5–10

Autoignition

SI engine (two valve)

 8–10

Knock and autoignition

SI engine (four valve)

 9–11

Knock and autoignition

SI engine (direct injection)

11–14

Knock and autoignition

Diesel (indirect injection)

18–24

Loss in efficiency and component stress

Diesel (direct injection)

17–21

Loss in efficiency and component stress

Figure 6.66 Compression ratios.

SI-engine e = 12,5 (geometrical)

1,2 [kJ/dm3] 1,0

12,5 12

0,8

Specific work

•• variable swept volume

11 0,6

10 9

0,4

8 7

0,2 0,0 1000

5

6 3000 5000 Engine Speed

[rpm]

7000

Figure 6.67 Effective compression ratio in an SI engine.

Internal Combustion Engine Handbook | 79

6606_Book.indb 79

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

The variable compression increases the efficiency of an SI engine, because its compression ratio is limited due to the gasoline’s tendency to knock under full load. If the compression ratio is increased at partial load, the internal effectiveness improves considerably. In the area relevant for the CSV bag analysis, a 10% savings in consumption can be achieved compared with engines with fixed compression. The improvements are even more significant in charged engines with variable compression, because an additional gain is realized by shifting the operating point in this case. In a given case, the compression of a charged engine was increased to a compression ratio of e = 13.5 for a charged engine under partial load, while the compression ratio at full load was e = 8. The consumption gain in this case was >20% during CVS testing at the same driving performance. Supercharging with up to 100-KW/L displacement can yield up a 30% reduction in fuel consumption in the New European Driving Cycle [25]. Systems with variable compression have not yet prevailed in series vehicles, due to the effort and high costs involved. Amongst others, the following systems have been examined as examples:

upward or downward. The rotational movement of a pivotable crankshaft axis is transmitted to the fixed axis of the gearbox input. This technically very expensive solution increases the engine mass only marginally [6-17], [6-18], [6-19], [6-20], [6-27]. •• An inclination of the cylinder head, designed so that the separation plane between the head and the block is shifted “down,” that is, the block height is reduced compared with a conventional engine [6-17], [6-21]. This is a very expensive solution as well. Figure 6.68 and Figure 6.69 demonstrate the pivot mechanism by which the cylinder head can be pivoted by up to 4°, enabling a change in the compression ratio from 8:1 to 14:1.

•• Piston with variable compression height. The disadvantage here is the large mass of the piston causing high mass forces at high speeds. •• Enlargement or reduction of the combustion chamber by shifting, for example, a cylinder in the cylinder head. This concept results in deteriorating combustion conditions due to a fissured combustion chamber. Displacement of the crankshaft axis by a parallel crank gear, for example. An eccentric unit would reposition the crankshaft

14:1

Figure 6.68 Mechanisms for turning the cylinder head (source: MOT). See color section page 1066.

8:1

Figure 6.69 Longitudinal section through the Saab Variable Compression engine (source: MOT). See color section page 1066.

80 | Internal Combustion Engine Handbook

6606_Book.indb 80

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6.3 Variability of Compression and Swept Volume

Another option to design the compression ratio to be variable is designing the connecting rod with an eccentric wristpin bearing [6-28], [6-29]. This eccentric bearing in the small conrod eye (Figure 6.70) enables a length-variable connecting rod that uses the crank gear forces for the shifting action. We have a number of options with which to realize a variable compression. Figure 6.71 demonstrates some other theoretical possibilities. Which procedure will ultimately be implemented in series production cannot yet be finally evaluated at this point in time. Some essential issues certainly are important for series production, in addition to a reliable functioning: •• package capability (space) •• production costs •• transferability to other engine models •• engine mass.

Figure 6.70  Variable compression ratio conrod (source: MTZ/Pischinger). See color section page 1066.

Figure 6.72 shows an assessment of the advantages and disadvantages of individual concepts for compression variation.

Vertical shift of cylinder block

Vertical momentary combustion chamber due to secondary piston

Piston with adjustable compression height

Conrod bearing in eccentric crank pin

Eccentric crankshaft bearing (VCR principle)

Power transmission with gear-wheel drive

Second movable articulation point of the conrod (1)

Second movable articulation Second movable articulation point of the conrod (2) point of the conrod (2)

Figure 6.71 Schematic of a variable compression (source: MOT). See color section page 1067.

Internal Combustion Engine Handbook | 81

6606_Book.indb 81

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 6 Crank Gears

Figure 6.72 Comparison of various systems for the variable setting of compression ratios (source: MTZ/Pischinger).

Bibliography

6-1. Bauder, A., Krause, W., Mann, M., Pischke, R., and Pölzl, H.-W. 1999. “Die Neuen V8-Ottomotoren von Audi mit Fünfventiltechnik.” MTZ 60, 1: p. 16. 6-2. Dorsch, H., Körkemeier, H., Peiters, S., Rutschmann, S., and Zwickwolf, P. 1989. “Der 3,6-Liter-Doppelzündungsmotor des Porsche Carrera 4.” MTZ 50, 2. 6-3. Biezeno, C. B. and Grammel, R. 1953. Technische Dynamik, 2. Nachdruck. Springer: Berlin.

6-16. Nissen, P.-J., Heidingsfeld, D., and Kranz, A. 2000. “Der MTD—Neues Dämpfungssystem für Kfz-Antriebsstränge.” MTZ 61, 6. 6-17. Blumenstock, K. U. 2004. “Ungenutzte Potenziale.” MOT, 14. 6-18. Schwaderlapp, M., Pischinger, S., Yapici, K. I., Habermann, K., and Bolling, C. 2001. “Variable Verdichtung—eine konstruktive Lösung für Downsizing-Konzepte.” Aachener Kolloquium Fahrzeug- und Motorentechnik. 6-19. Guzella, L. and Martin, R. 1998. “Das SAVE-Motorkonzept.” MTZ 10.

6-4. NN. Kolben für Pkw- und Nkw-Motoren. Kapitel 1: Grundlagen. Kolbenschmidt: Hsg.

6-20. Fraidl, K. G., Kapus, P., Piock, W., and Wirth, M. 1999. “Fahrzeugklassenspezifische Ottomotorenkonzepte.” MTZ 10.

6-5. Riedl, C. 1937. Konstruktion und Berechnung moderner Automobil- und Kraftradmotoren, 3. Aufl. R.C. Schmidt: Berlin. pp. 224–231.

6-21. Bergsten, L. 2001. “Saab Variable Compression SVC.” MTZ 62, 6.

6-6. Krüger, H. 1990. “Sechszylindermotoren mit kleinem V-Winkel.” MTZ 51, 10. 6-7. Krüger, H. 1993. “Der Massenausgleich des VR6-Motors.” MTZ 54, 2. 6-8. Kraemer, O. 1476. “Kurbelfolge günstigsten Massenausgleichs 1. Ordnung.” ZVDI 81, 51: p. 1476. 6-9. Kraemer, O. 1963. Bau und Berechnung der Verbrennungsmotoren, 4. Aufl. Springer: Berlin. p. 74. 6-10. Gumpesberger, M., Landerl, C., Miritsch, J., Mosmüller, E., Müller, P., and Ohrnberger, G. 2006. “Der Antrieb der neuen BMW F800.” MTZ 67, 6. 6-11. Neukirchner, H., Arnold, O., Dittmar, A., and Kiesel, A. 2003. “Die Entwicklung von Massenausgleichseinrichtungen für PKW-Motoren.” MTZ 64, 5. 6-12. Gruber, G., Prandstötter, M., and Hollnbuchner, R. 2008. “Integriertes Ausgleichswellensystem des neuen Vierzylinder-Dieselmotors von BMW.” MTZ 69, 6. 6-13. Lang, O. R. 1966. Triebwerke schnelllaufender Verbrennungsmotoren. Springer: Berlin. p. 66. 6-14. Anisits, F., Borgmann, K., Kratochwill, H., and Steinparzer, F. 1998. “Der neue BMW Sechszylinder Dieselmotor.” MTZ 59, 11. 6-15. Pilgrim, R. and Gregotsch, K. 1989. “Schwingungstechnischakustische Entwicklung am Sechszylinder-Triebwerk des Porsche Carrera 4.” MTZ 50, 3.

6-22. Schäfer, F. and Basshuysen, v. R. 1993. Schadstoffreduzierung und Kraftstoffverbrauch von Pkw-Verbrennungsmotoren, Die Verbrennungskraftmaschine, Band 7. Springer Verlag: Wien, New York. 6-23. Basshuysen, R. and Schäfer, F. (Hrsg.). 2006. Lexikon Motorentechnik. Vieweg Verlag. 6-24. Kreuter, P., Gand, B., and Bick, W. 1989. Beeinflussbarkeit des Teillastverhaltens von Ottomotoren durch das Verdichtungs-verhältnis bei unterschiedlichen Hub-Bohrungs-Verhältnissen, 2. Aachener Kolloquium Fahrzeug- und Motorentechnik. 6-25. Gand, B. 1986. Einfluss des Hub-Bohrungs-Verhältnisses auf den Prozessverlauf des Ottomotors, Dissertation. RWTH Aachen. 6-26. Bick, W. 1990. Einflüsse geometrischer Grunddaten auf den Arbeitsprozess des Ottomotors bei verschiedenen Hub-BohrungsVerhältnissen, Dissertation. RWTH Aachen. 6-27. Pischinger, F. 1990. Gedanken über den Automobilmotor von morgen, Vortrag VW-AG. 6-28. Pischiner, S., Wittek, K., and Tiemann, C. 2009. “Zweistufiges Verdichtungsverhältnis durch exzentrische Kolbenbolzenlagerung.” MTZ Jahrgang 70. 6-29. Wittek, K. 2006. Variables Verdichtungsverhältnis beim Verbrennungsmotor durch Ausnutzung der im Triebwerk wirksamen Kräfte, Dissertation. RWTH Aachen.

82 | Internal Combustion Engine Handbook

6606_Book.indb 82

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7 Engine Components 7.1 Pistons/Wristpins/ Wristpin Circlips 7.1.1 Piston 7.1.1.1 Requirements and Functions The functions carried out by the piston include accepting the pressures created by the ignition of the fuel and air mixture, transferring these forces via the wristpin and the connecting rod to the crankshaft, and, in addition, providing guidance for the small conrod eye. As a moving wall that, working in conjunction with the piston rings, transfers power, the piston has to reliably seal the combustion chamber against gas escaping and lubricant oil flowing by in all operating situations. Increases in engine performance have caused parallel increases in the demands on the pistons. One example for piston loading: when a gasoline engine is running at 6000 rpm, every piston (D = 90 mm) at a peak cylinder pressure of 75 bar, fifty times a second, is subjected to a load of approximately 5 tons. Satisfying the various functions—such as adaptability to various operating situations, security against the pistons seizing while at the same time achieving smooth running, low weight at sufficient strength, low oil consumption, and low pollutant emissions—results in requirements for engineering and materials that in some cases are contradictory. These criteria have to be weighed carefully against each other for each type of engine. Consequently, the solution that is ideal in any particular instance may be quite different. Compiled in Figure 7.1 are the operating situations for the pistons, the resultant requirements for their design, and the requirements in terms of engineering and materials. 7.1.1.2 Engineering Designs We find that, given the operational requirements of the various internal combustion engine designs (two-cycle, four-cycle,

gasoline, and diesel engines), the aluminum silicon alloys are, as a rule, the most suitable piston materials. Steel pistons are used in special cases, but they then require special cooling measures. In the interest of weight reduction, a carefully worked out engineering design for the pistons is necessary, combined with the requirement for good piston cooling. Important terms and dimensions used to describe the geometry are shown in Figure 7.2 and Figure 7.3. The increase in the engines’ specific performance is affected in part by increasing engine speed. The strong rise in the mass inertias that results in the reciprocating engine components is largely compensated for by reducing the compression height (CH) and optimizing the weight in the piston engineering design. Particularly in smaller, high-speed engines, the total length of the piston (GL), referenced to piston diameter, is shorter than in larger engines running at medium speeds. The compression height CH influences the overall engine height and most decisively the weight of the piston. The engineer thus strives to keep this dimension as small as possible. Consequently, the compression height is always a compromise between demands for a short piston and for high operational reliability. The values given in Figure 7.3 for the head thickness s apply generally for pistons with a flat and level head, as well as for those with a convex or concave crown. In the case of pistons for diesel engines with direct injection, with deep recesses, the head thicknesses, depending on the maximum cylinder pressure, lie between 0.16 and 0.23 times the maximum recess diameter (DMu). We learn from the guideline values in Figure 7.3, with regard to the wristpin diameter BO, that the higher working pressures in diesel engines require larger wristpin diameters. The piston ring zone, with the piston rings themselves, represents moving seals between the combustion chamber and the crankcase. The length of this zone depends on the number and thickness of the piston rings used and the lengths

Internal Combustion Engine Handbook | 83

6606_Book.indb 83

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Operating Conditions

Requirements for the Piston

Engineering Solution

Materials Solution

Mechanical loading a. Piston head/combustion recess  Gasoline engines:  Ignition pressures 50–120 bar  Diesel engines:  Ignition pressures 80–230 bar b. Piston skirt: Lateral force: approximately 6%–8% of maximum ignition pressure c. Piston bosses: Permissible surface pressure, temperature dependent

High static and dynamic strength at high temperatures. High surface pressure in the bores in the bosses. Little plastic deformation.

Sufficient wall strength, stable engineering design, uniform “power flow” and “heat flow.” Boss bushing, Ferrotherm piston heads made from steel or one-piece steel piston

Various Al–Si casting alloys, with heat exposure (T5) or hardening by precipitation (T6), cast or forged special brass, bronze, tempered steel

High temperature in combustion chamber: Mean gas temperature approximately 1000°C At piston head/edge of recess: 200–400°C For ferrous materials: approximately 350–500°C  At the wristpin boss: 150–260°C  At the piston skirt: 120–180°C

Strength must be maintained even at higher temperatures. Indicator values: Hot hardness, permanent strength, high thermal conductivity, resistance to scale (steel)

Sufficient thermal convection cross sections, cooling channels

As above

Acceleration of piston and conrod at higher speeds: In some cases, far above 25,000 m/s2

Low weight, resulting in small inertial forces and moments of inertia

Lightweight construction with maximum utilization of material capabilities

Al–Si alloy, forged

Sliding friction in the ring grooves, Low friction resistance, high wear at the skirt, in the wristpin bearings. resistance (influences service life), Unfavorable lubrication situation in low tendency to seize some cases

Sliding surfaces of sufficient size, uniform pressure distribution. Hydrodynamic piston shapes in the skirt area. Armored grooves, oil supply

Al–Si alloys,skirt tinned, graphited,coated;groove reinforcement by ring carriers cast in place

Change of contact from one side of the cylinder to the other (above all at top dead center)

Low play when running,elastic skirt design with an optimized piston shape, offset bores in the bosses

Low coefficient of thermal expansion. Eutectic or supereutectic Al–Si alloys

Low noise, low “piston slapping” with engine cold and warm, little susceptibility to cavitation, low impact pulses

Figure 7.1 Operating conditions and the resulting demands on the piston as well as for solutions based on the engineering design and materials selection.

Gasoline Engines

Diameter D (mm)

Diesel Engines (Four-Cycle)

Two-cycle

Four-cycle

Passenger car diesel

30–70

65–105

65–95

Overall length GL/D

 0.8–1.0

 0.6–0.7

 0.8–0.95

Compression height CH/D

 0.4–0.55

 0.30–0.45

 0.5–0.6

Wristpin diameter BO/D

 0.20–0.25

 0.20–0.26

 0.32–0.40

Fire landF (mm)

 2.5–3.5

 2–8

 4–15

D Dmax F

s

ST

DL

KH

GL

BO SL UL

1. First ring land St/D*  0.045–0.06

 0.040–0.055

 0.05–0.09

Groove height for first ring (mm)

 1.0–1.75

 1.75–3.0

 1.2 and 1.5

Skirt length SL/D

 0.55–0.7

 0.4–0.5

 0.5–0.65

Boss clearance AA/D

 0.25–0.35

 0.20–0.35

 0.20–0.35

Head thickness s/D or s/DMu**

 0.055–0.07

 0.06–0.10

 0.15–0.22

*Values for diesel engines are applicable to pistons with ring carriers, depending on peak combustion pressure. **For direct-injection models ~0.2 × combustion recess diameter (DMu).

AA

F s ST KH DL GL

Fire land Bottom thickness First ring land Compression height Strain length Overall length

BO SL UL AA D Dmax

Boss bore-∆ (bolt-∆) Skirt length Lower length Boss clearance Piston diameter max. recess diameter

Figure 7.3 Important terms and dimensions at the piston.

Figure 7.2 Major dimensions for lightweight metal pistons and passenger cars.

84 | Internal Combustion Engine Handbook

6606_Book.indb 84

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.1 Pistons/Wristpins/Wristpin Circlips

Control Piston

Without Regulating Strips Modern Lightweight Pistons

Piston Designs

Hydrothermik

Hydrothermatik

Al Piston

Operating principle

Gasoline

Gasoline and diesel

Gasoline (two-cycle)

Diesel

Gasoline (four-cycle)

Installation play (nominal dimension range)

0.3–0.5

0.6–1.3

0.7–1.3

0.3–0.5

Upper end of skirt

0.6–1.2

1.4–4.0*

1.8–2.4

1.7–2.2

1.8–2.2

*Only for single-ring designs and maximum performance engines (end of skirt near the fire land).

Figure 7.4 Normal installation play dimensions for light-alloy pistons in vehicular engines (as ‰ of nominal diameter; installation in gray cast engine block).

of the lands between the rings. The compression ring set, with just a few exceptions, comprises two compression rings and an oil control ring. The three-ring piston is the standard design today. The length of the first ring land is selected in accordance with the ignition pressure occurring in the engine and the temperature of the land. The lengths of the lands located below are shorter, which is because of the falling temperature and loading because of gas pressure. The piston skirt is used to guide the piston within the cylinder. It transfers to the cylinder wall, in sliding fashion, the lateral forces occurring because of the deflection of the conrod. With sufficient skirt length and close guidance, the so-called “piston slapping,” occurring at the moment when contact shifts from one side of the piston to the opposite side (secondary piston motion), is kept to a minimum. This is important for smooth engine running and to reduce wear at all the piston’s sliding surfaces. The piston bosses must transfer all longitudinal forces from the piston to the wristpin and must therefore be well supported against the head and the skirt. Sufficient distance between the upper face of the boss bore and the inside of the piston head favors a more uniform distribution of stresses at the cross section for the support area. At high loads, particularly careful design of the support area is thus required. To avoid fissures forming at the bosses, the mean calculated surface pressure in the boss bore (dependent on the boss and wristpin configuration and particularly dependent on the boss temperature) should not exceed values of between 55 and 75 N/mm2. Attaining higher values is possible only by adopting special measures to increase the strength at the boss bore. The distance between the two bosses AA depends on the width of the small-end eye. This value has to be optimized in the interest of lower deformation values for the piston and wristpin. Only with the smallest possible boss clearances, ideal support can be achieved and the oscillating masses kept small. 7.1.1.3 Offsetting the Boss Bore Offsetting the axis of the wristpin in relation to the piston’s longitudinal axis optimizes the contact properties for the piston at the change of sides. The impact pulses can be influenced decisively with this measure. The location and amount of offset to the piston’s longitudinal axis can be optimized by calculating for the piston movement. Thus, a reduction of the piston running noise and minimization of cavitation hazard at the cylinder liner is achieved.

7.1.1.4 Installation and Running Plays One attempts to keep installation play at the piston skirt as small as possible so that uniformly smooth running is achieved in all operating situations. When working with light-alloy pistons, this objective can be achieved only with special engineering efforts. This is because of the high coefficient of thermal expansion for lightweight alloys. In the past, steel strips were often cast in place to influence expansion in response to heat (“control piston”). Figure 7.4 provides an overview of the amount of play found at the skirt and fire land for various piston designs. The amount of play at the wristpin, inside the wristpin boss, is important for smooth piston running and low wear at these bearing points. When determining the minimum play (Figure 7.5), it is necessary, in the case of gasoline engines, to determine whether a floating wristpin is used or whether it is fixed in the small-end eye by shrink fit. The floating wristpin is the standard design and the version that can handle the highest loads in the piston bosses. The “shrink-fit conrod,” which according to statements by some engine builders is more economical, is used only in gasoline engines. The shrink-fit conrod design is not suitable for modern diesel engines and for turbocharged gasoline engines. Floating Wristpin

Shrink-Fit Wristpin (Fixed Pin)

0.002–0.005

0.006–0.012

Figure 7.5 Minimum wristpin play in gasoline engines, in millimeters (not for racing engines).

7.1.1.5 Piston Masses The piston and its accessories (rings, wristpin, and circlips) form, with the oscillating share of the conrod, the oscillating masses. Depending on the engine design, free mass inertias and/or free moments occur; in some cases, these can no longer be compensated for or may be compensated for only with considerable effort. It is because of this phenomenon that, above all in the case of high-speed engines, the need to achieve the lowest possible oscillating masses arises. The piston and the wristpin account for the largest share of the oscillating masses. Consequently, weight optimization has to start here. Approximately 80% of the piston weight is located between the center of the wristpin and the upper surface of the head. The remaining 20% is located between the center of the wristpin and the end of the skirt. Of the major dimensions previously discussed, the determination of the compression height obtains decisive significance; with the determination

Internal Combustion Engine Handbook | 85

6606_Book.indb 85

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

of the compression height, approximately 80% of the piston weight is predetermined. When dealing with direct-injection gasoline engines, the piston head is used to deflect the stream and is shaped accordingly (see Figure 7.6). The pistons are both taller and heavier. The center of gravity shifts upward. In jet-controlled injection processes, the heads become more level.

[°C]

301.0 290.0 280.0 269.0 259.0 248.0 237.0 227.0 216.0 206.0 195.0 184.0 174.0 163.0 153.0 142.0

Figure 7.8 Temperature distribution at a piston for a gasoline engine. See color section page 1067.

Figure 7.6 Piston for a gasoline engine with direct injection.

The piston’s masses GN can best be compared when one references them to the comparison volume V – D3 (without piston rings and the wristpin). The mass indices GN/D3 (without rings and the wristpin) for proven piston designs are shown in Figure 7.7.

[°C]

375.0 359.0 343.0 327.0 311.0 295.0 279.0 263.0 247.0

Material

Operating Principle

GN/D3 (g/cm3)

Aluminum alloys

Four-cycle gasoline engines*

0.40–0.55

Two-cycle gasoline engines*

0.5–0.7

199.0

Four-cycle diesel engines

0.80–1.10

183.0

231.0 215.0

* Intake manifold injection.

167.0

Figure 7.7 Mass indices for passenger car pistons <100-mm diameter.

151.0 135.0

7.1.1.6 Operating Temperatures An important factor regarding operational reliability, safety, and service life is the component temperature for both the pistons and the cylinders. The piston head, exposed to the hot combustion gases, absorbs varying amounts of heat, depending on the operating situation (engine speed and torque). These volumes of heat, where the pistons are not oil cooled, are given off to the cylinder wall primarily through the first piston ring and, to a far lesser degree, through the piston skirt. When piston cooling is affected, by contrast, a major part of the heat volume is transferred to the motor oil. Because of the material cross sections determined by the engineering, there heat flows appear which result in characteristic temperature fields. Figure 7.8 and Figure 7.9 show typical temperature distributions at pistons for gasoline and diesel engines.

Figure 7.9 Temperature distribution at a piston with cooling channel for a diesel engine. See color section page 1067.

Severe thermal loading, on the one hand, reduces the durability of the material from which the piston is made. The critical points in this regard are the zenith of the boss and the edge of the recess in direct-injection diesel engines, and the transitional area between the hub connection point and the piston head in gasoline engines. On the other hand, the temperatures in the first piston ring groove are significant with regard to oil carbonization. Whenever certain limit values are exceeded, the piston rings tend to stick and, as a result, are limited in their functioning. In addition to the maximum temperatures, the dependency of piston temperatures on engine operating conditions (such as engine speed, mean pressure, ignition angle, and volume injected) is of significance. Figure 7.10 shows typical values

86 | Internal Combustion Engine Handbook

6606_Book.indb 86

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.1 Pistons/Wristpins/Wristpin Circlips

for gasoline and diesel engines used in passenger cars, in the area around the first piston ring groove, depending on the operation conditions. Change in Piston Temperature at Groove 1 (°C)

Engine Conditions

Change in Engine Conditions

4–8

Water cooling

Water temperature 10°C 50 % antifreeze

+5 to 10

10°C

1–3

Lubricating oil temperature (without piston cooling)

Injection nozzle in conrod big end

−8 to 15 on one side

Normal injection nozzle (stationary nozzle)

−10 to 30

Cooling channel

−25 to 50

Cooling oil temperature 10°C

4–8 (also at edge of recess)

Mean pressure (n = constant)

0.1 MPa

5–10 (15–20 at edge of recess)

Engine speed (pe = constant)

100 L/min

2–4

Ignition point, start of injection

1° kW

1.5–3.5

Fuel-to-air ratio, lambda

Lambda = 0.8–1.0

Little influence

Piston cooling with motor oil

Figure 7.10 Influence of engine operating conditions on the piston groove temperatures.

7.1.1.7 Piston Cooling Focused piston cooling becomes increasingly important in gasoline engines because of the increased engine output and charging. 7.1.1.7.1 Spray Cooling One version often found is a nozzle located at the lower end of the cylinder, through which motor oil is sprayed onto the inside contours of the piston. The cooling effect is dependent upon the volume of cooling oil and the surface area available for heat transfer. In this way, temperature reductions of up to 30°C can be attained at the first groove and the boss. A simpler version is a bore through the big-end eye, which is provided with oil from the conrod bearing lubrication system. In addition to a lesser cooling effect, the part of the stream of oil that meets the cylinder running surfaces provides better lubrication, which, in turn, offers greater security against fuel friction. 7.1.1.7.2 Pistons with Cooling Oil Cavities A more complex but more effective option for piston cooling is to provide cavities in those areas at the piston head and the ring grooves that are subjected to severe thermal loading. An annular cooling channel is supplied with oil, through a feed opening, by a spray nozzle; after taking on heat (⊗T up to approximately 40°C), the oil passes through a discharge opening on the opposite side of the piston and returns to the oil sump. The recommended specific masses for cooling oil

come to approximately 5 kg/kWh. A cooling channel cast directly at the ring carrier (“cooled ring carrier”) provides ideal effectiveness with regard to groove cooling. Figure 7.11 shows the typical application ranges for various piston designs. Operating Principle

Loading No piston cooling

Piston with spray cooling

Forged piston with spray cooling

Low ≈40 kW/L

Medium ≈65 kW/L

High ≥60 kW/L

Spray cooling

Cooling channel piston

Cooled ring carrier

Low ≤35 kW/l

Medium ≈35–60 kW/l

High >45 kW/l

Gasoline

Passenger car diesel

Figure 7.11 Survey of cooling variants.

7.1.1.8 Piston Designs Ongoing piston development has produced many designs, the most important of which, having proven themselves in practice, are presented here. In addition, new directions for development are being pursued, for example, pistons for engines with an extremely low profile, pistons made of composites with local reinforcing elements, or pistons with a variable compression height (VKH pistons), which permit variable compression ratios. Modern gasoline engines use lightweight designs with symmetrical or asymmetrical oval skirt shapes and, if indicated, differing wall thicknesses for the contact side and the opposite side. These piston designs are distinguished by optimized weight and particular flexibility in the center and lower skirt areas. It is for the reasons mentioned here that the control piston is becoming less and less common. Older designs are also discussed briefly in the interest of completeness. 7.1.1.8.1 Pistons with strip inserts to regulate thermal expansion, for installation gray cast iron engine blocks The primary objective in regulating piston design, and for many inventions in this sector, was and is the effort to reduce the relatively large differences in the coefficients of thermal expansion between gray cast engine blocks and aluminum pistons. Known solutions range from Invar strip pistons to the Hydrothermik or Hydrothermatik pistons. 7.1.1.8.2 Hydrothermik piston Hydrothermik pistons (Figure 7.12) are designs with a skirt profile formed in accordance with hydrodynamic aspects. They are installed in gasoline engines for passenger cars. The pistons are slotted at the transition from the piston head to the skirt, at the level of the third groove. These pistons are characterized by particularly smooth running and long service lives. The strips cast in place between the skirt and the wristpin bosses, made of non-alloyed steel, in conjunction with the lightweight metal that surrounds them, form regulation elements that reduce the thermal expansion of the skirt in the direction that is important for guidance within the cylinder.

Internal Combustion Engine Handbook | 87

6606_Book.indb 87

1/19/16 8:30 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

In the Hydrothermatik piston, the transition from the head area to the skirt is not slotted; the transitional cross sections are dimensioned so that, on the one hand, the flow of heat from the piston head to the skirt remains relatively unhindered, while, on the other hand, the effect of the steel strips, because of the connection of the skirt with the rigid head section, is not affected in any essential way. Thus, this piston design joins the high strength of the non-slotted piston with the advantages of the design using regulation strips. The Hydrothermatik piston is also suitable for use on naturally aspirated diesel engines.

Figure 7.12 Hydrothermik piston.

7.1.1.8.3 Hydrothermatik piston Hydrothermatik pistons (Figure 7.13) operate on the same expansion regulation principle as the Hydrothermik pistons.

7.1.1.8.4 Asymdukt piston This modern piston design is distinguished by very low weight, optimized support, and a boxlike, oval-shaped skirt section. It is outstandingly suited for use in modern gasoline engines for passenger cars. It is suitable for both aluminum engine blocks and gray cast engine blocks. With the flexible skirt design, the differences in thermal expansion between the gray cast block and the aluminum pistons can be excellently compensated for within the elastic range. The pistons may be either cast or forged. The forged version is used above all in high-performance sport engines or in heavily loaded, turbocharged gasoline engines. 7.1.1.8.5 Evotec piston The Asymduct piston is further developed for weight optimization under the name Evotec piston, which is characterized by a significant advantage in the oscillating masses without compromising the load-handling capabilities. 7.1.1.8.6 Piston for race cars These are always special designs (Figure 7.15). The compression height (KH in the illustrations) is very short, and the piston as a whole is superbly optimized for weight. Only forged pistons are used here. Weight optimization and piston cooling are decisive criteria for the design of these pistons. In Formula 1 engines, specific output of more than 200 kW/L and engine speeds exceeding 18,000 rpm are common. The service life of the pistons is matched to the extreme operating conditions.

Figure 7.13 Hydrothermatik piston.

7.1.1.8.7 Pistons for two-cycle engines In the two-cycle piston (Figure 7.16), the thermal loading is particularly high because of the more frequent exposure to heat; there is one ignition event for each rotation of the crankshaft.

Figure 7.14 Asymduct and Evotec pistons.

88 | Internal Combustion Engine Handbook

6606_Book.indb 88

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.1 Pistons/Wristpins/Wristpin Circlips

In addition, it has to close or open the inlet and outlet channels in the cylinder during its upward and downward strokes. This means that it has to control the exchange of gases. The result is severe thermal and mechanical loading.

Figure 7.17 Ring carrier piston with boss bushings made from special brass.

Figure 7.15 Forged Formula 1 piston.

The ring carrier material is made of a non-magnetic cast iron with a coefficient of thermal expansion similar to the material used for the piston itself. This material is particularly resistant to friction and impact wear. The groove that is most seriously endangered and the piston ring seated in it are effectively protected in this way against excessive wear. This is particularly advantageous where high operating temperatures and pressures are encountered, such as those found in diesel engines in particular. 7.1.1.8.9 Cooled pistons There are various types of cooling channels and cooling spaces to achieve particularly effective heat dissipation in the area near the combustion chamber and to combat the elevated temperatures resulting from performance increases. The cooling oil is generally delivered through fixed nozzles mounted in the crankcase. In the cooling channel piston (Figure 7.18), the ring-shaped cavities are created by inserting salt cores during casting. These cores are dissolved and removed with water introduced at very high pressure.

Figure 7.16 Piston and cylinder for a two-cycle engine.

Two-cycle pistons are equipped with one or two piston rings and, with regard to their outward design, can vary from the open “windowed” piston to the full skirt piston version. This depends on the design of the overflow channels (long or short channels). In this case, the pistons are normally manufactured from the MAHLE 138 supereutectic Al–Si alloy. 7.1.1.8.8 Ring carrier piston In the case of ring carrier pistons (Figure 7.17), introduced to mass production as early as 1931, the topmost ring groove and, in some cases, the second ring groove lie in a so-called ring carrier or groove insert that is joined permanently with the piston material by a metallic bond.

Figure 7.18 Cooling channel piston with ring carrier for a passenger car diesel engine.

Internal Combustion Engine Handbook | 89

6606_Book.indb 89

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.1.1.8.10 Piston with cooled ring carrier Another cooled piston variant is the piston with a “cooled ring carrier” (Figure 7.19). The “cooled ring carrier” permits much improved cooling of the first ring groove and the edge of the combustion recess, which is subjected to extreme thermal loading. The intensive cooling of the first ring groove makes it possible to use a rectangular ring instead of the double trapezoid ring normally employed.

Figure 7.20 Ferrotherm piston.

Figure 7.19 Passenger car piston with cooled ring carrier.

7.1.1.8.11 Piston with bushing in the boss bore One of the most heavily loaded areas of the piston is the wristpin bearing area. There the piston material is subjected to thermal loads of up to 240°C, and thus enters the temperature range in which the strength of the aluminum alloys declines. When dealing with extremely heavily loaded pistons, measures such as shaped bores, relief pockets, and oval boss bores are no longer sufficient to increase the load-carrying capacities of the boss. That is why reinforcing was developed for boss bores into which shrink-fit bushings made of a higher strength material (e.g., CuZn31Si1) are inserted.

With this design, the Ferrotherm piston offers not only greater strength and temperature resistance but also low wear values. Its constant, low oil consumption, its small dead space, and its relatively high surface temperature offer good prerequisites for complying with low exhaust emission limit values. 7.1.1.8.13 Monotherm piston The Monotherm piston 4 (Figure 7.21) grew out of development work for the Ferrotherm piston. This new design is a one-piece piston made of forged steel and highly optimized for weight.

7.1.1.8.12 Ferrotherm piston In the Ferrotherm piston (Figure 7.20), the guidance and sealing functions are separated one from another. The two sections, that is, the piston head and the piston skirt, are joined one with another in a movable fashion by the wristpin. The piston head, made of forged steel, transfers the ignition pressure through the wristpin and the conrod to the crankshaft. The lightweight aluminum skirt handles only the lateral forces that are created by the angular positions of the conrod and, being of the appropriate shape, guarantees the oil cooling necessary for the piston head. In addition to this “shaker cooling” via the skirt, enclosed cooling cavities can also be integrated into the piston head. For this purpose, the outer cooling cavity for the steel piston head is closed off with split tabs made of spring steel (Figure 7.20). Figure 7.21 Monotherm piston for utility vehicle engines.

90 | Internal Combustion Engine Handbook

6606_Book.indb 90

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.1 Pistons/Wristpins/Wristpin Circlips

At smaller compression heights and with machining above the clearance for the eye (on the inside), the piston weight, with the wristpin, can almost match the weight of a comparable aluminum piston with its wristpin. In the interest of improving piston cooling, the external cooling cavity is closed with two halves of a spring steel plate. The Monotherm piston is used primarily for utility vehicle engines subjected to heavy loading. 7.1.1.9 Piston Manufacture The latest in casting and machining equipment, in conjunction with an integrated quality management system, guarantees maximum quality across the entire product range. 7.1.1.9.1 Die casting Pistons made of aluminum alloys are manufactured, in the main, using the gravity die-casting process. The molds, made of ferrous materials, cause quick solidification of the molten metal; a fine-grained structure with good strength properties is formed at short casting cycle times. Optimized mold casting, in conjunction with carefully designed riser and gating technology, is necessary to achieve the most errorfree and dense casting possible. Graduated solidification aligning with the differences in wall thickness from the thin skirt to the thick piston head as mandated by the design lead to optimzation. Multi-part casting forms and casting cores provide great latitude in laying out the piston geometry so that even undercuts inside the piston, for example, can be realized. To increase wear resistance at the ring grooves, ring carriers made of austenitic cast iron with intermetallic bonding (Al–fin bonding) can be cast in place with as little trouble as for expansion-regulating steel struts or other engineering elements. By casting around cores made of compressed salt, which are then dissolved and removed with water, hollow cavities can be formed for piston cooling purposes. To do justice to high demands for quality and economy, multi-cavity molds and casting robots are used in mass-production operations. 7.1.1.9.2 Centrifugal casting The centrifugal casting process is used to manufacture the ring carriers used to reinforce the piston ring grooves. Tubes made of austenitic cast iron with flaked graphite are cast in rotating molds, and the ring carrier rings are then made up from the tubes. 7.1.1.9.3 Extrusion casting This process is known for use with wrought alloys—primarily for bars, ingots, and blocks. MAHLE has further refined this process, in which the extrusion is cooled with water immediately after leaving the mold so that it can be used with standard piston alloys. The high solidification speed has beneficial effects on the internal structure. The extrusions are cast in various diameters and serve as the feedstock material for forged pistons or piston components. 7.1.1.9.4 Forging (pressing) Forging or warm flow pressing is used to manufacture pistons and piston skirts (assembled pistons) from aluminum alloys for engines subject to heavy loading. Sections of extrusion castings

are normally used as the feedstock material. Reforming results in much higher and much more uniform strength values than those can be achieved with casting. A further option is found in using semi-finished products made of blast-compacted materials or those made up in a powder metallurgical process. This process technology makes it possible to employ extremely heat-resistant materials for high-performance (racing) pistons, which could not be manufactured with hot metal technology. 7.1.1.9 Liquid Pressing (Liquostatik, Squeeze Casting) Squeeze casting differs from gravity die casting by the pressure applied to the molten material (up to and beyond 100 MPa), which is maintained until the casting has fully solidified. The extremely good contact of the molten material with the mold walls as it solidifies makes for very fast solidification. In this way, a very fine structure, advantageous in terms of material strength, is created. Squeeze casting makes it possible to manufacture pistons that are reinforced locally with ceramic fibers or porous metallic materials at the piston head or in the areas around the ring grooves or bosses. These cast-in-place components are penetrated completely by the piston alloy owing to the pressure applied to the molten metal. 7.1.1.9.6 Tempering Lightweight alloy pistons, depending on their alloy and the manufacturing process used, are subjected to single-stage or multi-stage heat treatment. In this way, the hardness and strength of most alloys can be increased. In addition, the remaining changes in volume (“growing”) and the dimensional changes that would otherwise occur under the influence of operating temperature are pre-empted. 7.1.1.9.7 Machining Leading piston manufacturers themselves develop manufacturing concepts and special equipment for machining pistons. The distinguishing features are found in •• complex shapes at the exterior of the pistons and close tolerances in piston diameter •• complex piston head shapes (round, oval, or special shapes) and close boss bore tolerances •• high surface quality and geometry in rectangular and trapezoidal grooves in aluminum piston alloys as well as in ring carriers made of Niresist •• close compression height tolerances. Thus, complex exterior piston shapes are machined on user-programmable shaping lathes whose Computerized Numerical Control (CNC) guarantee great flexibility and high quality. Irregular piston shapes that may, for example, be discovered empirically in engine test series can be easily manufactured in volume. The same applies to the machining of the boss bore. Using a precision drill press, which is also user programmable, differing boss bore shapes are possible along the direction of the boss bore axis and at the circumference of the boss bore.

Internal Combustion Engine Handbook | 91

6606_Book.indb 91

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Machining the piston grooves in the ferrous material making up ring carrier type pistons places particularly high demands on machinery capabilities. 7.1.1.10 Protection of Running Surfaces/Surface Protection The materials that have been highly developed to date, and the precision machining processes used for pistons, ensure high wear resistance and good running properties. In spite of applying protective coatings to the piston skirt, offering special emergency running properties is advantageous for the break-in phase and unfavorable dry-running operating conditions following frequent cold start attempts, temporary loading, and insufficient lubrication. Under certain circumstances, wear protection finishes may be required in the groove area. Severe thermal loading at the piston head must be counteracted with additional local protective measures. The coatings and finishes described below have proven their suitability for the various tasks in many applications. With the use of automated machinery engineered, especially for surface treatment, pistons may be finished by •• tin plating the entire piston surface •• applying phosphate and graphite (spray process) •• applying graphite (screen printing) with and without phosphate a. piston skirt b. piston shaft and ring section •• partial iron plating of the piston skirt (in conjunction with cylinder running surfaces made of aluminum) •• hard anodized finishing a. first groove b. piston head (complete or partial). 7.1.1.10.1 Improving slip properties A thin plating of tin, which is applied by a chemical process to the lightweight metal piston, protects against seizing during cold starts and during break-in at unfavorable lubrication conditions. The layer is approximately 1-μ m thick. Where there are narrow installation tolerances and very high requirements for protection against seizure, the GRAFAL running surface is used. This finish comprises a graphitefilled synthetic resin that adheres permanently to the piston running surface. This layer is generally 20-μ m thick. Pistons for passenger car engines are typically finished with the GRAFAL 255 version, applied in a screen-printing process, while the sprayed GRAFAL 240 or the screen-printed GRAFAL 255 version is used on pistons for utility vehicle engines and industrial engines. In aluminum pistons, the pairing of the wristpin and the boss is normally not critical in terms of sliding processes, and they require no special coatings assuming the correct shapes and tolerances. In steel pistons, on the other hand, special protective measures are required. As an alternate to boss bushings, a slip phosphate coating becomes more important here.

7.1.1.10.2 Increased wear protection FERROSTAN pistons are paired with non-coated SILUMAL cylinders or other non-coated, Al–Si-based cylinder materials. The skirts of FERROSTAN pistons are iron-plated to a thickness of 6 ⎧m and hardness of 350–600 HV. The iron layer is galvanically precipitated out, to precise dimensions, from special electrolytes. To conserve and improve slip properties, the iron-plated piston is finished with an additional layer of tin, 1-⎧m thick. Something new in technology is the application of layers, containing iron particles, using a screen-printing process. Known as FERROPRINT layers, they have been successfully introduced into mass production. Owing to increased thermal and mechanical loads, wear and fretting effects are more frequently seen along the flanks of the first groove in gasoline engine pistons. Hard anodizing for the endangered area has been introduced in volume production as an effective countermeasure. When hard anodizing aluminum alloys, a zone near the surface of the aluminum substrate is transformed by electrolytic means into aluminum oxide. The layer created here is ceramic in nature, with a hardness of approximately 400 HV. In this application, a layer of approximately 15-⎧m thick is specified, and the process parameters are optimized so that the layer roughness is relatively moderate, eliminating the need for subsequent machining of the groove flanks. 7.1.1.10.3 Thermal protection Pistons for diesel engines are subjected to severe temperature fluctuation stresses in the area at the top and in the combustion recess. The result may be fissures resulting from temperature fluctuation. A hard oxide layer at the top of the aluminum piston, shown in Figure 7.22, typically approximately 80-⎧m thick, improves resistance to the effects of temperature fluctuation, and thus prevents fissuring at the edge of the recess and/or in the top. Cutouts along the direction of the wristpin make sense to avoid notch effects in the area where maximum tensile strain occurs.

Figure 7.22 Hard anodized piston head.

92 | Internal Combustion Engine Handbook

6606_Book.indb 92

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.1 Pistons/Wristpins/Wristpin Circlips

7.1.1.11 Piston Materials 7.1.1.11.1 Aluminum alloys Pure aluminum is too soft and too susceptible to wear for use in pistons and for many other purposes. That is why alloys have been developed that are matched particularly to the requirements found in piston engineering. They combine, at low specific weight, good heat strength properties with a low tendency to wear, high thermal conductivity and, in most cases, a low coefficient of thermal expansion as well. Two groups of alloys have come into being, depending on the primary additive silicon or copper. Aluminum–silicon alloys: •• Eutectic alloys contain from 11% to 13% silicon and smaller amounts of Cu, Mg, Ni, and the like. Included in this group of piston alloys, the ones used most frequently in engine construction, is MAHLE 124, which is also used for cylinders. For many applications, they offer an ideal combination of mechanical, physical, and technological properties. The MAHLE 142 alloy, with a greater proportion of copper and nickel, was developed for use particularly at high temperatures. It is distinguished by better thermal stability and considerably improved strength when heated. A further step in this direction is the nearly eutectic MAHLE 147+ alloy. •• Supereutectic alloys contain from 15% to 25% of silicon and use copper, magnesium, and nickel as additives to deal with high temperatures; examples include MAHLE 138 and MAHLE 145. They are used for pistons wherever a need for reduced thermal expansion and greater wear resistance is in the forefront. The MAHLE 147 (SILUMAL) alloy is used for cylinders and/or engine blocks without any special treatment for the running surfaces. Figure 7.23 and Figure 7.24 show characteristic values for the materials.

Designation

MAHLE 124

MAHLE 138

MAHLE 142

Young’s modulus E (N/mm2)

20°C 150°C 250°C 350°C

80000 77000 72000 65000

84000 80000 75000 71000

84000 79000 75000 70000

Thermal conduction coefficient λ (W/mk)

20°C 150°C 250°C 350°C

155 156 159 164

143 147 150 156

130 136 142 146

Mean, linear thermal expansion α (1/k · 10–6)

 20– 100°C  20– 200°C  20– 300°C  20– 400°C

20 21 21.9 22.8

18.6 19.5 20.2 20.8

19.2 20.5 21.1 21.8

Density ρ (g/cm3)

 20°C

 2.70

 2.68

 2.77

Figure 7.23 Physical properties of MAHLE aluminum piston alloys.

Aluminum–copper alloys: To a lesser extent, alloys containing copper but almost no silicon and just a small amount of nickel as an additive are used for their good heat strength. In comparison with the Al–Si alloys, they exhibit greater thermal expansion and less wear resistance. While the Al–Si alloys can be both cast and reformed when warm, the Al–Cu alloys are more suitable for warm reforming. 7.1.1.11.2 Lightweight alloy bonded materials The introduction of bonded materials technology opened a number of different options for significantly increasing the load-bearing capacities of lightweight metal pistons. Here reinforcement elements such as ceramics, carbon fibers, or porous metallic materials are arranged in closely defined

Strength values are applicable to test bars made up separately. Designation

MAHLE124 G

MAHLE 124 P

MAHLE 138 G

MAHLE 142

Tensile strength Rm (N/mm2)

 20°C 150°C 250°C 350°C

200–250 180–230 100–150 40–65

300–370 250–300 110–170 40–70

180–220 170–210 100–140 60–80

200–280 180–240 100–160 50–70

Elongation limit Rp0.2 (N/mm2)

 20°C 150°C 250°C 350°C

190–230 180–220 70–110 20–30

280–340 230–280 90–120 10–30

170–200 150–190 80–120 20–40

190–250 180–220 80–120 40–60

Ductile yield A (%)

 20°C 150°C 300°C 400°C

 0.1–1.5  1.0–1.5  2–4  9–15

 1–3  2.5–4.5  8–10  31–35

 0.2–1.0  0.3–1.2  1.0–2.2  5–7

 0.1–0.5  0.2–1.0  1–3.5  5–13

Fatigue strength at reversed bending sbw (N/mm2)

 20°C 150°C 250°C 350°C

 80–120  70–110  50–70  15–30

110–140 90–120 60–70 15–25

 80–110  60–90  40–60  15–30

 90–130  70–110  50–70  30–50

 0.9

 0.95

Relative wear index

1

Brinell hardness HB 2.5/62.5

 90–130

100–150

Figure 7.24 Mechanical properties of MAHLE aluminum piston alloys.

Internal Combustion Engine Handbook | 93

6606_Book.indb 93

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

positions in regions of the piston that are subject to particularly high loading. The bonded material is manufactured by infiltrating the reinforcing elements with lightweight metals such as aluminum or magnesium using the squeeze casting process. High price and unfavorable creep properties are the primary reasons that magnesium is not yet used in largescale production. Among the many options available, reinforcing aluminum pistons with short ceramic fibers made of aluminum oxide is the one most widely adopted for series production. Following a washing process to remove components that are not fiber shaped, the fibers are processed to create mold components that can be cast (preforms) with fiber content of between 10% and 20% by volume. In this way, considerable improvements in strength can be achieved at the edge of the recess in directinjection diesel pistons, for instance. A reinforcing element made of porous sintered steel with uniform porosity from 30% to 50% was developed for ring grooves. The Porostatik material offers favorable wear properties and a sure bond with the surrounding aluminum material. It is suitable, for example, for reinforcing ring grooves that are at an extremely high location, leaving hardly any room to cast around on the side toward the piston head.

7.1.2 Wristpins 7.1.2.1 Function The wristpin makes the connection between the piston and the connecting rod. It is subjected to the extreme, alternating loads exerted by the pressure of the exploding gas and the mass inertias. Because of the small relative motions (rotary motions) between the piston and the wristpin and between the wristpin and the conrod, the lubrication situation is unfavorable.

The essential aspects for the design of the wristpin are •• sufficient wristpin strength (operating safety) •• reverse effect on piston loading •• weight (mass inertia) •• surface quality and dimensional accuracy (running properties) •• surface hardness (wear). Today, the wristpin is usually dimensioned with the aid of three-dimensional (3-D) finite-element (FE) calculations, in some cases taking into account the shape of the lubricating oil film (pressure distribution) in the boss and the conrod. Solid knowledge in the dynamic behavior of the material is required to evaluate a material’s dynamic properties. Guideline values for selecting the wristpin diameter for the various application ranges can be found in Figure 7.25. 7.1.2.4 Materials The materials that are used primarily today are 17Cr3 and 16MnCr5 case-hardened steels. Nitrated steel alloy 31CrMoV9 can be used where higher loading is anticipated. Figure 7.26 shows characterizing values for the materials used in wristpins. Wristpins for racing use are manufactured in an electro-slag remelting process (Elektro-Schlacke-Umschmelzverfahren) to ensure a higher degree of purity in the material.

7.1.3 Wristpin Circlips

Because the wristpin is not held in the connecting rod by shrink fit, it has to be secured against wandering laterally from the holes in the boss and making contact with the cylinder wall. Used almost exclusively for this purpose, inside circlips (made of spring steel) are installed in grooves at the outer edge of the boss holes (see Figure 7.27).

7.1.2.2 Designs The wristpin with cylindrical inside and outside contours has been successful in most applications. To reduce weight and, with it, the mass inertias, the outer ends of the wristpins’ inside bore may be conical because the load is less there. Wristpins in passenger car gasoline engines are often held in the conrod with a tension because of shrinkage (“shrink-fit wristpin” and “clamp-type wristpin”). In more heavily loaded gasoline and diesel engines, the wristpin “floats” in the conrod. It is secured with circlips to keep it from wandering laterally and out of the piston (see Section 7.1.3). 7.1.2.3 Requirements and Dimensioning Under the influence of the forces described above, loading on the wristpins is very complex and is influenced, in addition, by the deformation of the piston and wristpin.

Ratio of Wristpin Outside Diameter to Piston Diameter

Ratio of Wristpin Outside Diameter to Wristpin Inside Diameter

Small two-cycle engines

0.20–0.25

0.60–0.75

Passenger cars

0.20–0.26

0.55–0.70

Passenger cars

0.32–0.40

0.48–0.52

Application Gasoline engines Diesel engines

Figure 7.27 Wristpin circlips.

Figure 7.25 Wristpin dimensions (guideline values).

94 | Internal Combustion Engine Handbook

6606_Book.indb 94

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.2 Connecting Rod

L (17Cr3) Tool Steel

Material Class

Chemical composition in wt. %

M (16MnCr5) Tool Steel

N (31CrMoV9) Nitriding Steel

C

0.12–0.20

0.14–0.19

0.26–0.34

Si

0.15–0.40

0.15–0.40

0.15–0.35

Mn

0.40–0.70

1.00–1.30

0.40–0.70

P

maximal 0.035

maximal 0.035

maximal 0.025

S

maximal 0.035

maximal 0.035

maximal 0.25

Cr

0.40–0.90

0.80–1.10

2.3–2.7

Mo

—

—

0.15–0.25

V

—

—

0.10–0.20

Surface hardness HRC

59–65 (vol. const. 57–65)

59–65

59–65

Core strength in N/mm2

From 700 to 1500, From 850 to 1350, 1000–1400 depending on wall thickness depending on wall thickness

Mean linear thermal expansion 1/K ⋅ 10–6 20–200°C

12.8

12.7

13.1

Heat conductivity index (W/m · K)

51.9 48.2

50.0 48.7

46.4 45.5 210,000

 20°C 200°C

Young’s modulus (N/mm2)

210,000

210,000

Density (kg/dm3)

7.85

7.85

7.85

Use

Standard material for wristpins

For heavily loaded wristpins

For heavily loaded wristpins (special cases)

Figure 7.26 Wristpin steels DIN 73 126.

Where the wristpin diameters are small, wound rings made from round wire are normally used. In engines that run at slower speeds, the ends of the snap rings may be bent inward to form a hook-like shape to facilitate installation. Such rings, when made up for racing use, are often bent outward at one end to keep them from rotating. If, in isolated cases, greater axial thrust is encountered in the wristpins, outside circlips may also be used. These circlips are mounted in grooves at the ends of the wristpins.

7.2 Connecting Rod The power system for reciprocating internal combustion engines uses a crank drive in which the connector rod end or the connecting rod joins the piston with the crankshaft. The connecting rod converts the reciprocating movement of the piston into rotary motion. Moreover, the connecting rod transfers forces from the piston to the crankshaft. A further function of the connecting rod is to accept channels used to supply lubricating oil to the piston bushing in cases where the wristpin is of a floating design. The weight and design of the connecting rod have a direct influence on the power-to-weight ratio, power output, and smooth engine operation. This is why connecting rods that have been optimized in terms of weight are gaining more importance in terms of engine running quality. Corresponding to the inverted attitude of the connecting rod in the early engines, those built in the nineteenth century, the lower section (at the piston) is sometimes referred to as the conrod foot, while the upper end (at the crankshaft) is called the connecting rod head.

7.2.1 Connecting Rod Design

The connecting rod has two so-called conrod ends [7-7]. It is at the small conrod eye that the connection to the piston is made by the wristpin. Because of the lateral deflection of the connecting rod as the crankshaft turns, the rod end has to be attached to the piston in a way that allows it to rotate. This is done with the help of a sliding bearing. For this purpose, a bearing bushing is pressed into the small conrod eye during assembly (Figure 7.28). Alternately, the bearing may be integrated into the piston. In this case, the wristpin is held in the small connecting rod eye with shrink fit. The split, large connecting rod eye is located at the crankshaft end of the rod. Proper functioning is ensured with a sliding bearing (rolling bearings are used less often) and by fixing and screwing down the conrod bearing cap. The conrod shaft joins the two connecting rod eyes. This section may have a special cross section, depending on the requirements at hand, for example, I-shaped or H-shaped. The connecting rod has to ensure sufficient slip in the bearings at both the small and the large ends. Grooves may be machined in the ends to improve lubrication at the large end and/or to lubricate the cylinder and the wristpin; these grooves facilitate lubricant feed. The wristpin bearing may be lubricated by means of a hole along the longitudinal axis of the shaft, through which oil is fed from the large end. This channel interferes with the structural relationships within high-performance conrod bearings. That is why, as an alternative to a longitudinal channel through the shaft, one or more holes may be drilled at the small end in the surface facing the piston (Figure 7.28). This solution is more economical.

Internal Combustion Engine Handbook | 95

6606_Book.indb 95

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

4

are influenced by the masses of the piston, the wristpin, and the conrod.

24 2

1

3

7

A A

23

19 20 21 22 23 24 25 26

5

8

9 25

20 16

l

l · cosb

6

b

14

l · sinb 19

21 13

r · cosa

13 14 15 16 17 18

Small conrod eye Conrod bushing Wristpin bore Oil bore Conrod length Support surface Skirt Shoulder Sprayed oil bore Boss Bolt Conrod nut (not provided) Conrod end 26 Conrod bearing shells Balancing mass Pin bore Conrod width Screw head contact surface Separation plane Retaining nose Conrod thickness Rib thickness Wall thickness Front face Groove in front face Large conrod eye

17

r+l

1 2 3 4 5 6 7 8 9 10 11 12

22

10 18

15

11

a

r

r · sina

24

Figure 7.28 Geometry and designations for a connecting rod with straight split (source: Federal Mogul).

7.2.2 Loading

The connecting rod is subject to a load exerted by the gas forces inside the cylinder and the inertia of the moving masses. Figure 7.29 shows the kinematic relationships in the crankshaft drive. The lateral deflection in the conrod oscillation plane generates centrifugal forces leading to bending that, however, can be neglected in the first approximation. The acceleratedly decelerated motion of the masses in the conrod and piston causes tensile strain in the shaft and at the transition from the shaft to the large eye. Thus, the conrod is subjected to alternating tensile and compressive forces; in diesel and turbocharged gasoline engines, the magnitude of the compressive force exceeds that of the tensile force. For this reason, resistance to buckling has to be examined carefully when engineering the conrod. The tensile forces are also decisive in today’s high-speed gasoline engines. The inertial forces generated during accelerated and decelerated motion within a reciprocating engine’s working cycle

Figure 7.29 Kinematics for the crank drive.

To simplify the determination of the resulting forces, the mass of the conrod is divided into rotating and reciprocating portions, assuming that the overall mass and the center of gravity for the conrod are retained unchanged. The masses concentrated in the large eye are assigned exclusively to rotational movement; those concentrated in the small eye are assigned to reciprocating motion. To determine the various shares of the overall mass, it is first necessary to find the center of gravity (SP in the equation) for the conrod. The share of mass for the small eye results from

mPl, small eye = mPl, total  ⋅

SP l

(7.1)

with l as the distance between the centers of the conrod eyes, which is defined as the conrod length. The difference between this and overall weight gives the share of mass for the large eye [7-6].

96 | Internal Combustion Engine Handbook

6606_Book.indb 96

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.2 Connecting Rod

The reciprocating masses for the conrod (and the piston with the wristpin and piston rings) influence, by the inertial forces they generate, the loading and the smooth running of the engine. These reciprocating forces can be fully compensated for only by providing additional compensating shafts. Thus, it is necessary to reduce the conrod mass and/or the conrod’s share of reciprocating mass. This can be done by optimizing the shape of the conrod shaft and, for example, by using a trapezoidal design for the small eye. The true movement situation for a particle of mass in a conrod and thus the force effects are far more complex than what is reflected in the breakdown described above, this being only an approximation. Essentially, each particle of mass between the small and large conrod eyes executes a reciprocating and a rotating movement. The reciprocal component declines in the direction of the large conrod eye. Suitable FE method (FEM) calculation processes make it possible to simulate this dynamic behavior and to assess the forces exerted. The elasto-hydrodynamic bearing calculations (see Section 7.19.2.3) have also demonstrated how essential the deformations of the conrod eyes are for the running behavior of the conrod bearing and the conrod bushing. For this reason, the conrod design should be optimized with the bearing calculation. The masses for various conrods are shown in Figure 7.30. Application

Mass (kg)

Mass-production diesel truck

1.6–5

b. guidance by small pins next to the bolts or the bushings that surround the bolts (Figure 7.31) c. milling toothed ridges into the parting plane d. guidance by the separation surface at split (cracking) (Figure 7.32). If pins, bushings, or fracture-split conrods are used, one may do without body-fit bolts. In this case, the structure at the parting surface or the pins and bushings offer sufficient resistance to relative motion between the upper and lower halves (see Section 7.22.3.3).

Figure 7.31 Fitting bearings and expanding bolt.

Material Forged steel

Mass-production Spark Ignition 0.4–1 (SI) engine car

Forged steel, gray casting, sintered steel

Sport use

0.4–0.7

Steel, titanium

Racing engine/F1

0.3–0.4

Titanium, carbon fiber

Compressor

0.2–0.6

Aluminum

Figure 7.30 Conrod masses for various applications.

7.2.3 Conrod Bolts

The upper and lower halves of the big end are held with the conrod bolts. These threaded connections must fulfill two functions [7-8]: •• The conrod bolts must prevent any gap forming in the separation plane between the lower half and the upper half of the big end. The forces that are effective on the conrod bolts include the inertial forces of the conrod and the piston, along with a transverse force resulting from off-center loading and the forces resulting from “crushing” the bearing shell protrusion. During engine assembly, the bolts, opposing the effective inertial force, are normally preloaded by controlled tightening to the offset limit or rotary torque plus rotation angle [7-9], [7-10]. •• The conrod and the cap have to be moved toward one another precisely and secured against shifting (offset). There are several options to choose from: a. guidance by the conrod bolts, the shoulder or grooves of which are in the parting plane and thus prevent the upper and lower halves from shifting

Figure 7.32 Fracture-split conrod.

7.2.4 Design

The following aspects are of significance with regard to conrod design: •• Dimensional stability of the areas that accept the two bearing shells. •• Oil channels for lubricating the small-end eye may be required (unusual in modern designs). •• Separation of the big end bearing for mounting on the crank shaft journal. •• Fixing and securing the conrod cap. •• Engineering the conrod web to optimize design and/or reduce masses. •• Design of critical zones in accordance with loading.

Internal Combustion Engine Handbook | 97

6606_Book.indb 97

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.34 Warm blanks (source: Krupp Gerlach). Figure 7.33 Stress analysis for a conrod with an angular split, with a trapezoidal small end (half model, Federal-Mogul). See color section page 1068.

To reduce the mass of the piston and/or the conrod, the small end may be flattened toward the top, creating a trapezoidal shape. This shape, for reasons associated with the loading (in turbocharged engines, for example), is advantageous because it permits close spacing to the wristpin bosses and thus reduced wristpin flexure. The big end of the conrod is split to permit assembly on the crankshaft and is held with two bolts. The big end is normally split perpendicular to the long axis of the conrod. As an alternative, to reduce the maximum width of the conrod, the big end can also be split at an angle. This angular version makes it possible to pass the conrod (without the cap mounted at the big end) through the cylinder for assembly. The angled split conrod has some disadvantages. The separation plane is forced to handle great lateral forces. The unsymmetrical structure of the big end in load direction results in further advantages because uneven static and dynamic deformations of the bearing housings are considerably impeded. Conrods with an angled split are used primarily in V-block engines and large diesel engines, which, because of the loading involved, have large-diameter crankshaft journals. The big end and small end are joined by the conrod web, which has an I or H cross section. This makes it possible to satisfy requirements for reduced weight at a high section modulus.

The lateral forces on the piston rise with the conrod ratio. This can, for instance, result in modified specifications for the piston’s engineering design. As the conrod ratio falls, the overall height of the engine rises as a result of the increase in cylinder block height. Finally, restrictions imposed by the manufacturing process (cylinder block height) may prohibit a change in the conrod.

7.2.4.1 Conrod Ratio The conrod ratio is a comparative geometric magnitude, based on the crank radius r and the distance l between the centers of the small-end and big-end eyes (Figure 7.34). It is defined as

l = r /l

(7.2)

In passenger car engines, this value is normally between 0.28 and 0.33 with the lower values applicable to diesel engines. The selection of conrod length is influenced by many factors such as the stroke/bore ratio, piston speed, engine speed, peak combustion chamber pressure, engine block height, and piston design.

Figure 7.35 Blasted blanks (source: Plettac).

7.2.5 Conrod Manufacture 7.2.5.1 Manufacturing the Blank The blank for the conrod may be manufactured in any of a number of different ways, depending on the particulars of the application.

98 | Internal Combustion Engine Handbook

6606_Book.indb 98

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.2 Connecting Rod

(a) Drop forging: The feedstock material for making up the blank is a steel bar with a round or rectangular cross section, which is heated to a temperature of between 1250 and 1300°C. A roll-forging process is used to effect a preliminary redistribution of the masses toward the big and small ends. As an alternative to roll forging, cross-wedge rolling may also be employed, improving the preliminary geometry for the blank.

Because the forging procedure in this manufacturing process is costly, developments are now being pursued with the goal of eliminating this by using new powder technologies [7-16], [7-17].

The major reforming process takes place in a press or a hammer unit. Excess material flows into flash, which is removed in a subsequent operation. Simultaneous with flash removal, the big eye and, in the case of larger conrods, the small eye are punched.

Powder

To achieve the required structural and strength characteristics, the conrod requires various treatment processes, the choice depending on the steel alloy used: •• hardening with the forging heat (VS) •• controlled cooling in an air stream (BY) •• conventional hardening. Then the scale on the blank is removed by blasting; here compression stresses of 200 MPa are generated near the surface. Additional procedures such as fissure inspections follow.

Preform, pressed and sintered

In most cases, the conrod web and the end are cast as a unit, and then separated during later machining. Depending on the conrod and the capacity of the available equipment, productivity can be boosted by tandem forging, that is, shaping two conrods simultaneously. (b) Casting: The starting point for making up the blank is a model made of plastic or metal, comprising two halves which, when put together, create a positive image of the conrod. Several such identical halves are mounted on a model plate and joined with the model for the casting and gating system. In a process that can be reproduced many times, the two model plates are imaged by compacted green sand. The sand molds represent a negative image of the corresponding model plate. Placed one above the other, they form a hollow cavity in the shape of the conrod being manufactured. This is filled with liquid casting iron that is melted in a cupola blast furnace or electric furnace with steel scrap used as the feedstock material. The metal solidifies slowly inside the mold. (c) Sintering: The manufacturing process begins with servohydraulic pressing of the powder, in its final alloy, to create a powder preform. Weighing follows to ensure that this preform is within narrow weight tolerances of ±0.5%. The sintering process, illustrated in Figure 7.36, takes place at approximately 1120°C in an electrically heated continuous charge furnace. The parts remain here for approximately 15 min. Subsequent forging merely reduces the height of the component to increase the component density to the maximum theoretical limit. Then ball blasting is used to relieve the strain in the surface to the desired level.

Forged and finished

Figure 7.36 Process—sinter-forged conrod.

7.2.5.2 Machining The blanks are machined down to the final dimensions. In mass production, this is done in fully automatic lines that are integrated into the engine manufacturing process. Machining centers with a lower degree of automation are available for smaller production runs. After machining, the finished part is weighed and classified. Conrods in a particular weight class are then installed in any given engine. If the blank was already manufactured to close weight tolerances, then it may be possible to do without this classification step. To achieve the specified weight for the finished conrod, tabs can be provided at the small and/or big end of the blank (Figure 7.37). During mechanical finishing, these tabs are ground down far enough that the specified weight value is attained.

Internal Combustion Engine Handbook | 99

6606_Book.indb 99

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Cam weight

the separation surfaces, which used to be standard, is eliminated. The two halves fit together exactly pursuant cracking and, with the irregular surface, are secured against relative movement, without the need for any additional guide elements. A further benefit is found in the use of a simplified conrod bolt because it does not need to carry out guidance or lateral fixing functions [7-12].

Figure 7.37 Conventional conrod, body and cap forged separately, with bolts and nuts.

In more modern manufacturing processes, the manufacturing parameters can be monitored exactly so that blanks can be made within adequate weight tolerances. Thus, grinding to remove excess material provided deliberately for this purpose is rarely seen today. The processing steps are described below, by example, for conrods that are split after manufacture (cracking):

Figure 7.38 Design differences between a fracture-split conrod (above) and a sawn conrod.

•• grinding the faces of the big and small ends •• prespindling the big and small ends •• drilling and tapping the bolt holes •• cracking, blowing off fracture waste •• bolting the cap to the upper half of the big end and—if necessary—inserting the guide bushing •• loosening bolts, opening cap, retightening bolts •• finishing final grinding, milling the trapezoid of the small eye •• drilling out the small eye

Notching by laser

•• spindling the big end and optional honing. The term “cracking” or “fracture splitting” describes the separation of the conrod web and the cap by breaking the latter away during processing. The prerequisites for this process are, in terms of the materials, a coarse-grain structure and, in terms of equipment, a cracking unit that can apply the required breaking energy at high speed. If the material exhibits a ratio of tensile strength to tensile yield strength (0.2 offset limit) that is near 2:1, then cracking can be carried out without any major deformation of the part. Blanks made with any of the modern manufacturing processes can be split by cracking [7-11]. The difference in the design of the conrod is shown in Figure 7.38. In preparation for cracking, notches are made in the side surfaces of the big-end eye by laser or broaching to achieve a deep notch effect at the desired separation plane (see Figure 7.39). The large eye is positioned over a two-part breaker drift and fixed in place. The breaker drift is spread at high speed, and the stresses created in the workpiece initiate breaks within the notches. These breaks then propagate radially outward. If this process runs optimally to conclusion, then the outof-roundness following cracking will be 30 μ m at the most. The advantage offered by fracture splitting is found, primarily, in reducing the number of processing steps. Machining

Separating by wedge fracturing

Figure 7.39 Conrod fracture splitting.

Fracture-split conrods are an economical alternative to conrods separated in a conventional fashion.

7.2.6 Conrod Materials

Depending on the particulars of the application and the resultant loads, any of a number of different materials may be used for connecting rods.

100 | Internal Combustion Engine Handbook

6606_Book.indb 100

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.3 Piston Rings

7.2.6.1 Cast Materials The casting materials used most widely for connecting rods are nodular cast iron (GGG-70) and black malleable cast iron (GTS-70). GGG-70 has both technical and economic advantages when compared with malleable cast iron. In particular, the specific oscillation resistance, which is important for conrods, is considerably greater for GGG-70. GGG-70 is an iron–carbon casting material; graphite inclusions that are largely spherical are introduced into a basic structure that is primarily pearlitic. The compact shape of the graphite gives the material an optimum strength and ductility. At the same time, the graphite is also responsible for the good casting properties. The required structure is created during the casting process without additional heat treatment. In the case of malleable iron, which is also an iron–carbon material, the structure is determined by applying heat subsequent to casting. 7.2.6.2 Forged Steel Most conrods are manufactured from steel in the drop forge process. In most cases, micro-alloyed steel such as 27MnVS6 BY or carbon manganese steels like C40 mod BY are used. Steel with high carbon content (C70 S6 BY) is used for forged and fracture-split (cracked) conrods. These materials attain a tensile strength of Rm = 1000 MPa [7-13]. Available for high-performance conrods is 34CrNi-Mo6 V (or 42CrMo4), a steel alloy that achieves a tensile strength of 1200 MPa. In this case, additional heat treatment (hardening) is required. New developments in steel have reached tensile strengths— even in materials used for cracking—of up to 1000 MPa at 0.2 offset limits in excess of 700 MPa. These steels are identified with the designation “C70+” in the table of materials [7-14]. To meet the demands for higher tensile strengths to reduce mass, the precipitation-hardened ferritic–pearlitic steel 36MnVS4 is integrated in the development. Similar to C70S6, this new steel has good “cracking” properties and exhibits 30% higher fatigue strength [7-18]. 7.2.6.3 Powdered Metal The material P/F-11C50 is now the most common for conrods made from powdered metal. The strength-increasing elements, 2% Cu and 0.5% C, achieve a tensile strength of up to 950 MPa after sintering and forging [7-15]. Further developments of

this sinter material by increasing the copper portion from 2% to 3% have resulted in a 10% improvement of the tensile strength and an increase of 22% in fatigue strength under alternating loads [7-19]. 7.2.6.4 Alternate Materials In addition to the obligatory materials used for conrods in mass production, explorations into using alternate materials pursue principally the objective of reducing conrod weight while maintaining load-handling capabilities. Carbon fiber reinforced aluminum or carbon fiber reinforced plastic is used for this purpose. Widely used in racing are titanium conrods, with which a considerable weight reduction is achieved. The disadvantage of the titanium conrods is the strong tendency for bores to expand during operations, which has a deleterious effect on the tightness of the seat for the bearing shells. Another drawback is the fact that titanium is not a good “friction partner” for steel. Consequently, slip coatings on the mating surfaces are needed to protect against scuffing (friction-induced damage) and/or on the bearing’s steel backing to prevent fretting. Common to all conrods made of these alternate materials and fabricated for individual engines are the high manufacturing costs that hinder greater use in mass production engines. The most important materials and their properties are summarized in Figure 7.40.

7.3 Piston Rings Piston rings are metallic gaskets whose functions are to seal the combustion chamber against the crankcase, to transmit heat from the piston to the cylinder wall, and to regulate the oil required, on one hand, to ensure a minimum oil film to create a hydrodynamic lubrication film on the cylinder sleeve and, on the other hand, to keep oil consumption as low as possible. It is necessary, for this purpose, that the piston rings be in close contact with both the cylinder wall and the flank of the groove machined into the piston. Contact with the cylinder wall is ensured by the spring action inherent to the ring itself, which expands the ring radially. Figure 7.41 shows the forces at a piston ring.

P/F-11C50 Cu2C5

HS150 Cu3C6

C70S6/ C70+

C38

42Cr

Al

TiAl4V4

cast

forged in open die

open die

forged and fracturable truck/car

BY

HT

cast

forged aircraft

Young modulus (GPa)

170

199

200

213

210

210

210

68.9

128

Fatigue strength (pull) (MPa)

200

320

390

300/365

430

420

480

50

225

Fatigue strength (push) (MPa)

200

330

395

300/365

430

420

480

50

309

Rp 0.2% yield strength (MPa)

410

550

700

550/650

750

550

>800

130

1000

Compressive yield strength (MPa)

—

620

—

550/650

700

620

850

150

—

Rm: tensile strength (MPa)

750

860

950

900/1050

950/1100

900

1050

200

1080

Conrod material density

7.2

7.6

7.8

7.85

7.85

7.85

7.85

2.71

4.51

Material Name

NCI

Process comment

36MnVS4

Figure 7.40 Properties of conrod materials (source: Federal-Mogul).

Internal Combustion Engine Handbook | 101

6606_Book.indb 101

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Gap width m Gap ends

Ring gap Gas pressure

uncompressed ring compressed ring

Gas pressure

Spring force

d Ring running surface

Friction force Gas pressure Inertial force Friction force

Piston movement direction

Figure 7.41 Forces at a piston ring.

Oil control rings are usually given further support with an additional spring. The gas pressure acting on the ring is significantly supported by both radial and axial contacts in the ring groove in the piston. Axial contact may alternate between the lower and the upper flanks of the groove because of the influences of gas, mass, and friction forces [7-20]. Trouble-free piston ring functioning depends on the partially, very dynamically changing thermal and mechanical loads generated by combustion, the engineering details, machining quality, and the choice of materials for the piston, piston rings, and cylinder. The quality of the rings themselves but also the precise matching of these components to each other has a decisive influence on their operating properties [7-20]. The number of rings per piston influences the friction inside the engine. The rings’ masses represent a part of the reciprocating mass forces. These reasons have driven the trend to fewer rings per piston. A three-ring combination of compression rings and oil control rings is standard. Two-ring arrangements to reduce friction losses are also found in massproduction models. Figure 7.42 shows the most important terms and concepts.

7.3.1 Designs

One may differentiate among the types of piston rings on the basis of their primary functions: •• sealing rings to seal the combustion chamber against the crankcase •• oil control rings to regulate the oil balance. 7.3.1.1 Compression Rings Among the types of compression rings available (Figure 7.43), one differentiates the following: Rectangular ring [Figure 7.43(a)] with its rectangular cross section: This piston ring is used for sealing purposes at normal operating conditions.

Ring back a

h Ring flanks

a = (radial) wall thickness h = (axial) ring height d = Nominal diameter Figure 7.42 Piston ring terms.

Bevel edge ring [Figure 7.43(b)] with a conic running surface that shortens the wear-in period: Because of its oilstripping effect, it also supports oil consumption control. Double trapezoid ring [Figure 7.43(c)]: The conical flanks of the ring significantly reduce “sticking” at the rings, because they are continuously freed of soot and combustion residues. It is used only in diesel engines. Single trapezoid ring [Figure 7.43(d)]: This ring has a sloped flank at the top. Similar to the double trapezoid ring, it reduces “sticking” and is mostly used in diesel engines. Ring with inner chamfer or inner shoulder, top [Figure 7.43(e)]: The interruption in the cross section by the inner chamfer or the inner shoulder of rectangular or bevel rings causes the ring to deform when it is installed, thus creating a concave shape. This forces the ring in all running phases to contact the cylinder wall only with the lower running surface edge and the lower flank of the groove only with the inner edge (a so-called positive twist). The conical running surface formed in this manner results in an improved oil-stripping effect. However, under gas pressure, the ring is pressed plane creating an additional dynamic stress during operation. Ring with inner chamfer or inner shoulder on the lower flank [Figure 7.43(f)] or the so-called negative torsion ring: This interruption in the cross section causes a negative twist after installation, that is, in the inversed direction of the positively twisting ring. To avoid a contact of the upper running edge with the cylinder wall, the conicity of the running surface must be designed larger than for the bevel edge ring with or without positive twist. L-shaped compression ring [Figure 7.43(g)]: This design is used primarily in small two-cycle SI engines as the so-called “head-land” ring; the end of the vertical arm of the “L” is flush with the piston head upper surface. Because of the gas

102 | Internal Combustion Engine Handbook

6606_Book.indb 102

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.3 Piston Rings

pressure effective behind the vertical arm of the “L,” this ring seals tightly even when in contact with the upper flank of the piston ring groove.

•• Oil control rings that practically are compression rings with particular oil-stripping effect as they are used in the second groove of SI and diesel engines [Figure 7.44(a)–(c)] •• One-part, self-expanding oil control rings for the lowest piston ring groove [Figure 7.44(d)–(f)]

a)

Rectangular ring

b)

Bevel edge ring

c)

Double trapezoid ring

d)

Single trapezoid ring

e)

Ring with inner chamfer or inner shoulder, top

f)

Bevel ring with inner chamfer or inner shoulder, top

g)

L-shaped compression ring

Figure 7.43 Compression rings.

7.3.1.2 Oil Control Rings Oil control rings are of particular significance in managing the engine’s oil supply [7-21] and consumption; they are subdivided into the following:

•• Multi-part, spring-expanded or spring-supported oil control rings, also for the lowest piston ring groove. Here, one differentiates between two-part [Figure 7.45(a)–(e)] and three-part systems [Figure 7.45(f)–(h)]. Shoulder and shoulder/bevel ring: Thanks to turned or relief groove in the area of the bottom ring running surface, the shoulder ring [Figure 7.44(a)] achieves an excellent oilstripping effect. To strengthen this effect, the running surface is additionally realized as a conic shape in the shoulder/bevel ring [Figure 7.44(b)]. Shoulder/bevel ring, closed at the gap: This special type of the shoulder/bevel ring [Figure 7.44(c)] is characterized by an improved gas seal because the shoulder is closed in the gap area and realized without relief groove. In some cases, it is inserted in the first groove. Oil slotted ring: Because of the reduced height of the running edges in the oil slotted ring [Figure 7.44(d)] compared to the overall height, a significantly higher surface pressure can be obtained with this self-expanding ring than with the rectangular ring. The flanks of the oil slotted ring are parallel with each other. Beveled-edge oil control ring: The beveled-edge oil control ring [Figure 7.44(e)] differs from the oil slotted ring by outsidechamfered running edges for an even higher surface pressure. Double-beveled oil control ring: In the double-beveled oil control ring [Figure 7.44(f)], the running edges are each chamfered at their flanks pointing to the combustion chamber. At the same surface pressure as the beveled-edge ring, a lower oil consumption can be achieved thanks to the improved oil-stripping effect. In addition to the surface pressure at the running surface, the capacity for filling of oil control rings is an important characteristic to ensure low oil consumption. The normal method used to combine both demands is the use of multi-part oil control rings with an additional spring which supports itself at the ends to press the optimized ring body against the cylinder wall. Springs supporting in the groove bottom are no longer used today because the rings would be forced to also transfer the lateral piston forces. Oil slotted, beveled-edge and double-beveled oil control rings with tubular spring: The ring types shown in Figure 7.45(a)–(c) feature a groove at the internal diameter to accept the tubular spring, in addition to the corresponding singlepart oil control rings. Beveled-edge ring with chromium-plated profile-ground running edges and tubular spring: The chromium-plated running surfaces ensure a high long-term stability. For this reason, this ring type [Figure 7.45(d)] is used mostly in diesel engines. Because of the profile grinding of the running edges, tight tolerances can be achieved at these important functional surfaces.

Internal Combustion Engine Handbook | 103

6606_Book.indb 103

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

a)

Shoulder ring

b)

Shoulder/bevel ring

c)

d)

e)

f)

a)

b)

Beveled-edge oil control ring with tubular spring

c)

Double-beveled oil control ring with tubular spring

d)

Beveled-edge ring with chromium-plated profile-ground running edges and tubular spring

e)

Nitrided profile steel ring

f)

VF-system

g)

MF-system

h)

SS50 system

Shoulder/bevel ring, closed at the gap

Oil slotted ring

Beveled-edge oil control ring

Oil slotted ring with tubular spring

Double-beveled oil control ring

Figure 7.44 Stripping or single-part oil-stripping rings.

Nitrided profile steel ring: This beveled-edge ring is made from a high-chromium alloyed steel and nitrided on all sides for wear protection [Figure 7.45(e)]. Three-part oil-stripping systems comprise two thin steel strip rings—also called rails or steel plates—and a distance spring which, on the one hand, keeps the rails in the desired axial distance to each other and, on the other hand, simultaneously presses them against the cylinder wall. For the spring types, three fundamental designs are established, which lend their names to these three-part systems. VF system: Figure 7.45(f) shows this three-part system that features a spring bent from a slotted steel strip into a U-shape with the opening pointing to the inner diameter. It is mostly made from simple C-steel or a nitridable CrNi steel. MF system: The spring consists of an axially crimped CrNi strip [Figure 7.45(g)]. Systems with untreated spring and chromium-plated rails and fully nitrided systems are used. SS50 system: Similar to the MF spring, the spring in the SS50 system is also manufactured from CrNi steel. Unlike the other version, the steel strip is radially crimped [Figure 7.45(h)].

Figure 7.45 Two- and three-part oil control rings.

7.3.2 Ring Sets

Piston ring designs are mostly determined by their functional demands which, in turn, depend on the technical and commercial marginal conditions of the SI engine passenger car, diesel passenger car, and diesel commercial vehicles sectors but also customer-specific application experience. Thus, the

104 | Internal Combustion Engine Handbook

6606_Book.indb 104

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.3 Piston Rings

specific requirements of the engine designs are always the defining factor in the optimization of each piston ring set. For this reason, Figure 7.46 and Figure 7.47 show only typical set samples for the individual market segments.

1. Groove

2. Groove

3. Groove

Rectangular ring, spherical running surface, material: Nitrided steel axial height: 1.0–1.2 mm Shoulder/bevel ring or bevel ring material: Gray-cast iron, uncoated running surface axial height: 1.2-1.75 mm

2. Groove

3. Groove

Double trapezoid ring, one-sided spherical running surface material: spheroid casting or steel, running surface coating from chromium-ceramic (CKS) or chromium-diamond (GDC) axial height: 2.5-4.0 mm

2. Groove

Bevel ring material: Gray-cast iron, chromium-plated running surface axial height: 2.0-3.0 mm

3. Groove

Oil control ring with tubular spring material: GG/GGG or steel profile, chromium-plated or nitrided running surface axial height: 3.0-4.0 mm

MF system untreated or nitrided spring, rails with chromium-coated running surface or nitrided surface axial height: 2.0 or 2.5 mm

Figure 7.47 (a) Ring set for passenger car diesel engines. (b) Ring set for utility vehicle diesel engines.

Alternatively: Two-part oil control ring with tubular spring material: Gray cast iron or steel profile, uncoated or nitrided running surface

For utility vehicle engines, the double trapezoid ring with adjusted axial height is standard but will be also realized in steel, unlike the passenger car model. Figure 7.47(b) displays a typical utility vehicle set.

Figure 7.46 Ring set for passenger car SI engines.

1. Groove

1. Groove

Rectangular or double trapezoid ring, one-sided spherical running surface material: spheroid casting, running surface coating from chromium-ceramic (CKS) or chromium-diamond (GDC) axial height: 1.75-3.5 mm

7.3.3 Characterizing Features 7.3.3.1 Tangential Force The tangential force Ft is that force which must be present at the ends of the ring, at the outside diameter, to compress the piston ring to the specified gap (Figure 7.48). This is the determinant factor for the contact pressure. The contact pressure influences the sealing function and is the force with which the piston ring presses against the cylinder wall.

Bevel ring or shoulder/bevel ring material: alloyed gray-cast iron, uncoated running surface axial height: 2.0-2.5 mm

Ft

Ft p = const. Ring gap

Oil control ring with tubular spring material: Gray-cast iron or steel profile, chromium-plated or nitrided running surface axial height: 2.0-3.0 mm

The ring set shown in Figure 7.46 is a typical design for passenger car SI engines. A typical combination for passenger car diesel engines is shown in Figure 7.47(a). At higher thermal loading, the ring of the first groove is a double trapezoid ring which otherwise exhibits identical characteristics.

Figure 7.48 Tangential force at a piston ring.

It is calculated as shown below, where p = contact pressure, d = nominal diameter, and h = ring height:

p=

2 ⋅ Ft ⎡ N/mm 2 ⎤⎦ d⋅h ⎣

(7.3)

Internal Combustion Engine Handbook | 105

6606_Book.indb 105

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.3.3.2 Radial Pressure Distribution Contact pressure can be set up to be constant around the circumference of the ring or to correspond to specified graduations in radial pressure, the so-called radial pressure distribution. Radial pressure distribution is important for the sealing function of the piston ring at the running surface of the cylinder wall. The further development of constant contact pressure resulted in inconstant radial pressure distributions as shown in Figure 7.49, for a focused effect on the functional behavior of the rings within the engine.

the ring external diameters is assumed which are offset in direction of ring gap/ring back and measured. 7.3.3.3 Installed Flexure Tension This is the flexure load to which the piston ring is subjected when installed in the cylinder. The maximum tension is found at the back of the ring and is calculated for the rectangular ring as follows:

db =

a ⋅E ⋅ 2 ⋅ k ⎡⎣ N/mm 2 ⎤⎦ d−a

(7.4)

and for an oil control ring as follows: a) Four-cycle characteristic (positive oval)



db =

xl ⋅E I +I ⋅ k ⋅ u s ⎡⎣ N/mm 2 ⎤⎦ Is d−a

(7.5)

where a = ring wall thickness ; d = nominal diameter; E = Young’s modulus for the ring material; k = piston ring parameter; xl = twice the distance of the center of gravity to the outside diameter; Iu = geometrical moment of inertia for the cross section without slotting; and Is = geometrical moment of inertia for the slotted oil control ring.

b) Constant pressure characteristics (circular)

c) Two-cylce characteristic (negative oval)

7.3.3.4 Opening Stress The greatest stressing of the ring occurs during assembly, as the ring must be stretched open sufficiently to enable the inner contour to slip over the outer diameter of the piston. The standard formula for the opening stress δ bü found in the literature makes the assumption that the closure stress and opening stress are equal. This assumption is only true in very limited marginal conditions and can lead to significant errors. Based on the mathematically exact, but complicated, approach to calculate the opening stress, more convenient formulas for the calculation of rectangular cross sections and those of oil control rings were derived. These extensively documented [7-20] formulas differentiate between slipping the ring with a purely tangential load and opening the ring by means of a sleeve. It must be noted that, as a rule, the purely tangential load leads to a maximum opening force in the back of the ring, while, when opening the ring by means of a sleeve, this load is mostly found in the 90°/270° area. 7.3.3.5 Piston Ring Parameter The piston ring parameter k characterizes the elastic property of the ring. For rings with a rectangular cross section, for instance, it is defined as



k = 3⋅

( d − a )2 ⋅ Ft h ⋅ a3

E



(7.6)

where tangential force is used Ft, or

Figure 7.49 Radial pressure distributions.

In the past, inconstant characteristics could be generated only by the additional requirement of a positive or negative ovality. As the dimension for the ovality, the difference in



k=

2 m ⋅ 3⋅p d − a

(7.7)

when using the gap width m (Figure 7.42). 7.3.3.6 Capacity to Fill the Space The capacity to fill the space is the capability of the piston ring to adapt even to cylinders that are out of round. Good

106 | Internal Combustion Engine Handbook

6606_Book.indb 106

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.3 Piston Rings

shape-filling capability ensures correct sealing against the gas and the lubricating oil. Taking ui as the radial, harmonic deformation of the cylinder of the ith order, the shape filling capacity QR at which the ring is just in contact with the cylinder wall, exerting radial pressure of p = 0, is calculated as follows: Qr =

ui k = 2 r ( i 2 − 1)

(7.8)

where r = (d − a)/2. Because with increasing ordinal i, the filling capacity decreases at nearly fourth power, cylinder warping of a higher power is extremely critical for the functioning of the piston rings. In compression rings, the gas pressure behind the ring increases the shape-filling capacity, while in oil control rings, the additional support by spring force acts so that the overall shape-filling capacity results in Qges = QR ⋅ ( 1 + x ) (7.9) pf pz with x = for compression rings and x = for oil control p p rings; pz = contact pressure by gas pressure; and pf = contact pressure by spring force. It must be noted that the simplified equations (7.8) and (7.9) enable only statements on the shape-filling capacity in the back of the ring, but not the local capacity at the ring circumference [7-22].

7.3.3.7 Ring Gap The ring gap is the space left between the ends of the ring after installation; this space is necessary to allow for thermal expansion in the piston ring (Figure 7.48). If the ring gap is too large, then gas loss (blow-by) will result; if it is too narrow, then ring expansion can exert pressure on the ends of the rings and cause ring failure. In this event, thermal expansion of the ring is obstructed by touching gap ends. The result may be ring break or seizing between ring and cylinder running surface because the contact pressure rises to impermissible levels. Ring gaps with straight end surfaces are normally used. Bevel joints and lap joints are not used for passenger car and utility vehicle engines and do not offer any advantages with regard to the tightness of the seal. Ring gaps with increased sealing quality (roll shaped or beveled) improve the sealing quality of the rings in comparison with the butt joint. These types are often used in hydraulic applications, but their testing in combustion engines leads to indifferent results

7.3.4 Manufacturing

Piston rings made of cast iron are manufactured in a single casting process as single, double, or multiple blanks, on mold plates following a mathematically determined model, and are cast in stack molding. Another manufacturing option is to make up cast bushings in stationary or centrifugal casting. Cold-drawn, profiled steel is preferred for manufacturing steel piston rings. Here, not only simple profiles may be

selected for the compression rings but also special profiles for oil control rings as well. 7.3.4.1 Shaping While conventional processes (face milling and lapping) are used to work the flanks of the rings, the outside contour, which determines piston ring characteristics, is shaped using the special processes of tandem turning (cast-iron rings) and winding steel rings. 7.3.4.1.1 Tandem turning Here the blank, the flanks of which have been ground, is worked simultaneously on the inside and outside using a copying lathe, ensuring uniform wall thickness all around the circumference of the ring. Once the section of the ring corresponding to the width of the gap has been removed, the ring exhibits the uncompressed shape that will develop the desired degree of radial pressure distribution once it has been inserted into the cylinder. The shape of the copying cam is determined mathematically, separately for each radial pressure distribution pattern. 7.3.4.1.2 Winding Winding is the process used for steel piston rings. The steel wire, having been drawn to the appropriate profile, is wound around a mandrel; The coil that is thus created is split lengthwise, separating the turns, and the rings that result are then mounted on a shaping mandrel and annealed to set the shape. The outside contour of the mandrel is calculated and designed for certain radial pressure characteristics. Profiling of the running surfaces of chamfered, shoulder and oil slotted rings, in particular, is carried out, depending on the ring design, on automatic lathes or profile grinding machines using special profile cutting tools before or after coating. 7.3.4.2 Wear-Protection Layers To diminish piston ring and cylinder wear, the ring running surfaces, in particular, are provided with wear-reducing protective layers [7-25]. The following types of protective finishes are used. •• Chrome plating: In the tribological system “piston ring/ cylinder wall,” electrochemically deposited hard chromium coatings on piston ring running surfaces are noted for their high wear resistance and the low cylinder wear they generate. Normal chromium plating is today used only on rings in the second groove and oil control rings. Special surface topologies to optimize the oil film treatment have been developed to protect from damage to the layer, such as burns and/or fatigue-related breakouts that mostly occur in the run-in phase. The special lap finish must be mentioned in this context [7-20]. Chromium ceramic coating (CKS): With demands on the load levels of modern internal combustion engines ever increasing, there is often a need to improve the thermal and/or mechanical load-carrying capacity of piston rings in the first groove and for oil control ring coatings beyond early life. The inclusion of ceramic particles (Al2O3) in an

Internal Combustion Engine Handbook | 107

6606_Book.indb 107

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

electrochemically generated hard chromium layer not only improves the scuff resistance for the entire service life of the layer but also provides resistance against scuffs, that is, against thermal overload (Figure 7.50). Chromium microdiamond coating (GDC): The CKS coating was further developed for even higher engine loads by embedding minute diamond particles in the hard chromium layer, instead of ceramic. The inherent coating wear is nearly halved and the scuff resistance further improved without significantly increasing cylinder wear [7-26]. Molybdenum coating: It is used above all because of its great resistance to scuffing. Molybdenum is applied to the piston ring running surface as a thermal spray layer, usually in a flame spatter process. The molybdenum layer’s great resistance to burns can be traced hypothetically to the material’s high melting point (approximately 2600°C) and its porous structure. •• Plasma spatter layers: Plasma spatter technology makes it possible to apply mixed metallic and/or metal-ceramic layers whose component materials exhibit particularly high melting points. The wear protection layers created in this way have even higher wear resistance than molybdenum layers and higher resistance to scuffing than chrome layers. •• High velocity oxy-fuel (HVOF) layers: HVOF coating, a high-velocity flame spray, is based on the plasma coating’s superior resistance to burns, further reducing inherent wear values and cylinder wear values. In HVOF coating, a supersonic flame is used to accelerate and heat the sprayed material. This creates a layer that is considerably denser and stronger than one applied with plasma spatter [7-23]. These fundamental advantages for the engine, when compared with plasma, can be achieved only when the coating materials are ideally matched to the properties of the process being used. The materials most frequently employed are metals with high carbide content. •• Nitriding and nitrocarburizing: Thermochemical treatment (diffusion) is used to introduce nitrogen and, in some cases, carbon into the surface of the piston rings (primarily in rings made of steel). This diffusion process creates extreme

surface hardness (approximately 1300 HV 0.025), which imparts high wear resistance to the layer. Layer hardness and thickness rise with the amount of alloying elements that form nitrides in the ring material (largely steel containing 13% or 18% chromium). In gasoline engines, this is used as an alternative to electroplated chromium layers and in part also to thermal spray layers, particularly at ring thicknesses of ≤1.2 mm. Additional advantages are dimensional trueness, which makes it possible to create sharp running edges at the piston ring, and coating on all the surfaces, providing additional protection against wear at the flanks. The burn resistance of these layers is similar to the chromium layers deposited with normal electroplating processes while that found in thermal spray layers is not reached. •• Physical vapor deposition (PVD) layers: Employing the modern technology used to vapor-deposit hard materials such as CrN gives wear protection layers that replicate exactly the contour of the surfaces. In this way, one can treat only the functional surfaces of the paired wearing materials, which may be advantageous. PVD layers are characterized in part by great wear resistance, high burn resistance, and low Top Dead Center (TDC) wear at cylinders in diesel engines. The layer thicknesses that can be created (50 µm) limit the life span expectations for truck applications, in particular. The expected improvement of the layer systems compared with the CrN systems mostly used today will result in expanded applications in the future. Additional applications are given because of the advantages relative to friction minimization; in this area diamond-like coating coatings will become more important in engines with Al cylinder tracks, in particular [7-27]. 7.3.4.3 Surface Treatments The surface treatments listed below are employed with piston rings primarily to protect against corrosion during storage, to cover up minor surface defects, to improve break-in properties, secondarily to reduce wear at the running surfaces and flanks, and to not increase burn resistance during the run-in period. •• Phosphating (zinc-phosphate and/or manganese-phosphate layers): The surface of the piston ring is transformed into

Network of fissures

Time

Fissure width

Layer thickness fissure depth

+ – Flow density

Figure 7.50 Schematic representation of chromium layers with solid particle inclusion.

108 | Internal Combustion Engine Handbook

6606_Book.indb 108

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.3 Piston Rings

phosphate crystals with chemical treatment. This phosphate layer is softer than the substrate material and thus wears away more easily, which accelerates ring wear-in. This layer’s thicknesses are between 2 and 5 µm. •• Tin and copper plating: Both these metallic layers are applied by electroplating. Because of their softness, they act somewhat like lubricants. Standard layer thicknesses are also between 2 and 5 µm. •• Black oxiding: Black oxiding is used mainly for coating the sides of rails made of carbon steel. These very thin iron oxide layers (thickness < 1 µm) provide a certain amount of corrosion protection. •• CPS and CPG: CPS (for nitrided steel rings) and CPG (for nitrided cast rings) are processes of chemical passivation which reduce the risk of the so-called micro-welding because of a focused change in the surface morphology [7-24]. Corrosion resistance and geometrical stability are also positively affected. 7.3.4.4 Materials for Piston Rings Determinants for the selection of the piston ring materials include demands for good running properties in both normal and emergency situations (wear behavior), for good elastic properties, for good heat conductivity and thermal expansion, and high corrosion resistance. Great strength is required whenever extreme conditions such as high engine speeds or high gradients in the combustion pressure are present. The following materials may be used [7-25]: •• Cast iron with flaked graphite, non-hardened: This is the “standard material” for piston rings, with good break-in and emergency running properties and satisfactory wear properties. The values for resistance to flexure with 350 N/ mm2 at minimum are relatively low. The standard material is used today only for the ring in the second piston ring groove and for oil control rings. •• Cast iron with flaked graphite, alloyed, hardened: The low strength properties of the “standard material” are improved by hardening. Resistance to flexures is at least 650 N/mm 2, and hardness is increased. This material is also for rings in the second piston groove. •• Cast iron with spheroidal graphite (nodular cast iron), alloyed, hardened. This type of cast iron is distinguished particularly by its great resistance to flexure, of at least 1300 N/mm2. Because of its high resistance to flexure, nodular cast iron is given preference for rings mounted in the first piston ring groove.

by the inward bending stresses imparted when the ring is compressed so that it can enter the cylinder. Dynamic loading also occurs, namely, an axial motion of the piston ring caused by the interactions of gas, mass, and friction forces. In extreme situations, uncontrolled axial and radial movements of the ring are caused which can lead, in SI engines in particular, to significantly reduced sealing capacity and thus to high blow-by losses at low mean pressure and high speeds. These extreme ring movements may also cause ring failure in some cases, as do extreme pressure rise rates at knocking combustion in an SI engine and pinging combustion in a diesel engine. Extraordinarily high loading on the ring can arise from soot collecting in the piston ring groove, which can cause sticking. Additional ring damage includes burn traces and seizure. The service life of the seal made at the piston rings is determined to a large degree by the amount of wear. This includes radial wear (wear on the running surface), axial wear (wear at the flanks, “microwelding,” and piston groove wear), and secondary wear at oil control rings (wear between the ring and the tubular spring and between the rails and the spacer spring). The tribologic system surrounding the seal created by the piston ring is extremely complex because most normal types of wears—abrasive, adhesive, and corrosive—occur to a greater or lesser extent and effect. The piston group accounts for approximately 40% of all the mechanical losses of an engine. The piston rings cause a bit more than half of this loss. The essential factors that influence piston ring friction include the surface pressure, ring thickness (width of the running surface), the rail height in oil control rings, the shape (crowning; various designs, see Figure 7.51) of the contact surface, the coefficient of friction for the running surface layer (only in mixed friction areas at TDC and BDC, where the piston speed is very slow), and the number of rings per piston required for sufficient sealing function. Measures taken to reduce piston ring friction must not interfere with ring functioning. The sealing effects of the ring set for both gases and the lubricating oil have to be maintained undiminished.

Ring with spherical running surface

a)

Ring with asymmetrically spherical running surface

b)

•• Steel: Because of its great breaking strength, steel is used, for example, at low ring heights (≤1.2 mm) for gasoline engines and in diesel engines with steep rates of pressure rise. Steel is also used for the rails and spacer springs in oil control rings, as well as in profiled oil control rings.

7.3.5 Loading, Damage, Wear, and Friction

The piston rings are loaded by outward stresses when they are stretched to pass over the cylinder and, when installed,

Ring with optimized spherical running surface

c)

Figure 7.51 Contact surface shapes.

Internal Combustion Engine Handbook | 109

6606_Book.indb 109

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.4 Engine Block The engine block is the component that encloses the cylinders, the cooling jacket, and the engine block shell.

7.4.1 Tasks and Functions

The primary functions that the engine block fulfills are •• Absorbing the gas and mass forces in the crankshaft bearings and at the cylinder head bolts. •• Accepting the energy conversion assembly, comprising the pistons, conrods, crankshaft, and flywheel. •• Accepting and connecting the cylinders or, in the case of multi-section engine blocks, connections to the individual cylinders or to the cylinder bank block or blocks. •• Carrying the crankshaft from, possibly, an intermediate shaft for the control drive and from one or two balance shafts for mass balancing. •• Accepting channels to convey operating media, primarily lubricants and coolants. Lubricants must be delivered to crankshaft and connecting rod bearings, possibly piston spray nozzles for piston cooling, any installed hydraulic chain tensioners, and the components arranged in the cylinder head. These are camshaft(s), push rods, rocker arms or cam followers, any installed hydraulic elements for automatic compensation of the valve play, and any installed adjustment mechanisms for control time adjustment. The lubricant return from the cylinder head(s) is usually designed as channels also arranged in the engine block. •• In liquid-cooled engines, the engine block also comprises the so-called water jacket around the cylinders and, possibly, other channels containing further coolants. Frequently, this is also the mounting point for the coolant pump. •• Integrating a system for crankcase venting. •• Connecting to the transmission and to the valve actuators (with cover) and carrying and guiding power transmission elements such as chains. •• Connecting to and mounting position for various auxiliary assemblies, such as the engine mounts, coolant preheating components, oil/water heat exchanger, oil filter, oil separator for crankcase venting, and a variety of sensors for oil pressure and temperature, crankshaft speed, knock detection, and so on. •• Isolating the crankcase from the outside world with the oil pan and—by way of radial shaft seals—at the point where the crankshaft passes through the engine block. Because of the variety of functions to be carried out, the engine block is subjected to different types of loads that are superimposed one upon another. It is exposed to tensile and compression loading, bending, and torsion as a result of mass and gas forces. Taken individually, these are •• ignition gas forces, which have to be absorbed by the cylinder head bolts and the crankshaft bearings •• internal mass moments (flexural moments), resulting from rotating and reciprocating mass forces

•• internal torsion moments (tipping moments) among individual cylinders •• crankshaft torque and the resulting reactive forces in the engine mounts •• free mass forces and moments, resulting from reciprocating mass forces, which have to be borne by the engine mounts. Working processes and operating limits determine the maximum forces occurring. For example, diesel engine usually requires larger dimensioned engine blocks than SI engines, because of their higher peak and mean pressures. The level of inertial forces occurring is determined by the maximum speed and the design of the crankshaft drive. The trend to charging diesel and SI engines and to downsizing with the same output as smaller working volumes increases the forces to be absorbed by the engine block. The effect of forces and the resulting moments, both inside the engine block and outside (engine mounts, mechanical vibrations, and noise emission), depend on the engineering design of the engine. The major parameters in the engine design that have effects on engine block loading are the number and arrangement of the cylinders, the arrangement of the crankshaft throws, and the ignition sequence. The loads occurring in the engine block influence the type of engine block selected and its specific design in view of achieving sufficient strength, minimum deformations, economical manufacturing, recyclability, noise emissions, engine block weight, and, with it, the total engine weight. The strength of the engine block is determined by the material used, by the choice of heat treatment (which depends on the material and the casting process), and by the engineering design (characterized by the type of engine block, ribs or fins, wall thickness, etc.). Common engine block materials, in comparison with vermicular graphite cast iron, and the most important material properties are shown in Figure 7.52. Engine blocks are characterized by the following major dimensions, which depend on the engine configuration, such as inline, V-block, or boxer (pancake) engine (Figure 7.53): •• length, measured from the front edge of the engine block to the transmission flange •• width, as maximum overall width •• height, measured from the center of the crankshaft, along the axis of a cylinder, to the top plate plane •• cylinder bore, expressed as the nominal inside diameter of the cylinders •• cylinder spacing, given as the distance between the centers of two adjacent cylinders •• cylinder offset in V, W, and boxer engines, specified as the distance between the centers of two cylinders located opposite each other in adjacent banks of cylinders •• cylinder length, measured from the top plate to the lower end of the cylinder. Dimensions and drilling pattern for cylinder head bolts: depending on their design, for example, four or six per cylinder.

110 | Internal Combustion Engine Handbook

6606_Book.indb 110

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.4 Engine Block

Materials

(standard materials for crankcases) Material group: Material: Remark:

Aluminum AISI6Cu4 Subeutectic

Material state:

Casting state

Casting technique: Elongation limit R p02(N/mm2) Tensile strength R m(N/mm2 ) Ductile yield A6 (%): Brinell hardness HB: Fatigue strength at reversed bending stresses (N/mm2) E-module (kN/mm2): Thermal expansion coeff. (20°–200° C) (10–4/K): Thermal conductivity(W/mK): Density (kg/dm3:)

AISi17Cu4Mg

Supereutectic

Supereutectic

Heat-treated Casting state

Sand and Sand and permanent Die casting permanent mold casting mold casting 90–100 150–170 1 60–75 60–80

140 240 1 80 70–90

190–320 220–360 0,5 90–150 90–125

73–76 21–22,5

75 22,5

100–110 2,75

100–110 2,75

AISi9Cu3

Subeutectic

Iron G3L-240 G3L-300 G3V Cast iron Cast iron Vermiwith flaked with flaked cular graphite graphite graphite

Casting state

Sand and Die casting permanent Die casting mold casting 150–210 260–300 0,3 25 70–95

90–180 150–170 1 60–75 60–95

150 220 1 80 70–90

165–228 250 0,8–0,3 180–250 87,5–125

195–260 300 0,8–0,3 200–275 105–150

83–87 18–19,5

83–87 18–19,5

74–78 21–22,5

75 21

103–118 11,7

108–137 130–160 11,7 11–14

117–134 2,75

117–150 2,75

110–120 2,75

110–120 2,75

48,5 7,25

47,5 7,25

240–300 300–500 2–6 160–280 160–210

42–44 7,0–7,7

Sources: Kolbenschmidt AG, Neckarsulm, Handbuch Aluminium – Gussteile, Heft 18 DIN EN 1706, Aluminium und Aluminiumlegierungen, Gussstücke, chemische Zusammensetzung und mechanische Eigenschaften DIN EN 1591, Gusseisen mit Lamellengraphit Porsche Technische Lieferungsbedingungen 2002 Vermiculargraphitguss (GGV) – Ein neues Material für den Verbrennungsmotor, Aachener Kolloquium Fahrzeug- und Motorentechnik 95, Prof. Dr. techn. F. Indra, Dipl. Ing. M. Tholl, Adam Opel AG, Rüsselsheim

Figure 7.52 Materials for engine blocks.

(f) height of the lower engine block section in two-section engine blocks. 6

3

Figure 7.53 shows the most important dimensions. The conrod executes a swinging motion with each revolution of the crankshaft. The path that it follows, determined by the outside contour of the conrod and the cranking radius, has a shape similar to the body of a violin (Figure 7.54). It is

8 1 2 3 4 5 6 7 8

5 1

Length Width Height Cylinder bore Cylinder spacing Cylinder length Dimensions drilling pattern Dimension from crankshaft center to flange to oil pan

4 Conrod pattern

7 2

Figure 7.53 Major dimensions of the engine block.

These dimensions from the center of the crankshaft to the oil pan flange are defined: (d) equal to zero where the oil pan separation plane is level with the center of the crankshaft (e) height of the deep skirts where the engine block side walls extend downward

Figure 7.54 Conrod pattern.

Internal Combustion Engine Handbook | 111

6606_Book.indb 111

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

necessary, when laying out an engine block, to ensure that there is sufficient clearance for this outline. The most critical close clearances between the engine block and the envelope for the conrod are normally

top plate. This has positive effects on top plate deformation, cylinder warping, and acoustic properties.

•• lower surface of the cylinder and in V-block, W-block, and boxer engines that of the opposite cylinder, too •• engine block sidewalls with channels located next to the conrod for oil return or for crankcase venting. The clearance is, as a rule, between 3.5 and 4.5 mm and is determined after having taken into account all the tolerances for the components involved, and to include the casting tolerances for the engine block itself.

7.4.2 Engine Block Design 7.4.2.1 Types of Engine Blocks The types of engine blocks can be classified according to the engineering design in the areas at the •• top plate •• main bearing pedestals •• cylinders. Because a separate section is devoted to the cylinders, they are not dealt with here. 7.4.2.1.1 Top plate A basic engineering feature, one that limits the selection of the casting process, is the engine block top plate. Here, one differentiates between closed-deck and open-deck designs. a)  Closed deck: In this version, the top of the engine block is largely closed in the area around the cylinders. In the top plate, there are always, depending upon the specifics of the design, openings for the cylinders, openings for the tapped holes for the cylinder head bolts, and bores and channels for oil feed and return (Figure 7.55). Here, except the cylinders, the top plate is penetrated essentially only by the smaller openings of appropriate cross sections to allow for coolant passage. These openings join the water jacket surrounding the cylinders (with the water jacket inside the cylinder head) through specified channel cross sections in the head gasket and at openings in the cylinder head combustion chamber plate. This design suffers disadvantages regarding cylinder cooling in the TDC area. Producing the engine block water jacket requires a sand core in the closed-deck version because the water jacket, in the upper area of the engine block, is largely sealed off by the top plate. Consequently, the water jacket cannot be created as a feature of the external casting mold for the engine block upper section; a core has to be inserted inside the casting mold. These bearing points are generally found in the finished engine block as casting “eyes” in the engine block side walls. The openings for the core inserts are closed off with sheet metal plugs. Once the engine is assembled, core insertion points such as this are an indication that it is a closed-deck engine block. The advantage of the closed-deck design in comparison with the open-deck version is the greater stiffness of the

Figure 7.55 Closed-deck design.

Selecting an engine block with closed-deck design does, however, limit the casting processes that can be used. The sand core required for the water jacket makes it possible to fabricate the closed-deck type only in sand casting and die-casting processes. The lost-foam process applicable for creating the water jacket core is used only in rare cases. Engine blocks made of gray cast iron, made in a sand casting process, are almost exclusively of closed-deck design. Engine blocks made of aluminum–silicon alloys in a closed-deck design are manufactured in mass production primarily as die castings, as low-pressure castings, and, more recently, in an automated sand casting process. b)  Open deck: In the open-deck version, the water jacket surrounding the cylinders is open at the top as shown in Figure 7.56. From the casting technology viewpoint, this means that no sand core and, thus, no core inserts are required to form the water jacket. The casting core for the water jacket requires no undercuts and may be made up as a steel mold. Compared with the closed-deck design, the water jacket open at the top enables better cooling of the cylinder’s hot upper section. There is less stiffness in the top plate of the open-deck design than that of the closed-deck version. A metallic head gasket is used to compensate for the increase in negative influence on the top plate by deformation resistance and cylinder warping. In comparison with the conventional soft-material cylinder head gasket, the metallic head gasket permits a lower preload value for the head bolts, thus reducing top plate deformation and cylinder warping.

112 | Internal Combustion Engine Handbook

6606_Book.indb 112

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.4 Engine Block

Figure 7.56 Open-deck design.

Manufacturing open-deck engine blocks enables the use of essentially all types of casting processes. The open-deck design made it possible to manufacture engine blocks from an aluminum–silicon alloy using the economical die-casting process. Over and above this, it enables the use of special techniques for the cylinders and cylinder sleeves. 7.4.2.1.2 Main bearing pedestal area The main bearing pedestal area in engine blocks is the area around the crankshaft bearings. The engineering design for this area is of particular importance because, among other things, the forces acting on the crankshaft bearings have to be taken up here. Options for further structuring the design of the engine block include selecting the location for the separation plane between the engine block and the oil pan, and the engineering of the main bearing caps. One uses this separation plane to distinguish between an oil pan with the flange level at the center of the crankshaft and one that is below the center of the crankshaft. In designing the main bearing caps, one distinguishes among individual main bearing caps, their integration into a longitudinal frame unit, and integration into the engine block lower section. Main bearing cap: The main bearing caps represent the lower boundary of the main bearing pedestals; the caps are affixed and bolted to the main bearing pedestals. The main bearing caps and the main bearing pedestals have essentially the same function, that is, absorbing the forces and torques imposed

upon the crankshaft, accepting the corresponding bearings including the thrust bearing (collar bearing or thrust washers), as well as accepting a radial shaft sealing ring at the transmission output end, at the final main bearing, to seal the rear end of the crankshaft. The main bearing caps and main bearing pedestals in the engine block are machined together and are joined during post-machining assembly procedures. The normal methods used for fixing these items are surfaces broached at the side in the main bearing pedestals or bores for guide bushings. Main bearing caps are manufactured exclusively as gray castings and are combined with engine blocks made both of gray cast iron and of aluminum alloys. Working the aluminum main bearing pedestal and the gray cast bearing cap simultaneously is not without its difficulties because of the differences in ideal cutting speeds (specific to the materials). This is the procedure used in mass production today. The combination of an aluminum main bearing pedestal and a cast iron main bearing cap has advantages resulting from the gray cast iron: the low coefficient of thermal expansion in the main bearing cap made of gray cast iron limits the amount of play in the crankshaft bearings that develops during operation. This reduces the amount of oil that passes through the main crankshaft bearings. Reduced main bearing play and greater stiffness in the cast iron bearing cap (Young’s modulus for gray cast iron is higher than that for aluminum) reduce noise generation and emissions in the area around the main bearing pedestals. The version previously most widely used in mass production was the engine block made of gray casting with main bearing caps of the same material. The engine blocks were engineered either with the oil pan flange level with the center of the crankshaft or as an engine block with side walls or skirts that extend downward. In V-block engines, one most commonly found an aluminum engine block combined with individual cast iron main bearing caps. Since the early/mid-1990s, new engine designs featured and feature the engine block increasingly as all-aluminum versions in mass production. Main bearing pedestal: The upper section of the crankshaftbearing surface in the engine block is referred to as the main bearing pedestal. Regardless of the engineering design of an engine block in the area around the crankshaft bearings, the main bearing pedestals are always a part of the casting for the engine block or for the upper section of the engine block (Figure 7.57). The number of main bearing pedestals for an engine block depends on the engine type and, in particular, on the number of cylinders and their arrangement. Today, for reasons associated with vibration phenomena, engine blocks are almost always made with a full set of bearings for the crankshaft. Crankshafts such as these have a main bearing journal next to each crankshaft throw. A four-cylinder inline engine thus has five main bearing pedestals, six-cylinder inline and boxer engines have seven main bearings, V-6 engines have four main bearings, V-8 engines have five main bearings, and so on.

Internal Combustion Engine Handbook | 113

6606_Book.indb 113

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

block lower section form the outer limits of the crankcase; the lower plane forms the flange to the oil pan. An engine block lower section offers essentially the same engineering design options as those described for the longitudinal frame concept. Because engine block lower sections are mass produced almost exclusively from aluminum alloys and in a die-casting process, additional functions can be integrated into it: Breakthrough Oil feed bore

•• oil removal, that is, radial stripping of the motor oil around the envelope for the crankshaft counterweights and the conrods

Main bearing pedestal

•• parts of the motor oil circuit such as the oil intake channel between the oil pump and the oil sump, the oil channel between the oil filter head and the oil pump, the oil filter head itself, the oil return channels, the main oil channel and oil channels to the individual main bearing points, partial integration of the oil pump housing

Main bearing cap

•• accepting shaft seal rings to seal the crankshaft. Engine block lower sections are used in mass-production all-aluminum engines and in racing engines.

Figure 7.57 Main bearing pedestal/main bearing cap.

The major functions of the main bearing pedestals are •• accepting axial and radial forces and moments impinging upon the crankshaft bearing system •• accepting the upper sliding bearing shell for the crankshaft radial bearings along with accepting the collar bearings or thrust washers in a main bearing pedestal, the so-called thrust bearing, for axial control of the crankshaft

Longitudinal frame concept: Similar to the situation where an engine block lower section is used, in the longitudinal frame concept, the individual main bearing caps are consolidated into a single component (Figure 7.58). In contrast to the engine block lower section, the longitudinal frame has no flange plane interfacing the oil pan.

•• accepting the threads, fixing holes, or fixing bushings used in attaching and fixing main bearing caps or longitudinal frames or the lower section of the engine block •• accepting oil feed bores and oil grooves used to supply the crankshaft main bearings with oil •• depending on the engine design, accepting the radial shaftsealing ring in the last main bearing pedestal, used to seal the rear end of the crankshaft. The main bearing pedestals often exhibit passageways to equalize pressures in the individual chambers in the crankcase area and, thus, reduce losses because of internal engine friction. Vertical holes or channels for oil return from the cylinder head or for crankcase venting through the main bearing pedestals are commonly found. These many functions require great care in the engineering and design of the main bearing pedestals and the components that interface with them—the main bearing caps or longitudinal frame or lower engine block section. Engineering for these assemblies is carried out today almost exclusively with the engineering aids now available, such as FEM calculations. Engine block lower section: Just as in the longitudinal frame design, the individual main bearing caps in the engine block lower section are combined into a single component. In contrast to the longitudinal frame, the engine block lower section does not lie within the engine. Instead, the sidewalls of the engine

Figure 7.58 Longitudinal frame concept.

Rather, the longitudinal frame lies inside the engine and, thus, is enclosed by the oil pan in the version where the oil pan flange is centered on the crankshaft or by the deep sidewalls

114 | Internal Combustion Engine Handbook

6606_Book.indb 114

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.4 Engine Block

of engine blocks that incorporate the same. The advantages of a longitudinal frame are •• greater stiffness in comparison with individual main bearing caps and thus better acoustic properties, and easier and faster installation •• almost the same degree of engineering freedom as the engine block lower section with regard to integrating functions •• more economical and lighter in weight than an engine block lower section. Longitudinal frames made of aluminum alloys can be manufactured using die casting. This also allows the integration of cast oil grooves to supply oil to the main bearings. In the areas around the individual bearing points, insets made of cast iron with spherical graphite (e.g., GJS 600) can be cast in place. This yields the same advantages (reducing the bearing play at the crankshaft, increasing stiffness of the longitudinal frame, and reducing noise radiation in the main bearing pedestal area) as for the combination of aluminum engine block and main bearing cap made of cast iron. In existing engine block designs with individual cast iron main bearing caps, these may be replaced with a longitudinal frame construction to increase stiffness and/or to improve acoustic properties without having to completely re-engineer the block. Also possible are combined solutions in which individual bearing caps are joined by bolting them to a separate cast part shaped like a ladder. The engine blocks are engineered either with the oil pan flange level with the center of the crankshaft or as an engine block with side walls or skirts that extend downward. Oil pan flange level with the center of the crank: A further design characteristic is the position of the separation plane between the engine block and the oil pan being level with the center of the crankshaft (Figure 7.59). In this version, the upper halves of the crankshaft bearing seats are integrated into the casting for the engine block as main bearing pedestals. The lower halves of the crankshaft bearing seats are engineered either as individual main bearing caps, as a longitudinal frame or as engine block lower section. The seal between the engine block and the oil pan is between the two flanges congruent with the separation plane. The seal for the crankshaft at the front and rear ends depends on the particular engine design. The front end of the crankshaft may be sealed by a radial shaft seal in the oil pump housing or in the front-end cover. The rear end of the crankshaft may be sealed by a radial shaft seal in the last main bearing pedestal or in a separate cover. Cast iron engine blocks, in which the separation plane for the oil pan is level with the center of the crankshaft and with individual main bearing caps, were often used for smalldisplacement (to approximately 1.8 L) four-cylinder, inline engines and in some V6 and V8 engines. The advantages of this design are found in favorable manufacturing costs. The disadvantages of this design, in comparison with engine blocks with deep skirts or a lower engine block section, are less stiffness and less favorable acoustic properties.

Center crankshaft

Main bearing cap Oil pan

Figure 7.59 Oil pan flange level with the center of the crankshaft.

Oil pan flange below the center of the crankshaft: With the separation plane between the crankshaft and the oil pan in this location, one differentiates between two types of engine block construction. a.  Design with upper engine block section and lower engine block section [Figure 7.60(a)]: In this version, the main bearing caps are joined to form a bearing case, the so-called engine block lower section. The separation plane between the upper and lower sections of the engine block is level with the center of the crankshaft. This means that here the component designated as the engine block upper section corresponds to the engine block of the type where the oil pan flange is level with the center of the crankshaft. The lower face of the lower engine block section forms the flange surface with which the oil pan mates. Depending on the engine design, the crankshaft is sealed at the rear end (toward the transmission) by a radial shaft seal in the last main bearing pedestal and at the front end with another radial shaft seal (located in the oil pump housing or front-end cover, for example). The advantages of this concept are great stiffness, good acoustic properties, and the engineering design options available for the lower engine block section as made clear at the description of the lower engine block section and longitudinal frame design (e.g., casting in place for inserts made of cast iron with spherical graphite in the area of the individual bearing points for lower engine block sections made of aluminum alloys and manufactured in a die-casting process). The disadvantages are higher manufacturing costs and, in some cases, slightly greater weight than if individual main bearing caps are used. This concept is built in mass production with the upper and lower engine block sections made of aluminum alloys. Because racing engines are often integrated into the frame as a load-bearing component in the overall concept for the

Internal Combustion Engine Handbook | 115

6606_Book.indb 115

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Top part Center crankshaft

Center crankshaft

Main bearing cap

Bottom part

Crankcase

Oil pan

Oil pan

 Figure 7.60  (a) Version with upper and lower engine block sections. (b) Engine block with side walls extended downward.

vehicle, racing engine blocks (because of the high degree of stiffness required) are designed almost exclusively using this engineering principle. b.  Engine block with side walls extended downward [Figure 7.60(b)]. In this version, the outside walls of the engine block are extended to below the middle of the crankshaft and end at the flange interfacing the oil pan. The separation of the main bearing pedestals continues to be centered on the crankshaft for reasons associated with the machining. Designs that have been realized exhibit both individual main bearing caps and main bearing caps that have been combined to form a longitudinal frame. The benefits of using a longitudinal frame are stiffness and acoustic properties similar to the concept with separate upper and lower engine block sections. The manufacturing costs for this method may be slightly lower, depending on the manufacturing volume.

7.4.3 Optimizing Acoustic Properties

Complying with noise emission regulations and satisfying owners’ expectations for quiet operation are key areas of attention in acoustic development for drive components. The acoustic properties and smooth running of an internal combustion engine depend on many parameters and are predetermined to a great degree by the selection of the design for the engine and engine block. Optimizing the acoustic properties for the engine block structure, such as increasing stiffness at the engine block sidewalls, taking into account the many and varied functional

requirements, is an important development target. This is achieved by low noise radiation, avoiding natural frequencies, and damping resonance-inducing vibrations. The loading on the engine block resulting from the nonuniform progression of torques in the crankshaft because of the free mass forces and moments causes mechanical vibration. Their exciter frequency is in a specific relationship to the rotational speed of the crankshaft, according to the orders of excitation for the free effects of ignition gas and mass. Mechanical vibrations are caused by low exciter orders, are at a low frequency, and are found primarily in the area of the main bearing pedestals and the crankcase. High-frequency vibrations in the engine block walls are induced by the combustion process itself, and are in part caused by pulse-like power transmissions in the valve actuators and by forces induced at the pistons. The high frequencies are in the audible spectrum and are referred to as acoustic vibrations. A part of the high-frequency acoustic vibrations is radiated from the sidewalls of the engine block. Low- and high-frequency vibrations exert their effects through the interface of the engine block with the engine mounts in the vehicle. Depending on the type of engine mount used, vibrations and structural noise may be transferred to the vehicle. To be taken into account in the acoustic optimization of an engine are the following: •• the above-mentioned causes for initiation of structural noise •• the structural noise propagation paths in the cylinder head, cylinders, pistons, wristpins, conrods, and crankshaft

116 | Internal Combustion Engine Handbook

6606_Book.indb 116

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.4 Engine Block

•• the design of the engine mounts and their connection to the engine block or to other engine and drive train components •• the structure of the engine block in conjunction with the engine block engineering concept selected. When optimizing the acoustic and vibration-associated properties, the interaction with the flanged transmission must be taken into account, as it affects not only the vibration of the entire engine-transmission assembly but also their transmission to the vehicle structure. Modern engine block development is undertaken in a closed Computer Aided Engineering (CAE) process chain. The 3-D Computer Aided Drawing (CAD) depiction and networking for the housing structure form the foundation for FEM calculations of strength, stiffness, and dynamic and acoustic properties. An experimental model analysis at the finished engine block provides additional information on the forms of its own vibrations. Both experience and the engineering calculation and analysis options available today support the basic claim that noiseoptimized engine block design requires the stiffest possible engine block and the stiffest possible combination of engine and transmission. This is achieved by measures that are independent of the selected engine block design and by exploiting advantages specific to a particular design, such as the following: •• Manufacturing engine block surface structures with reinforced areas and ribs or fins to reduce airborne noise propagation. •• Stiff top plate and a force engagement point for the head bolts that are well below the top surface of the top plate. They minimize deformations at the sealing surfaces and cylinder warping. The latter is a prerequisite for low piston play and, thus, low piston noise. •• Stiffness at the crankshaft’s main bearing pedestal configuration, which permits only slight play at the bearing. •• Stiff flanges interfacing with the oil pan and the transmission as a prerequisite for a stiff engine and transmission assembly. The various engine block designs have differing specific acoustic advantages: •• The closed-deck design has a stiff top surface with benefits, in comparison with the open-deck design, with regard to deformation at the sealing surfaces and cylinders. •• A design comprising upper and lower engine block sections gives a stiff engine and transmission group in comparison with an engine block with side walls extended below the center of the crankshaft in combination with individual main bearing caps. In the latter design, stiffness is increased by joining the individual main bearing caps to form a longitudinal frame. •• In solid aluminum engine blocks, composed of upper and lower sections, gray cast iron components that are cast into the engine block at the main bearing points reduce thermal expansion and, in turn, bearing play.

•• Using a cast aluminum oil pan with a flange interfacing with the transmission provides a stiff engine and transmission group.

7.4.4 Minimizing Engine Block Mass

The important objectives in engine development are the reduction of pollutants, the lowering of fuel consumption, and an improvement in vehicle performance. This target requires, in addition to other measures, consistent implementation of lightweight engineering techniques for all vehicle components. Reducing the engine block weight is one contribution to reducing weight for the entire drive train. Depending on the engine’s size, design, combustion principle, and engine block design, the engine block accounts for between 25% and 33% of the overall engine weight (as per DIN 70020 A). Reducing the engine block weight thus makes a vital contribution to reducing overall vehicle engine weight. The measures undertaken to reduce engine block weight can be subdivided into weight reductions attained by optimizing the structures and weight reductions specific to the materials. The trend for weight reduction is opposed by the trend for higher loads on the engine block because of downsizing with increased peak and medium pressure and (high) charging. 7.4.4.1 Reducing Weight by Optimizing the Structure The design of the engine block has a critical influence on total engine block weight. The engineering and calculation methods (such as CAD and FEM) that are commonplace today enable more closely targeted optimization of the design needs, along with loading and functional needs, than could be achieved in the past. This means that the wall cross-sections required to carry out important functions such as the exact position, number, and geometry of ribs (which increase stiffness and improve acoustic properties) can be designed using minimal amounts of material. Cylinders that are cast together and the integration of many functions into the engine block also contribute to reducing overall engine weight. 7.4.4.2 Weight Reductions Through Material Selection Until the early/mid-1990s, most engine blocks in mass production were gray castings. The necessity to reduce weight resulted in using aluminum silicon alloys more frequently for massproduced engine blocks for SI engines and even diesel engines. The weight reduction potential in engine blocks made from cast iron is lower than the weight reduction achieved by using Al–Si alloys. In cast iron engine blocks, a weight-saving potential of approximately 30% is available by optimizing the structure and using thin-wall casting and vermicular graphite cast iron (GJV). The advantage of GJV compared to cast iron with flaked graphite (GJL) is the Young’s modulus, while its disadvantage is higher material costs. In comparison with gray cast iron engine blocks, the weight of Al–Si alloy engine blocks may be reduced by 40%–60%, depending on the size of the engine. This weight reduction is less than the ratio of the specific weights of Al–Si alloy and cast iron, because the different

Internal Combustion Engine Handbook | 117

6606_Book.indb 117

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

material properties, fatigue strength, and Young’s modulus must be considered during the design. Figure 7.61 shows the data for some materials used for engine blocks. 0.2 % Offset Limit (N/ mm2)

Density (g/cm3)

Young’s Modulus (kN/mm2)

Die-cast magnesium alloy

140–160

1.8

 45

Die-cast Al–Si alloy

140–210

2.75

 74–78

70–90

7.2–7.7

100–160

85–210

Material

Gray cast iron

Fatigue strength (N/mm2)

Figure 7.61 Materials for engine blocks.

Magnesium exhibits an even lower density than aluminum and becomes interesting for this reason. The low specific weight supports using magnesium alloys for engine blocks. Disadvantages in comparison with the Al–Si alloys normally used today in mass production are the high costs of the material, the lower material strength, and lower resistance to corrosion. •• Because of the lower material strength, it is established that engine block from Mg alloys cannot be lighter at a proportion corresponding to the ratio of their specific weights, compared with Al alloys. In working out an engineering design that is in line with the loading, the differences in the material properties must be taken into account. In comparison with an engine block made of an Al–Si alloy, using a magnesium alloy in a comparable design can cause weight savings in an order of magnitude of 25%. •• The corrosion resistance of magnesium alloy components, where no additional protective measures are adopted, is lower than that for components made from Al–Si parts; their natural surface or skin after casting already provides sufficient corrosion resistance. Surface and contact corrosion will occur at unprotected surfaces. Contact corrosion arises when parts made of magnesium alloys come into contact with components made of other metals or alloys. It results from the differing positions of the various metals along the electrochemical series. Contact corrosion may arise at threaded connectors and at holes for fixing elements such as alignment bushings and pins. •• The cost advantage for Al–Si alloys in comparison with magnesium alloys is an order of magnitude of about the factor three and results essentially from the absence of a recycling market for magnesium. While Al–Si alloys are available at low cost in the form of secondary alloys from components that have been melted down, it is necessary to draw upon the costly primary alloys for magnesium alloys. The modern design of a 3-L SI engine in series design compensates for these disadvantages by casting around an aluminum insert containing the cylinder sleeves, the crankshaft main bearing pedestals, and the coolant system. The coolant cannot touch the magnesium jacket of the engine block and corrosion is thereby avoided. Weight savings of 57% compared with a comparable gray cast block and 24% compared with an aluminum engine block are claimed for this design.

7.4.5 Casting Process for Engine Blocks

Engine blocks for automotive engines are manufactured from cast iron or aluminum–silicon alloys. The costs, numbers produced, and engineering design are the main criteria applied when selecting the casting process. 7.4.5.1 Die Casting Permanent molds made of hardened, hot-work tool steels are used in the pressure die-casting process. The sections of the mold have to be treated with a parting agent before each casting is made. In contrast to sand casting and die casting, no cores can be inserted into the mold because the lightweight metal melt is introduced into the casting form at high pressure and high speed. The pressure level depends on the size of the casting and is between 400 and approximately 1000 bar. The pressure is maintained during solidification. In larger castings, the two halves of the form are cooled, allowing a directional solidification of the cast component. In contrast to sand casting and die casting, pressurized die casting provides the most precise reproduction of the hollow cavity in the mold and thus for the greatest precision in the cast component. Thin-walled castings with close dimensional tolerances, great exactness of shape, and superb surface quality can be fabricated. Casting eyes, holes, passages, and lettering to exact dimensions eliminates the need for subsequent machining and casting in place bushings such as the cylinder sleeves made of gray cast iron. Pressure die casting, when compared with sand casting or die casting, offers the highest productivity because almost all the casting and mold movement processes are fully automated. The drawbacks include the limited engineering freedom for the cast component, because undercuts are not possible. Air or gas bubbles that might be trapped in the casting preclude double heat treatment, as for sand casting and die casting. Engine blocks made of aluminum–silicon alloys, in particular with special cylinder sleeve technologies, are produced to an even greater extent in pressure die-casting. 7.4.5.2 Die Casting A die is a permanent metal mold made of gray cast iron or hot-worked steels and is used to manufacture cast components from lightweight metal alloys. Just as in sand casting, sand cores are inserted into the casting mold, offering the benefit of greater freedom in the engineering design. Undercut areas are possible, which is in contrast to pressure die casting. The die-casting process makes it possible to use each mold for many casting cycles, unlike in sand casting where new sand cores are required for each cycle. Again, in contrast to the sand-casting mold, solidification of the metal melt in the die is fast and directional. Closely defined cooling of the die is possible, and this option is often used. The die has to be protected against the lightweight metal melt by applying a parting agent.

118 | Internal Combustion Engine Handbook

6606_Book.indb 118

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.5 Cylinders

In comparison with sand casting, castings taken from the die exhibit a finer inner structure, greater strength, greater dimensional accuracy, and better surface quality. Double heat treatment is possible for die-cast components. In addition to carefully defined control of cooling for the cast component inside the die, the first heat treatment undertaken is often a further heat treatment, controlled cooling. In die casting, one differentiates between gravity die casting and low-pressure die casting. The difference lies essentially in the way in which the melt is introduced to the mold. In the low-pressure casting process, the molten metal is introduced into the die from below at a gauge pressure of from 0.2 to 0.5 bar, which is then maintained during solidification. The almost perfectly directional solidification of the casting that results is one of the fundamental reasons for the high quality of low-pressure cast components. In gravity die casting, by contrast, the mold is filled at atmospheric pressure, using the force of gravity acting upon the molten metal.

During the casting process, the hollow cavities between the outside mold and the cores are filled with molten metal. Following the filling process and after the metal has solidified, the casting is removed from the sand mold. The sand mold is destroyed when doing so. The casting is then reworked to remove traces of the gating, sprues, casting skin, and flash. In sand-cast components made of Al–Si alloys, double heat treatment to increase the strength is possible. The first heat treatment phase is found in the controlled cooling period for the casting inside the sand mold. The second heat treatment occurs during time- and temperature-controlled storage of the casting in a kiln. The sand mold can be used to produce only a single casting. Sand casting is the process traditionally used for engine blocks made of gray cast iron. Engine blocks in Al–Si alloys are also produced in large numbers using a precision sandcasting process. Further applications for sand-casting are creating prototypes and performing short production runs.

7.4.5.3 Lost-Foam Process This is a special variation of the sand-casting process. A plastic model is made of the piece to be cast, using expanded polystyrene, by foaming in place and, if necessary, by gluing individual segments together. The expanded polystyrene model is coated with a water-based parting agent. The model, coated and dry, is placed in a casting shell in which pure quartz sand (without any binders) is filled using vibratory compaction. In this fast casting procedure (taking 15–20 s), the molten metal is directed to the plastic model as a so-called full-mold casting. The heat in the molten metal degrades the plastic model: the liquid and gaseous components are absorbed by the casting sand. Following cooling and deforming, a flash-free casting is obtained. The particular advantages of this process are found in the capabilities of making up plastic models that replicate casting geometries not possible with conventional sand-casting processes because of technical limitations on the latitude for mold fabrication. The disadvantage of this casting process is that larger wall thicknesses are required than die casting, for example. The lost-foam process is suitable for making up both gray castings and lightweight metal alloy castings.

7.4.5.5 Squeeze Casting The squeeze-casting process represents a combination of lowpressure die-casting and the die-casting process. Permanent metal molds are filled from below with molten lightweight metal at a gauge pressure of from 0.2 to 0.5 bar. This is followed by solidification under high pressure at approximately 1000 bar. The excellent density attained when filling the mold also makes it possible to use high-strength alloys with less favorable properties. The solidification of the melt while under high pressure imparts a very fine internal structure to the cast component. Slow filling of the mold and solidification under high pressure give a structure virtually free of pores. As a result, the material is capable of enduring high strength against alternating loads along with great resistance to temperature changes in comparison with both low-pressure casting and die casting. As in die casting, the use of sand cores is not possible when utilizing squeeze casting. Because undercuts cannot be created, the same engineering restrictions apply to squeeze casting as for die casting. In contrast to die casting, double heat treatment is possible in squeeze casting, because there are virtually no pores in the structure. Thus, squeeze casting combines the advantages of die casting, low-pressure casting, and pressure casting.

7.4.5.4 Sand Casting Models and core boxes made of hardwood, metal, or plastic are used to replicate the later engine block casting inside the sand mold. The casting molds are normally made of quartz sand (either natural or synthetic sand) and binders (synthetic resin and CO2). The sand is introduced, using “sand shooting machines,” to make the cores. Combining individual cores to form a core package, and assembling this core package and the outer casting mold, is handled mechanically and fully automatically, even when producing only moderate numbers of castings. Model, core, and mold parting in various planes and inserting cores in the casting mold make it possible to produce complex cast components with undercut areas.

7.5 Cylinders The piston group is mounted in the cylinders. With their surface and the material used—and working in concert with the piston rings—the cylinders also support slip and sealing functions. Over and above this, they contribute, depending on the design, to heat dissipation via the engine block or directly into the cooling water.

Internal Combustion Engine Handbook | 119

6606_Book.indb 119

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.5.1 Cylinder Design

Both engineering and materials aspects have to be taken into account when designing the cylinder and the cylinder running surface. Both aspects are linked to each other. Taking the materials as the starting point, the designs for the cylinders and engine block may be subdivided as follows: •• monolithic design •• insert technique •• bonding technology. 7.5.1.1 Monolithic Design Typical representatives for a monolithic design are engine blocks made of cast iron alloys in which the cylinders are an integral part of the engine block. The required surface quality is achieved by machining in several steps, including preliminary and precision reaming and honing. Monolithic engine blocks made of Al–Si alloys are found in two versions: •• Manufacturing the engine block casting from a hypereutectic Al–Si alloy. Al–Si alloys are deemed to be hyper-eutectic if their silicon content exceeds 12%. The primary silicon precipitated out in the cast component, following the machining of the engine block at the cylinder running surfaces, is exposed by a chemical etching process or with special mechanical honing. A hard, wear-resistant cylinder running surface (referred to as non-reinforced) is created; it has to be mated with an iron-plated piston. Because of the high share of silicon in hypereutectic Al–Si alloys, workpieces made of this material cannot be machined as readily as cast components made of hypoeutectic alloys. The primary silicon crystals precipitated out in the cast part are damaged and splinter during mechanical processing. This results in undesirably short chips. In hypereutectic Al–Si alloys and a closed-deck design, this monolithic cylinder block or engine block design can be manufactured in a low-pressure process, and in hypoeutectic Al–Si alloys and open-deck design, in a pressure casting process. When using this latter process, the primary silicon grains are found, which are far smaller than with low-pressure casting processes. This improves machining properties significantly. Because of the reduced tendency to splinter, the smaller silicon crystals can be worked faster while better cutting results are achieved at the same time. •• Manufacturing the engine block from a hypo-eutectic Al–Si alloy in combination with a finish for the cylinder running surface. The finish may be applied either by electroplating or with a thermal spatter process. In the meantime, cylinder running surfaces that are remelted or plated using lasers are in the development phase. Used exclusively in mass production to date is the quasimonolithic engine block design in which a nickel dispersion layer is electrodeposited on the cylinder running surface. This layer consists of a nickel matrix into which silicon carbide particles are inserted at uniform distribution. Cylinder surfaces finished in this way exhibit excellent running properties and low wear. Moreover, they may be combined

with pistons and piston rings made of conventional materials and platings. The layer is, however, to a certain extent, sensitive to corrosion when using fuels that contain sulfur. The nickel dispersion layer is better known under trademarked names such as NISKAL, GALNICAL, and GILNISIL. The nickel dispersion layer can be combined with both closed-deck and open-deck designs. Because even minuscule porosity in the cylinder running surface can cause plating problems in that the plated layer spalls off, the selection of the casting processes that may be used for Al–Si alloy engine blocks is restricted. The conventional pressure die-casting process, for example, cannot be used without special techniques such as vacuum support. Nickel dispersion plated cylinders are used frequently for single-cylinder motorcycle engines. Multi-cylinder engine blocks for automotive use, incorporating nickel dispersion finished cylinders, are massproduced only to a very limited extent. 7.5.1.1.1 Cylinder engineering for monolithic designs One differentiates between cylinders that are cast as a single unit along the engine block’s longitudinal axis and those that are not cast together. In the past, engine blocks in both closed-deck and open-deck designs made of either cast iron or aluminum–silicon alloys were executed with cylinders that were not joined together along the engine block’s longitudinal axis. This was done to achieve the most uniform possible temperature distribution in the cylinders (by coolant present between the cylinders) and the smallest possible degree of cylinder warping (by preventing mutual influencing of neighboring cylinders). The greater length of the engine block that this involved proved to be detrimental. Today, suitable engineering measures can be employed to ensure that cylinders that are cast as a single, solid unit along the engine’s longitudinal axis can exhibit almost uniform temperature distribution in spite of the absence of coolant between the cylinders. This eliminates any appreciable warping problems and the concomitant functional problems such as excess oil consumption or blow-by. The advantages of unitized cylinders include greater engine block strength, a shorter engine block, and lower engine weight. Today, the reduced engine length is a dominant criterion in view of transverse engine mounting and in light of the ever-declining amount of space available for installing drivetrain components. Depending on the particular engine design (inline, V-block, or boxer engine), designing the engine block with cylinders cast as a unit results in differing degrees of reduction in length and weight. The lower limit for joining cylinders is represented by the web remaining between the cylinders. Regardless of the material used, engine blocks manufactured in mass production incorporate cylinder webs thinner than 5.5 mm. This was made possible by employing metal head gaskets with low compression and setting properties, thus requiring lower preload values at the head bolts. In addition to perfect sealing at the cylinder web, cylinder deformation is reduced to a minimum because of the lower preload values in the aggregate comprising the cylinder head and the engine block.

120 | Internal Combustion Engine Handbook

6606_Book.indb 120

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.5 Cylinders

7.5.1.2 Insert Technique Insertion techniques are normally used for cylinder sleeves in automotive engines in conjunction with aluminum engine blocks. Sleeves made of any of a variety of materials are inserted into the engine block. Following differentiation by function into wet and dry cylinders, one distinguishes whether the sleeve is cast in place, pressed in place, shrink fit, or slid in place in the engine block. Moreover, one may distinguish according to the material used for the sleeve, possibly cast iron or aluminum. 7.5.1.2.1 Wet cylinder Wet cylinders are slid into the engine block, mating with mounting areas machined and prepared accordingly. The water jacket around the cylinder is formed between the engine block and the sleeve (Figure 7.62). Water space

Cylinder sleeve Housing

Figure 7.62 Wet cylinder.

The hanging cylinder sleeve features a collar at its upper end; this collar is clamped between the engine block and the head gasket or cylinder head. The sleeve is centered in the engine block at the collar itself or at a diameter below the collar. Using the collar for centering offers the advantage of good cooling for the top end of the cylinder sleeve, which is subjected to severe thermal loading. The disadvantage is the heavy loading at the fillet in the engine block. Centering the sleeve at a point below the collar causes less satisfactory cooling at the upper end of the sleeve, but does relieve the heavy loading at the fillet. O-rings are used at wet sleeves that are suspended from the top to seal against coolant at the top and against oil from the crankcase at the bottom. In the standing wet cylinder, the support and centering functions are at the lower end of the sleeve or approximately the center (the so-called “mid-stop” design). This sleeve concept requires particularly careful engineering to keep down cylinder deformation. Sealing is by the head gasket at the top and a flat gasket at the bottom, below the sleeve support surface or by O-rings. Misalignment of wet cylinder sleeves at the top plate, resulting in protrusion or depression, can be a problem. This

has a negative effect both on the surface pressure applied by the head gasket around the cylinder and on cylinder deformation. Consequently, sleeve protrusion or depression must be reduced to a minimum. Inserting fully machined, wet cylinder sleeves into an engine block after the top plate has been finished is done by imposing extremely close tolerances on the relevant sleeve dimensions. When installing standing sleeves, shimming is a commonly used technique. A further option is final machining of the engine block top plate and the sleeves after the latter have been installed. Wet sleeves are normally manufactured from gray cast iron. The less common aluminum sleeves may be made of either hypereutectic or hypoeutectic Al–Si alloys. For economical reasons, wet sleeves are preferably installed in engine blocks with open-deck design and are manufactured in a die-casting process. The advantages in using wet sleeves are freedom in selecting the material for the sleeve, flexibility with regard to the cylinder bore, and thus, displacement specified by combining the appropriate sleeves with one and the same engine block. Further benefits are simple interchangeability and repairs. Unfavorable factors are the higher manufacturing costs when compared with monolithic concepts. Wet aluminum sleeves are found almost exclusively in lightweight metal engines for sports cars or racing cars, where lower weight and better heat transfer are given preference over cost considerations. 7.5.1.2.2 Dry sleeves Dry sleeves are pressed, shrink fit, or cast in place in the engine block (Figure 7.63). When cast in place, the sleeves are inserted in the engine block mold, and the molten aluminum alloy is cast around them. In contrast to wet sleeves, the water jacket is not between the sleeve material and the engine block casting, but, as in the monolithic design, is a component of the engine block casting.

Water space

Cylinder sleeve Housing

Figure 7.63 Dry sleeve.

Consequently, no sealing is required between the dry sleeve and the engine block.

Internal Combustion Engine Handbook | 121

6606_Book.indb 121

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Any protrusion of the dry sleeve—pressed or cast in place—in relation to the top plate level is corrected by machining the deck plate and the inserted sleeve together. Dry sleeves may be made of either gray cast iron or (hypereutectic) aluminum alloys; sintered sleeves made of powdered metal materials are another option. The running surfaces for dry sleeves made of gray cast iron or aluminum are treated the same way as wet sleeves and, thus, exhibit the properties listed there. Dry sleeves, regardless of the material, may be used for both open-deck and closed-deck designs and can be combined with all the casting processes that are normally used for engine blocks. Aluminum engine blocks are found in mass production and are characterized by the closed-deck design, made in die-casting or low-pressure casting, with gray cast iron cylinder sleeves pressed in place; also seen is the opendeck design made with pressure casting where gray cast iron sleeves are cast in place. To obtain geometrically locking connections during the casting in place, the outer surface of the sleeve is profile machined or cast with a rough surface (so-called rough-cast sleeve) [7-39]. The advantages of dry cylinder sleeves are freedom in the selection of the materials, the easy repair option (in the case of gray cast sleeves) by reaming out to oversize dimensions, separate manufacture (in aluminum sleeves) of the cylinder running surface, and the option for combining sleeves with an engine block made of a different aluminum alloy. One embodiment of this is the cast-in-place sleeve made of a hypereutectic, spray-compacted aluminum alloy with the trademarked designation SILITEC. A disadvantage in this concept, inherent to the concept, is poorer heat transfer between the cylinder running surface and the water jacket. Regarding manufacturing costs, there may be advantages or disadvantages when compared with a monolithic design depending on the number of units produced, the casting process selected, and the engineering details of the engine block and sleeve. Particularly, when manufacturing large numbers of units in pressure die-casting or automated sandcasting processes, cast-in-place gray cast sleeves can be very economical in terms of overall costs. 7.5.1.3 Bonding Technology Bonding technology can be used only in engine blocks made of aluminum alloys. In aluminum engine blocks incorporating bonding technology—and in contrast to aluminum engine blocks in classical monolithic design—an inseparable unit comprising the engine block and the cylinder running surface is created. There are two fundamental embodiments. In the first, shaped cylindrical bodies, so-called preforms, made of a bond of suitable metallic and ceramic materials, are inserted in the casting molds and are infiltrated by the molten aluminum alloy at high pressure during the casting process. Alternatively, sleeves made up of several layers and/or of several metallic materials are joined with the engine block by an inter-metal bond during the casting operation. The bonding technology limits the choice of casting process to pressure die casting and processes derived from pressure die casting, such as squeeze casting or the “new die-casting” process

developed by Honda. The limitations that the technology places on die-casting and casting-related processes make it necessary to adopt an open-deck design when implementing a bonding process. Cylinders that are cast together as a unit and those that are cast separately can be realized. Regarding the use of preforms, one may differentiate between two bonding technology processes: •• Honda Metal Matrix Composite (MMC) process: This metal matrix composite process has been in volume production for some years now. It is similar in principle to the KS Lokasil process. Fiber preforms are inserted into the mold prior to casting. The preforms consists of a bond of Al2O3 fibers and carbon fibers and, in the new Honda die-casting processes, are infiltrated by the molten aluminum alloy. •• KS Lokasil process: A high-porosity, cylindrical body made of silicon is infiltrated by a liquid aluminum alloy at high pressure during the squeeze-casting process. The cylinder running surface is prepared with three honing phases. In preliminary honing using diamond strips, many of the silicon crystals in the surface are damaged. Intermediate honing using silicon carbide removes this damaged silicon crystal layer. The third honing phase, using grains bound up elastically in the honing strips, exposes the silicon grains. Similar to the silicon crystals that are exposed in the monolithic version by etching the hypereutectic aluminum alloy, these crystals form a hard and wear-resistant cylinder running surface. An iron-plated piston is required for use as the mating material. As a rule, a set of piston rings similar to that used with cylinders made of gray cast iron is sufficient. When using bonding technology with metallic sleeves, it is possible, in one embodiment, to cast in place (during die casting) a sleeve made up completely by thermal spraying and comprising various materials in multiple layers (GOEDEL technology). The inter-metallic bond between the sleeve and the molten aluminum alloy is ensured by the appropriate choice of materials (normally an Al–Si alloy similar to that used for the engine block) and the special surface at the outside face of the thermal spray sleeve. Regarding the material used for the cylinder running surface, with its influence on tribologic properties, the thermal spray process allows a broad choice of FE-based alloys. Machining for the cylinder running surface in each case (as a rule, by a two- or three-step honing process) is chosen to suit the material selected. The same applies to the selection of the mating materials in the pistons and piston rings. In another embodiment, the desired inter-metallic bond can be achieved by applying the thermally sprayed outer layer of the GOEDEL sleeve on a conventional gray cast iron sleeve (referred to by the manufacturer as the HYBRID sleeve). Thus, the usual situations apply to machining the cylinder running surface and selecting the pistons and piston rings for the gray cast iron running surfaces. This option is thus particularly economical and has been used for mass production. This is different from other bonding technology solutions that are limited essentially to sports cars and other high-performance engines.

122 | Internal Combustion Engine Handbook

6606_Book.indb 122

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.5 Cylinders

Bayonet connection

Forcing lever Measuring air intake

Double cone

Tool body Guide strips

Strip holder

Air measuring nozzles Honing strips

Return spring for strip holder

7.5.2 Machining Cylinder Running Surfaces

The cylinder running surface in internal combustion engines is the tribologic mating material and sealing surface for the pistons and piston rings. The properties of the cylinder running surfaces have a determinant influence on establishing and distributing an oil film between the mating components. There is a strong interrelation among cylinder roughness, oil consumption, and wear inside an engine. The final machining of the cylinder running surface is effected by precision boring or turning and subsequent honing. During the honing process, rotational and alternating translational motions are superimposed upon each other to create a cutting motion. In this way, a cylinder shape of less than 10 μm and uniform surface roughness can be achieved. The scoring arising from the cutting motion includes the so-called honing angle. This processing, as shown in Figure 7.64, should be as gentle on the material as possible to avoid breakout, pinching at the edges, and the formation of burrs. The material is cut with the assistance of honing strips running under a water-based coolant/lubricant or special honing oil [7-35]. At the prescribed surface pressure or advancing speed, material removal of 100 μm in diameter is achieved in less than a minute.

Figure 7.64 Multi-strip honing tool with air measurement system (Mfg. Nagel).

7.5.2.1 Machining Processes In standard honing in a single-stage or multi-stage machining process, a surface structure exhibiting normal distribution is created so that in the roughness profile, there are as many valleys as peaks. Plateau honing, on the other hand, levels peaks with a supplementary machining step, creating a plateau-like slide surface with deep scoring that retains oil. Helical slide honing is a further refinement of plateau honing. It differs from plateau honing primarily by the reduced roughness (and, in particular, the peak roughness) and a very large honing angle of from 120° to 150° for the deep scoring. Very uniform surface roughness is achieved using special honing strips that follow the shape of the bore. Laser texturing offers almost unlimited freedom in surface design through carefully defined removal of material by the laser beam [7-36]. The cylinder running surface is textured in the TDC area and is otherwise made as smooth as possible. Textures and structures such as helically arranged slots and pits, as well as cupping, are possible in addition to conventional, uniform cross-scored textures. The roughness profiles for various honing processes are shown in Figure 7.65. The influence of the running surface

5µm

A B

C D

E 0

2,5

5

7,5

Total measuring path l m in mm

10

12,5

Figure 7.65 Roughness profile for (a) standard honing, (b) plateau honing, (c) helical slide honing, (d) laser-imparted texture, and (e) smooth standard honing.

Internal Combustion Engine Handbook | 123

6606_Book.indb 123

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

topology on oil consumption and particle emission has been clearly established at the example of a Direct Injection (DI) diesel engine [7-37]. A complex variation of honing, in which free grinding grains are used, is lapping. Here loose grain is used to give the cylinder running surface a random high/low structure. Solid strips press this hard lapping agent in part into the surface, and a plateau surface is created. In brush honing, the surface texture is rounded and deburred following standard honing; a brush coated with a carbide material is used for this purpose. Fluid blasting is yet another process used to remove the metal “frost” (sometimes also referred to as spangle) from the surface and to flush out pores present in the surface. In this process, the entire cylinder running surface is blasted with a water-based coolant/lubricant at a pressure of approximately 120 bar. Also used is a process in which gray cast cylinder running surfaces are exposed to excimer laser beams, which opens the graphite precipitates and simultaneously melts on the surface to improve running properties [7-38]. Exposure honing for aluminum cylinder running surfaces employs specially designed honing strips to depress the soft aluminum matrix in comparison with the reinforcing fibers or particles. The particles can also be exposed by etching. The purpose is to depress the aluminum, which has a tendency to weld, by 0.5–1 μ m. The oil retention spaces created by the suppression of the aluminum improve the running properties of the surface. Plasma or flame-sprayed cylinder sleeves can be smoothed to a great extent, similar to inductance-hardened gray casting. The oil retention spaces created by the material’s porosity guarantee good running properties.

Other elaborate special processes are nitriding and phosphatizing the honed cylinder running surfaces. Nitriding creates a very rough and hard layer that is not suitable for use as a cylinder running surface without supplementary treatment such as phosphatizing. Phosphatizing is also used without nitriding and has a smoothing effect while also acting as a solid lubricant. The 3-D surface images after honing gray cast iron and aluminum cylinder running surfaces are shown in Figure 7.66 and Figure 7.67.

Figure 7.67 The 3-D surface image of an aluminum cylinder running surface with exposed reinforcing particles.

7.5.3 Cylinder Cooling

Figure 7.66 The 3-D surface image for a honed, gray cast iron cylinder running surface with metal frost (white marbling) and the honing angle sketched in.

7.5.3.1 Water Cooling With just a very few exceptions, today’s automotive engines are water-cooled. In contrast to air-cooled cylinders which are fitted with cooling fins, the cylinders are surrounded by a water-filled cavity, the water jacket. An important engineering dimension is the water jacket depth, defined as the distance from the top plate plane to the lowest point in the water jacket. In modern cast iron engine block designs, the water jacket ends in the area swept by the lower piston ring, that is, in the area between the first compression ring and the oil control ring when the piston is at BDC. The water jacket is even shorter in aluminum engine blocks. The water jacket depth corresponds to approximately one-third of the length of the cylinder running surface. This is made possible by the greater thermal conductivity of aluminum alloys in comparison with cast iron materials and by pistons with ever shorter compression heights. A short water jacket reduces the coolant volume in the engine and, thus, the engine weight. The smaller coolant volume and thermal capacitance

124 | Internal Combustion Engine Handbook

6606_Book.indb 124

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.6 Oil Pan

shortens the engine’s warm-up phase, with positive effects in terms of unburned hydrocarbon emissions and the response time for the catalytic converter.

applied. This gives an inter-metallic bond between the cylinder sleeve and the rib jacket and results in uniform heat flow.

7.5.3.2 Air Cooling Only a very few manufacturers still use air-cooled cylinders in automotive engines today. Heat removal in air-cooled cylinders is dependent upon the thermal conductivity of the cylinder fins and of the cylinder materials, shape of the cooling fins, and the way in which cooling air passes across the fins.

In engines subjected to less severe loading, cylinder designs were also used in which a lightweight metal rib jacket was cast around a cast iron sleeve without any particular bonding, or a cast iron sleeve was shrink fit into a prepared aluminum fin jacket.

7.5.3.2.1 Shape of the cooling fins In air-cooled cylinders, cooling fins are located on the cylinder outside walls to increase the effective surface area for heat transmission. In theory, fins with a triangular cross section are the most effective. In cast cylinders, the particulars of the process result in slightly trapezoidal fins with rounded edges being formed; these are hardly any less effective than fins with a triangular cross section. Heat transfer at cooling fins can be increased by •• increasing the fin surface area by, for example, lengthening and by greater fin height •• increasing cooling air velocity •• converting from random to directed cooling air flow by installing air baffles and deflectors, for example •• using a material with the highest possible heat transmission capacity for the cylinder and fins, such as aluminum alloys instead of gray cast iron. The fins on motor blocks carry out other functions in addition to heat removal:

7.5.3.2.2 Cooling air flow In air-cooled automotive engines, forced or positive cooling using a fan is implemented without exception. Here the cooling air is routed from the fan housing, through baffle plates that surround the cylinder and cylinder heads, and is thus directed onto and between the cooling fins. The more favorable the flow characteristics and the thermal conductance values in each case, the lower the required amount of cooling air and fan output.

7.6 Oil Pan The oil supply for passenger engines today is provided almost exclusively with a wet sump lubrication design. In such engines, the oil pan forms the bottom termination for the engine block (Figure 7.68). The design and construction of the oil pan is strongly determined by the installation situation (package) in the actual vehicle. Under certain circumstances, the same engines may have different oil pans, depending on whether they are installed in a longitudinal or in a transverse direction.

•• increasing the stiffness of the engine block side walls, which improve acoustic properties •• optimizing force transmission from less stiff areas into load-bearing areas of the engine block structure •• optimizing the casting process to achieve better flow of the molten metal to areas in the engine block.

Sheet metal oil pan

Modern calculation methods make it possible to optimize fins with regard to weight, structural strength, and heat removal. The thermal conductivity of aluminum alloys is almost three times that of cast iron materials. That is why cast iron cylinders, once suitable running surface finishing technologies had been developed for aluminum cylinders, are increasingly replaced by cylinders made of aluminum alloys. In air-cooled gasoline and diesel engines that are in harsh service the dimensional stability of the pure lightweight metal cylinder may in some cases not be sufficient. It was for that reason that cylinders were used in which a cast iron or steel liner is surrounded by a jacket of fins made of lightweight metal. These so-called bonded cast cylinders were made up in two processes: •• Casting around a gray iron cylinder sleeve (with a roughened outer surface) a rib jacket made of a lightweight metal alloy in a pressure casting process.

Al die cast oil pan with dipstick, engine 944 turbo

Figure 7.68 Oil pan.

•• Casting around a steel or cast iron sleeve to which, prior to the casting process, a thin iron–aluminum coat had been

Internal Combustion Engine Handbook | 125

6606_Book.indb 125

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

The most important of the oil pan functions are as follows: •• Serving as a container to receive the motor oil when oil is first installed and as the collecting basin for motor oil returning from the bearings and lubrication points. •• Enclosing the crankcase, it serves, in specially engineered oil pan types, at the same time in stiffening the engine and transmission assembly. This improves the acoustic behavior of the engine and transmission assembly in the low-frequency range. •• It drives the oil through oil guide fins to the oil intake point, contains slug plates to steady the oil, and enables the separation of air from the oil. •• Taking the threads for the oil drain plug and the dipstick guide tube and often housing, in addition, an oil level gauge showing the oil fill in the vehicle.

7.6.1 Oil Pan Design

In mass production engines, the oil pan is normally a singlelayer, deep-drawn component made of sheet steel. To improve acoustic properties, a design is now used that incorporates two layers of sheet steel with a plastic film between them. Used in conjunction with large-displacement engines incorporating cast iron or aluminum engine blocks are oil pans made of Al–Si alloys, manufactured by die casting or pressure casting. This is achieved with a stiff design for the oil pan side walls and, primarily, with an integral flange at the clutch end of the engine, as the connection to the transmission flange. This design makes a significant contribution to stiffening the engine and transmission group and, consequently, to better acoustic properties. This design is used in about half of European engine concepts. Oil pans made of aluminum alloys are made in single- and two-component versions. Two-part oil pans compose an upper section made of a lightweight metal and a lower section made of sheet steel which is bolted to the upper section. The steel component can be changed more economically in case of deformation (if the car bottoms out). In comparison, an oil pan made entirely of aluminum would have to be completely replaced if deformation occurs. Today, this advantage is only of slight significance because of the underbody claddings being used more frequently to enclose the engine. A more recent development is the use of fiber-glass reinforced polyamide in oil pan construction. This application leads to special demands on the synthetic material because of the long-lasting contact with air and oil at high temperatures. For this reasons, the aging behavior is tested with oil wetting at 160 °C for more than 5000 h. Mechanical demands on the synthetic material arise also from lifting and lowering the engine, for example, by forklifts in the body shop. From a design point of view, it is differentiated between purely synthetic oil pans and synthetic oil pans with aluminum die-cast components (hybrid). They allow again the support of the transmission. The sealing of the plastic oil pan is similar to the one of cylinder head shrouds; T- or I-profile seals are used. Sealing with liquid silicone is also possible. Oil pans made from synthetic materials have some advantages:

•• Weight reductions of approximately 30% with the hybrid solution and approximately 60% with the purely plastic design, compared with, for example, an aluminum type weighing 2.2 kg. •• The high integration density enables packaging advantages. For example, not only oil channels and channels of the crankcase ventilation system can be integrated but also an oil filter, an oil pressure regulator, and an oil cooler. With suitable dimensioning, this can also result in a reduction of the pressure loss in the oil channels and the drive power required for the oil pump.

7.7 Crankcase Venting Because of the limited tightness of the piston rings during reciprocating engine operation, gases (the so-called blow-by gases) from the combustion chambers pass through the gap between the cylinder and the pistons and into the crankcase. Other causes for leak gas flows into the engine block are leaks at the valve guides and the shaft bearings of turbochargers but also the vacuum pump used in diesel and SI engines with direct fuel injection (brake servo pump). The high pressure differences present at the piston rings, in particular, in conjunction with high temperatures cause a portion of the engine oil adhering to pistons and cylinders to atomize in a fine oil droplet mist which then flows, with the blow-by gases, into the engine block. Separate from this, mostly larger oil droplets are thrown off the moving transmission components (crankshaft, pistons, camshaft, and timing chain) when the engine is running. The injection oil cooling of the pistons common in high-performance engine must also be regarded as a major source of oil mist; furthermore, condensation processes of previously vaporized, low-boiling lubricating oil components contribute to oil particle formation in addition to the lubricating oil atomization. Figure 7.69 shows blow-by oil paths and transport mechanisms for oil mist and oil wall film in the ventilation system of an exhaust gas turbocharging engine. To avoid an unwanted pressure rise in the crankcase during engine operation (risk of gas and oil leaks, environmental hazards, and shaft sealing ring failure), a crankcase venting system must constantly discharge the blow-by gas from the crankcase.

7.7.1 Regulatory Marginal Conditions

In the early decades of engine construction, the crankcase venting gases were discharged directly into the atmosphere. The high portions of hydrocarbons in the blow-by gases of SI engines were the backdrop for the first voluntary (and later stipulated by law) introduction of closed crankcase venting systems (closed crankcase ventilation (CCV) systems). These were the first to return the blow-by gases into the intake tract in the 1960s, initially in California and then the other U.S. states [7-44]. Corresponding legal provisions were subsequently adopted in all important markets and extended to include passenger car and truck diesel engines [7-45]. However, for

126 | Internal Combustion Engine Handbook

6606_Book.indb 126

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.7 Crankcase Venting

TurbochargerTurboscharger

Oil

Exhaust

Intake silencer with air filter

Leak gas flow from the turbocharger shaft Intercooling

Separator unit with crankcase pressureregulating valve

Leak gas flow at the valve guides

Fine oil generation via atomizing mechanism in the piston ring gaps

Coarse oil in form of oil entrained at the walls

Coarse oil as droplets and splashes, generated by shearing, swirling and thrown oil at moving crank gear component and piston spray cooling.

engines for off-road applications and heavy utility vehicles, venting the crankcase into the atmosphere is still permitted in principle (open crankcase ventilation systems). Figure 7.70 shows a summary of the relevant statutory milestones in selected markets. According to the Directive 70/200/EEC, Annex V, Test Type III applicable for passenger cars in the European Union, evidence must be provided that no gauge pressure is present in the crankcase in three different engine operating points. The compositions and mass flows (in European applications up to 1% of the intake air mass flow) of the crankcase ventilation gases depend, to a significant extent, on the combustion process, the crank gear design, the operating state (effective mean pressure, speed, and coolant temperature), and the wear condition of the crank gear. Correspondingly, the gases must be continuously returned into the engine intake system. Compared with the gas recirculation and the return of the activated carbon filter purge gases of the tank ventilation system (SI engines) where the mass flows can, principally, be freely applied according to the requirements in the map, the blow-by mass flows imprinted by the crank gear significantly complicate a pollution-optimized mixture formation and thus compliance with strict exhaust emission limits in SI engines in particular. The tougher requirements

Fine oil generation via condensation of vaporized lubricant portions

Figure 7.69 Illustration of blow-by gas paths and transport mechanisms for oil mist and oil wall film in a exhaust gas turbocharged SI engine.

in respect of oil consumption of the engines, contaminations and deposits in the intake system components, the combustion chambers, and the exhaust gas treatment systems (in particular, in combinations of cooled EGR systems with CCV), in particular, require that lubricating oil components present in the venting gas must be separated as completely as possible to prevent soot deposits in tacky HC components.

7.7.2 Technical Requirements

The statutory marginal conditions result in many technical requirements for crankcase venting systems. Essential requirements concern maintaining a specific pressure in the crankcase, the separation of oil components from the venting gas, and the recirculation of the separated oil into the oil sump of the engine. To maintain the prescribed pressure in the crankcase, throttles or flow-limiting valves are arranged in the intake and discharge flow line to the crankcase or, in modern crankcase venting systems, prevalently used pressure control valves (PCVs). In conventional venting systems, the pressure differential between the crankcase and the intake system is used to transport the venting gases into the engine intake system and to separate the lubricating oil particles. As can be seen in the intake manifold pressure and blow-by

Internal Combustion Engine Handbook | 127

6606_Book.indb 127

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Region

Regulation, Limit Value

Legislation

In Force From

U.S. California

Voluntary installation of so-called semi-open systems

Recommendation: Vehicle Combustion Products Committee

01.01.1961

“Positive Crankcase Ventilation” (PCV) systems for passenger cars with SI engines

Californian Health and Safety Code

01.01.1964

Installation of systems installed in California

Voluntary measure

Year of manufacture 1963

100% elimination of hydrocarbons emitted from the crankcases of SI engines

Code of Federal Regulations (40 CFR)

Year of manufacture 1968

Japan

Use of PVC systems in all new passenger vehicles with SI engines

Code of Federal Regulations for Road Vehicles (SRRV) Chapter II, Article 31(12)

01. 09. 1970 (new models), 11.01.1971 (current production)

Sweden

Closed system for crankcase ventilation in SI engines

F12-1968

01.01.1969

Adoption of the US Federal legislation

Code of Federal Regulations (40 CFR)

1976

Canada

Adoption of the US Federal legislation

Code of Federal Regulations (40 CFR)

1971

Germany

Limit for HC emission from crankcase ventilation: 0.15% of consumed fuel

Adoption of Directive 70/220 EEC as Annex XIV to § 47 StVZO in German national legislation

01.10.1970

EU (passenger car)

Limit for HC emission from crankcase ventilation: 0.15% of consumed fuel

ECE/R15 (published 1.8.1970)

01.10.1971

EU (utility vehicle)

General scope of validity: “Emission of gaseous pollutants and airpolluting particles” (requires recycling of crankcase gases)

U.S. Federal

70/220/EEC (published 4.4.1970) 88/77/EEC Annex I

09.02.1988

Figure 7.70 Introduction of regulations for the limitation of emissions from the crankcases of vehicles (Data derived from [7-44] and information provided by Berg-Automotive, Stuttgart).

gas maps of an exhaust gas turbocharged SI engine with fuel direct injection in Figure 7.71 and Figure 7.72, the blow-by mass flows and pressure drops for recirculating the venting gases into the engine intake system vary significantly. The map shows difficult conditions for effective oil separation when adverse relationships between blow-by gas flows and present intake pressures during, for example, full load at low engine speeds (large blow-by gas flows, low pressure differentials than can be used for oil mist separation, and small particle

size distribution spectra as a rule). Further influences result from pressure pulses at the crankcase and intake manifold side which are usually amplified at certain engine speeds because of pipe or acoustic oscillations. The quantity regulation of conventional SI engines with stoichiometric or “rich” (for a short time) air–fuel ratios causes significant portions of high-boiling fuel components conveyed with the blow-by gas into the crankcase and the water generated during combustion to condensate in the crankcase.

Figure 7.71 Pressure differential map, intake manifold pressure— atmospheric pressure of a turbocharged, quantity-controlled 1.8-L passenger SI engine with fuel direct injection. See color section page 1068.

Figure 7.72 Blow-by gas flow map, intake (without ventilation) of a turbocharged, quantity-controlled 1.8-L passenger SI engine with fuel direct injection. See color section page 1068.

128 | Internal Combustion Engine Handbook

6606_Book.indb 128

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.7 Crankcase Venting

These condensates can accumulate in the crankcase during lasting low-load and short-drive operation, in particular in an engine not having reached operational temperature. Apart from the deterioration of the lubricating oil quality and in consideration of a general tendency of extending the oil change intervals, frozen condensate in winter can cause severe engine problems or engine damage by blocking the crankcase ventilation system or the supply with lubricating oil. The problem shown concerns, in particular, charged SI engines with direct gasoline injection when large fuel quantities must be processed for combustion in short time intervals. A non-optimal formation of mixture in particular has the danger that, mostly in high load map points, large volumes of liquid fuel are conveyed into the crankcase within relatively short times. In addition to a reduction of the lubricant viscosity, very large portions of fuel in the oil (up to 20%), can cause the failure of proven sealing elastomers (severe swelling, “weeping” seal) which must be considered when defining the specification and stipulating materials for the development of an engine. To avoid the accumulation of water and fuel condensation in the crankcase, a small part of the intake air in SI engines is often conveyed as purge air through the crankcase (PVC systems: positive crankcase ventilation). It is essential for the functioning of PVC systems that water and fuel condensates are not separated into the oil mist separator and recirculated into the crankcase, but are, for example, conveyed in a gaseous state through the separator, which is promoted by a low pressure and high temperature level in the separator. The high hydrocarbon concentrations in the venting gas in the warm-up phase also in engines at operating temperature cause, under the consideration of increasing blow-by gas flow with increasing engine operational performance, essential influences on the mixture formation and the lambda control and, therefore, on the exhaust gas emissions. To minimize the influences as much as possible, the inlet point(s) of the venting gas flow into the intake system must be arranged and designed so that a very even distribution of the blow-by gas on all cylinders of the engine (uniform air conditions and knocking limits in the individual cylinders) can be attained. The examinations shown at the example of the tank ventilation system in [7-46] with respect to the influence of the introduction of the gas flows loaded with hydrocarbons into the intake system demonstrate that good mixture conditions and a good distribution can be realized with an introduction distributed across the channel circumference in the area of the throttle (high air speeds and turbulences). For crankcase ventilation systems for future vehicle engines, the significantly increased requirements result mostly from the general trend to extending or eliminating maintenance intervals, the toughening of emission standards and the extension of the same to high driving performance, and the verification of compliance with these requirements by on-site monitoring or on-board diagnosis (OBD) [7-47]. Meeting these requirements is generally made difficult by the fact that, because of increasing dethrottling of the entire intake systems in modern SI and diesel engines in critical map areas in particular (full load and low engine speeds), the pressure differentials that

can be used for fine oil separation become lower. In charged SI engines, the changing pressure levels in the intake system make it necessary to arrange two or, in some charging concepts, even three inlet points for the introduction of venting gases into the intake system and to use the corresponding check valves. This leads to a high degree of complexity of the venting systems. At the same time, the limited space conditions in the engine compartments of modern vehicles force compact, space-saving designs of individual components and the integration of different functional scopes into modules. It should also be mentioned that the continuing cost pressure in the automotive sector forces the development of less expensive components, modules, and systems despite higher demands for functionality and reliability.

7.7.3 System Structure of Current Crankcase Venting Systems

To return blow-by gases into the engine intake system, many different ventilation concepts are used in practical applications: important differentiation characteristics are the use or the omission of a crankcase PCV and the question as to whether the crankcase is only vented or also ventilated. Figure 7.73 provides an overview of various ventilation systems for charged SI engines [Figure 7.73(a)–(d)] and diesel engines [Figure 7.73(e)–(g)] with each system’s advantages and disadvantages [7-48] (corresponding to the overview of naturally aspirated engines in the third and fourth editions). In modern vehicle SI engines, mostly ventilation concepts with PCVs are used to ensure sufficient oil mist separation efficiency. In charged diesel engines, the venting gases are introduced upstream of the turbocharger’s compressor.

7.7.4 Oil Mist Separation

An essential function of the crankcase ventilation system is the separation of the oil portions carried with the blow-by gas. The gas flow conveyed through the ventilation system from the crankcase is a multi-phase flow with which, in most cases, lubricating oil droplets of different sizes and an engine oil wall film are conveyed with the gas in the direction of the intake manifold. Following term definitions from process engineering [7-49], oil mist particles in the blow-by gas are often designated with xP < 10 μ m as fine oil and xP < 1 μ m as super-fine oil (nanoparticles). For oil components in this flow that are visible with the naked eye, the term “coarse oil” has become accepted. Because the separation capacity drops with large volumes of coarse oil in most oil mist separators used for fine oil separation and the oil quantities to be returned into the crankcase against a high pressure differential should be minimized, coarse oil separators (such as volume separators, baffle plate separators, wire mesh separators, or coarse oil cyclones) are usually installed upstream of fine oil separators. Some of the criteria used in the selection of a coarse oil separator are space conditions or space requirement, costs, scalability and pressure loss. These coarse oil separators must reliably prevent a transport of surge oil through the ventilation system (oil entrainment) which may occur under adverse

Internal Combustion Engine Handbook | 129

6606_Book.indb 129

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Single-strand throttle RSVcontrolled crankcase ventilation with introduction upstream turbocharger and downstream DK, map focus of ventilation in engine partial load

Single-strand throttle RSVcontrolled crankcase ventilation with introduction upstream turbocharger and downstream DK, map focus of ventilation in engine partial load

Good pressure control process, good crankcase ventilation behavior in cold operation

Good pressure control process, god crankcase ventilation behavior at full load lubrication

Expensive

Expensive

Single-strand throttle RSVcontrolled crankcase ventilation with introduction upstream turbocharger and downstream DK, map focus of ventilation in engine partial load

Two-strand throttlecontrolled crankcase ventilation with flow reversal in the venting line at full load

Limited expense Application and tolerance sensitive, limited ventilation effect

Simple structure

Uncontrolled crankcase ventilation with introduction upstream turbocharger Possibly uncontrolled crankcase ventilating with discharge into ambient (OCV system)

Controlled crankcase ventilation with introduction upstream turbocharger Good pressure control behavior

Simplest structure

2 oil separators required

Low possible pressure differentials for oil separators

Low possible pressure differentials for oil separators Low tolerable intake resistance

Controlled crankcase ventilation with RSV-controlled introduction up- and downstream of turbocharger (usability depending on charger tuning) Use of maximum available vacuum Increased expense

Intake air filter Compressor Pressure control valve (PCV) Auxiliary oil separator (AOS) Throttle valve (TV) Check valve (CV) Throttle Intercooling (IC)

Figure 7.73 Overview of various ventilation systems in (a)–(d) charged SI engines and (e)–(g) charged diesel engines with each system’s advantages and disadvantages.

130 | Internal Combustion Engine Handbook

6606_Book.indb 130

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.7 Crankcase Venting

conditions. The decisive factor for a reliable function of a coarse oil separator is that the flowing venting gas does not exercise high pushing forces on the oil wetting the surfaces of the ventilation system which would cause previously separated oil to be carried away. While the oil consumptions caused by the crankcase ventilation can be lowered to 1 g/h (approximately 1%–2% of the direct oil losses caused by piston rings and valve guides) in passenger cars with well-designed coarse and fine oil separating systems, the entrained oil volumes carried away by the venting system can be 10–2000 g/h. Aside from a purposeful design of the coarse oil separator and the corresponding oil recirculation, the risk of oil entrainment can be reduced with these measures: •• selection of a favorable removal point screened from sprayed oil for the blow-by gas flow in the crankcase or cylinder head •• sufficient dimension of the ventilation and oil recirculation cross sections •• limitation of the venting gas flows from the crankcase (by using an optimized piston ring arrangement, for example) •• screen of the rotating crank gear components from the oil sump by using baffle plates or oil strippers •• limitation of the engine oil filling levels •• limitation of the oil temperatures •• limitation of the reduction in viscosity of the used engine oils •• reduction of pulsations. The ventilation flows, pressure pulsations in the crankcase and the intake manifold which strongly fluctuate during engine operation, different phase distribution states (illustrated in [7-50]) of coarse oil in pipes, and the pressure differentials limited in individual map areas for use in separation, combined with increasing requirements for the separation quality overall, present high demands on the separator system used. For the separation of oil mist from the gas flow of the crankcase ventilation, many different oil mist separators are available as a rule, as shown in [7-51], which utilize the pressure differential between intake manifold and crankcase, mechanical drive energy, or electric energy for the separation. The essential selection or evaluation criteria include here the separation efficiency, the pressure losses across the separator, the costs, maintenance requirements, and the principal feasibility of each separation principle in a given vehicle engine. To assess the fine oil separation capacity of a separator, often the separation efficiency for a medium droplet size spectrum of xP = 1 μ is used. This is a particle size range in which special measures are necessary to realize effective separation according to the physically effective separating mechanisms [7-52]. The wire mesh separators frequently used in the past have the disadvantage of limited separation capacity for fine oil in particular. In comparison, fiber separators allow high to ultrahigh separation efficiency. According to their functional principle, fiber separators can be classified into through-flow fiber separators and impact mats. Through-flow fiber separators have relatively high space requirement, but, if correctly dimensioned, are significantly more efficient with very small particle sizes, compared with impact mat separators. Their

disadvantage in diesel engines in particular is their tendency to leave deposits on the separating media which will often require a replacement of the separator insert during the vehicle service life. Furthermore, the capacity of the separating medium to hold condensate has the inherent risk of the separator freezing during the winter months. To avoid a harmful crankcase overpressure because of blockages, emergency valves are frequently used in through-flow fiber separators which will open a bypass when a blockage is detected. For the aforementioned reasons, mostly cyclone separators are used in the crankcase ventilation systems of modern vehicle engines, representing a compromise among capacity, space requirement, and costs. They are not particularly sensitive to contamination and do not require replacement. Many detail optimizations at several small-dimensioned, parallel-connected cyclones have substantially increased the fine oil separation capacity in recent years, relative to the same separator pressure losses. A fundamental disadvantage is that, under the marginal condition of limited pressure gradients available for oil mist separation, cyclone separators, but also other oil mist separator type, deliver optimal separation efficiency only at a gas throughput determined by the dimensioning of the separator. One approach to mitigating this advantage is to install a pressure-limiting valve in parallel to the relatively small dimensioned separating unit as shown in Figure 7.74. When a defined pressure differential is exceeded, this pressurelimiting valve will open an impact separator as a bypass and thus prevents an unwanted buildup of overpressure in the crankcase in the event of large blow-by mass flows. The design of the pressure-limiting valve as impactor prevents an unwanted decrease in separation efficiency of the separator with a pressure-limiting valve flown through in parallel. Figure 7.75 displays the relevant pressure loss characteristics and separation efficiencies of a cyclone separator unit with and without pressure-limiting valve. If a pressure-limiting valve is omitted, a very large dimensioned fine oil separator must be used to avoid impermissible gauge pressure levels in the crankcase at large blow-by mass flows (worn crank gear). This has the effect of fine oil separation capacity being low in the low blow-by mass flows that are present in large map areas (Figure 7.75, cyclone variant II). Alternative to the mentioned mat, winding and cyclone separators, some engine manufacturers use a coarse oil centrifuge for oil mist separation, which is driven by either camshaft or the mass balancing shaft. With these centrifuges, larger quantities of the coarse oil conveyed with the blow-by gas flow can be reliably separated. However, the fine oil separation efficiency of these centrifuges is relatively low because of the limited speeds of camshafts or balance shafts, limited component dimensions, and relatively high conveying speeds of the venting gas through these separators. High fine oil separation levels can be obtained with fine oil centrifuges which have their own advantages and disadvantages [7-53]; they are mechanically driven via the crankshaft, hydraulically via pressure oil, pneumatically, or electrically. The hydraulic drive with the engine pressure oil allows an

Internal Combustion Engine Handbook | 131

6606_Book.indb 131

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Raw gas discharge Purge air inlet to crankcase Check valves Pure gas discharge

Core-less shaped cooling water channels

Crankcase pressure control valve

Figure 7.74 Sectional view of a crankcase ventilation module for a turbocharged direct fuel-injected SI engine with coarse oil separator, fine oil separator, discontinuous oil recirculation through collection tanks and check valves, crankcase PCV, and check valves at the pure-gas side (Hengst).

Multi-cyclone unit Pressure-limiting valve

Oil drain valve

Collection tank for separated oil

Degree of separation

Differential pressure [hPa]

Coarse oil consumer

Blow-by-Volumetric flow rate [l/min] Differential pressure behavior, Cyclone I without DBV Differential pressure behavior, Cyclone I with DBV or bypass Differential pressure behavior, Cyclone II without DBV Degree of separation, Cyclone I without DBV Degree of separation, Cyclone I with DBV Degree of separation, Cyclone I with non-separating bypass Degree of separation, Cyclone II without DBV

DBV: Pressure-limiting valve (impactor valve)

Figure 7.75 Pressure loss characteristics and separation efficiencies of a cyclone separator unit with and without pressurelimiting valve.

132 | Internal Combustion Engine Handbook

6606_Book.indb 132

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.7 Crankcase Venting

Degree of fraction separation [%]

inexpensive and reliable application. However, the usable capacity is limited by the permissible shut-off volume flow, the discharge cross sections, and the oil pressure provided in the engine map. Changes in oil viscosity (temperature and oil condition) represent a challenge in the design of such drives. The necessary hydraulic interfaces limit universal application options. With respect to the overall efficiencies and the required separation capacities, an electric drive is the more advantageous solution, compared with a hydraulic drive. This drive would be also independent from the operating state of the combustion engine and would even allow operation as needed up to targeted run-on of the centrifuge after shutting down the combustion engine (start/stop in hybrid engines). The OBD monitoring demanded by tough emission rules can be achieved very easily. Finally, standardized components can be installed relatively universally in small series applications. The achievable higher continuous speeds enable the miniaturization of separating centrifuges. The development of inexpensive, long-lasting and temperature-resistant brushless electric motors is challenging. The disk separator sectionally shown in Figure 7.76 represents a classic type of such separating centrifuges which have been used in process engineering for more than 100 years. The core of this separator type is the rotor in which a number of hollow cone-shaped disks are axially clamped after each other with the shaft. By flowing through the narrow gaps between the disks, the centrifugal force radially moves the oil mist droplets to the outside, because of their higher density, where they impact on the lower side of each axially arranged plate above and form a wall film there. At the outer radius of the disks, the separated oil is thrown off in the form of larger drops which are no longer carried away by the gas flow and gather at the inner wall of the centrifuge housing. This oil accumulates at the bottom of the centrifuge housing and can be returned to the crankcase. Figure 7.77 shows the dependency of the separation capacities to the rotor speed for various mid-sized particle spectra for a separation centrifuge

Medium particle size [µm]

of the disk type shown in Figure 7.76. The illustration shows the rise in separation efficiency with increasing rotor speed as it is typical for separator centrifuges. As a rule, disk separators allow the realization of highest separation efficiencies, in particular for small particle sizes. Conceptually, the gas can flow through disk separators from outside to inside, but also from inside to outside.

Intake manifold Crankcase

Figure 7.76 Sectional view of a disk separator for oil mist separation (schematic).

A flow from inside to outside has the advantage of a negative differential pressure in many operating ranges. The conventional design of the disk separator with individual disks clamped with the shaft has challenges with respect to tolerances, balancing and manufacturing costs. In principle, centrifuges with alternative rotor designs, such as honeycomb, may be used as fine oil centrifuges in crankcase ventilation. In the case of very high demands on oil mist separation efficiency including ultrafine particles and low pressure gradients usable for separation, the use of E separators [7-54] may be considered. These separators utilize the forces on a charged particle in an electric field [7-52] for separation. For pipe separators, an emission electrode is centrally arranged

Figure 7.77 Dependency of the fine oil separation efficiencies of a disk separator from the rotor speed and the mean particle size.

Internal Combustion Engine Handbook | 133

6606_Book.indb 133

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

in a pipe-shaped channel, with the channel wall acting as collecting electrode. By applying high voltage between the emission and collecting electrodes of approximately 10 kV, an electric field is generated. Free electrons in the immediate vicinity of the negatively charged emission electrode are accelerated in the direction of the positive electric potential of the collecting electrode and generate further electrons and cations when they collide with neutral gas molecules. The corona discharge generated by this process is stabilized by cations ejecting further electrons from the emission electrode. Excited gas ions also generate further charge carriers by photoionization. On their way to the emission electrode, the generated electrons attach to the oil droplets which are then separated on the emission electrode because of the forces acting in the electric field. In diesel engines, E separators enable fine and ultrafine oil separation efficiencies of nearly 100% at minimum pressure losses. The disadvantages of the current conceptual approaches are mostly the comparatively high costs but also the risk of deposits in the separator. The risk of combustion of blow-by gases with high fuel content in the event of a flashover opposes the use in SI engines.

dependency of the crankcase pressure on each intake manifold pressure and the blow-by mass flow.

Figure 7.78 Sectional view of a crankcase PCV.

7.7.5 Crankcase Pressure Control

To transport the venting gases into the intake tract and to separate the lubricant oil particles, the pressure differential between crankcase and intake system is used in conventional crankcase ventilation systems. This requires a small underpressure (∆pamb. CC < approximately 30 hPa) in the map for strongly fluctuating intake manifold pressures and different blow-by mass flows. With the introduction of CCV systems in the 1960s, this was resolved with PCV systems in (quantity-controlled) SI engines. In these systems, the venting gas volume flowing from the crankcase into the intake manifold is dosed by a fixed throttle cross section or, alternatively, by a cross-section-variable spring-loaded flow control valve. To prevent an impermissible pressure drop in the crankcase at low blow-by gas volumes in partial load, fresh air is transported into the crankcase through another conduit. In high blow-by gas volumes and high absolute pressures in the intake manifold (full load), the direction of flow reverses in the fresh-air line to introduce the venting gases upstream of the throttle valve under these conditions. In current crankcase ventilation systems for SI and diesel engines, PCVs [Figure 7.73(a), (b), (f), and (g)] inserted in the conduit system between the crankcase and the intake manifold ensure that the pressure differential between crankcase and environment remains mostly constant, even for widely varying blow-by gas flows and intake manifold pressures. Figure 7.78 shows a sectional view of a PCV with conventional essential characteristics. The flow cross section through the valve formed by the membrane and the corresponding flow-off port regulates according to the force balance among the regulating spring, the ambient pressure acting on the top side of the membrane, and the pressure acting on the bottom side of the membrane in the in-flow space and the flow-off port of the valve. Figure 7.79 shows a schematic comparison of the characteristics of two different PCVs. The curve of valve A displays a large

Crankcase pressure large Gauge pressure Intake manifold pressure small

Vacuum pressure

Figure 7.79 Characteristics comparison between two crankcase PCVs (schematic). Valve A: unfavorable characteristics. Valve B: favorable characteristics.

Such a characteristic has the inherent danger that an unwanted overpressure is generated in the crankcase at low absolute pressures in the intake manifold and large blow-by gas flows. The problem shown is intensified by the fact that, with large blow-by gas flows, significant pressure losses are generated in the oil mist separator which is usually installed upstream of the PCV. Accordingly, the pressures present in the crankcase under these conditions are even increased by the value of the pressure losses in the oil mist separator. In comparison, the curve of valve B exhibits a desired far-reaching independence of the pressure regulating characteristic from the blow-by mass flow and the pressure in the intake manifold. Favorable pressure-regulating characteristics can be achieved with large effective membrane surfaces or a two-stage design. A less expensive alternative approach with fewer tolerance issues can be seen in Figure 7.78. In this solution, a pin is concentrically arranged over radial fins in the flow-off cross section from the membrane. This pin is either flush with the face surface of the flow-off cross section and the radial fins or protrudes some tenths of a millimeter. When the membrane approaches the face surface, the elastic membrane first touches on the pin or the fins. This

134 | Internal Combustion Engine Handbook

6606_Book.indb 134

1/19/16 8:31 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.7 Crankcase Venting

causes a substantial portion of the pressure forces acting on the membrane in the area of the flow-off cross section during high intake manifold pressures to be supported directly on the valve housing when the valve is mostly closed. In this manner, a premature closing of the valve at high intake manifolds pressure can be avoided. The solution shown allows, at the same regulating characteristic and same pressure loss over the fully opened valve, a significant reduction of the component dimensions or, conversely, in unchanged component dimensions, a substantial improvement of the regulating characteristic. Apart from that, the components of PCVs must have an unrestricted medium resistance against blow-by gases. The risk of functional faults caused by the valve, also the conduits and oil mist separator, freezing in winter can be minimized by positioning it in the warm areas of the crank gear and designing and arranging the components so that unrestricted condensate drainage is possible. A fundamental problem when using conventional PCVs in the crankcase ventilation is that, because of the throttling of the venting gas flow in the PCV, only a fraction of the large pressure differentials between intake manifold and crankcase—prevalent, in particular, in large map areas of conventional SI engines—can be used for fine oil separation. The oil mist separator insert shown in Figure 7.80 is an example for the effective use of the pressure energy throttled in the PCV for fine oil separation in wide map areas and without significant additional effort. At the PCV shown (see

also EP 1723480), concentrically progressing narrow fins are arranged with a radial distance of approximately 1.3 mm to each other in the flow-off cross section arranged at the front side of the membrane. When the valve is in control position, the elastomer membrane approaches the fin geometry of the flow-off cross section depending on the gas throughput and the pressure differential at the valve. The venting gas that radially flows at high speed into the gap between membrane and fin geometry is sharply deflected during axial flow into the concentrically progressing flow cross sections between the fins so that large portions of fine oil droplets, which are unable to follow the gas flow because of their inertial mass, are separated at the side surfaces of the fins and, with the remaining coarse oil, are recirculated into the crankcase. In large venting gas throughputs or low pressure differentials, the membrane of the PCV withdraws from the fin geometry of the flow-off cross section, reducing the deflection of the flow and thus the separation of fine oil in this area. In these operating points, the focus of fine oil separation moves to two fine oil cyclones installed in series with the PCV and a bypass valve installed in parallel to the cyclones for fine oil separation. Figure 7.81 shows the fine oil separation gradients of this separator unit depending on the differential pressure over the separator for various volumetric flows. Because of the principle of adaptation of the separator geometry to the gas throughput in the impactor PCV in particular, high or very high fine oil separation gradients can be achieved over

Diaphragm

Pure gas

Raw gas Oil drain

Oil drain

Figure 7.80 Sectional view of a fine oil separator insert with impactor PCV, fine oil cyclones, fine oil separating bypass valve, and oil recirculation via oil collection tank and check valve (Hengst).

Internal Combustion Engine Handbook | 135

6606_Book.indb 135

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Degree of separation T_1,0 µm [%]

Chapter 7 Engine Components

Only cyclone Cyclone impactor combination 30 l/min Cyclone impactor combination 50 l/min Cyclone impactor combination 80 l/min

Differential pressure [mbar]

broad map areas of the engine at relatively little constructive effort and will remain maintenance free for the entire service life of the engine. A further optimization potential of this separation concept consists in diameter enlargement of the impactor fin geometries in the flow-off port of the PCV (enlargement of the PCV membrane) and the use of high-efficient small fine oil multi-cyclones and a fine oil separating bypass valve (see Pat. WO 2007000281 Fig. 24, Impacting flow control valves) which sharply accelerates the flow prior to deflection, regardless of the throughput.

7.7.6 Modules and Valve Bonnet Integration

In particular, the continuous efforts in the automotive sector to reduce component and assembly costs, restrictions with

Figure 7.81 Fine oil separation gradient curves of the fine oil separator insert shown in Figure 7.80, with dependence on the pressure differential over the separator at different venting gas volume flow rates.

respect to the available spaces and attempts to reduce the design and logistics interfaces in car production, form the backdrop for function integration and module creation in various functional systems in the vehicle. With respect to the crankcase ventilation system, a corresponding trend is to unify the pressure control, coarse and fine oil separation, and the check valves for returning separated oil into one module. The integration of PCVs and components for coarse oil separation in the valve bonnets of SI and diesel engines has been standard practice for many years. More recently, it is also required to integrate an effective fine oil separation in such valve bonnets. Figure 7.82 presents an example of a valve bonnet model in a passenger car diesel engine with integrated coarse separator, PCV, cyclone separator, and oil return line for the separated oil. Figure 7.83 shows

Cleaned blow-by gas discharge

Blow-by gas inflow

Coarse oil separator

Pressure-regulating valve

Cyclone aerosol separator

Figure 7.82 Sectional view of a valve bonnet with integrated coarse oil separator, PCV, cyclone separator, and oil return line for the separated oil (Woco/ Hengst).

136 | Internal Combustion Engine Handbook

6606_Book.indb 136

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

Pressure control valve Cleaned blow-by gas discharge

Filter bypass valve Lubricating oil supply for cylinder head

Pressure-limiting valve Cyclone aerosol separator

Oil return

Blow-by gas inflow

Lubricating oil filter insert

Oil drain valve for service Oil pressure sensor connection Oil/water heat exchanger

Cooling water preheating connection

Return shut-off valve Pressure-limiting valve Pure oil discharge Return to oil sump

Figure 7.83 Sectional view of a multi-function module with integrated cap oil filter, oil/water heat exchanger, oil diverting valve, temperature and pressure sensors, oil mist separator, and crankcase PCV (Hengst).

a multi-function module for the engine in a utility vehicle with integrated cup oil valve, oil/water heat exchanger, oil diverting valve, temperature and pressure sensors, and an oil mist separating unit with PCV.

7.8 Cylinder Head Great importance is attached to cylinder head design and engineering during engine development. The cylinder head determines, like no other subassembly in the engine, operating properties such as performance level, torque, exhaust emissions, fuel consumption, and acoustic properties. The section which follows provides insights into development work and on current cylinder head design. The key issues dealt with during cylinder head development and in the manufacturing processes are discussed in sequence below. Because of the scope of the material involved, the discussion is limited to passenger car engines; two-cycle engines will not be included.

7.8.1 Basic Design for the Cylinder Head

The engineering designs for the cylinder head have been continuously developed and refined over the past 100 years of engine history. Even today new developments require decisions on what shape and which cylinder head components

should be used in the new design. Current technologies such as variable valve actuation or direct-injection combustion concepts in gasoline and diesel engines take a prominent place in the discussion accompanying the new development of any engine. Not every company in the automotive industry follows the same path, because of differing requirements and the “signposts” set within the companies themselves. As was the case approximately 100 years ago, this is the reason for employing a variety of designs in passenger car engines. The cylinder head contains the fundamental elements used for mechanical control of gas exchange and combustion. The valve timing concept is of particular importance here. In the last 20 years, it is in this sector, in particular, that the techniques and components used in valve timing have become far more sophisticated. The two-valve engines, in which two valves are used for each combustion chamber, have been largely displaced by the more modern, multi-valve engines. In particular, the great increase in volumetric efficiency achieved in recent years demands refined geometries for the charge exchange process. The features inherent to multi-valve technology, such as the use of two camshafts, provide greater freedom in engine management. Variable valve timing is used in almost all modern gasoline engines [7-56]. 7.8.1.1 Layout of the Basic Geometry A number of technical requirements have to be satisfied when laying out the basic geometry for the cylinder head. At the

Internal Combustion Engine Handbook | 137

6606_Book.indb 137

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

beginning of any new development for a cylinder head, it is still possible to influence the individual parameters for a gasoline engine such as valve angle, cylinder head exterior dimensions, location of the gas ports, and the location of the spark plugs. Once the main geometries for these items have been established, the developer’s choices with regard to the remaining cylinder head geometries are limited. Shown in Figure 7.84 are the factors that influence the shape of the cylinder head. If, at the beginning of a new development project, only the overall engine type has been determined (inline or V-block), then it is necessary to find a compromise that considers several factors. These factors include the space available in the engine compartment, installation of the complete engine in that space, other influencing factors (such as the valve train components and their dimensions, and the shape of the gas ports), and requirements stemming from manufacturing, such as the technologies available for casting and mechanical machining. A great deal of experience is required to identify the compromise that culminates in improvements in the objectives set for the engine, such as reducing consumption and exhaust emissions. Not all the paths taken while developing a new cylinder head lead toward the defined goal. This may be the reason why the engines produced in the course of a series exhibit differing cylinder head designs. For example, the number of valves per multi-valve cylinder may vary between three and five in mass-production gasoline engines. Traditionally, the two-valve cylinder head is the most economical solution. Its valve train components are limited to a minimum, with just one intake valve and one exhaust valve. The number of moving parts is small, and its contribution to friction loss is commensurately low. The cylinder head can

Thermodynamics

Engine type, series or V engine for example

Valve train components

be kept compact with regard to its outside dimensions. There is great latitude in selecting the shapes for the gas exchange ports. In addition to, and because of, this design freedom, the component geometries can be better controlled in mass production with respect to the casting models and the shape of the cores. That is why the two-valve engines continue to be widely used in the standard engines, both gasoline and diesel, offered by many car makers. The exact design of the cylinder head varies widely, in accordance with the general design—inline, boxer, or V-block engine—because of the fact that many engine components are mounted on the cylinder block and the cylinder head. These include the intake manifold, exhaust system, camshaft drive, and vacuum and pressure pumps. Only rarely are engineers successful in using one cylinder head, with all its complexity, for a four-cylinder or for a V-8 engine. As a rule, a unique cylinder head has to be developed for each engine variant. It is thus in the pursuit of controlling costs that an attempt is made to use as many identical parts as possible in assembling the various cylinder heads. 7.8.1.2 Determining the Manufacturing Processes The casting process used for the cylinder head should be established very early in the proceedings. It is advisable, once the casting technique has been selected, to take into account the knowledge and expertise available in the model shop and casting department when laying out the basic cylinder head design. Not all the geometries that the engineer might want can be realized with each and every casting process. The development team often faces a daunting challenge in its attempt to boost product quality in the highly complex cylinder head casting while at the same time realizing the complex geometries in the head. In this scenario, it is important to

Number of valves

Costs

Combustion process, SI or diesel

Manufacturing process Variable valve train

Figure 7.84 Factors influencing cylinder head design.

138 | Internal Combustion Engine Handbook

6606_Book.indb 138

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

continuously refine the casting processes suitable for producing cylinder heads. Also considered early in the development phase is choosing the techniques to be used for machining the cylinder head; this depends in part on the numbers to be produced. Here, new designs in particular are subject to severe cost pressures. 7.8.1.3 Gas-Switching Element Design The shape and location of the intake and exhaust ports and the shape of the combustion chamber contribute to the determination of the overall cylinder head geometry. Many studies in this regard are carried out either empirically, through experimentation and trials, or calculated based on 3-D simulations. Flow trials in the ports, carried out using rapid prototyping models, serve to determine flow values. Fabricating singlecylinder engines during preliminary development work makes it possible to respond flexibly to developments at the ports. Depending on the combustion process, either gasoline or diesel, a wide range of basic investigations are conducted prior to defining the geometries. These basic studies are also performed in parallel to cylinder head development. The concept for a diesel should, for example, identify a favorable shape for a swirl-inducing intake port. When exploring a new combustion concept, such as is the case when developing a direct-injection multi-valve diesel engine, it is necessary to test many versions. Only in the course of overall cylinder head development are all the geometries for the components in the cylinder head determined. 7.8.1.4 Variable Valve Actuation As a rule, the implementation of variable valve control concepts makes it necessary to develop new cylinder head concepts. Using camshaft shifters in modern gasoline engines requires adaptation work only at the camshaft drive and for the oil management concept in the cylinder head. Fully variable valve control such as that implemented by BMW in its “Valvetronic” system [7-57] makes it necessary to use cylinder heads developed entirely from the ground up. The

components needed to vary valve stroke length are novel, and extensive adaptations have to be made in the cylinder head geometry. The amount of development work associated with this concept is considerable; several cylinder head construction stages have to be tested before the overall concept can go into volume production. The parameter studies required to optimize gas exchange ports, valve diameters, combustion chamber variations, and timing, as well as the control of valve stroke lengths, are very extensive.

7.8.2 Cylinder Head Design

The cylinder’s bore and spacing determine the basic layout for the cylinder head. As a rule, the number of valves per combustion chamber has already been specified for new engineering. The minimum wall thickness required by manufacturing constraints and the necessary degree of stability narrows the space available for installing valve train components. Because the number of camshafts is specified at the outset of engineering work, it is then necessary to specify the locations and arrangement of the valve train components, taking the geometry of the gas exchange elements such as the ports and combustion chamber, into account. Studies then follow to determine how the rough dimensions of the cylinder head change when parameters such as the valve angle, unrestricted valve flow area, or design of the gas exchange ports are modified. 7.8.2.1 Rough Dimension Determination One method to determine the cylinder head geometry is the creation of a rough design for valve train components. For this purpose, CAD files are used, enabling the parameterization of the individual geometrical dimensions of the components. By varying the dimensions such as the valve angle, valve spring installation dimensions, position of camshaft, or spark plug bearings, the geometrical effects on the overall concept can be roughly analyzed. Figure 7.85 displays rough dimensioning for a parameter study used in the design of a five-valve cylinder head with pushrod [7-58]. The cylinder head is designed with

124,5

80,3

129 a4,5

5 4

1

a4,5 = 20,2°

5

2

4

3

a1,3 = 21,6° a2 = 14,9°

130

139,5

139,5

a2 a1,3

1

4

3

a1,3 = 20,5° a2 = 14,5° a4,5 = 0°

5

2

a1,3 = 23°

a4,5 = 19,5°

1 2 3

a2 = 15,9°

Figure 7.85 Study on the basic geometric design of a five-valve cylinder head [7-58].

Internal Combustion Engine Handbook | 139

6606_Book.indb 139

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

three intake and two exhaust valves. The spark plug is shown at the center of the combustion chamber. Indicated beneath the cam geometry shown, there is the installation space required for the pushrods. The locations of the head bolts, which also require a certain amount of free space for installation, restrict the latitude for varying the valve angle. Accessibility to the head bolts after the head has been completely assembled is mandatory for almost all engines because of manufacturing and maintenance requirements. Illustrated in the figure at the center, for example, is the situation in which, with a vertically suspended exhaust valve having a valve angle of 0°, the head bolts are located outside the camshaft axis for accessibility. In a V-block engine, this type of cylinder head design provides more space on the exhaust valve side for the design of the exhaust components. Exhaust routing in the manifolds could be optimized. These studies help in cylinder head development by allowing better evaluation of the overall effects on the engine. Using parameterized assumptions in the CAD system can, particularly in this development phase, make it possible to examine the basic cylinder head geometry with regard to its effects on the engine as a whole. Concept comparisons between pushrod and cam follower designs can also be carried out very well in this way. One criterion for selecting the valve angle and the location and size of the valves is the determination of the unrestricted flow area around the valve disk. This is the unrestricted area available for gas exchange, as a function of the valve stroke, as described by Dong [7-59]. To influence engine breathing, an attempt is made, in coordination with the remaining potential geometric configurations for valve train components and gas exchange runners, to make this area as large as possible. Structural requirements and values resulting from experience—such as the width of webs between the runners—have to be maintained. In basic examinations of the geometric layout to preassess the situation regarding valve angle geometry, it is

7.8.2.2 Combustion Chamber and Port Design The geometry of the combustion chamber is of major significance in cylinder head engineering. For this purpose, technical calculations are executed simultaneously in the early development phase. The geometries to be developed for the combustion chamber variant are determined prior to the actual concept determination. In coordination with the portion of the combustion chamber volume accounted for by the bowl at the top of the piston, extensive basic examinations are performed. Concepts such as charge stratification in directinjection gasoline engines are assessed in conjunction with the port and combustion chamber geometries and are tested on real-world models. Three examples for the development of a two-valve concept with various combustion chamber

View

Anwev

A

Dzk Xzk

Aevr

Hev,x

Hev

Hav

A

Ae v

Dav ax

ay

Hav,y

ae

v Aa

aa

Hzk

v VLa

VLe v

Aavr

Dev

Ae vzk

vzk Aa

Dnw

Anwav

possible to compare variations with one another both quickly and simply [7-60]. Concept studies using various numbers of valves can be carried out quickly and easily. To ensure that these studies can be completed quickly, in the early phase of cylinder head conceptualization, simple PC programs should be used, as is mentioned in [7-58] and [7-60]. Depicted in Figure 7.86 are examples of the parameters that are pertinent to the basic design of a six-valve cylinder head. Minimum web widths between the valves have to be maintained for both cooling and cylinder head strength. One objective here is to incorporate the largest possible valve diameters. The results of such examinations are geometric magnitudes such as the utilization of available surface areas. This term is understood to be the quotient of the total intake or exhaust surface to the surface area for the cylinder bore. The results vary in dependency on the cylinder bore; when interpreted, this information gives differing numbers of valves. This phase of cylinder head development is particularly exciting because specifying the number of valves at a predetermined cylinder bore has decisive impact on cylinder head design.

Aevav

Figure 7.86 Study to establish geometry for the valve cross section [7-60].

140 | Internal Combustion Engine Handbook

6606_Book.indb 140

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

Figure 7.87 Combustion chamber variants for a two-valve cylinder head.

designs are shown in Figure 7.87. The rough geometry of the combustion chamber is also determined by the variation of the valve angle. In this example, in the interest of better comparison, the same cam follower design was used for all three embodiments. Among the matters examined was the extent to which the charge volume available burned most favorably. The overall influence becomes clear in specific consumption values, the amenability to leaning out the mix, and, in particular, raw NOx and hydrocarbon emissions in the exhaust gases. The situation shown in the right-hand depiction proved to be advantageous. The spark plug, extending well into the combustion chamber, is arranged so that it is totally surrounded by the mix drawn into the combustion chamber. In the design selected here, approximately 70% of the combustion chamber volume is inside the cylinder head and 30% is in the piston. The interdependencies described here between combustion chamber geometry and the effects on the engine are to be found again in direct-injection gasoline engines now under development where fully variable valve control is used. The development effort required there is considerable. The parameter studies to be defined for combustion chamber trials demand a great deal of experience and development discipline by the thermodynamics engineers. Four-valve cylinder heads with the spark plug at the center offer the fundamental advantage of short combustion paths in the combustion chamber. Because of the valve head’s large share of the total surface that defines the combustion chamber, the casting contour has only a slight influence on the volumetric tolerances, which can be kept very narrow at, in one example, 0.5 cm3. To reduce thermodynamic losses during combustion, one strives to achieve the lowest possible ratio of combustion chamber surface area to combustion chamber volume. One key thrust in development is optimizing the geometry of the squish surface. Here the location varies in relationship to the valves’ shape and size. An excessive share of squish area has proven to be detrimental because of the increase in the surface-to-volume ratio and the associated heat losses. Using the example of the four-cylinder engine shown here, a squish

surface share of 7% proved to be favorable. In modern, fourcylinder engines with external fuel mix blending, the trend is toward flat piston heads with the bulk of the combustion chamber located inside the cylinder head. With the development of new ignition concepts such as direct-injection gasoline and diesel engines, the development of the ports has become a science in itself. Attaining specific, reproducible charge flow is the subject of many basic research projects that are taking place parallel to overall cylinder head development. The design of the port has to be seen in conjunction with the designs for the intake and exhaust manifolds. This topic is dealt with primarily through trials and flow simulations. Here the engineer pays attention to finalizing these geometries early in the work because changes at the port can often trigger major changes in the cylinder head. Often so many thermodynamic interactions occur while defining the geometries for ports and combustion chambers that it is difficult to estimate how much time engineering will take. Figure 7.88 represents possible port arrangements for a diesel engine with direct injection [7-61]. In diesel engines, the incoming air is forced into a swirl motion to intensify the mixture formation. There are two basic options for intake port design that may be drawn upon here: •• helical (swirl or helical port) •• sloped port configuration (tangential port). In selecting the shape for the port, one pursues the objective of achieving the required swirl characteristic and the best possible flow throughput. This effect is to be preserved to mass production. In the swirl port shape, the port imparts the swirling motion on the incoming air. This results in smaller swirl deviation at relatively less favorable flow throughput values. In the tangential design, in contrast to the above, the incoming air is set in rotation by the cylinder wall, because of the port’s off-center location. Typical here are high throughputs at good cylinder fill. Combining a swirl chamber with a downstream tangential port is thus a very good compromise in the conflict of goals between throughput and swirl stability. The “helical port design, oriented vertically from the top to the combustion chamber,” as shown in Figure 7.89, improves

Internal Combustion Engine Handbook | 141

6606_Book.indb 141

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Exhaust side

Intake side

Figure 7.88 Intake and discharge port variants for a four-valve diesel engine [7-61].

port quality when compared with an arrangement at the side. Additionally, the glow plugs can be situated on the colder side of the cylinder head, where the thermal load is less. Because of the short run of the discharge ports in the cylinder head, heating is limited to a low level [7-61]. The described port configuration also allows a symmetrical valve image with positive effects on the arrangement of the valve train (Figure 7.89).

Figure 7.89 Arrangement of gas exchange ports in the cylinder head [7-61].

7.8.2.3 Valve Train Design As to which valve train concept is the best for a particular engine will not be discussed at this point. The engines’ requirements profile—which depend on its use—results in differing engineering strategies and, in turn, in differing valve train concepts. One does observe, however, a trend toward roller-actuated cam followers or rocker arms in passenger car series engines. These designs have the lowest friction valves for the individual valve trains. But these solutions, in comparison with sliding cam follower concepts, are heavier; consequently, they are not used in sports car engines, for example. The masses in motion are to be kept as small as possible and elasticities are to be minimized. For this reason, concepts with mechanical valve play adjustment are used in sports car engines. The design of the valve train takes high priority in cylinder head development. In new developments, pushrod concepts have proven their superiority to cam follower concepts. The installation situations for the valves are different. Different valve guide lengths have been worked out through time for pushrod and cam follower heads. Cam follower timing requires a better, and thus longer, valve stem guide than a pushrod concept because the pushrod itself has a guide. The valve length, in turn, results from the installation length required for the valve spring. During new developments, these mutual interdependencies result in increased employment of simulation techniques during the pre-development phase to keep the number of prototypes required for testing as low as possible. Taking the refinement of the valve train for a BMW six-cylinder engine as an example, Figure 7.90 shows the development steps undertaken over several model years in efforts to reduce the weight of the valve train. To keep the valves from lifting at high engine speeds, the valve spring has to be built for a minimum force of F1, and the shape of the cam lobe has to be selected to suit. The required spring force and the associated spring geometry determine the minimum installation space for the springs. To limit spring force F2 at maximum valve stroke, the primary

142 | Internal Combustion Engine Handbook

6606_Book.indb 142

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

Year of manufacture 1990

Year of manufacture 1993

Year of manufacture 1995

Spring

Ø 35, 80g 15 g 1,5 g Cylindrical double spring

Ø 35, 65g 11,1 g 1g Cylindrical single spring

Ø 33,48 g 7,9 g 1,0 g Conical single spring

Valve Sum

69 g x ½ = 34,5 g Ø Skirt 7, 58g 189,5 g

51 g x ½ = 25,5 g Ø Skirt 6, 46g 148,6 g

40,5 g x 1/3 = 13,5 g Ø Skirt 6, 46g 116,4 g

Cup Spring collar Valve keeper

Base: 2,5-l-Engine at intake valve side

thrust in valve train engineering is to keep down the masses that act on the valve. 7.8.2.4 Cooling Concepts In discussing cooling for the cylinder head, differentiation is made where water cooling is used, among cross-flow cooling, longitudinal flow cooling, and a combination of these two types. In cross-flow cooling, the coolant flows from the hot exhaust valve side to the intake valve side; in longitudinal flow cooling, the coolant flows parallel to the long axis of the cylinder head. The objective in cooling is to equalize temperature distribution within any cylinder head segment at a low level and to create uniform cooling conditions for all the cylinder segments. Moreover, the top of the combustion chamber and the valve webs are to be generously supplied while, at the same time, keeping pressure loss throughout the cylinder head flow pattern as small as possible. The coolant passes from the engine block, through several transfer ports and the head gasket, into the lower face of the cylinder head. The shape, location, and size of these transfer ports have to be harmonized appropriately. The coolant flow calculations described in Section 7.8.2.9 represent the state of the art. Only by simulation, problem areas such as the webs between the exhaust ports or the area around the spark plugs can be engineered for complete reliability. 7.8.2.5 Oil Management Motor oil under pressure, used to lubricate the cylinder head, is generally delivered by the oil pump in the engine block, through transfer ports in the head gasket. The oil passes through lateral bores or special supplementary lines to the points served, such as the camshaft bearings, hydraulic valve lifters, hydraulic valve play compensating elements, camshaft shifters, or oil spray nozzles above the cams. The pressurized motor oil supply for the cylinder head is managed by the cross sections of the supply tubing and specially provided choke points to keep the oil consumption to an absolute minimum. To keep the hydraulic valve play adjusters and the camshaft shifters from running dry, check valves are provided in the lines supplying these elements. Multi-valve cylinder heads, because of their greater number of lubrication points, are more difficult to coordinate and involve greater oil requirements.

Figure 7.90 Development steps for reducing weight in valve train components [7-62].

A more powerful oil pump is often required where camshaft shifters are employed. In spite of this, it has been possible in recent years to keep the total oil volume, even in multi-valve engines, within reasonable limits. This goal was met with higher precision in machining to minimize play, through more precise tuning of the oil circuit and through technical calculations. The oil flows back to the sump through return bores of appropriate size, located between the cylinder head and the engine block. These returns are situated at the lowest possible point, which depends in part on the engine’s placement when mounted in the engine compartment. The rotation of the camshafts in some cases slings the oil so severely that it foams. Accordingly, sufficient cross sections are also provided in the area below the camshafts to ensure draining toward the engine block. Particularly in boxer or V-block engines, it is necessary, because of the installation placement for the cylinder head, to engineer the design to ensure sufficiently large drain cross sections. 7.8.2.6 Engineering Design Details The cylinder head bolts are normally bolts with collars. Here the collar, because of the surface pressure to be transferred between the bolt contact surface area and the cylinder head, is broader than the bolt head itself. In monolithic cylinder heads, this can impose limitations on the camshaft arrangement. The diameter of the tool used to tighten the bolts or the outside diameters of the bolts themselves thus determine the locations of the camshafts if the latter are to remain in place inside the cylinder head while the cylinders are being installed. In some cases, the cylinder heads are made in two or more sections, and the valve timing elements are borne by one or two separate cast components. In this case, the design of the lower cylinder head section is simpler, as is the casting technology. Because of cost considerations, monolithic cylinder heads are used in most passenger car engines. Depending on the combustion process, appropriate space must be provided in the cylinder head to accommodate spark plugs, glow plugs, or injection nozzles and the diameters of the tools used to install and remove them. Wherever possible, spark plugs should be selected which use commonplace thread diameters and wrench sizes. In diesel engines or

Internal Combustion Engine Handbook | 143

6606_Book.indb 143

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

direct-injection gasoline engines, the arrangement of the cylinder head components is tight, particularly where a multivalve concept is used. It is for these reasons that the number of valves per combustion chamber is limited to four. The space required for these components can be modeled by assigning parameters using 3-D CAD when defining the basic layout for the cylinder head. This makes it easy to depict potential geometric arrangements. The wall thicknesses required around these components in the rough cylinder head casting reduce the overall installation space for the valve assembly or the camshafts. The cross sections required for cooling are also limited in the same fashion. Modern multi-valve engines incorporate camshaft shifters in the cylinder head. The systems in mass production are all located on the camshaft drive and are driven by the crankshaft by either a timing belt or a timing chain. Suitable oil supply lines have to be built into the cylinder head to serve the shifter. This is simpler when a cylinder head is developed from scratch. The space required by the shifter is not particularly great for the vane-type system normally used today. With these shifters, the adjustment angle of the camshaft can be continuously variably rotated to the crankshaft [7-56]. The diameter size of the drive sprockets of the camshafts is responsible for the minimum camshaft distance. Particularly where the camshafts are driven directly by the crankshaft, this distance has great influence on the cylinder head design. Often, and in multi-valve engines too, the camshafts are driven by intermediate gearing. When using camshaft shifters, however, drive directly at the end of the cylinder head is the most economical. In this camshaft drive concept, the clearance between the camshafts is of appropriate size or an intermediate gear is used between the crankshaft and the camshaft. The most widely used arrangement is with the camshaft drive at the forward end of the engine, that is, at the end opposite the clutch. Drives centered between the cylinders are seldom used in passenger car engines, while they are being seen more frequently in motorcycle engines. Drives at the clutch end of the engine are also unusual. 7.8.2.7 Engineering in Construction Steps It is impossible to predict all the influences that are encountered while engineering the cylinder head, particularly when new combustion processes are being developed. Computer assistance in the basic design or the calculation processes used in simulation technologies do, indeed, help to generate a great deal of information in advance. The mutually influencing factors on cylinder head development are very complex, however, so that there is much to be said for using several construction stages in cylinder head development. Moreover, testing the engine’s thermodynamic and mechanical properties delivers many findings that also cannot be predicted in advance (Figure 7.91). When developing entirely new cylinder head concepts, it may make good sense to obtain cylinder heads as prototypes—quickly and economically—for use in preliminary development. When building these prototypes, it is often advisable to use manufacturing techniques that differ from

those used in mass production. Thus, small numbers of cylinder heads to be used as prototypes may be built in a low-pressure sand-casting process.

Project release Project organization Cost analysis Manufacturing specification

Concept determination Concept selection

Construction-related preliminary examinations

Concept analysis and determination

Initial testing Calculations

Design version I

Design version II

Calculations

Head manufacture with series tooling

Prototype construction Version I Mechanical testing

Thermodynamic testing

Mechanical testing

Thermodynamic testing Endurance testing Release

Initial endurance tests, if appropriate

Pilot series manufacturing Series manufacturing

Figure 7.91 Example of development steps for cylinder heads in two construction steps.

Figure 7.92 shows an example of the flowchart for fabricating cylinder head prototypes by this procedure. Smaller companies have specialized in this particular field and are able to deliver initial prototypes, quickly and economically. 3D CAD data cylinder head

Laser-sintered sand cores

3D CAD model cylinder head blank

Solidification simulation

Construction: Models, molds, cores, core packages

Milling programs

Casting unit manufacture

CNC machining

CNC machining Quality assurance

Mechanical machining

Figure 7.92 Example of a flowchart for making up cylinder head prototypes according to Becker [7-63].

To reduce the overall cylinder head development time, the goals to be met in any given construction stage must be defined exactly. The project management work required here is of vital importance. As a rule, the development of

144 | Internal Combustion Engine Handbook

6606_Book.indb 144

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

the second construction step is commenced while the first construction step is still being tested. Here the manufacturing processes foreseen for mass-production use should be employed. Particularly, the rough cylinder head casting should be made up using the casting process selected for volume production. The development of a cylinder head to readiness for mass production in a single step is possible for designs based on existing heads and exhibiting only minor modifications. 7.8.2.8 Using CAD in Engineering Because of the multiple uses of CAD data, cylinder heads are modeled in the CAD systems in complete, 3-D renderings. The specifications for the model and the casting equipment can be derived from these data. The geometries can also be used for simulation calculations. When engineering a new cylinder head, interdependencies among its components can be parameterized. This makes it possible to carry out basic studies simply and quickly. Model builders and casting specialists should be consulted continuously during detailed engineering work, beginning as soon as the rough cylinder head concept has been finalized with the definition of the internal components and the major dimensions. In this way, manufacturing considerations are accounted for in the process early. Engineering methods vary, depending on which CAD system is used. It makes sense, for example, to limit parameterization of the cylinder head to a few parameters to maintain flexibility when changes are made to the model. All the engineers involved in the project should use identical software with identical default settings. Because of the complexity of the CAD methods, one person on the development team should be responsible for adherence to the methods. Because the cylinder head involves many interfaces to adjacent components, transfer conditions to these components have to be defined. The consistency of the CAD process provides many advantages. Data become more reproducible, can be used more easily for series of cylinder heads, and largely preclude any inaccuracies between engineering and manufacturing. Cylinder head engineers who prepare the overall concept for a new component need a great deal of practical experience. Today the designs are generated completely using CAD. 7.8.2.9 Computer-Assisted Design A number of calculation methods are now used to dimension the cylinder head geometries [7-64]. With calculation efforts in the early stage—beginning with the concept phase—the calculation results are applied in the initial cylinder head prototypes. This makes the steering of subsequent development steps more effective, and in this way, the number of components used in testing can be reduced. Ongoing verification of calculations against test results continues to be necessary. Computer support ranges from rough component dimensions and detailed design to optimization and simulation calculations. The target criteria for new engines—improved environmental compatibility, reduced exhaust emissions and fuel consumption, improved performance, product quality, and ride—can be better satisfied through technical calculations. Before the first prototypes are fabricated, the calculations are devoted primarily to specifying the valve, combustion

chamber, and gas exchange port geometries. To a greater extent, 3-D CAD data for the head geometry, once it has been prepared, can be used directly for technical calculations. During the development of a cylinder head in construction stages—understood to be differing cylinder head component development stages—technical calculations start right at the outset of development. During the course of development, the largest share of calculations is performed in the first construction stage. The goal here is to provide support in identifying and defining the concept for the main cylinder head geometries. During testing in subsequent construction steps, technical calculations are used more to lend precision to the concept and to specify details. Calculation activities decline the closer the design gets to mass-production launch. At this point, we mention briefly only a few activities that play a vital part in dimensioning the cylinder heads. Technical calculations contribute to making it possible to interpret, in a more understandable fashion, the complex processes involved in cylinder head development. The PROMO10 program [7-65] is used to calculate the gas charge exchange. Here dynamic gas flows in the intake and exhaust systems of aspirated and turbocharged systems are calculated. The gas exchange components in an engine, with its intake and exhaust systems, are assembled to form a virtual model. Events associated with flow, such as pressure fluctuations or mass flows, can be analyzed at various points in the engine. The program provides information on the characterizing values to be expected for the engine, such as charging efficiency, maximum torque, or power output for a particular engine configuration. The core for calculations is embedded in an interactive graphic user interface from which data record conditioning and result evaluation are undertaken. By establishing the geometry for the ports in the cylinder head, the PROMO program is particularly well suited for initial dimensioning of the gas exchange components in the early concept definition phase and, in particular, for laying out the timing. Cost-intensive trials can be minimized in, for example, the development of cylinder head concepts with variable valve actuations. For the engine development, the program also provides conclusions about •• intake manifold dimensions •• concepts for switching and resonance intake manifolds •• evaluation of cam lobe contours and timing •• estimating the potentials of various concepts for variable valve timing •• evaluating different port shapes •• exhaust manifold design with regard to pipe length and diameter. In addition, 3-D flow simulations are conducted to design the intake and exhaust ports and the combustion chambers in the cylinder head and pistons. The charge motions are simulated on the basis of the CAD description of the port and the combustion chamber surfaces. The calculations provide insights into the flow situation in the intake and exhaust ports,

Internal Combustion Engine Handbook | 145

6606_Book.indb 145

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

as well as for the charge as it flows into the cylinder. Solving the equations makes it possible to simulate the complex flow processes for static situations and for those which change through time. When dealing with transient calculations (i.e., for states that change through time), the calculation network to be prepared is modified at each timing phase in accordance with momentary valve and piston positions. The results of the simulation—which include pressures, velocities, turbulence, and blending values—have to be assessed with an eye toward perfect combustion. Figure 7.93 shows an example of the calculation results for an intake valve at center stroke position; reproduced there is the velocity distribution for the charge as it flows into the cylinder (here 90° after intake TDC). The 3-D flow simulation is helpful, particularly when developing new combustion processes. Swirl or tumble effects can be better analyzed and further refined in accordance with the findings.

head with cross-flow cooling. The cylinder head receives the coolant through transfer bores in the cylinder head gasket. Their graduated diameters ensure nearly identical distribution of coolant to the various cylinders. About two-thirds of the coolant passes into the cylinder head at the exhaust valve side. The coolant flow passes across the top of the combustion chamber and past the exhaust ports to the spark plug well. Behind the spark plug, the flow continues along a center coolant collector channel that runs longitudinally through the cylinder head. Figure 7.94 shows an example of the results of a simulation calculation, in the depiction of the convective heat transfer coefficients in the area around the exhaust port, which is subjected to severe thermal loading. The dark areas correspond to a high thermal transfer coefficient, a result that is achieved by the optimized position and selection of the diameters for the transfer bores in the head gasket. By optimizing the cylinder head cooling pattern, with the support of simulation calculations, the temperature level at all the cylinders can be kept constant, with only minor deviations. This method makes a major contribution to cylinder head development, which could be achieved using conventional techniques only with extensive effort in trials and testing.

Figure 7.93 Flow simulation at an intake valve [7-64].

The design of the valve lifting lobes and the simulation of the valve train dynamics take a prime position in cylinder head development. The findings here have a direct impact on cylinder head design. Geometries such as the pushrod diameter, valve length, valve stem diameter, valve spring dimensions, and finger follower geometry are determined by these calculations. Imaging of the entire valve train in mechanical models also makes it possible to determine precisely the dynamic properties. These findings are directly applied to the geometry of the camshafts or valve train components [7-66]. An essential contribution to the design of the coolant water chamber of the cylinder head is provided by the 3-D flow simulation of the entire coolant circuit [7-64]. This method is integrated in a larger calculation framework serving in the optimization of the entire engine cooling system, including water pump and cooler design. The geometry of the cylinder block and head cavities through which coolant flows is modeled and then compiled in a calculation matrix. Figure 7.94 shows the section of the water jacket as an example of coolant flow simulation in a five-valve cylinder

Figure 7.94 Section of the water jacket for coolant flow simulation [7-64]. See color section page 1068.

Strength calculations represent a major area where technical calculations are used in engine development to determine the dimensioning and geometries of cylinder heads and their components. To keep cylinder heads at as low a weight as possible and, at the same time, sufficiently stiff, FE calculations of the entire cylinder head are performed [7-64], [7-67]. The structural strength of camshafts and their bearings can, for example, be examined with respect to the shape and position of the camshaft bearings. Wall thicknesses can be minimized by using strength analysis. Stiffening ribs are provided to increase structural strength. Thus, designs with a favorable effect on force flow can be predetermined in detail. Figure 7.95 displays a section of the FE model of a complete cylinder

146 | Internal Combustion Engine Handbook

6606_Book.indb 146

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

high

low

head [7-67]. The stress variables used for the calculations are the spring and mass forces of the valve train, belt and chain forces at the camshaft end, and the forces arising from cylinder head bolt. Figure 7.95 shows the comparative stresses according to Mises at the deformed cylinder head when subjected to thermal loading at nominal power. Because of intense demands for reliability and smooth running at the valve train, the design of the lobe contour is of great importance. In addition to the purely kinematic design of cam contours, various computer programs are used to ensure good dynamic behavior in the valve train. To conduct the simulation calculations, the valve train structure is expressed as a multi-body oscillating system with adjustable coupling conditions for friction, stiffness, damping, and degrees of freedom in movement. The dynamic simulation for the entire valve train is obtained by calculating the design of individual valve systems to better evaluate the interactions of individual components with one another. The valve train is actuated by the lobe contour. Stiffness is determined on the basis of measurements made at the actual components or by using FEM calculations. The damping values are primarily experience values that are determined by comparing calculations and measurements. The valve spring, as the main vibrating or oscillating element, is broken down into many oscillating subsystems. One goal in dynamic calculations is demonstrating rotation speed strength for the valve springs at the smallest possible valve spring forces, to keep the overall valve train friction as low as possible. Simulation calculations make it possible to estimate, even at a very early development stage, the interactions among individual components. Closely defined changes in component properties make it possible to influence the overall structure of the cylinder head and its components in such a way that the components’ own shape properties are manageable within the valve train’s excitation spectrum. Suitable tuning for the actuation itself, which is determined primarily by the cam contour, can also bring about a marked reduction in the dynamic effects at the valve train.

Figure 7.95 Strength analysis at the cylinder head [7-67]. See color section page 1069.

Oil circuit calculations can be conducted to fine-tune oil management in the cylinder head [7-64]. Calculations for subsystems, such as for oil management at the cylinder head, make it possible, by simulating the entire motor oil supply system, to minimize the amount of oil required. This, in turn, keeps the amount of power consumed by the oil pump as low as possible. To do this, all the components in the engine in which oil is found are modeled in a virtual hydraulic system. The objective is to optimize by simulating the oil using points in the cylinder head, such as the push rods, camshaft bearings, camshaft shifter, and oil spray nozzles. The calculation models are further refined, incorporating the results of basic experiments. These preliminary calculations make it possible to predetermine with considerable accuracy the cross sections for the oil passages; this reduces the number of costly trials that would otherwise have to be undertaken using the complete engine.

7.8.3 Casting Processes

Cylinder heads for internal combustion engines place considerable demands on the mechanical properties of the materials in a temperature range beyond 150°C. The design latitude for the geometries in the cylinder head is severely limited by the components to be used in the cylinder head. Particularly when developing new cylinder heads for direct-injection diesel engines, the complexity in the type and magnitude of the stresses occurring during operation has risen considerably. To satisfy these more exacting requirements, the materials available for use have to be optimized and further developed. Any of a variety of materials may be used for cylinder heads, depending on the requirements profile and the casting process used. In addition to aluminum, cast iron materials are also used for industrial engines and utility vehicle power plants. In passenger car engines, aluminum is used almost exclusively, with just a few exceptions. Cylinder heads may be manufactured both from primary alloys—aluminum extracted from ore at the refinery—and from recovered alloys—that is, recycled aluminum following melting and purification; these

Internal Combustion Engine Handbook | 147

6606_Book.indb 147

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

may be delivered as ingots or as liquids. Aluminum casting alloys are also used for heavily loaded diesel engines with direct injection, but not all the available casting techniques may be used with these cylinder heads. At ignition pressures exceeding 150 bar, it is necessary to use alloys that satisfy the following stringent demands: •• high tensile strength and high creep resistance between room temperature and elevated temperatures of approximately 250°C •• great thermal conductivity •• low porosity •• high ductility and elasticity at great resistance to thermal shock •• good casting properties at low susceptibility to heat fissuring. The central area of the cylinder head near the combustion chamber and, in particular, all the webs located near the exhaust ports are subjected to severe thermal loading in a range of approximately 180–220°C, in addition to mechanical loading [7-68]. The casting technique should be determined as soon as the concept for a new cylinder head is finalized. An early evaluation by the model shop and the casting department helps to avoid errors in the engineering phase. The task of the casting department is to influence cylinder head design to optimize casting for the rough component. The filling and solidification processes in the casting procedure are assessed largely from simulation. These 3-D calculations give the casting department valuable information on problematic areas that might be anticipated right from the conceptualization phase. The geometry of the cylinder head can be modified to accommodate these areas before the first prototype is built. Considerable cost savings can be realized in the development process in this way. The basic casting techniques used for engine blocks can also be used for cylinder heads. A brief review of the most commonly used casting techniques is provided below. 7.8.3.1 Sand Casting Models and core boxes made of hardwood, metal, or plastic are used to replicate the later engine block casting inside the sand mold. The casting molds are normally made of quartz sand (either natural or synthetic sand) and binders (synthetic resin and CO2). The sand cores are formed in core-casting machines into which the sand is introduced under pressure; the mix of sand and resin is compacted to create the core by applying heat. It is advisable to use laser sintering processes when making sand cores in the prototype phase. Combining individual cores to form a core package and assembling this core package and the outer casting mold are handled mechanically and fully automatically, even when producing only limited numbers of castings. Model, core, and mold parting in various planes, and inserting cores in the casting mold, make it possible to produce complex cast components with undercut areas. During the casting process, the hollow cavities between the outside mold and the cores are filled with molten metal. Following the filling process and after the metal has solidified, the casting is removed from the sand mold. The sand mold is

destroyed here (which is why this is referred to as a “lost mold” process). Following casting, the rough part is cleaned, and the gate and risers are separated. In mass-production operations, these steps are fully automated. Sand-cast components made of Al–Si alloys permit double heat treatment. The first heat treatment phase is found in the controlled cooling period for the casting inside the sand mold. The second heat treatment occurs during time- and temperature-controlled storage of the casting in a kiln. These heat treatments increase the strength of the cast component and relieve inherent stresses created during the cooling process. The geometry of the components may include undercuts because the lost mold is used for only a single casting. One advantage of sand casting is that the fabrication equipment can be set up quickly and economically when making small numbers of units. Cylinder heads for special types of engines, such as sports car engines, can be quickly realized; implementing changes during development is relatively simple and economical because plastic positives are used. The low-pressure sand-casting process is suitable for prototypes and short production runs. Here the melt is introduced from below, through a riser, and into the sand mold; pressure at approximately 0.1–0.5 bar is applied to the molten metal (Figure 7.96). This pressure is maintained during casting. Because solidification under pressure is almost directional, the structures in the cylinder heads are very fine.

Mold cavity Sand space Nozzle

Gas pressure Riser Warming oven

Figure 7.96 Low-pressure sand-casting process.

The Cosworth low-pressure sand-casting process is also used for cylinder heads, because of its great dimensional accuracy and strength, a compact structure, and freedom from pores. In accordance with the specifications for the process, an aluminum alloy, in the form of assayed ingots, is melted in a resistance electric furnace under a blanket of inert gas (Figure 7.97). The melts are buffered in a generously dimensioned holding kiln, once again blanketed with

148 | Internal Combustion Engine Handbook

6606_Book.indb 148

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

Shipping

Heat treatment

Mold assembly

Cleaning Casting station

Cast parts

Mold and core manufacturing

Cooling path Emptying station

Electromagnetic pump

Aluminum block metal

Melting and warming oven

Dry comminution of the sand

Fluid bed sand recycling

Figure 7.97 Casting process developed by Cosworth.

inert gas. Casting is affected with an electromagnetic pump that moves the molten aluminum upward to the sand mold, where it flows from below into the mold cavity. Just as in the low-pressure die-casting process, the pressure on the molten metal is maintained during solidification. Programmable regulation of pump output makes it possible to set a delivery rate suitable to the particular shape of the cavity. Casting can be automated to a great extent; the finished molds are moved one after another to the casting station, above the electromagnetic pump. The core package process has been used to manufacture cylinder heads for approximately 20 years now. In this sandcasting process, a closed sand core package is assembled from several individual sand cores. Adhesives are normally used to hold these together, but screws may also be used. Core packages are used for cores of complex design, which cannot be made up in a single piece. In its original embodiment, the core package process, based on the low-pressure die-casting principle using an electromagnetic pump, was limited to short production runs for cylinder heads because of its low productivity. The latest approaches also point out perspectives for using this process in mass production once manufacturing facilities have been modified appropriately. The cast components do not fall below a temperature of approximately 500°C after casting through the complete removal of the sand. Thus, they are cast virtually free of strains, giving the parts superior dimensional accuracy. Because each part is cast in a new, cold mold, practically no dimensional deviations are found such as those that occur in die casting, where the permanent molds are subject to wear. 7.8.3.2 Die Casting Approximately 90% of the cylinder heads made in Europe are manufactured by the die-casting process. The dies are permanent metal molds made of gray cast iron or hot-work tool steels and used to manufacture cast parts from lightweight

alloys. Just as in sand casting, the sand cores are positioned inside the casting mold. Die casting can be subdivided into the gravity and low-pressure processes. In gravity casting, the mold is filled solely with the force of gravity acting on the molten metal and at atmospheric pressure. The casting process is used in partially or fully automated casting systems. In this casting process, in contrast to sand casting, the dies can be used many times. It is necessary only to make new sand cores for each casting cycle, which is referred to as lost-core casting. Because of the use of sand cores, die casting, like sand casting, offers the advantage of greater freedom in the engineering design. Undercut areas are possible, in contrast to pressure die casting. Using steel as the die promotes fast and directional solidification of the molten metal; this is not the case in sand casting. The die is protected against the lightweight metal melt by applying a parting agent, also referred to as a refractory coating. In comparison with sand casting, the die-cast components exhibit a finer internal structure, greater strength, improved dimensional accuracy, and better surface quality. Both die castings and sand castings can be further processed with double heat treatment. In addition to the advantages of carefully-defined control of cooling inside the die, which is the first heat treatment, additional heat treatment is often implemented. Unlike sand casting, there may be no undercuts in the permanent molds because they are used over and over. Most of the cylinder heads of the Volkswagon (VW) group, for instance, are manufactured using this process. The combustion chamber side of the cylinder head is cooled by inserting one steel die per cylinder. The sprue is at the upper side of the cylinder head, and the molten metal fills the mold as it flows downward from this point. The area around the combustion chamber cools faster because of the cooled combustion chamber dies, and this increases strength in that specific area. The casting process takes place on a turntable system with several stations; this reduces mass-production manufacturing

Internal Combustion Engine Handbook | 149

6606_Book.indb 149

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

costs to a minimum. The standard alloy used for this purpose is G-AlSi7MgCu0.5. Smaller runs are outsourced to suppliers. Similar processes are used there, and in some cases, the cylinder heads are cast from below using special runners. The results are comparable in terms of the quality found in the final product. Many cylinder heads are also produced with low-pressure casting, as is the case at HONSEL in Meschede, Germany. This is one of the processes the casting department at BMW uses for their diesel engines and a majority of their gasoline engines. In much the same way as described above, the inductance-heated melt is pressed into the mold through a riser at a pressure of approximately 0.1–0.3 bar. The combustion chamber, at the bottom of the mold, is filled from below. Here, too, the combustion chamber plate is cooled with air or water. The cavities for water and lubricating oil and the geometry required for the camshaft timing chain are formed with sand cores. The remainder of the geometric forms in the cylinder head are shaped with dies. Thanks to the low-pressure casting process, the surfaces at the cylinder head are densely compacted. This process is particularly good for diesel cylinder heads that are subjected to heavy loading. A technique developed by VAW Mandl and Berger is known as the Rotacast process. The entire mold is rotated during the casting process. This process is intended to achieve turbulencefree mold filling. The form is filled from below and, during filling, is rotated through 180° within a period of 15 s. The charge passes into the mold through several, variable openings. Metallurgical studies have revealed that—with this process and the G-AlSi7Mg0.5 containing 0.19% iron—very good and highly reproducible structures are achieved, particularly in the area around the combustion chamber. When using the “LM Rotacast T6” alloy, the mechanical properties with the 0.2 offset limit (Rm) in the combustion chamber area, at 272 MPa, are better than for the G-AlSi7MgCu0.5 alloy (gravity casting) at 260 MPa. The exact values depend upon the casting process used and subsequent heat treatments. Isuzu, for example, manufactures cylinder heads using the Rotacast technique. 7.8.3.3 Lost-Foam Process (Full-Mold Process) The full-mold (or lost-foam) process is used for mass production in the United States. At BMW’s Landshut plant, this process was used for the first time in a six-cylinder, inline gasoline engine. The lost-foam process may also be considered a special form of the sand-casting process. The key steps in manufacturing a cylinder head are schematically shown in Figure 7.98. First, the polystyrene granulate is warmed, expanded to approximately thirty times its original volume, dried, and stored. In the first step in the casting process, the contours, from which the cylinder head is assembled in various layers, are foamed using the polystyrene material. For dimensional stability, the foaming tools are cooled with water. Grippers remove the foam blank, which then cures on a conveyor belt. The sum of the foamed contours represents the exact geometry of the cylinder head, taking into account thermal shrinkage. The individual contours are now joined at two stations with hot-melt adhesive. The positive model of a cylinder head consists of five polystyrene layers glued together in this way.

Sand

Black wash

Foam model

Gluing

coating

Drying

Sand

Sand

Sand

Water

Condensing by vibration

Casting

Emptying by suction

Removing black wash

embedding

Separating

Figure 7.98 Lost-foam process.

Two cylinder head models are glued with the sprue and runners to form a cluster. At the third station, this cluster is immersed in a water-soluble ceramic refractory coating. The unit is rotated to better homogenize the application of the refractory coating. At the fourth station, the cluster is dried in a stream of dry, heated air. This extracts the water to form a dense, gas-permeable refractory coating. In the next step, the cluster is inserted into the casting frame, and non-bonded quartz sand is filled loosely around it. The sand is compacted at the sixth station by vibration. Casting then follows. A charge of molten aluminum is prepared and poured into the mold automatically using a casting ladle. The polystyrene retracts and becomes gaseous during filling. At the eighth station, the mold is removed, that is, the sand is taken out of the casting frame. The refractory coating is removed in a water bath, and in the final step, the individual cylinder heads are separated from the cluster. The casting process itself demands a high degree of expertise with this technique. There is a great range of freedom in the engineering design for the cylinder head. Bores in the cylinder head, down to a minimum wall thickness of 4 mm, can be cast directly with the cylinder head. Changes in the course of the production run can be implemented at the tool relatively easily and thus at favorable costs, because the tooling is made of aluminum. At a cycle time of four heads in 3 min, this system offers a production capacity of approximately 330,000 cylinder heads per year (Figure 7.99). In consideration of the high strength requirements for direct-injection diesel engines, this process has not been used as of now in mass production for this particular application. Figure 7.99 shows the polystyrene casting cluster used by BMW for the first time in Europe to cast a cylinder head in the lost-foam process. The material used here is G-AlSi6Cu4 (aluminum alloy 226). A thermally decoupled, secondary air channel was integrated into the exhaust side for the U.S. versions.

150 | Internal Combustion Engine Handbook

6606_Book.indb 150

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

moving sliders, is opened and the casting is demolded with ejector pins. This process can be used only for air-cooled cylinder heads such as those used with small engines. In contrast to sand casting and die casting, pressurized die casting provides the most precise reproduction, as well as the greatest precision in the cylinder head geometry. Thin-walled castings with close dimensional tolerances, great exactness of shape, and superb surface quality can be fabricated. Exact casting of eyes, bores, mating surfaces, and other surfaces is often possible without subsequent machining. Pressure die casting, when compared with sand casting, die casting, and low-pressure die casting, offers the highest productivity because almost all the casting and mold movement processes are fully automated. Drawbacks include the limited engineering freedom for the cast component, because undercuts are not possible. Air or gas bubbles that might be trapped in the casting preclude double heat treatment, as for sand casting, die casting, and low-pressure die casting. This process is not suitable for mass production of water-cooled passenger car engines.

7.8.4 Model and Mold Manufacturing

Figure 7.99 Lost-foam cylinder head casting model, BMW.

This process allows •• casting oil channels in virtually any desired shape •• obtaining water cavities with elaborately shaped flow control fins •• implementing curved intake and outlet ports •• achieving markedly narrower tolerances in the combustion chamber area •• using only a single foaming die for the duration of the production run •• significantly reducing the amount of post-casting machining work required for the cylinder head. 7.8.3.4 Pressure Die-Casting Process Permanent molds made of hardened, hot-work tool steels are used in the pressure die-casting process. The plug has to be coated with a parting agent or refractory coating before each casting cycle, also known as a “charge.” In contrast to sand casting and die casting, no cores can be inserted into the mold since the lightweight metal melt is introduced into the casting form at high pressure and high speed. The pressure level depends on the size of the cast component and as a rule ranges from 400 to approximately 1000 bar. As in low-pressure casting, this pressure is maintained during solidification. When casting larger components, the two halves of the casting mold are cooled for directional solidification and quicker cooling of the casting. Once the casting has solidified, the mold, comprising fixed and moving components and possibly

When making the casting models, cores, dies, and all the casting tooling, almost all the parts are generated as models on the basis of 3-D CAD data throughout the Computer Aided Design/Computer Aided Manufacturing (CAD/CAM) process chain. Thus, the geometry data are more reproducible and the response to change is more flexible. With the creation of the cylinder head design, all the CAD models required for model making can be derived, from the CAD rough casting model to the fully machined component. Here a carefully designed data management system is required to maintain transparency so that everyone participating in the project is kept informed of changes and changes at the CAD cylinder head component are reflected in all the data records required for model and tool manufacturing. The model shop specifies all the traditional details, such as mold sectioning, drafts, casting shrinkage, supplements for manufacturing, and any deformations that might be expected in casting; these are taken into account in the CAD model. An early and lively exchange of experience with the cylinder head engineers pays off in the long run. The model building activities vary, depending on whether designing for prototypes or mass production and upon the choice of casting processes. The low-pressure sand-casting process used by Becker [7-63] is superbly suited for small production runs and prototypes. Figure 7.100 shows a core plug (above) and the package used for a water jacket core (below). The rough casting contour plus the allowance for shrinkage (of the metal during solidification) serves as the starting point for forming the model. Here, areas of a cast component that align with a given demolding axis are represented as a positive model in the so-called core mold tool. These areas in the cylinder head include, for instance, the crowned head of the combustion chamber, the ends, the intake and exhaust channel sides, the intake and exhaust channels, the camshaft bearing area, and

Internal Combustion Engine Handbook | 151

6606_Book.indb 151

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

the interior contours for water and oil flow. All the core tools are fitted with sealing surfaces and the so-called core markers that make possible exact alignment and sealing of the cores. These cores are Computer Numerical Control- (CNC) milled in a special plastic resin in only a few days, on the basis of the 3-D data. In the casting department, these cores are filled with sand to which bonding resin has been added; this cures in a short period of time without further treatment. The sand core thus removed from the reusable core mold tool now exhibits the negative contour of the ultimate cast part. A special version here is the so-called sand laser-sinter core, which can be made, layer by layer, directly from 3-D CAD data. No core mold tools are required here. Cores for detailed interior contours such as the water jacket or oil-carrying cavities are especially suitable for this process because manufacturing a core mold tool for these cores is both costly and time consuming. Finally, all these core segments (both conventional and sand laser-sintered) are assembled to create the core package, and molten material is then poured around it in the low-pressure casting process. A core package can be used only for a single casting.

Figure 7.100 Core mold tool and package for a water jacket core, by Becker [7-63].

A section of the overall core is shown in Figure 7.101 for a cylinder head made by BMW for an eight-cylinder engine, using the low-pressure casting process. All cores are made of sand. The core frames required for this purpose are made of steel for mass-production work. The spaces between the core segments are filled with molten aluminum. During the development stage, the sand cores are made as rapid prototyping models, used to evaluate the overall geometry. In the lower section of the illustration, one sees the combustion chamber plate, shown in dark gray. To the right of that is the core for the timing chain case. In the foreground is the exhaust channel core package, which projects into the water jacket core. Located above this is the core for the oil cavity.

Figure 7.101 Model of a cylinder head for an eight-cylinder BMW engine.

7.8.5 Machining and Quality Assurance 7.8.5.1 Mass-Production Manufacturing Cylinder heads are machined in mass-production operations on transfer lines or at linked machining centers, which make it possible to respond more flexibly to changes. A trend toward machining at sequential machining centers is emerging. Here, the rough component passes through several machining stations, one following another. It is necessary for each station to adhere to the prescribed cycle time. To limit the high overall investment costs, as many machining phases as possible are implemented at any given station. When developing a cylinder head, manufacturing planners should be integrated into the project following the tenets of simultaneous engineering to take into account the needs of manufacturing at an early date, all with the goal of achieving economical production. Changes to the cylinder head that have to be implemented retroactively at transfer lines are expensive and time consuming because the entire manufacturing process has to be interrupted. Because of the needs attributed mass production, it is often necessary to adopt compromises at cylinder heads that restrict developers’ design latitude. 7.8.5.2 Prototype Manufacturing Machining centers are normally used to work small production runs and prototypes. Often these individual stations are standardized machine tools that can be flexibly programmed and allow changes in the cylinder head to be implemented quickly. The machining costs are higher in comparison with mass production. The combustion chambers are in some cases machined to achieve better uniformity in the combustion processes. It is possible to machine the transitional areas from the gas exchange ports to the combustion chamber and the complete port shapes. 7.8.5.3 Quality Assurance for Cylinder Heads Failure of a cylinder head in the field often results in complete destruction of the engine. The goal for both the casting and the machining is to achieve a high quality standard for the customer, so the entire cylinder head is tested 100% for leaks.

152 | Internal Combustion Engine Handbook

6606_Book.indb 152

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

Spot checks by measuring components are standard procedures in quality control. It is imperative to minimize the reject rate in manufacturing. Computer-assisted tomography, known from the field of medicine, can be used to examine cylinder heads and to check the wall thicknesses, slice by slice, for compliance with the specified shapes and dimensions. Such examinations are standard, particularly for thin walls in a range of approximately 2.5 mm, as required in racing engines for reducing weight; see Figure 7.102.

Figure 7.102 Computer tomogram section of a cylinder head [7-63].

Figure 7.103 shows verification measurements for a cylinder head using a coordinate measurement unit. This makes it also possible to measure channel inside geometry. The channel surface can be traced point-by-point to form clusters of individual points. Deviations from the geometry described in the CAD data sets can be detected. Using the points transmitted to CAD systems makes it possible to apply reverse engineering methods to establish surface areas based on the cluster of points, which can also be used for 3-D flow simulations. These techniques are particularly valuable in association with direct-injection engines because even slight dimensional deviations can have considerable effects on the engine.

Figure 7.103 Digitizing an intake channel [7-69].

7.8.6 Shapes Implemented for Cylinder Heads 7.8.6.1 Cylinder Heads for Gasoline Engines Only four-cycle engines are discussed here. The cylinder heads illustrated here provide a selection from the multitude of valve train concepts found on the market, which have considerable influence on head geometry. The first example in Figure 7.104

Figure 7.104 Two-valve cylinder head for the BMW V-12 engine with roller-type cam follower.

Internal Combustion Engine Handbook | 153

6606_Book.indb 153

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

shows a two-valve cylinder head with roller cam followers made by BMW. This compact cylinder head concept is used in four- and twelve-cylinder engines. The head for the V-12 engine shown here is designed to be reversible and thus is identical for both cylinder blocks. To minimize friction, roller cam followers made of precision castings are used. This choice reduces friction in the valve train by as much as 70% when compared with the cylinder head with sliding cam followers, previously used. For weight restrictions, a hollow camshaft was developed using the process devised by Süko Company. Pushrods with hydraulic adjustment are often used in mass-production engines. Figure 7.105 shows, as an example, a four-valve cylinder head crafted by BMW for use in a V-8 engine. Longitudinal bores are provided in the single-component cylinder head to supply oil to the valve lifters; after casting, these channels are drilled into, from the outside, near the valve lifter bores. In V-block engines with hydraulic pushrods, the oil requirements in the cylinder head and the danger of oil foaming because of camshaft rotation are considerable so that drains of sufficient cross section have to be provided for oil to return through the engine block and to the oil pan. In this cylinder head, six return ports are provided for each bank of cylinders. The diameters of the intake valve disks are 32 mm for the 3-L engine and 35 mm for the 4-L engine; the exhaust valves measure 28.5 and 30.5 mm in diameter, respectively. The valve shaft diameters are just 6 mm. The angle between the port and the valve is 39°45′ on the intake side and 55°45′ on the exhaust side. The intake and exhaust valves form an included angle of 39°30′ and thus make possible a very compact, crowned combustion chamber. The spark plug is located at the center, between the valves. The valve cover is mounted elastically and thus is largely acoustically decoupled. The combustion chambers inside the cylinder head are machined

Figure 7.105 Four-valve cylinder head with pushrods made by BMW.

throughout to maintain close tolerances for the volume. The longitudinal-flow cylinder head is cast from aluminum alloy 226. For weight limits, the head is not designed to be reversible in this eight-cylinder engine. Both variants of the cylinder head are manufactured at a single production line and arrive fully assembled at the final installation point. Figure 7.106 shows a four-valve cylinder head concept using push rods in a multi-section design. Separate bearing strips are provided for the camshafts and the pushrods, both on the intake and on the exhaust sides.

Figure 7.106 Multi-section, four-valve cylinder head made by BMW.

Thus, the cylinder head, when in mass production, can be made using die casting because there are no undercuts in the upper area of the cylinder head. An example of a four-valve cylinder head with roller cam followers is depicted in Figure 7.107. This cylinder head, made by BMW, is a further refinement of the head illustrated in Figure 7.106. The objective in reworking the valve train was to reduce friction in the cylinder head, which was previously

Figure 7.107 Four-valve cylinder head with roller cam follower made by BMW.

154 | Internal Combustion Engine Handbook

6606_Book.indb 154

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

fitted with pushrods. Hydraulic compensation is affected here by static adjustment elements. Positioning the play adjustment unit in the stationary part of the valve train makes possible lower spring forces, because of the reduced oscillating masses, even though the valve stroke and opening period are retained. At the start of engineering, manufacturing operations had specified that the existing production line was to be retained. Thus, the valve angles and positions and the camshaft bearings were kept from the previous design. The scope of changes is thus limited to eliminating the bearing strips with the pushrod bores, the mounting bores for the compensators, which were arranged in a cloverleaf pattern around the spark plug, and the oil supply. Casting the camshaft bearings in place also lent stiffness to the cylinder head. The intake and exhaust ports and the combustion chamber were taken over without modification from the previous cylinder head. Three-valve cylinder head concepts are used on the V-block engines made by Daimler-Chrysler (Figure 7.108). These cylinder heads use an overhead camshaft (OHC) and roller rocker arms for valve actuation. Two spark plugs are used in each combustion chamber for faster burn propagation.

the interest of reducing fuel consumption. Four cylinders are shut down in the eight-cylinder version and six in the twelve-cylinder version. Positioning a camshaft shifter is made more difficult by this single-camshaft solution. Because of the relatively heavy rocker arm, this cylinder head concept is not suitable for concepts involving high engine speeds. The overall concept is, however, more economical than a four-valve arrangement with two camshafts. In 1994, with the introduction of its A4 series, Audi built for the first time a five-valve cylinder head in passenger car engines. This cylinder head has been adopted throughout the VW Corporation for four-, six-, and eight-cylinder engines (Figure 7.109). Except the eight-cylinder engine that uses roller cam followers, these engines employ pushrods with hydraulic compensation. For geometric reasons (the valve stem centerline would otherwise intersect the camshaft), the angle for the center intake valve differs from the other two.

Figure 7.109 Five-valve cylinder head made by Audi [7-58].

The valve angle for the outer intake valves is 21.6°, that for the center valve is 14.9°, and the exhaust valve angle is 20.2°. To improve force transfer at the head bolts, a bushing is screwed into the cylinder head; thus, the collar on the head screws can be kept small. This effect helps alleviate the tight geometric situation at the cylinder head. In addition, the camshaft clearance can be kept at 129 mm because the bolts pass close by the camshafts. This is a one-piece cylinder head made up in gravity die casting. Similar five-valve designs had been used prior to their debut at Audi in one-, two-, and four-cylinder motorcycle engines made by Yamaha.

Figure 7.108 Three-valve cylinder head made by Daimler-Chrysler [7-70].

In its eight- and twelve-cylinder engines, Daimler-Chrysler incorporates cylinder cutout in this rocker arm concept, in

7.8.6.2 Cylinder Heads for Diesel Engines Presented as the first example of an engineering design is the cylinder head for a two-valve engine with swirl chamber. These diesel engine concepts have dictated cylinder head design ever since diesel power plants were introduced in passenger cars. Seen in the cross section through the cylinder head in Figure 7.110 is the prechamber with the injection valve and the glow plug. The hollow-cast camshaft actuates the intake and

Internal Combustion Engine Handbook | 155

6606_Book.indb 155

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.110 Lateral and longitudinal sections with the installation situation for a two-valve diesel cylinder head made by BMW.

exhaust valves (with diameters of 36 and 31 mm, respectively) via pushrods with hydraulic compensation. In passenger car engineering, this design has been used in mass production at BMW since 1983. With the introduction of direct-injection diesel engines by Audi in 1989, the share of diesel power plants in passenger cars has risen distinctly, primarily in Europe. Four-valve technology was introduced to a greater extent to achieve even higher power densities in diesel engines, too. Because of the greatly increased ignition pressures, maximum demands for strength and durability are made on today’s diesel engine cylinder heads. Roller cam followers can be employed to minimize friction losses in the cylinder head (Figure 7.111).

Figure 7.111 Four-valve cylinder head with roller cam follower for a six-cylinder engine.

This example shows a six-cylinder engine made by BMW, which employs this head technology in its four- and eightcylinder engines, too. The cylinder head is fitted with swirl ports; here the air is introduced from above, through the cylinder head. The cylinder head is cast from an alloy produced in primary refining. The timing chain case is cast into the front end of the eight-cylinder head. This lends the component significant additional strength. An exhaust gas return channel is integrated into the rear section. The camshafts are driven by straight-tooth spur pinions, while the intake camshafts are driven by chains. The common rail injection technology used here requires two rails attached at the side of the cylinder head to supply fuel to the injection valves, which are positioned at the center of the cylinder head. The coolant flows inside the cylinder head from the exhaust side to the intake side. To ensure that crosswise flow is maintained, the cylinder units are separated one from another by partitions inside the cooling cavity and have a water collector manifold cast as a unit on the intake side. A further process used for diesel direct injection is the pump nozzle technique developed by VW. A separate injection pump, actuated by the camshaft, is provided for each cylinder, and this naturally has a major impact on the overall cylinder head concept (Figure 7.112). This two-valve cylinder head is equipped with pushrods with hydraulic valve play compensation. Located to the side, above the camshaft, is a bearing axis for rocker arm actuation of the pump nozzle elements. Used as the timing element is a synchronous belt that has to be fabricated from a high-strength material because the moments that the pump nozzle drive induces at the camshaft are very high. Fuel is supplied to the pump nozzle elements inside the cylinder head by one each of supply and return rails. A vane pump driven by the camshaft delivers the required feed pressure. These pump nozzle elements can now achieve injection pressures exceeding 2000 bar. This makes it possible to resolve the conflict of interest between low pollutant emissions and higher specific output because, even with small nozzle

156 | Internal Combustion Engine Handbook

6606_Book.indb 156

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.8 Cylinder Head

orifices and high injection pressures, it is possible to achieve a short injection period and rated output. Eliminating the distributor-type injection pump with its mounting bracket, drive components, and injection lines allows unification of auxiliary component arrangements on gasoline engines.

Figure 7.112 Pump nozzle cylinder head made by VW [7-71].

7.8.6.3 Special Cylinder Head Designs In the VR series of engines made by VW, five- and six-cylinder engines are made with a V-angle of 15° for a very compact configuration, in fact, a sort of synthesis of the inline and V-block engines. The one-piece cylinder heads are quite wide [7-72]. Placing the intake and exhaust runners on either side of the cylinder head mandates differing intake and outlet port lengths for the two cylinder banks. Concepts with symmetrical gas exchange ports are also possible, but they require a minimum of three camshafts instead of the two used here [7-73]. Figure 7.113 shows two sections through the mass-production, fourvalve cylinder head, illustrating the various lengths of the gas ports. The cylinder head is provided with a camshaft shifter and uses precision-cast roller cam followers in the valve train. The design selected here, with its two camshafts, permits the spark plugs to be located at the center by adjusting the valve lengths. The difference in valve lengths is 33.9 mm. The valve diameters are 31 mm for the intake valves and 27 mm for the exhaust valves; the valve stem diameter is 6 mm. The combustion chambers of the two cylinder banks are almost mirror images of one another. The angle between the intake and exhaust valves is 42.5°. The cross-flow concept used in the VR cylinder head requires differing angles for the valves in relation to the cylinder centerline: 34.5° for the long ports and 8.0° for the short ports. In addition, the angles differ in relation to the port axes.

Figure 7.113 Sections of the type VR four-valve cylinder head made by VW.

To achieve uniform combustion in both cylinder banks, the short and long intake ports have to be tuned to achieve uniform flow-through and tumble effects. Air-cooled cylinder heads are very rare in passenger cars. The two-valve cylinder head for a six-cylinder boxer engine made by Porsche and shown in Figure 7.114 has been supplanted in the current series by water-cooled, four-valve cylinder heads.

Figure 7.114 Air-cooled cylinder head made by Porsche [7-74].

Internal Combustion Engine Handbook | 157

6606_Book.indb 157

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

To handle the degree of heat dissipation required at the cylinder head, cooling fins with large surface areas are needed, in addition to the existing cooling fan. In this example, a ceramic port liner is cast in place in the cylinder head. Its insulating effect limits the amount of heat transferred into the cylinder head. In addition, this keeps the exhaust temperature high, accelerating catalytic converter warmup following a cold start. High engine speeds and very lightweight valve train components are required for sports engines with their extremely high volumetric efficiency. The masses in motion are kept as small as possible. Here it is advisable to eliminate heavy hydraulic valve play compensation elements. An example of such a design has been realized by BMW in a six-cylinder engine with precision-cast cam followers and mechanical valve play adjustment. These sliding-type cam followers are very lightweight and rest on a shaft inserted into the cylinder head. In selecting the lever ratio for the cam follower, component stiffness was given precedence over minimizing the amount of installation space required. The rocker arm drive uses a 1:1 lever ratio not to induce any bending loads (Figure 7.115).

Direct-injection gasoline engines are now being developed in readiness for mass production all over the world. It is necessary to free up space in the cylinder head, next to the spark plug, to accommodate the injection nozzle, a situation similar to that seen in direct-injection diesel engines. Space is tight even for a four-valve concept. Developing and fine-tuning the locations and shapes of the gas exchange ports are extremely difficult because of the charge stratification associated with the combustion process. Figure 7.116 shows a section of a four-valve cylinder head used in mass production for a directinjection engine built by Mitsubishi. The introduction of air through the intake port is effected, from the intake manifold, at the top surface of the cylinder head to induce a defined tumble effect in the flow, matched to the shape of the bowl in the piston crown. The valves are actuated by roller cam followers. The injection nozzle is located toward the side, at the cylinder head. The spark plug is located at the center of the cylinder head.

Figure 7.116 Cylinder head for a direct-injection gasoline engine made by Mitsubishi [7-75].

7.8.7 Perspectives in Cylinder Head Technology

Figure 7.115 Four-valve BMW cylinder head with sliding cam followers.

The four-valve cylinder head used here is a one-piece unit cast in a steel die. A cross-flow cooling concept is used here. Integrated into the cylinder head is an air distribution line into which supplementary air is blown. Splitting off from this line, which is 12 mm in diameter, are 4 mm bores leading directly into the exhaust port, next to each exhaust valve.

The gas charge and the combustion process itself are controlled in the cylinder head. Further development in cylinder head technology is targeted on lightweight construction, higher strength materials, and more economical manufacturing processes while at the same time improving all the engine targets. Multi-valve cylinder heads have made a breakthrough on all fronts to include diesel engines. Their introduction and better gas control make it possible to realize higher specific outputs per cylinder. Adopting advanced cylinder head concepts leads to downsizing concepts that result in high-performance engines with favorable emissions and consumption levels in a range of engines available for the customer’s choice depending

158 | Internal Combustion Engine Handbook

6606_Book.indb 158

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.9 Crankshafts

on power requirements. In the recent past, direct-injecting SI engines are increasingly available in series vehicle using lambda 1 and lean concepts. The newly introduced combustion processes require the adaptation of the combustion chamber in the piston which is usually given recesses to support the mixture preparation. In jet-controlled direct injection, the injection nozzle is moved into the cylinder head center. SI engines are increasingly used as turbo engines—sometimes with two-charge systems—to attain higher performance and to implement a downsizing strategy with more efficient consumption. The objective is clearly to lower CO2 emissions. For the attachment of the injection nozzles for direct injection, either a lateral position is selected (Mitsubishi engine, Figure 7.116) or the central position in the cylinder head which means a position inclined to the cylinder bore must be selected for the spark plug. Both positions are chosen by manufacturers for lean-burn engine concepts and for λ = 1 operation [7-76]. The increase in the specific output of the engines represents higher demands on the cylinder head materials and their durability. The remaining effects of direct injection on the cylinder head design mostly relate to determining a suitable position for the injection nozzle. Concepts with a central position of the injection nozzle in the cylinder head have more effects on cylinder heads with respect to geometry, thermal issues, and resistance. Furthermore, the cylinder head technology has gained a new dimension with the use of fully variable timing in gasoline engines. Thanks to a throttle-free load control, the load change work and, nearly proportional, the specific fuel consumption has been significantly reduced. Now, purely mechanical concepts have been implemented in mass-production engines [7-57], [7-77]. Since BMW uses the Valvetronic system in various models of four-, six-, eight-, and twelve-cylinder engines, Toyota, Nissan, and Mitsubishi have also implemented mechanical variable valve actuation systems in series engines. The cylinder head geometry has been completely redesigned with the introduction of these systems. The additional components in the cylinder head require new valve train concepts in which an adjustment gear to adjust the valve elevation must be additionally housed in the cylinder head. The combination of direct injection with fully variable valve actuation—the Valvetronic system—has been realized in a V12 engine by BMW [7-78] (Figure 7.117). This trend will continue and the first designs of this type with turbocharging are expected to be seen in series engines. It will be exciting to observe the degree to which throttleless control will provide greater degrees of freedom in valve timing and thus in engine control. Studies with an additional variability on the exhaust valve end indicate further potential for lowering CO2 emissions. The full potential for lowering fuel consumption by using a fully variable valve actuation has not yet been attained. Another focus in the development of new cylinder head concepts is lowering valve train friction and reducing the lubricants required for the valve train components. Now, intensive studies on camshafts on roller bearings and ultracompact valve train components are conducted [7-79].

Figure 7.117 Cylinder head of the V12 BMW engine with direct injection and Valvetronic. See color section page 1069.

The demands on cylinder heads will change with valve actuation systems which might incorporate electromechanical or other innovative actuation principles. As was the case approximately 100 years ago, the diversity in cylinder head concepts in series engines will continue to be diverse. This presents a major challenge to engine developers, who must keep pace with demands as they continue to develop.

7.9 Crankshafts 7.9.1 Function in the Vehicle

Despite the need to reduce average CO2 emissions and the related efforts to develop alternative drives, the combustion engine still dominates in vehicles, preferably designed as a reciprocating piston engine. This will be the case in the coming years. To achieve the emission targets, charged engines with fewer cylinders are increasingly used as a result of the so-called downsizing and use of hybrid drive systems. 7.9.1.1 Crankshafts in the Reciprocating Piston Engine The piston’s linear movements are converted, with the intervention of the conrod, into rotational movement at the crank of the crankshaft, thus making torque available for use at the wheels. The functional elements of a crankshaft are schematically shown in Figure 7.118. Because of the strains involving forces that change in both time and location, with rotational and flexural torques and the resulting excitation for vibrations, the crankshaft is subject to very high, very complex loads. 7.9.1.2 Requirements The crankshaft’s service life is influenced by (a) resistance to flexural loading (weak points at the transition from the bearing seat to the web) (b) resistance to alternating torsion (the oil bores are often weak points) (c) torsion alternation behavior (stiffness and noise)

Internal Combustion Engine Handbook | 159

6606_Book.indb 159

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

(d) wear resistance of the crankshaft main bearing, for example (e) wear at shaft seals (leaks and motor oil escaping).

1

2

3

Passenger Cars

Forged

Cast

Western Europe

11.3

3.1

8.2

USA

 6

0.35

5.65

Japan

 8.5

4.75

3.75

Global

34.1

*)

*)

Trucks and buses

Forged

Cast

Western Europe

 1.25

0.85

0.4

USA

 4.9

4.0

0.9

Japan

 2.75

2.1

0.65

Global

13.1

*)

*)

*) Figures not available.

Figure 7.119 Passenger car crankshafts, by manufacturing process (in millions of units, 1993) [7-80].

7.9.2.1.1 Casting There are several processes available for casting crankshafts; they are listed in Figure 7.120.

4

5

6

7

Figure 7.118 Functional elements of a crankshaft, schematic representation: 1 = flange, 2 = conrod bearing, 3 = counterweight, 4 = main bearing, thrust bearing, 5 = oil channels, 6 = concave radii, and 7 = stop collars.

For ecological reasons, the trend is toward high-torque engines with fewer cylinders that will develop high moments even at low engine speeds. In these engines, the crankshaft is subjected to far greater loading in all the aspects mentioned above than is the case in conventional, aspirated engines. In addition to the torque, the engine design is an essential criterion for the stresses on the crankshaft. For example, at same engine output, V6 crankshafts are usually substantially more burdened than R6 crankshafts [7-93].

7.9.2 Manufacturing and Properties

In 2002, 16.5 million passenger cars and light utility vehicles were produced in Western Europe. The global production was 57.2 million vehicles. The demand for crankshafts was correspondingly high. 7.9.2.1 Processes and Materials Crankshafts are either cast or forged. The shares accounted for by the individual manufacturing processes in 1993 are shown in Figure 7.119. In the past years, the portion of forged shafts has increased in Europe because of the increasing demand for diesel engines and the development trend toward higher torques. The necessity for reducing CO2 emissions and fuel consumption is leading increasingly to turbocharged gasoline engines, which now are preferably fitted with forged crankshafts.

Process

Position in the Mold

Forming Process

Green sand IMD

Horizontal

Automatic system with form box

Green sand

Horizontal

Automatic system with form box

Shell mold

Vertical

Croning sand crust in boxes, backfilled with steel shot

Shell carrier

Vertical

Croning sand crust in steel support crusts

Water glass-CO2 process

Vertical

Double-sided molds from automatic molding machine, horizontally formed = 1 package. Gas application in vertical position

SF process with cold box sand

Horizontal

Automatic molding machine with box, wear sand is renewed

Lost foam

Vertical

Styrofoam model in boxes, backfilled with sand

Figure 7.120 Survey of casting processes used to manufacture crankshafts.

Based on the evaluation of the various processes, we find that, because of better dimensional stability, there are advantages for the green sand IMD process [7-84]. But the technique most commonly used in practice is the shell mold process. The further development of casting processes [7-81] goes in the direction near net shape geometry, high-strength, harder casting materials, and the realization of oil channels during casting. 7.9.2.1.2 Forging Two companies in Germany concentrate on making forged crankshafts for road vehicles [7-80]. In 1993, the share of forged shafts was 28% in Western Europe. Because of technological considerations, the trend toward forged crankshafts is continuing. Crankshafts for passenger car engines reach their operational strength by the basic strength of the material and after-treatment close to the end of the production process (heat treatment and surface hardening). In forged materials, a higher basic strength is available, in addition to the higher Young’s modulus. For flexible manufacturing, the subsequent

160 | Internal Combustion Engine Handbook

6606_Book.indb 160

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.9 Crankshafts

expansion of an engine generation is less problematic when the higher operational strength can be achieved with the basic strength and not a modification of the after-treatment. 7.9.2.1.3 Advantages and disadvantages of forged and cast crankshafts •• Cast crankshafts are considerably more economical than forged units. •• Casting materials respond well to surface finishing processes used to boost oscillation resistance. Thus, for example, the resistance to flexural loading can be increased considerably by rolling the radii at the transition between the journals.

It is possible, however, to compensate for this advantage: •• By larger diameters in the bearing area, which is not possible for existing engine concepts and which is not desirable in new concepts because of greater friction losses and the associated rise in fuel consumption. •• With complex vibration dampers, which increase system costs, however, and can offset the cost advantage of cast crankshafts in comparison with forged versions. •• With a very stiff design for the engine block, joined with crankshaft bearing bridges, cast oil pans, and a stiff linkage with the transmission.

•• Cast crankshafts can be hollow and thus may be as much as 1.5 kg lighter in weight.

•• By using the Integrated Starter Alternator Damper system under development [7-83].

•• Cast crankshafts of the same design offer a weight advantage of approximately 10% compared with steel, which is because of the lower density of the nodular cast iron.

•• By eliminating the separate starter and alternator, this system offers the option of damping engine oscillations using “alternating reactive power.” Here any crankshaft overspeeding is braked by kicking in generator action and any lag is compensated by applying the energy stored in capacitors.

•• Machining the cast crankshaft is generally simpler. It is possible to work with smaller supplements for later machining, the mold parting flash is reduced and no longer needs to be removed, and the slopes in the webs can be specified more closely. In fact, it is often possible to do without any machining of the webs at all. •• Thanks to the higher design degrees of freedom in casting processes, trigger wheels for adjusting ignition timing or drives for accessories with similar tasks can be integrated more easily. 7.9.2.1.4 Disadvantages of cast crankshafts compared with forged crankshafts •• Cast materials have a lower Young’s modulus and a lower basic strength than steel. As a consequence, cast crankshafts are less stiff and exhibit different vibration properties. •• Measures implemented to increase drive train stiffness are even more necessary when aluminum blocks and crankcases are used, as this material has a far lower Young’s modulus and thus less material stiffness. •• Cast crankshafts, when compared with steel, may exhibit less favorable wear characteristics at the bearing journals, which is because of the micro-voids in the surface (exposed spheroliths), lower fundamental hardness, and less enhancement of hardness in the usual hardening processes [7-86].

In torque-optimized engines, the cast crankshafts have more physical problems with torsional oscillations, because of Young’s modulus and less stiffness, which can lead to an unacceptable noise level in the vehicle. The Young’s modulus for steel is 210 kN/mm2 and that for cast iron with nodular graphite is 180 kN/mm2. Consequently, forged crankshafts are used now in engines that develop high torques at low engine speeds. This applies, in particular, to engines with more than four cylinders, as well as to diesel engines. 7.9.2.2 Material Properties for Crankshafts Crankshaft material properties are shown in Figure 7.121 and Figure 7.122. The development in forging steels for crankshafts moves toward AFP steels (i.e., age-hardened ferritic–pearlitic steels) [7-92], [7-96]. These steels do not require tempering to achieve their basic strength. The ferritic–pearlitic cast material GJS-700-2 is a material type commonly used for crankshafts. Engine manufacturers sometimes have their own company specifications for spread of hardness values. After-Treatment Potential

Proof Stress Rp0.2 (N/mm2)

Elongation at Fracture A (%)

Machinability

Nitrided

Rolled, RollerBurnished

Inductive Hardened

Steel

Tensile Strength Rm (N/mm2)

C38+N2++

780–900

>450

>12

Good

Good

Good

Very good

C38mod++

820–1000

>550

>12

Good

Good

Good

Very good

37Cr4

880–1030

>620

>11

Demanding

Good

Low

Very good

37Cr4V

850–950

>650

>14

Demanding

Good

Low

Very good

42CrMo4V

980–1100

>850

>12

Demanding

Good

Low

Very good

++

Designation Gerlach-Werke in BY quality. (BY = with controlled cooling from melt temperature. No further heat treatment is required to achieve the basic strength.)

Figure 7.121 Properties of steel forging materials for crankshafts [7-93].

Internal Combustion Engine Handbook | 161

6606_Book.indb 161

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

After-Treatment Potential [7-93]

Cast

Tensile Strength Rm (N/mm2)

Proof Stress Rp0.2 (N/mm2)

Elongation at Fracture A (%)

Hardness HB 30

Machinability

Nitrided

Rolled, RollerBurnished

Inductive Hardened

GJS-600-3

600

370

3

200–250

Very good

Good

Very good

Low

GJS-700-2

700

420

2

230–280

Very good

Good

Very good

Good

GJS-800-2

800

500

2

250–300

Good

Good

Very good

Good

Figure 7.122 Properties of nodular cast iron (GJS); minimum values according to DIN EN 1563.

As a rule, the dimensions of engines become smaller, while the output stays the same. To reduce weight, light metals are being increasingly used for crankcases and cylinder heads. This raises the demands on the crankshaft on the one hand and, on the other hand, the space conditions leave less and less margin for material- and manufacturing-appropriate design.

7.9.3.2 ADI—Austempered Ductile Iron (Ausferritic Ductile Iron) The conversion in an austenitic–ferritic structure can be used to increase the basic strength of cast iron materials. This material is manufactured with an additional heat treatment process and owns properties (Figure 7.124) such as high strength, good elongation, and good hardness, but poor machinability. Aside from the considerably higher costs, “nearly finished” [7-81] manufacturing does nothing to alleviate the basic problem associated with nodular graphite casting materials: the Young’s modulus cannot be raised to above the values for normal GJS even with the extensive heat-treatment processes employed to impart greater strengths.

7.9.3.1 Hollow-Cast Crankshafts In general, cast crankshafts of a comparable design weigh approximately 10% less than a forged unit because of the lesser density. Hollow cast crankshafts offer a further weight reduction of up to 1.5 kg (Figure 7.123).

7.9.3.3 Increasing Component Strength Through Post-Casting Treatment The static properties say little about crankshaft service life. Component strength, heavily influenced by sufficient vibration resistance, is achieved only through supplementary treatment

New developments also move toward increasing the basic strength and optimizing after-treatment without adversely affecting the remaining properties [7-95].

7.9.3 Lightweight Construction and Processes for Increasing Strength

 Figure 7.123 Cast crankshaft for a four-cylinder motor using GJS-600-3 (hollow version weighing 10.6 kg at the right and solid cast version weighing 12 kg at the left).

Material Code

Tensile Strength Rm (N/mm²)

0.2% Offset Proof Stress Rp0.2 (N/mm2)

Hardness HB 30 (Standard Elongation at Fracture A (%) Values)

EN-GJS-800-8

≥800

≥500

≥8

260–320

EN-GJS-1000-5

≥1000

≥700

≥5

300–360

EN-GJS-1200-2

≥1200

≥850

≥2

340–440

EN-GJS-1400-1

≥1400

≥1100

≥1

380–480

Figure 7.124 Mechanical properties of ausferritic (bainitic) spheroidal graphite (ADI) according to DIN EN 1564.

162 | Internal Combustion Engine Handbook

6606_Book.indb 162

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.9 Crankshafts

Vibration resistance of crankshafts Vibration resistance 0

Achievable vibration loading ± sa (N/mm2, Nominal bending tension)

Fatigue resistance 200

150

with induction-hardened wristpin and radii

0

Hardened without surface Without surface treatment treatment

Jonitrided

with compacted radii

with compactedblasted radii

hardened with compacted radii

100

50

0

Ck-45

GJS-700-2

GJS-400-18

GJS-700-2

GJS-700-2

GJS-700-2

Ck-45

Figure 7.125 Influence of post-casting treatment on crankshaft vibration resistance.

processes; this is true for both castings and steel (Figure 7.125). The critical radii of the conrod, and main bearings in particular, must be strengthened for their operational use by a strength-increasing process. 7.9.3.3.1 Radius rolling Rolling the radii is the standard process [7-88], [7-89], [7-90], [7-91] used to enhance fatigue strength under reversed bending stress for both cast and steel crankshafts. Here pressureinduced self stresses are created at the transitions from the bearing journals to the webs; this improves long-term strength considerably in this heavily loaded area. 7.9.3.3.2 Inductive hardening of radii with and without journals This process is used in some cases on crankshafts for diesel engines to increase the bearing journals’ resistance to oscillation and wear. A combination of inductive hardening and radius rolling [7-88] is also possible.

7.9.4 Calculating Crankshafts

For the design of crankshafts, the knowledge of material characteristics and the possibility of influencing the properties with post-casting treatment are of vital importance. The continuous further development of FEM and multi-body simulation (MBS) programs also contributes to a weightoptimized design of crankshafts. An integrated simulation process allows not only dynamic and acoustic analyses but also assessments of service life. Integrated simulation process [7-94]: •• The crankshaft is considered as a linear and elastic structure. Any non-linear movements and non-linear coupling between the components are analyzed within MBS (Figure 7.126).

7.9.3.3.3 Nitriding In this process, too, pressure-induced self stresses are induced in the journals and radii areas with a positive effect on enduring resistance to vibration and wear. But nitriding is used less and less because it cannot be integrated into the manufacturing line and disposal of the salts is difficult. 7.9.3.3.4 Ball calibration This process is employed to boost resistance to torsional vibration by strengthening the oil bores in the bearing journals. It must be noted that the post-casting processes must be individually optimized for each material. This must be considered when planning a change in material. Figure 7.126 MBS model with flexible crankshaft and EHD bearing model.

Internal Combustion Engine Handbook | 163

6606_Book.indb 163

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

780.0

1.0

702,4

0.9

624.8

0.8

547.2

0.7

469.6

0.6

391.9

0.5

314.3

0.4

236.7

0.3

159.1

0.2

81.5

0.1

3.9 1020.7 1306.6 1592.4 1878.3 2164.1 2450.0 2735.9 3021.7 3307.6 3593.4 3879.3 Rotational speed [U/min]

The dynamic properties of the lubrication film in the main bearings of the crankshaft drive are mapped using the elasto-hydrodynamic (EHD) oil film model and integrated in the MBS. A non-static engine run-up is simulated with the MBS. The entire speed range is traveled to analyze that all system resonances are occurring. Using Campbell diagrams of relevant measuring variables (Figure 7.127), one obtains detailed information regarding any potentially occurring effects (bending and/or torsional vibrations, flywheel wobbling, etc.).

Maximum protection from fatigue fracture

•• The dynamic simulation is followed by service life prediction. The load distributions of the individual mode shapes (modal loads) from the FE analysis are used as load data sets. The result is the speed dependency against fatigue fracture (Figure 7.128).

Non-stationary Stationary

1000

1500

2000 2500 3000 Crankshaft speed [U/min]

3500

4000

Figure 7.128 Example of a distribution of minimal fatigue fracture protection across speed.

Normalized amplitude

Frequency [Hz]

Chapter 7 Engine Components

0.0

Figure 7.127 Campbell diagram of an engine run-up. See color section page 1069.

•• In high-speed engines, in particular, where dynamic effects often have a decisive influence on component safety, this integrated simulation process provides an efficient development tool to detect weaknesses at an early stage. It is possible to perform variant and material comparisons at an early development stage or to adjust add-on components (such as torsion dampers).

7.10 Valve Train Components 7.10.1 Standard Valve Train

In recent years, a trend has emerged in passenger car engines toward OHCs and double OHCs (DOHC) while engines with camshafts located below [overhead valve (OHV)] continue to be used, particularly in large-displacement, V-block engines. Engines with OHCs are developed so that valve trains can be engineered to withstand the high speeds required in higher performance engines. DOHC concepts give the engineer the option of mutually independent timing for the intake and exhaust camshafts using camshaft shifters. OHV and OHC concepts are characterized by compact shapes and sizes and by economy in their manufacture. For diesel engines for utility vehicles, one sees a trend toward four-valve concepts. Rocker arms or double-rocker arms fitted with mechanical valve play adjustment and driven by pushrods and camshafts located below—as is the case in two-valve designs—affect valve lift. OHC concepts are employed, in addition to OHV versions, in smaller utility vehicle engines that utilize engine braking effect, and hydraulic valve lifters are used increasingly to compensate for valve lash. 7.10.1.1 Direct Drive Valve Trains This category embraces valve trains with hydraulic (Figure 7.129) or mechanical valve lifters as well as so-called “bridge”

164 | Internal Combustion Engine Handbook

6606_Book.indb 164

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

solutions in which components, guided by columns, lift multiple valves by direct actuation with a single camshaft. A subgroup within the latter solution is represented by the bridge that interfaces with two hydraulic valve lifters (Opel direct-injection diesel engines).

with this version, it is not absolutely necessary to remove the camshaft. Such units are, however, considerably heavier and require more installation space than the other version (at identical valve lift). The basic body for the valve lifter is made of ductile steel. Aluminum is found in only two applications (Toyota Lexus V-8 and Jaguar V-6 and V-8). The shims are usually made of steel that can be hardened. When bodies made from deep-drawn sheet steel and small hydraulic elements (11-mm O.D.) are used, hydraulic valve lifters achieve very low masses that, at identical lobe contact diameters, are far lighter than mechanical valve lifters with shims at the top.

Figure 7.129 Hydraulic valve lifters.

Direct drive always offers very good stiffness values with relatively modest masses in motion. This is the prerequisite for trouble-free valve train operation even at very high speeds (loss of contact force and premature valve seating). Thus, efficient, high-speed engines can be developed particularly by employing valve lifters. In the interest of reducing the masses in motion, preference among mechanical valve lifters is given to those with graduated crown thickness (Figure 7.130) or those with adjustment shims located at the bottom.

Figure 7.130 Mechanical valve lifter with graduated crown thickness.

For service work (adjusting valve play), rods with an adjustment shim at the top (Figure 7.131) are preferable because

Figure 7.131 Mechanical valve lifter with adjustment shim at the top.

The sliding contact with the lobe requires careful machining at the camshaft—stone finishing following cam lobe grinding has proved to be the most favorable. Over and above this, the camshaft material has to be matched to the loading situation to avoid wear. The versions that have been found to be particularly advantageous are hard-cast camshafts and camshafts made of gray cast iron with a remelted surface. The valve lifters and shims should rotate to achieve uniform wear for the cam lobe contact surfaces. This is achieved by shifting the cam in relation to the shim (toward the camshaft centerline) or, with offsetting and an angular grind for the cam lobe, at the point where the lobe contacts the valve lifter. Valve trains with valve lifters and the mechanical versions, in particular, offer the advantage of lower cylinder head height in DOHC designs. Valve lifters are found in many different applications, for example, two- and four-valve SI engines and diesel engines. The Volkswagen group uses a valve lifter incorporating a special hydraulic element designed to prevent increases in contact force while it sweeps the lobe’s circular segment in all of its pump-nozzle diesel engines. 7.10.1.2 Indirect Drive Valve Trains Included in this group of valve trains are as follows:

Internal Combustion Engine Handbook | 165

6606_Book.indb 165

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• Cam follower valve trains with stationary valve play compensation elements; the cam follower rests on the spherical upper end of the hydraulic element. •• Rocker arms that pivot on a shaft. •• OHV concepts comprising the cam follower (flat or roller valve lifter), pushrods, and rocker arms. There is a clear trend in cam follower drive trains toward cam followers that are made of sheet metal and are fitted with a rolling bearing at the point of contact with the camshaft. Cam followers made from cast steel in a precision-casting process give the engineer greater design leeway (stiffness and moment of inertia). The cost advantages for the sheet metal cam are so great, however, that precision cast cam followers are used only in exceptional cases (Figure 7.132). When compared with plain cam followers or valve trains with valve lifters, the use of the rolling bearing effects a reduction in friction, particularly in the lower speed range that is relevant to reducing fuel consumption. This reduction in friction losses is, however, paid for with a significant reduction in damping of torsional vibrations at the camshaft, which has consequences for the timing chain or belt.

is given to valve train geometries in which the roller is positioned approximately at the center between the valve and the hydraulic element. Here the camshaft is located above the roller. This arrangement makes it possible to keep the hazard of “pumping up” under control (see hydraulic valve play compensation). This configuration, with the cam lobe offset from the valve stem centerline, makes the cam follower concept interesting for four-valve, direct-injection diesel engines because, in these units, the valve stems either are parallel or have only a very small included angle (Figure 7.133). Only with the use of cam followers is there sufficient clearance between the camshafts. Using cam followers also makes it possible to serve “inverted” valve arrangements (e.g., DCC OM 668).

Figure 7.132 Roller cam follower with hydraulic support element.

Figure 7.133 Valve train for a direct-injection diesel engine with cam followers [7-100], Wiener Motorensymposium (2001).

Mass moments of inertia and stiffness are highly dependent on the shape of the lever. Short levers cause low mass moments of inertia, with masses on the valve side that are lower than for valve lifters. Seen as a whole, roller cam followers are inferior to valve lifters with regard to stiffness. The profiles for the lobes in valve trains incorporating roller cam followers differ significantly from those in valve trains that use valve lifters (greater radius at the apex, shorter lobe stroke—depending on the lever ratio, and concave flanks). To keep the concavities of the cam narrow enough that they can still be ground with mass production technology, preference

As opposed to cam followers, rocker arms are mounted on shafts. One differentiates between rocker arms in which the pivot point is toward the center of the lever (Figure 7.134) and those that pivot at one end; the latter are also known as cam followers. The camshaft is located below one end of the rocker arm, while cam motion is transferred via either a plain, sliding surface or a cam roller. To achieve low friction losses, needlebearing cam rollers are used in most modern rocker arms. The valve is lifted at the opposite end of the lever, via a hydraulic

166 | Internal Combustion Engine Handbook

6606_Book.indb 166

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

valve play compensating element or a setting screw used for mechanical valve play adjustment (Figure 7.135).

relatively high surface pressures at the end of the valve stem. Where surface pressures are excessive, hydraulic elements are employed which incorporate a pivoted foot at the point of contact with the valve. The contact itself is at a virtually flat surface, while the pivoted foot executes a movement around a ball mounted on the hydraulic element (Figure 7.136).

Figure 7.134 Typical rocker lever valve train.

Contact radius min. 105 mm surface pressure at 1000 N load is 616 N/mm2

Contact radius min. 15-30 mm surface pressure at 1000 N load is 2254-1420 N/mm2

Figure 7.136 Hydraulic elements for rocker arms, 11-mm O.D.

Figure 7.135 Side view and section through an aluminum rocker arm.

The contact surface at the rocker arm has to be angled to maintain unbroken contact between the adjustment element and the end of the valve stem as the arm executes its rocking motion. Because neither the hydraulic compensation element nor the mechanical adjustment screw is mounted in the rocker arm to specify a direction, the contact surface at the valve actuation element is crowned. This geometric design leads to

Aluminum, preferably manufactured in a die-casting process, or cast steel is used here. Oil is supplied to the hydraulic compensator elements from the rocker arm shaft. From this point, bores in the rocker arm lead to the hydraulic elements. Support shims with a little play in the guide, which are always used in aluminum rocker arms, permit the escape of air that, for instance, can get into the hydraulic element when the engine is started. Either shims such as this or very tiny bores are used to vent steel rocker arms. Starting at the oil supply bores in the rocker arm shaft, bores in the rocker arm can be used to spray the cam roller or the cam sliding surface. Rocker arms of this shape and design are found in diesel and gasoline engines. Using rocker arms makes it possible to set up two-, three-, or four-valve arrangements with just a single camshaft. Where valve trains with two intake or exhaust valves are used, double or twin rocker arms can be used to lift two valves simultaneously with a single cam. Valve play is compensated for individually, however, with the aid of hydraulic elements. It is even possible to actuate three valves (Figure 7.137). Audi uses a triple finger follower in the valve train for its V-8 engines, incorporating three intake valves. Actuation force flows from two cam lobes to two rollers in the rocker arm and then to three hydraulic compensators.

Internal Combustion Engine Handbook | 167

6606_Book.indb 167

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

valve play compensation is employed in all engine classes and even in high-speed engines such as those used by Ferrari and Porsche. The hydraulic elements consist of an outer casing in which a plunger with an integrated check valve is installed. These two parts can slide one inside the other and, at the contact surface, form a leak gap only a few micrometers wide. A spring on the inside keeps the two components apart. During the valve stroke, the valve spring and mass forces impose load on the hydraulic element. High pressure is developed in the space defined by the casing and the plunger (with the check valve closed). A small amount of oil escapes through the very narrow gap and is passed to the reserve space inside the plunger. In the following phase, while contact is made with the lobe’s circular segment (valve closed), the inside spring pushes the hydraulic element apart until the valve play is once again fully compensated for. The differential pressure thus arising causes the check valve to open; the amount of oil required for compensation can flow in. Thus, the length of the hydraulic element can change in both directions. The advantages of hydraulic compensation for valve play include

Hydraulic element support

•• simple mounting of the cylinder head (no measurement or adjustment work since the hydraulic element compensates for all tolerances) Rocker arms

Figure 7.137 Triple finger follower for the Audi V-8 engine.

In addition to the solutions previously mentioned, where the rocker arm actuates the valve directly, there are also rocker-arm valve trains that use bridges, either guided on posts or free moving, to lift two valves simultaneously. In four-valve diesel engines, including those with an inverted valve arrangement, it is possible to actuate all the valves with just a single camshaft while, at the same time, maintaining the space needed for the injection nozzles. Stiffness values in rocker arms are low, because of the geometry and, particularly, because of great distance between the cam contact point and the valve contact point, the relatively large number of contact points, and the shaft, which have to be taken into account, in addition. The much more direct force flow in cam follower designs produces far better stiffness values. 7.10.1.3 Hydraulic Valve Play Compensation For many years, one goal of engine designers has been to keep the adjustment and service work of the engine to a minimum. Thus, it is hardly a surprise that the first engines with hydraulic—and thus automatic—valve play compensation were produced well before World War II. These were, however, large-displacement engines that ran at moderate speeds. Higher engine speeds were attained in the 1970s in the Mercedes Benz V-8 engines with hydraulic screw-in elements (cam follower system). A further milestone reached in the 1970s was the introduction of hydraulic valve lifters in the V-8 engine used in the Porsche 928. Today hydraulic

•• freedom from service requirements •• constant timing at all throttle settings and at all times (no need to adjust time to account for thermal effects or wear in valve train components) •• low noise level (thanks to low opening and closing ramps at the camshaft and low opening and closing speeds). Achieving this places certain demands on the oil circuit (oil pressure and foaming). It is also necessary to observe close shape tolerances when machining the circular lobe segment. The elements could become compressible in the event of a deficiency in the oil supply (air in the high pressure chamber), which would result in insufficient valve lift and, consequently, would induce noise or changes in dynamic response at high engine speeds. The hydraulic element recognizes loss of contact force as valve play, and this could result in an undesired lengthening of the element, with the result that the valves would not close completely. 7.10.1.4 Mechanical Valve Play Adjustment Valve play is adjusted with •• screws •• adjusting shims of graduated thicknesses •• valve lifters with graduated crown thickness (only for valve trains incorporating valve lifters). Common to all three options is finite adjustment precision, which needs to be taken into account in the design of the lobe ramps for opening and closing the valves. It is necessary to measure and adjust for valve play when mounting the cylinder head. The increase in valve play resulting from wear at valve train components can be corrected by adjustments made during

168 | Internal Combustion Engine Handbook

6606_Book.indb 168

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

service work; changes in play resulting from temperature development in the engine cannot be automatically corrected. The effects enumerated here include the potential for a wide spread in the amount of play and necessitate steep ramps with great opening and closing speeds. This wide spread implies critical changes in timing and thus has negative effects on exhaust gas quality; rapid closing causes valve train noise. The advantages of mechanical valve play adjustment (compared with comparable hydraulic valve train components) include

The large valve lift can then be optimized for further increase in performance.

•• greater stiffness •• lower friction losses (by eliminating friction at the lobe’s circular segment and through modified valve spring characteristics) •• lower component costs. 7.10.1.5 Variable Valve Drive Trains Building upon the systems already explained (Sections 7.10.1.1–7.10.1.4), it is possible to respond to the needs of engine designers and the wishes of thermodynamics engineers to apply differing lift curves, selectively, to an engine valve. This is done by introducing a shifting capability into the transmission path for the valve train. Valve lift cutout and switching systems using defeatable force transmission elements such as valve lifters (Figure 7.138), roller tappets (Figure 7.139), or rocker arms have already been implemented in various series applications. A separate cam has to be provided to initiate the stroke for each additional and alternate valve stroke length—unless the alternate stroke is no lift at all.

Figure 7.138 Switchable valve lifter.

7.10.1.5.1 Stroke changeover Stroke changeover allows the operating point-dependent use of, at least, two different valve lifts. A smaller valve lift specifically adjusted for partial load is used to improve the torque progression and reduce consumption and emissions.

Figure 7.139 Switchable roller tappets.

The small valve lift with less maximum stroke and shorter event length enables a reduction of the load change work (Miller cycle) because of a significantly earlier “inlet closes” time and dethrottling in the intake tract. Similar results are possible with the Atkinson cycle, that is, an extremely late inlet closing. An optimal filling of the combustion chamber causes an increase in torque in partial load. Other consumption benefits are achieved by increasing the residual gas compatibility by, for example, asymmetric valve lifts (different small valve lifts for both inlet valve to generate the swirl) and by masking the inlet valves for a targeted charge motion. Other applications are defined with respect to new combustion processes such as HCCI which require a targeted residual gas control. Implementing this system with camshaft shifting lets us achieve thermodynamic optimization at many of the engine’s operating points, and this will be reflected in a significant drop in fuel consumption. This technology was introduced by Porsche in 1999 with the switching tappet for example, with the so-called VarioCamPlus system (Figure 7.140). In various series applications, consumption savings of up to 6% have been achieved in the NEFZ. Positive effects with respect to friction have also been proven because the partial stroke operation causes lower valve spring forces at smaller valve lifts.

Internal Combustion Engine Handbook | 169

6606_Book.indb 169

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.140 VarioCamPlus System Porsche [7-101], Wiener Motorensymposium (2000), Porsche AG, Dr. Neußer.

7.10.1.5.2 Cylinder cutout Cylinder cutout (Zylinder Abschaltung System—ZAS) is mostly used in large-volume, multi-cylinder engines (such as eight, ten, or twelve cylinders). Selected cylinders are shut down by stroke cutout at the intake and discharge valves, and the cam lobe stroke is fully decoupled. Equidistant ignition sequences make it possible to “convert” standard V-8 and V-12 engines to V-4 or six-cylinder inline engines, respectively. The purpose of cylinder cutout is to minimize the gas exchange losses (pumping and throttle losses) and/or to shift the operating point enabling higher thermodynamic efficiencies. Series applications such as the Chrysler HEMI-V8 engine demonstrate that the use of a cylinder cutout enables a fuel savings potential of approximately 10% in standard driving cycles. The previously mentioned positive effect of a reduction in friction loss in the valve train is even more prominent for the shutdown operation in cylinder cutout. 7.10.1.5.3 Technical designs By now, several different technical solutions for stroke changeover and shutdown are present in the market place. One variant is to decouple the element that follows the cam lobe from the engine valve. This “lost” motion lends its name to what is sometimes called a “lost motion” stroke; the negative mass forces here have to be absorbed by a “lost motion” spring, because the valve spring is no longer actuated. In the event of a stroke changeover, the cam now determines the valve stroke curve for the partial-load stroke. Adapted technical solutions with the same functional principle are available for the various valve train types such as lifter or cam follower drives. Another variant to use an electromechanical actuator to control the following stroke by axially shifting the cam or cam package on the shaft; this system is used in the Audi Valve Lift system, for example (Figure 7.141).

Figure 7.141 Audi valve lift system [7-102].

7.10.1.5.4 Switching component design The switching tappet consists of two coaxially arranged tappets, the interior and the outer housing. The anti-twist mechanism ensures the correct alignment of the crowned cam contact surfaces of the tappet to the cam. The advantage of a curved cam contact surface is that a larger stroke is possible, compared with a flat tappet ground. In unlocked state, the “lost motion” spring, designed as compression spring, prevents the outer housing from lifting off the cam. Because of the locking mechanism, interior and outer housings can be coupled or decoupled (Figure 7.142). Depending on the required switching strategy, the locking mechanism can be realized as depressurized locked or depressurized unlocked. The hydraulic valve clearance compensation element (Hydraulic Vacuum Assist—HVA) is in the interior housing. Switchable roller tappets as they are used in OHV engines and switchable support elements (Figure 7.143) generally have the same structure as described above; however, in these components, only a stroke shutdown or cylinder cutoff is possible. The switchable cam follower consists of two levers that can be coupled with each other and may be designed as sliding or rolling contact. The “lost motion” spring is a helical torsion spring in most cases. Figure 7.144 shows a variant being used for valve stroke change or shutdown, depending on the camshaft design. The transmission of the small stroke is realized by sliding contact in this case. A variant for a rolling contact of the small stroke is shown in (Figure 7.145). The locking mechanism of the aforementioned switching components is usually actuated via the engine oil pressure available in the location of the switching element, via the HVA supply or a separate passage. In this case, a 3/2-directional switching valve (Figure 7.146) controls the pressure buildup in the oil passages for actuating the switching elements or fast pressure reduction for switching back.

170 | Internal Combustion Engine Handbook

6606_Book.indb 170

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

Locked operation

Unlocked operation

Locked operation

Figure 7.144 Switchable cam follower for valve stroke change or shutdown.

Unlocked operation

Figure 7.142 Switching positions, coupling mechanism [7-101], Wiener Motorensymposium (2000).

Figure 7.145 Switchable cam follower for valve stroke change.

Figure 7.146 3/2-directional switching valve. Figure 7.143 Switchable support element.

Internal Combustion Engine Handbook | 171

6606_Book.indb 171

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.147 Oil circuit for the actuation of the switching elements.

The switching valve is electrically controlled by the Electronic Control Unit (ECU) to a stored map. There are alternative technical solutions for a direct electromagnetic control of the locking mechanisms which now pose major challenges in their arrangement in the cylinder head. Switching times between 10 and 20 ms can be achieved in a well-tuned hydraulic system which allows a switchover within one camshaft rotation in the customarily required torque range. When designing the switching oil circuit in the engine (Figure 7.147), the position and geometry of the passages must be carefully considered to realize a system that is as hydraulically stiff as possible, to prevent air accumulations, and to avoid throttling locations. Optimal response behavior is ensured by continuous spraying of the switching passage at reduced oil pressure. This process quickly purges any accumulated air bubbles and ensures a low pressure level for the preload of the hydraulic oil circuit. 7.10.1.5.5 Fully variable valve trains Among the fully mature embodiments of fully variable valve trains is the BMW Valvetronic concept. It offers great benefits in terms of consumption as well as in retaining stoichiometric operation with all its advantages and, in addition, can be used all around the world, regardless of fuel formulations (sulfur content). The Valvetronic achieves engine operation without the need for a butterfly throttle valve. Cylinder fill at partial load is regulated by the intake valve lifting stroke and opening period. The intake and exhaust camshafts are driven by variable cam adjustment. To achieve stepless adjustment of intake valve stroke, an intermediate lever, backed against an eccentric shaft, is inserted between the camshaft and the cam follower. The contour of the contact surface between this intermediate

lever and the roller cam follower defines the valve lifting curve. Rotating the eccentric shaft moves the fulcrum for the intermediate lever and thus—steplessly—changes the lever ratio and, consequently, the relationship between the cam lobe stroke and the valve stroke. In this way, it is possible to achieve valve excursion from approximately 0.3 mm at idle to 9.7 mm at full throttle [7-103]. Greater freedom is possible with electro-hydraulic valve trains, such as the UNIAIR system (Figure 7.148) being developed by Schaeffler KG. Here, a valve lifter or lever with rolling contact transfers the cam contour to a pump unit. The oil is moved into a pressure chamber which can be sealed at the other end by a switching valve. If this is the case, pressure built up in this chamber actuated the engine valve using a piston. The switching valve now provides the option to open and close the high-pressure chamber at any time. This allows, in addition the full stroke in which the switching valve remains closed for the entire cam stroke, a number of variations of the valve stroke design (Figure 7.149). If the switching valve is opened during the engine valve lift, the pressure in the high-pressure chamber drops and the engine valve spring closes the engine valve [early valve closing (EVC)]. The closing action is thus not cam-controlled but is similar to a ballistic flight. To reduce the valve contact speed, a hydraulic brake delays the engine valve by approximately 1.5 mm before contact. The oil escaping from the high-pressure chamber during EVC is pushed through a check valve in the mean pressure curve isolated from the oil supply system into a pressure receiver. When the high-pressure chamber is being refilled, the pressure receiver returns the energy stored in its spring to the camshaft. If the switching valve is closed only after the cam stroke has started, the pump first pushes oil from the high-pressure chamber into the pressure receiver. Upon closing of the

172 | Internal Combustion Engine Handbook

6606_Book.indb 172

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

switching valve, the lift of the engine valve starts and the full-stroke curve is thus shifted downward (late valve opening). In this case too, as in all modes, the valve is not closed camcontrolled, but the hydraulic brake determines the contact. If the switching valve remains open for the entire cam stroke, the engine valve does not lift. Thus, using the UNIAIR system, the valve stroke can be individually controlled for every engine valve, being limited by the full-stroke curve determined by the cam geometry. The required control electronic system communicates with the engine control unit and actuates the switching valves, depending on the current demands on the valve stroke and sensor data such as oil temperature. This fully variable valve control can be used equally in gasoline and diesel engines and supplied from the existing engine oil circuit. In SI engines, it enables a throttle-free load control, improvement in mixture formation, and the generation of load motion. In diesel engines, the regulation of the internal exhaust-gas recirculation and the option of a homogeneous combustion (control of the air mass in the cylinder) can be named as benefits, in addition to the generation of the air movement in the combustion chamber (swirling). This creates potential for lowering consumption, increasing performance and torque, and improving the emission behavior. The first series application of the FIAT powertrain (called “MultiAir”) features an intelligent architecture that enables the actuation of both exhaust valves and the pump drive with a single camshaft. The pump drive simultaneously influences both intake valves over a so-called hydraulic bridge (Figure 7.150).

Figure 7.148 Concept representation of the UNIAIR system. See color section page 1070.

solenoid status

Solenoid Switching

0

20

40

LVO

60

80 100 120 angle camshaft [°]

140

160

180

200

EVC

UniAir Valve Control Modes

10

valve lift [mm]

8 6 4 2 0

0 full lift

20

40

60

80 100 120 angle camshaft [°]

140

160

180

200 no lift

Figure 7.149 Valve lift modes, UNIAIR system. See color section page 1070.

Internal Combustion Engine Handbook | 173

6606_Book.indb 173

1/19/16 8:32 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Ideal adjustment of belt tension is not possible in such systems because they cannot compensate for belt tension fluctuations caused by temperature changes, aging, or for dynamic effects (belt vibration, reverse influences from the valve train, etc.). Compensating for such fluctuations and effects is absolutely necessary in modern synchronous belt drives because only in this way, one can achieve the targeted system service lives (corresponding to engine service life). The influence that a fixed idler (tensioner) pulley has on static belt preload is illustrated in Figure 7.151. Change of the static belt preload value as function of the motor oil temperature

Figure 7.150 Hydraulic bridge.

7.10.2.1 Introduction For 40 years now synchronous belts have been used successfully in mass production to drive camshafts and balancer shafts in internal combustion engines. The first application was in a four-cylinder engine made by Glas without any auxiliary components such as idler or deflection pulleys. In later designs, the preload in the toothed belt was generated either with a component that was driven by the toothed belt (such as the water pump) and mounted on an eccentric bracket or by fixed idlers (eccentric tensioning pulleys or the like).

Stat. belt preload F

7.10.2 Belt Tensioning Systems, Idler, and Deflection Pulleys

Area of risk of impermissible toothed belt chatter Belt force spread when using an automatic belt tensioner

Belt force spread due to setting tolerance when using a rigid tensioning pulley

Automatic belt tensioner

Rigid tensioning pulley

Area of risk of impermissible toothed belt oscillations 0

20

40

60

80

100

120

Engine oil temperature [°C]

Figure 7.151 Changing the static belt preload value using motor oil temperature as the lead variable; comparison between fixed tensioning idler and automatic tensioning system.

Figure 7.152 Development of synchronous belt profiles.

174 | Internal Combustion Engine Handbook

6606_Book.indb 174

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

Using an automatic belt tensioning system makes it possible to considerably reduce the spread in preload values at initial assembly and to keep preload values nearly constant across the engine’s full operating temperature range. Automatic tensioners have been used with synchronous belt drives for internal combustion engines since the beginning of the 1990s and, for the reasons mentioned above, have forced fixed systems almost completely off the market. 7.10.2.2 Automatic Belt Tensioning System for Synchronous Belt Drives The primary requirements for automatic tensioning systems are derived from the conditions enumerated above and are the following: •• setting specified belt tension at initial installation and after service (compensating for belt, diameter, and positioning tolerances) •• maintaining the most constant belt tension possible at all operating states across the entire required system service life (compensation for thermal elongation, belt stretch, and wear, taking account of crankshaft and camshaft dynamics) •• ensuring ideal noise levels while at the same time reducing belt vibration •• preventing tooth jump. The parameters shown in Figure 7.153 have to be considered when specifying the working range for a tensioning system such as this.

Belt preload value FT [N]

Influenced by: Operating temperature range of the engine Camshaft and crankshaft dynamic Belt stretching and wear

A1

N

A2 Angular travel

Figure 7.153 Mechanical synchronous belt tensioning unit—sample operational chart with influencing parameters.

Of the various styles for synchronous belt tensioning systems (with hydraulic damping, linear action with reversing lever; with hydraulic damping and rotating; and with mechanical damping and rotating), the rotating mechanical systems are most widely used for reduced costs and less space required. The temperature-based tensioning systems using wax thermostats, employed in some engines in the past, never made a breakthrough. The basic design of a mechanical tensioning system such as this, using the so-called double-eccentric principle, is shown in Figure 7.154. Here the adjustment eccentric compensates for the tolerances in all the components present in the belt drive; its setting is fixed after initial adjustment. The working eccentric mounted

movably on the adjustment eccentric compensates for the temperature-induced changes in length at all the components used in the belt drive, for belt stretch and wear, and for dynamic effects originating at the crankshaft and camshaft. Lever spring

Adjustment eccentric

Adapter plate

Tensioning pulley

Plain bearing Working eccentric Setting screw

Figure 7.154 Mechanical synchronous belt tensioning unit with doubleeccentric design.

The lever spring is designed in accordance with ideal belt preload. Damping is affected with the slide bearing, and the tensioner’s geometry is matched to the requirements of the belt drive. 7.10.2.3 Idler and Deflection Pulleys for Synchronous Belt Drives It is for the foregoing reasons that fixed tensioning pulleys are found only rarely in modern engines. Deflection pulleys used, for instance, to calm critical sections of the belt, to avoid collisions with adjacent components, or to increase the wrap angle at neighboring pulleys have to satisfy the same requirements regarding service life and noise development. High-precision, single-row ball bearings with enlarged grease reserve spaces have proven their suitability for this application; for improved belt guide, double-row angular ball bearings, also with optimized reserve grease spaces, may be used. These bearings are normally packed with high-temperature rolling bearing grease and fitted with suitable sealing rings; standard bearings are less suitable for this purpose. These bearings serve as the centers for pulleys that match the prevailing geometric requirements. Exemplary embodiments with plastic and steel pulleys are depicted in Figure 7.155; these pulleys may be equipped with flanges on one or both sides to guide the belt. 7.10.2.4 INA Durability Drive INA Durability Drive (ID²) represents a system for the reduction of forces and vibrations in synchronous belt drives. By using pulleys perfectly matched to each belt drive system, these effects can be attained: •• reduction of the forces occurring in the belt drive by up to 30% •• reduction of the camshaft vibration angle by up to 60% •• noise optimization by lowering the power level. The two diagrams below (Figure 7.156) illustrate the effects—the curves in black show the initial state, while the

Internal Combustion Engine Handbook | 175

6606_Book.indb 175

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

green curves display the optimization options. The following options are the result: •• cost-optimized synchronous belt drive systems by downsizing the belt drive components •• significantly longer service life of the synchronous belt drives, compared with standard designs •• representation of critical drive configurations with synchronous belts.

Belt force [N]

Figure 7.155 Deflector pulley with single- and double-row ball bearings and pulleys in plastic and steel.

7.10.2.5 Prospects for the Future Modern synchronous belt drives for internal combustion engines are no longer conceivable without automatic tensioning systems because only with their help the required system service life can be achieved. Because of cost and space considerations, hydraulically damped systems are being replaced to greater extent by those that are mechanically damped. Key areas of emphasis in development work at the present are mechanical tensioning systems for the heavily loaded synchronous belt systems used in diesel engines, systems designed to facilitate

1400 1200

Series

1000

ID²

800 600 400 200 0 800

1200 1600 2000 2400 2800 3200 3600 4000 4400 4800 5200 5600

Angular displacement [°Cam]

Crankshaft speed [1/min]

2 1,8 1,6 1,4 1,2 1

Series

0,8 0,6

ID²

0,4 0,2 0 800

1200

1600

2000

2400

2800

3200

3600

4000

4400

4800

5200

5600

Crankshaft speed [1/min]

Figure 7.156 Effects of the ID² system. See color section page 1071.

176 | Internal Combustion Engine Handbook

6606_Book.indb 176

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.10 Valve Train Components

installation, and mechanical tensioning systems (with either open-loop or closed-loop control) for ideal matching of the preload force to engine operating conditions. Friction optimization is also possible, compared with chain drives. In oil environments (belt in oil) in particular, significant saving potentials are anticipated.

7.10.3 Chain Tensioning and Guide Systems 7.10.3.1 Introduction In addition to the synchronous belt tensioning and deflection systems described in Section 7.10.2, timing chains have become established in the control drive and various secondary unit drives such as oil pump drives, balance shaft drives, and more. Continuous further development of the individual components and technical detail optimization help to satisfy the continuously increasing customer requirements, although the chain as a drive in the combustion engine has, in principle, a long history—the first applications were developed more than 100 years ago. Chain drives are characterized by

7.10.3.2 Chain The chain itself is one of the most important components to be considered when designing a timing drive. Depending on the moments and engine speed to be transmitted, the partition and the chain type (single- or multiple-chain, roller, bushed roller, or silent chain) must be selected. Chains in combustion engines must always be designed with even link numbers. The partition is of utmost importance for the system size, and it also influences the noise behavior (polygonal effect). In addition to considering the number of limit teeth (for example, eighteen in passenger car crankshafts), a constant chain line (geometric progression of the chain in the engine) is important for the limit value analysis of tolerances and chain length. For the dynamic design of the chain drive system using simulation programs, chain data such as stiffness, mass assignment, and tooth profiles are of critical importance. Even similar components can have major differences here (Figure 7.158).

•• little space requirement •• maintenance free (no replacement interval) •• high stability •• reliable transmission of large torques •• far-reaching adaptability of their geometry to the given space. In addition to the chain as the most important element in a timing drive, the chain tensioning element is very important. In contrast to belt drives, the use of free span lengths in chain drives is very limited so that the tensioning and guide rails used to guide the chain attain great significance and, in exceptional cases, sprockets. All of these components will be discussed in more detail below. The terms used here are explained in Figure 7.157 on the basis of a DOHC timing drive configuration.

Figure 7.158 Timing chains with 9.525-mm partition of different types.

The selection of the chain also affects the design of subsequent systems such as the camshaft shifter. Here, the sprocket of the camshaft is a fixed component of the shifter and affects so the size of this element. The design of both elements must therefore be matched geometrically and dynamically. 7.10.3.3 Chain Tensioning Element The principal task of the chain tensioning element or chain tensioner is the control of the high-dynamic vibrations of the chain drive caused by the uneven rotation of the crankshaft and the alternating drive moments of the camshaft. The tensioning element must ensure, at all times and under all operating conditions, problem-free transport of the chain over the tensioning rail. It must also meet further requirements: •• compensating the tolerances in the chain drive by adjusting the chain line

Figure 7.157 Two-art chain drive.

•• preloading the chain drive to keep the chain from “climbing” or jumping on the sprockets (at high speeds and/or a chain lengthened by wear)

Internal Combustion Engine Handbook | 177

6606_Book.indb 177

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• adapting the chain line for axle distance changes caused by heat-induced expansion

valves. These valves are intended to enable a targeted relief of the chain drive at specific chain loads or vibration frequencies. Under certain circumstances, all other components may be designed with smaller dimensions, thus saving space and costs. The correct tensioner adjustment is usually determined by simulation programs which are verified by measurements during dynamic engine testing. It is becoming increasingly important in the design of modern chain drives and must, of course, consider the effects on other systems such as the camshaft adjuster or valve train (see also the note for Section 7.10.3.2). This task becomes interdisciplinary and requires knowledge of the overall system beyond the chain drive. Mechanical chain tensioners are usually used when a chain drive comprises solely units with low change moments. Targeted damping is usually omitted in these tensioners. They frequently consist of a spring—often designed as a lever spring—and a tensioning rail. If a low spring force is sufficient for controlling the chain drive, inexpensive plastic components may be used.

•• compensating for the chain wear occurring during the engine’s service life. As a rule, chain tensioners are designed as hydraulic components. The most familiar type of a hydraulic chain tensioner is the speed-proportional seepage gap damper with directional damping. Such elements are connected to the engine oil circuit via an oil supply bore. When the slack span in the chain drive is relieved, the plunger preloaded by the pull-back spring is extended from the housing and presses the tensioning rail to the chain. The check valve opens, and oil is drawn into the high-pressure chamber of the tensioner. The extension movement of the plunger is thus undampened. If the load in the slack span is reversed, the check valve closes, the plunger is loaded, and oil is forced from the high-pressure chamber through a small ring gap (seepage gap) between the piston and housing of the tensioner. The chain vibration is so dampened. Hydraulic tensioning elements are characterized by

7.10.3.4 Tensioning and Guide Rails Tensioning and guide rails are used to guide the chain along the spans. This prevents the chain from vibrating in the drive plane—the so-called transversal vibration. Guide rails are solidly connected with the engine via screw or plug connections. Tensioning rails, on the other hand, can be designed as fulcrum rails and are also called tensioning levers. The chain tensioner and its piston form the second abutment for the rail, while the first is the fulcrum.

•• low component wear (often from hardened steel) •• precisely adjustable damping because of leak gap design •• targeted damping by using a check valve •• little space requirement •• cost-effective production thanks to multiple use of components. In addition to the described simple elements, there are also tensioners with auxiliary functions designed as overpressure

Piston position after extended operating time Oil supply bore Leak gap

Piston position at new engine Transport circlip

Housing High-pressure chamber Tensioner rail Piston Oil reservoir

Sealing screw Return spring Valve unit

Figure 7.159 Schematic representation of a hydraulic tensioning element in installed state.

178 | Internal Combustion Engine Handbook

6606_Book.indb 178

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.11 Valves

When designing the rails, which is usually done by use of finite element calculations followed by functional and costrelated optimization, influencing parameters such as load, space, application temperature, and costs must be considered. Available are (Figure 7.160) •• single-component rails from plastic •• two-component rails with a sliding coating of non-reinforced plastic and bearing bodies from aluminum, sheet metal or plastic (usually fiber-reinforced).

Figure 7.161 Sprockets.

7.11 Valves

Figure 7.160 Chain tensioner for camshaft drives.

In addition to the tensioning rails described above, there are also tensioning shoes solidly connected to the piston of the tensioning elements and executing a translatory motion. They are used in very short chain drives (such as camshaft drives). 7.10.3.4 Sprockets Sprockets are used to transmit the rotatory movement of the engine shafts to the translatory motion of the chain. Low-wear and quiet operation are achieved by accurately fitting the shaft hub connection and a tooth profile optimally matching the chain. While the tooth shapes in roller and bushed roller chains are mostly following standards, toothed chains may feature a multitude of different profiles which frequently differ only in minor details, but are adjusted for the chain used and often protected by patents. Taking economical factors, the connection geometry to be achieved and, above all, the loads occurring into account, several options are possible. As a rule, one differentiates by •• manufacturing: precision blanking, chip removal, forging or flow pressing, powdered metallurgy (sintering) •• configuration: single- or multi-web, roller, bushed roller, or toothed chain sprocket •• partition: 6.35 mm (1/4″), 7 mm, 8 mm, 9.525 mm (3/8″), or special partitions.

Gas exchange control is of critical importance for the reliability of a piston engine. The disk valve has been found to be very suitable with respect to function and operational reliability. With respect to thermodynamic and fluid mechanics, however, the disk valve is not optimal in all areas. For example, the valve stem leading through the gas channels represents an obstacle in reducing the cross section, and the geometry of fillet and seat may be only a compromise between mechanical or tribological requirements and fluid-mechanical ideal. In addition, the dynamics of traditional valve trains quickly reaches its limits in the upper speed range in particular. A faster clearance of the opening cross section would be advantageous to the maximally achievable performance and fuel consumption. The attempt to avoid the dynamic and fluid-mechanical issues of the disk valve control resulted in the past in the design of various slider controls following a common goal: achieving large time cross sections at total independence of the engine speed. Problems with heat removal, sealing, and operational reliability, however, make the slider control attractive only when the engines are not required to have long service lives. Hence, current developments have the objective to create fully variable valve actuations on the basis of the disk valve. Despite all of its problems, the disk valve has a decisive advantage: under internal pressure, it seals automatically. No other process is capable of doing so with the same reliability in combustion engines [7-104].

7.11.1 Functions and Explanation of Terms and Concepts

Intake and exhaust valves are precision engine components used to block gas flow ports and to control the exchange of gases in internal combustion engines. They are intended to seal the working space inside the cylinder against the manifolds. An example of a valve in place in the engine is shown in Figure 7.162. The intake valves, which are not subjected to such extreme thermal loading, are cooled by incoming gases and mainly by thermal transmission at the seat. Exhaust valves, by contrast, are exposed to severe thermal loads and chemical corrosion. These two types of valves are manufactured using different materials, matched to the demands to be placed on them. It may be assumed that, during an engine’s service life, the valves will execute about 300 million operating cycles, many

Internal Combustion Engine Handbook | 179

6606_Book.indb 179

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

at very high temperatures. The most important terms used to describe valves are depicted in Figure 7.163.

Figure 7.162 Valve in place in the engine.

7.11.2 Types of Valves and Manufacturing Techniques

Gas exchange valves may be manufactured either in the hot extrusion process or by upsetting processes. The starting point in the hot extrusion process is a section of rod whose diameter is about two-thirds of the final disk diameter; its length corresponds to the volume of the blank to be manufactured. This rod is heated and reformed to make the blank in two forging steps. During the upsetting process, a ground section of rod, the diameter of which is slightly greater than the ultimate valve stem diameter, is first heated at the end; the rod is then forced forward to form a “pear,” which is then reformed in a die to create the valve head. Valves may be subdivided essentially into three major groups: monometallic valves, bimetallic valves, and hollow valves. 7.11.2.1 Monometallic Valves Monometallic valves are manufactured in a single piece in the aforementioned processes. The most commonly used material is X45CrSi93 (1.4718).

7.11.2.2 Bimetallic Valves Bimetallic valves permit the combination of materials, each matched exactly to the needs of the valve stem and the valve head. Here one works on the basis of a heat-reformed head that is made in the process described in Section 7.11.2 and then attached to the stem by friction welding (Figure 7.164). The preferred materials for the head piece are X50CrMnNiNb219 (1.4882), X60CrMnMoVNbN 2110 (1.4785), NiCr20TiAl (2.4952), and X45CrSi93 (1.4718) for the stem. The welding seam is positioned so that, when the valve is closed, the seam is inside the guide by half of the valve lifting stroke and/or about 6 mm above the sweeping shoulder (see also Section 7.11.3.3). Here, for reasons dictated by the

Figure 7.163 Terminology used for valves.

manufacturing technology, it is necessary to ensure that the length of the cylindrical section on the head itself is at least 1.5 times the stem diameter before welding. Austenitic heads in bimetallic valves are usually not sufficiently wear resistant and, thus, require suitable reinforcement of the seat (see also Section 7.11.3.2). 7.11.2.3 Hollow Valves This version is used primarily for the exhaust valves and, in certain special circumstances, on the intake side as well, to lower the temperatures primarily in the concave area at the back of the head and at the disk area and for weight reduction. If one employs hollow valves to reduce temperatures, then approimately 60% of the volume inside the hollow space will be filled with metallic sodium. The sodium can move freely in the hollow cavity of the valve stem.

180 | Internal Combustion Engine Handbook

6606_Book.indb 180

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.11 Valves

Figure 7.164 Bimetallic valve.

This liquid sodium (melting point 97.5°C) is shaken inside the valve cavity to an extent corresponding to the engine speed. It transports heat from the valve head into the valve stem and from there through the valve guide into the cooling circuit. The degree of temperature reduction at perfect thermal energy flow and the smallest possible working clearances is in the range from 80 to 150°C.

Figure 7.165 Hollow valve.

7.11.2.3.1 Hollow valve variants •• “Tube on solid metal” version (Figure 7.165): The head, which is drilled from the stem end (forming the tube), is attached by friction welding to a (solid) stem end section, which is alloyed so it can be hardened. •• “Closed-off” version: This version is far more elaborate in its manufacture than the version described above. The basic body is also drilled from the stem end. The bore is closed with inductive heating and subsequent closing forging. The stem end section is attached using friction welding. Such closed-off hollow-stem valves are used primarily in highperformance engines and aviation applications. •• Hollow valve (Figure 7.166): This valve represents a further measure taken to reduce weight and enhance heat transfer away from the center of the valve disk. These valves, in contrast to the above-mentioned designs, are drilled and machined from the disk end. The opening is closed by welding-in a capping plate. These valves, more expensive to manufacture, are used primarily in racing engines. Hollow valves may be made at stem diameters of 5 mm and upward. The diameter of the internal bore is approximately 60% of the stem diameter. To avoid exposing the valve stem seals to excess temperatures, the bore inside the valve has to end about 10 mm away from the contact range for the sealing lip. Changing of the play between valve stem and valve guide because of higher temperatures must be considered when comparing with solid valves.

Figure 7.166 Hollow valve.

Hollow valves may be of a single metal, but bimetallic valves with the following combinations of materials are more common: Head: X50CrMnNiNb219 (1.4882) and NiCr20TiAl (2.4952); stem: X45CrSi93 (1.4718).

7.11.3 Embodiments 7.11.3.1 Valve Head The theoretical diameter of the valve seat is the basis for the engineering design of the valve.

Internal Combustion Engine Handbook | 181

6606_Book.indb 181

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

The overall disk height depends on the combustion pressure and the average valve component temperature. This height is established by FEA. Practice has shown that values of from 7% to 10% of valve head diameter are required. The thickness of the edge of the disk determines the stiffness of the valve head and is coordinated with the valve seat angle; at 45°, it is approximately 50% of the overall disk height, and at 30°, approximately 50%–60% of the overall disk height. The valve seat angle is generally 45°. Seat angles of 30° and 20° may, however, also be selected to reduce valve seat wear. Small seat angles are indispensable in gas-fired engines to keep seat wear within permissible limits. Manufacturing technology requires a difference of at least 5° between the valve seat angle and the valve face angle (Figure 7.163). The differential angle between the valve seat and the seating ring achieves initial sealing along a line of contact, thus creating better seal of the face against the combustion chamber (Figure 7.167).

Figure 7.167 Differential angle and valve seat width.

The valve seat width is always specified larger than the seating ring contact width in the cylinder head. The overage between valve seat and seating ring should be realized in a ratio of 1/6 free valve seat, 2/3 overage, and 1/6 free valve seat to ensure sealing during seat wear (Figure 7.167). Curved depressions on the valve face are provided to reduce valve weight, to influence combustion chamber shape, and to distinguish between intake and outlet valves or similar valves. The ideal shape of the concave area is determined according to the existing mechanical stresses and fluidmechanical considerations.

7.11.3.2 Valve Seat The seat for the exhaust valve is heavily impacted by heat and corrosion, which is why, as a rule, it is hard-faced with special alloys. In isolated cases, this is also done for the intake valve even though here martensitic hardening is normally used because of the material selected. Hard facing can be used to reduce wear and enhance the sealing effect. The following processes are used for valve hard facing: •• fusion welding, in which the hard-facing material in rod form is melted and applied by means of an oxyacetylene flame •• electrical plasma-transferred arc process in which the pulverized hard-facing material is melted in a plasma arc and applied to the workpiece. These hard-facing techniques are used for hollow valves, bimetallic valves, and occasionally for monometallic valves as well. To keep any reduction in hardness at the inductively hardened valve seat within acceptable limits, it is necessary to ensure that valve temperatures have sufficient safety relative to the starting temperature of the material used (for example, the maximum continuous operating temperature of X45CrSi93 (1.4718) is approximately 500°C). 7.11.3.3 Valve Stem This component is used to guide the valve inside the valve guide and is defined by the first keeper slot provided to mate with the conical keeper and by the sweeping shoulder and/ or the transition to the concave area at the back of the head. To limit the formation of soot on the end toward the gas port, a sweeping shoulder is made by narrowing the stem diameter (Figure 7.163). When the valve is closed, this shoulder should be inside the valve guide by approximately one-half of the valve-lifting stroke. If, during the valve closing phase, bending is induced because of cylinder head warping or coaxial errors of the centerlines, then it is desirable for the welding seam to be inside the valve guide. This is why the friction-welding seam for bimetallic valves is positioned to at least one-half stroke length inside the valve guide. Depending on the tribologic situation, it may be necessary to protect the valve stem surface against wear by using chrome plating or nitriding. As a rule, the valve stem is cylindrical in shape. To take account of the variations in expansion because of the temperature gradation between valve head and end of stem, the valve stem may be tapered between 10 and 15 μm, depending on the length and diameter of the stem. Valve stem ends featuring multiple keeper slots (Figure 7.168), the purpose of which is to support unrestricted valve rotation, is always inductively hardened in the area where the keeper makes contact to avoid wear. Valves with a single keeper slot are seldom hardened because there is no relative movement between stem and keeper (see also Section 7.11.6.1). Depending on the valve actuation acting on the stem end surface, it may be necessary to additionally protect the stem

182 | Internal Combustion Engine Handbook

6606_Book.indb 182

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.11 Valves

end surface from wear. Inductive hardening is the preferred method of achieving this. If this measure is found to be insufficient or the stem is made from a material that cannot be hardened, hard surfacing may be used or platelets from a hard or hardening-capable material may be welded on the stem end.

Single-row slot

Triple-row slot

Figure 7.168 Types of keeper slots in valve stems.

7.11.3.3.1 Valve guide The valve guide ensures the centering of the valve on the valve seat and the heat removal from the valve head through the valve stem to the cylinder head. This necessitates an ideal clearance between the guide bore and the valve stem. If there is insufficient clearance, then the valve tends to stick. Too much clearance interferes with heat dissipation. One should strive to achieve the smallest possible valve guide clearance. Depending on the stem diameter, this may be 0.03–0.08 mm for outlet valves and 0.01–0.07 mm for intake valves. In addition, it is necessary to ensure that the end of the valve guide does not protrude unprotected into the exhaust port as otherwise there is a danger of the valve guide dilating and combustion residues entering the valve guide. As a rule of thumb, the length of the valve guide should be at least 40% of the length of the valve. To ensure perfect valve functioning, it is essential that the offset between the centerlines of the valve shaft and the seating ring be kept within certain limits (0.02–0.03 mm in a new engine). Excessive misalignment can cause, above all, serious bending of the valve disk in relationship to the stem. This excessive loading can lead to premature failure; other consequences may also be leaks, poor heat transmission, and high oil consumption.

7.11.4 Valve Materials

The demands made on a valve include endurance strength at elevated temperatures, wear resistance, resistance to high-temperature corrosion, as well as oxidation and corrosion resistance. The standard valve materials are as follows: •• Ferritic–martensitic valve steels: X45CrSi93 (1.4718) is the standard choice for monometallic intake valves and is used exclusively as the material for the stem in bimetallic valves.

X85CrMoV182 (1.4748) is a higher alloy and is used as an intake valve material where the thermal and mechanical loading does not permit the use of the Cr–Si material. •• Austenitic valve steels: Here, the austenitic Cr–Mn steels have proven to be an economical solution. A widely used choice is the X50CrMnNiNb (1.4882) alloy, which is deemed to be the classic exhaust valve material—for hollow valves, too. •• Valve materials with high nickel content: If the Cr–Mn steels no longer satisfy thermal requirements, then a transition to materials with high nickel content (such as NiCr20TiAl—2.4952) is the solution. They are necessary where maximum operational reliability, providing resistance to spalling and corrosion, is needed. Application examples include aviation engines, racing engines, highly turbocharged diesel engines, and for engines using heavy oil as the fuel. 7.11.4.1 Heat Treatment Closely defined heat treatment makes it possible to further improve the technical characteristics of the valve steels. In many cases, this can obviate the need for going to higher quality alloys. Martensitic valve steels are generally hardened. The hardness and strength of austenitic steels can be boosted by so-called structural (precipitation) hardening. 7.11.4.2 Surface Hardening The following techniques may be used: •• Hard chromium plating for the valve stem: The manufacturing process, choice of materials, and operating conditions may make it necessary to chromium-plate standard valves at the contact area along the stem. In standard bimetallic valves, the chromium layer, from 3 to 7 μ m thick, covers both valve materials. Thicker applications of chromium, up to 25 μ m, may be employed in truck or industrial engines where there are high load levels or where there is more severe wear. •• Abrasive polishing: In all cases, the stem has to be polished whenever the valve is chromium plated to remove any chromium nodules still present and to level out any unevenness. Roughness after the polishing operation is a maximum of Ra 0.2 (maximum Ra 0.4 for non-plated), which has a very favorable effect on valve guide wear and thus permits engineering for minimum clearance. •• Nitriding the valves: Bath immersion and plasma nitriding are used. The nitriding layers, approximately 10–30 μ m in thickness, are extremely hard at the surface (approximately 1000 HV 0.025) and are particularly insensitive to wear. Like chromium-plated valves, immersion-nitrided valves are abrasively polished to finish them.

7.11.5 Special Valve Designs 7.11.5.1 Valves with Material-Related Low Mass In view of the theoretically possible savings in consumption, the interest in lightweight valves has significantly increased in past years. The idea of using lightweight valve materials

Internal Combustion Engine Handbook | 183

6606_Book.indb 183

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

is not exactly new, however. In high-speed racing engines in particular, lightweight valves have been successfully used for a long time. The main focus in the past was to achieve very high speeds and, thus, high performance. Today, new emission legislation has added to the requirements. To reduce the emission of pollutants and to lower fuel consumption at the same time, ever higher peak temperatures and pressures become necessary. For this reason, the efforts in developing lighter, highly heat-resistant valve materials have been intensified. Until now, very different approaches have been followed. In addition to ceramic materials such as silicon nitride, inter-metallic titanium aluminide, aluminum, and titanium alloys present themselves as metallic materials for lightweight valve concepts. Starting with the stress profile of the intake and discharge valves, ceramic and TiAl valves are well matched as exhaust valves because of their high-temperature resistance. Titanium valves can be used in a limited fashion only at the exhaust end. Al alloys are suitable as inserts for intake valves and present a less expensive solution. 7.11.5.2 Exhaust Control Valves 7.11.5.2.1 Turbocharger regulation valves (overrun control valve) The overrun control valve limits the charging pressure developed by the exhaust turbocharger and, in gasoline engines, can be exposed to temperatures of approximately 1050°C; the thermal load in the diesel engine is approximately 850°C. This is the criterion used when selecting engineering materials. Diesel engines can usually get along with the X50CrMnNiNn (1.4882), while a material that can withstand high temperatures, such as NiCr20TiAl (2.4952), is used in gasoline engines. Two designs are commonly used. Typical designs of a type similar to a disk valve are shown in Figure 7.169. The overrun control valve is secured with screws or rivets. In modern passenger car engines, this design is less common however, and is usually replaced by the flap valve shown in Figure 7.170.

Figure 7.170 Overrun control valve with flap mechanism.

7.11.5.2.2 Exhaust gas return valve Exhaust gas return valves have to cope with temperatures of up to approximately 800°C. Of the valve materials available for use, X50CrMnNiNb (1.4882) has been found to be sufficient for this application because the valves are subjected to thermal stress only; they have moderate exposure to corrosive effects and very little mechanical loading.

7.11.6 Valve Keepers 7.11.6.1 Tasks and Function The purpose fulfilled by valve keepers is joining the valve spring collar with the valve in such a way that the valve spring always keeps the valve in the required position. Cold-embossed valve keepers are state of the art for valve stems up to 12.7 mm in diameter. The C10 and/or SAE1010 material qualities are used. The valve keepers are classified according to their function as follows: •• clamping connection creating a frictional connection among the valve, the valve keeper, and the valve spring collar •• Non-clamping connection, which allows for unrestricted valve rotation. 7.11.6.1.1 Clamping connections Clamping valve keepers transfer force through a frictional connection. To achieve this, it is necessary that a narrow gap be maintained between the two halves of the valve keeper. That is why valve keepers with conical angles of 14°, 15°, and 10° are used. Valve keepers with smaller conical angles bring about far more intensive clamping action. They are suitable particularly for engines that run at extremely high speeds. Where the clamped connections are heavily loaded, the use of case-hardened (480–610 HV1) or nitrided (> 400 HV1) valve keepers is recommended

Figure 7.169 Valve designs of overrun control valves.

184 | Internal Combustion Engine Handbook

6606_Book.indb 184

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.11 Valves

Figure 7.171 shows an example of a clamping valve keeper in its installed position.

Valve Valve spring collar

Valve Valve spring collar

Valve cone No gap Valve cone

Gap

Figure 7.172 Installation principle for valve keepers with non-clamping connection. Figure 7.171 Installation principle for valve keepers with clamping connection.

7.11.6.1.2 Non-clamping connections A non-clamping connection is achieved by using valve keepers with a conical angle of 14°, 15′. Because of the fact that the two halves of the valve keeper, when installed, rest against each other at flat surfaces, they provide clearance between the valve keeper halves and the valve stem. This allows the valve to rotate in the spring collar. Rotation is supported by vibration, by eccentric contact between the rocker arm and the end of the valve stem, and by the impetus provided by valve lifter rotation. When a non-clamping connection is used, the forces along the axial direction are transferred by the three or four beads inside the valve keeper. That is why case hardening of the valve keeper is indispensable. Figure 7.172 shows an example of a clamping valve keeper in its installed position. 7.11.6.2 Manufacturing Techniques Valve keepers are cold-pressed from profiled strip steel. Multislot valve keepers are always case hardened and ground at their mating planes. Other versions may be used without hardening or with case hardening or they may be nitrided, as desired. Manufacturing may require that the outside jacket, about halfway down the side, be made concave by as much as 0.06 mm, the amount depending on the exact design. The outside jacket should never be convex.

In freely rotating multi-slot valve keepers, the correct valve stem clearance is achieved by dimensioning an inner diameter of the combined keepers 0.06 mm smaller than the external diameter of the valve stem. The conical section in the spring collar has to be long enough that the valve keeper does not hang over at either end when installed (Figure 7.171 and Figure 7.172). The conical jacket may, in no case, be convex and should serve as the reference surface for the dimensional and positioning tolerances in the spring collar.

7.11.7 Valve Rotation Devices 7.11.7.1 Function Regular rotation of the valve is of critical importance to its ideal functioning. In this way, valve head temperatures are stabilized and leaks because of warping are avoided. Deposits at the valve seat and one-sided wear are also reduced. Positiveaction rotation devices are used in industrial engines, for example, wherever natural valve rotation is not sufficient. 7.11.7.2 Designs and Functioning Valve rotators function according to one of two principles: •• Rotation during the valve opening stroke: The system comprises a round base featuring several oblong slots along its circumference. Mounted in each slot are a ball and a coil spring that force the ball to the upper end of an inclined

Internal Combustion Engine Handbook | 185

6606_Book.indb 185

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

race. A flexible washer is located around the base’s center hub, and this is topped by a collar (Figure 7.173).

location, its functioning is less likely to be affected by grime (Figure 7.174).

Cover Cover

Cup spring

Cup spring

Ball race

Tangential spring Ball

Needle cage

Basic element

Ball

Tangential spring Basic element

Figure 7.174 Valve rotation during closing stroke.

Rotator function here is the reverse of the situation where the valve is rotated during the opening stroke.

Figure 7.173 Valve rotation during opening stroke.

When the valve opens, the stroke is transferred to the collar (because of rising valve spring force), and the collar then flattens the flexible washer. This washer forces the balls in the slots to roll downward along the inclined races; the washer itself rolls downward on those balls. The contact with the balls causes the pressure exerted by the flexible washer on the hub to be reduced, causing slippage at this point. The collar and the flexible washer, however, are joined one with the other by friction, thus preventing rotation. When the rotator is located below, the relative rotation between the base and the unit formed by the collar and the flexible washer is transferred to the valve via the collar, valve spring, flexible washer, and keeper. When the valve closes, the flexible washer is relieved and the coil springs move the balls, which do not roll in this phase, back into their initial position at the top of the inclined races. When a valve with rotator opens, it is rotated by the function of the rotator mechanism but also by the torsion of the compressed valve spring. While the unloaded valve spring returns to its original position, except for a small residual angle, when the valve closes, the rotation of the rotator mechanism is retained. The effective angle of rotation per stroke is thus the sum of both values, assuming that the direction of rotation is the same. •• Rotation during the valve closing stroke: If at all possible, the rotator should be located at the top, because in this

Either type may, in principle, be employed in the versions situated above or below. In high-speed engines, preference is given to location at the bottom to avoid increasing the masses in the valve train. In the version at the top, the rotator replaces the spring collar. It is used in slow-running engines as well as when the version at the bottom cannot be used because of space limitations. What is important here is continuous valve rotation in dependency on engine speed.

7.12 Valve Springs The purpose of the valve spring is to close the valve in a controlled fashion. This requires maintaining constant contact among valve train components during valve movement. In the “valve closed” state, the spring force F1 must be great enough to keep the valve from “bouncing” on the valve seat immediately after closing. In the “valve open” state, it is necessary to prevent “fly-over,” that is, the valve stem lifting off and breaking contact with the cam at maximum deceleration. The kinematics are such that the required spring force F2 is the product of the valve’s mass and the maximum valve deceleration amax [7-107]. When engineering the valve springs, additional, and sometimes conflicting, objectives have to be achieved: •• Reducing spring forces: Among other factors, fuel consumption can be influenced by the engine’s internal friction. The friction losses occurring in the valve train are proportional to the required spring forces. The maximum required spring forces are determined by the inertia of the moving valve train components, from the cam lobe to the valve. Consequently,

186 | Internal Combustion Engine Handbook

6606_Book.indb 186

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.12 Valve Springs

the mass of the spring, the cam lobe contour, and maximum camshaft rotation speed are influencing factors. A reduction in spring mass can be influenced by increasing vibration resistance and optimizing the shape of the valve spring.

applied as is shown in Figure 7.175(a), two shearing strains are induced in the longitudinal and transverse sections. According to Mohr’s circle, these shearing strains can be assigned to two primary direct stresses σ 1 and σ 2 at less than 45°. Whereas pure shear stress load is induced in the torsion bar, the situation in a coiled spring is different. Because of the spring’s geometry and the potential deviation of the effective force axis from the spring’s centerline, the bending moment Mb, the lateral force Q, and the standard force N can generate additional stresses under load. Moreover, because of the curvature of the wire, the strains along the circumference are not uniform. Maximum load tensions thus occur on the inside of springs made of round wire. The equations used to calculate helical compression springs are given in DIN 2089. The following situations apply to the spring rate R, the force F, and the torsional strain τ :

•• Reducing height: Reducing the height of the assembly can also have a positive effect on fuel consumption. On the one hand, this provides greater latitude for the design of the hood and improving vehicle aerodynamics. On the other hand, reducing the height of the assembly is another key to reducing engine weight. The design of the valve spring and an increase of its fatigue limit can have a favorable influence on assembly height. •• Ensuring minimum failure rates: The increased demands on the valve springs unavoidably lead to an increase in operational strength. In the course of an engine’s service life, at approximately 200,000 km, the spring has to withstand up to 300 million loading cycles. At the same time, only a minuscule spring failure rate is acceptable. The use of multi-valve technology makes it necessary to further reduce the failure rate for individual springs. If one assumes, for example, a failure rate of just 1 ppm for the valve springs and if one is building a twenty-four-valve engine, then the result is that, at maximum, only one engine in 40,000 will fail as a result of valve spring failure. Ensuring low failure rates imposes stiff demands on valve spring design, materials, and production.



F = s⋅R



t=



k=



(7.11) (7.12)

w + 0.5 w − 0.75

(7.13)

The strains on the inside of the spring when under load are thus as follows [7-108]: t = k⋅



7.12.1 Determining Strain under Load

8 Dm ⋅ ⋅ F . p d3

(7.14)

The shear stresses determined analytically do not take into account the additional load strains previously mentioned,

Fundamentally, the loading on a coiled compression spring is that of a rod subjected to torsion. When torsional moment Mt is Forces and moments on a spring winding

8 Dm ⋅ ⋅F p d3

(7.10)

The approximation formula developed by Bergsträsser is among the techniques used to correct the strain values resulting from the curvature of the wire:

•• Economy of product improvement: The demands presented here have to be economically justifiable, that is, the benefit associated with any given measure has to be greater than any additional costs that might be incurred. This challenge has been taken up by valve spring manufacturers in the face of increasingly tougher competition.

Forces acting on the valve spring F

G d4 ⋅ 8 Dm3 n

R=

Tensions in the wire cross-section

y

FQ

f

Mb

x Mf

FN

Dm y t(f) t(x)

F

d

x

Figure 7.175a Forces, moments, and strains on valve springs.

Internal Combustion Engine Handbook | 187

6606_Book.indb 187

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

which result from the bending moment and the transverse and standard forces. In addition, the spring’s natural vibrations at high engine speeds cause dynamic overshooting, to values that can exceed by as much as 50% the load strains determined in static testing. These dynamic effects can be ascertained either with multibody simulation programs or by metrologically using strain gages. The experiments are usually performed on specially prepared engine mockups [7-109]. The resulting tracing shows the load strain plotted against engine speed and crankshaft angle. Depending on the loading and the limitations imposed by available installation space, the shapes shown in Figure 7.175(b) have been developed. The standard shape is the symmetrical, cylindrical spring. In this spring, the distances between the turns are symmetrical at both ends of the spring, and the diameter of the turns is constant. Progression in the spring characteristics is achieved by the partial contact of the turns across the spring deflection path. Depending on the progression engineered into the spring, the spring rate and the spring’s natural frequency may change across the spring deflection path. The dynamic excitation of the spring thus becomes broader in spectrum, and dynamic overshooting is reduced. The spring may be wound asymmetrically to keep the spring masses in motion as small as possible. This means that the closely spaced turns required for progression are located toward the cylinder head. The disadvantage of the asymmetrically wound spring is that additional measures have to be implemented to ensure that the spring is properly oriented for mounting in the cylinder head. The conical valve spring offers the advantage that, on the one hand, the moving masses are smaller than for a cylindrical spring and, on the other hand, the fully compressed height is slightly shorter. Furthermore, a conical spring permits

the use of a smaller spring collar at the valve, which in turn has a positive influence on the masses. A disadvantage is that a conical spring often exhibits less progression than a cylindrical spring. The so-called “beehive spring” comprises a cylindrical section fixed in place and a conical section in contact with the spring collar. This shape combines the advantages of the cylindrical and the conical spring. In this way, the masses in motion can be reduced significantly by employing a spring collar smaller than the one used with a cylindrical spring. The required degree of progression can be determined in the cylindrical section. Round and multi-arc (“egg-shaped”) wires are the shapes normally used. With the multi-arc wire one has, in addition, the benefits of reduced installation height and more uniform distribution of strains across the wire’s cross section. This is in contrast to round wire, which, as mentioned above, is subjected to the greatest stress on the inside of the spring. Ideal utilization of the material’s properties is achieved with the [7-110] proposed wire cross sections. This cross section provides, on the one hand, the equivalent diameter of a round wire and the axial ratio of the two primary axes. Thus, a 3.8 MA 25 designates a multi-arc wire whose axial ratio is 1:1.25 and whose polar geometrical moment of inertia corresponds to that of a round wire 3.8 mm in diameter. The low failure rates required here place the maximum demands on the material used to make valve springs. The primary reasons for valve failure are found in nonmetallic inclusions in the spring wire or in mechanical damage to the surface. The Cr–V steels which were often used in the past can no longer satisfy the demands for tensile strength as found in heavily loaded valve springs. They have largely been supplanted in Europe by Cr–Si alloys. Cr–Si steels, in

Designs Cylindrical, symmetrical

Cylindrical, asymmetrical

Conic

Bee hive shaped

Wire profiles Round

Ovoid

Figure 7.175b Valve spring shapes and wire profiles [7-107].

188 | Internal Combustion Engine Handbook

6606_Book.indb 188

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.13 Valve Seat Rings

Fatigue Strength Factors Manufacturing Step

Degree of Purity

Molten steel

Smelting and refining

•

Slab ingot/block

Casting

•

Product Form

Surface

Billet

Hot rolling

•

•

Hot rolling

•

•

Cold drawing

• •

•

Turning

•

•

Oil tempering Valve spring

Intrinsic Stress

•

Patenting Valve spring wire

Structure

•

Rolled rod wire

Peeling

Mech. Values

•

•

•

•

•

•

Stress-relief annealing

•

Planar-grinding the spring ends Shot peening

•

(·)

Heat setting

• •

Figure 7.175c Factors affecting valve spring fatigue strength [7-107].

comparison with Cr–V steels, exhibit fewer nonmetallic inclusions and greater tensile strength. Being used to an even greater extent is HT (high-tensile) wire alloyed with Cr–Si–V or Cr–Si–Ni–V. The wire rod is peeled prior to cold drawing to achieve a wire that is free of surface defects. The required degree of strength is attained by a hardening process; this is usually oil tempering, but inductive hardening may also be used. Following hardening, eddy current sensors are used to check the wire for surface defects. Any faulty areas are marked, and the wire is rejected before it goes to the spring manufacturing process. After the spring has been turned, it is stress-relief annealed to reduce the internal strains in the turns. Finally, the ends of the spring are ground flat to ensure that they are parallel with their mating surfaces. The spring may be chamfered, depending on the specifics of the application. The ball blasting process may be used to compact the surface and to introduce residual compressive force in the areas near the surface. The tensile stresses occurring during operation are superimposed on these intrinsic compression stresses and prevent fissure spread. To further increase fatigue strength, springs which are subject to severe loading, are hardened as well. In this way, the amount of stress that can be handled is increased significantly, by approximately 10%, when compared with conventional springs [Figure 7.175(c)]. Moreover, valve springs are nitrided for some applications and then shot-peened once again. Because of the costs associated with this process, it has not yet been used in either Europe or North America.

7.13 Valve Seat Rings 7.13.1 Introduction

Valve seat inserts and valve guides are important components within the valve train and are essential to perfect ignition

and combustion within the cylinder. With the valve, these components must ensure complete sealing of the combustion chamber so that the required compression and ignition pressures can be generated inside the cylinder. Excessive wear causes changes in the combustion parameters and thus degrades engine performance and emission data. Figure 7.176 shows a drive train with hydraulic valve lifters and an overhead camshaft.

Cam Cup Valve stem seal Valve guide Valve stem Valve Valve seat ring

Figure 7.176 Valve lifter drive with camshaft at the top.

The valve seat and the valve guide are components that are typically produced in large numbers. Figure 7.177 provides an overview of passenger car engines produced in 2006 and 2007 [7-111]. This represents a demand for more than one billion components. Around the world, thirteen companies manufacture valve seat inserts, and they can be subdivided by materials groups into cast materials and powdered metal materials (PM), which account for a 90% share of the market.

Internal Combustion Engine Handbook | 189

6606_Book.indb 189

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Passenger Car Engines Produced Market

2006

2007

Europe

21,394,826

22,890,157

NAFTA

15,877,161

15,425,600

Mercosur

3,043,135

3,517,469

Asia

26,762,274

29,118,108

Rest of the world

1,041,325

979,097

Total

68,118,871

71,930,431

Figure 7.177 Worldwide production of vehicle engines.

7.13.2 Demands Made on Valve Seat Inserts

More than 99% of all aluminum cylinder heads are fitted with separate valve seat inserts because the properties of aluminum and its alloys are not adequate for making up valve seats. The valve seat insert, with the valve itself, forms a tribologic system that has to ensure sealing capacity even after several million operating cycles. Thus, specifications for modern engines mandate maintenance-free operation of the mechanical valve train without compensation for clearance, for mileages of up to 300,000 km (<1 μ m/1000 km). All this takes place in an extremely demanding operating setting. The major factors influencing wear at valve seat inserts are discussed below. 7.13.2.1 Load on Valve Seat Inserts Varying loads are encountered in the valve seat contact area, depending on the specific engine design. The method used for adding fuel, the compression ratio, the ignition pressure and the associated specific forces, as well as the temperatures prevailing in the contact area, all vitally influence wear and deformation in the tribologic system comprising the valve and valve seat insert. The wear factors that thus arise are summarized below. (a) Mechanical loading at the valve seat area: This load comprises the spring preload (Ff), the valve’s closing force (FB) and the pressure exerted by combustion (FP). Figure 7.178 provides a survey of the percentages for the various types of loading imposed on the valve seat in an overhead camshaft engine. This load is subdivided into forces exerted perpendicular and parallel to the seating surface; the split varies with the valve seat angle. The parallel forces are the primary factor in wear and deformation at the valve seat. The sizes of the forces and the distribution of the loads they generate depend on the engine design and the current operating status (e.g., electromagnetic valve operation and engine braking). Share in Overall Load (%) Spring preload

1–3

Closing force (maximum 1500–7900 m/s2)

2–17

Combustion pressure

80–97

(b) Dynamic loads exerted on the valve seat because of valve motion relative to the valve seat insert: One portion of the motion is the rotation of the valve. This depends on engine speed and, in valves actuated conventionally, may be as much as 10 rpm or, when using the so-called Rotocaps, up to 45 rpm. This motion is desirable because, on the one hand, it ensures uniform valve temperature and, on the other hand, it has a cleaning effect on the valve seat. A further dynamic load on the seat results from valve disk deflection, which occurs automatically when the pressure in the combustion chamber impinges upon the valve head. This effect is reinforced by a differential between the contact angles at the valve and valve seat, between 0.5° and 1°, which is referred to as the differential angle (Figure 7.179). In this way, a narrower seat diameter and thus higher pressure at the sealing surface, with enhanced sealing effect, is achieved where ignition pressures are low. When the pressure is increased, the bearing portion of the contact surface increases because of bending at the valve disk, resulting in reduced surface pressure at the valve seat.

Figure 7.179 Differential angle at the valve.

(c) Lubricating the seat contact area: The wear rates at the tribologic system formed by the valve and valve seat insert are greatly influenced by intermediate lubricating layers. The effects at the intake and exhaust sides differ, depending on the composition of the fuel mixture. Figure 7.180 compares the influence of the types of fuels on wear between the valve and the valve seat insert. These effects are essentially superimposed by further phenomena. In particular, the potential enrichment of the mix resulting from introducing crankcase vapors at the intake must be mentioned. Additionally, oily components can pass through the valve stem seal and along the valve stem to the seat contact area. (d) The partner in wear—the valve: When designing the valve train, it is important to ensure that the valve contact surface is harder than the mating surface at the valve seat insert. This is necessary to achieve proper distribution of wear—one-third at the valve and two-thirds at the valve seat insert. This wear ratio is necessary because, in the opposite case, the valve disk could gradually be weakened. Consequently, the valve could slip into the valve seat, causing engine damage. Typical hardness values are summarized in Figure 7.181.

Figure 7.178 Load distribution at the valve seat [7-112].

190 | Internal Combustion Engine Handbook

6606_Book.indb 190

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.13 Valve Seat Rings

Inlet Gasoline Wear rate 1–5 μ m/1000 km

++

–

Exhaust

Liquid lubrication in aspirated and turbocharged engines. No lubrication in diesels using SI cycle because only the intake air passes through the intake port.

+

Solid lubrication with deposits from the combustion gases

Diesel fuel Wear rate 1–5 μ m/1000 km

–

No lubrication by fuel because only the intake air passes through the intake port.

++

Solid lubrication with deposits from the combustion gases

Alcohol Wear rate 1–10 μ m/1000 km

O

Liquid lubrication in aspirated and turbocharged engines but with corrosive components. Effect will vary with alcohol content (critical from E50)

O

Little solid lubrication, increased water content. Effect will vary with alcohol content (critical from E50) Problematic in thermally highly loaded turbo engines.

CNG Wear rate 1–50 μ m/1000 km

–

No lubrication because only the gas mixture passes through the intake port.

—

Little solid lubrication because of minimal combustion residue. Problematic in thermally highly loaded turbo engines.

LPG Wear rate 1–70 μ m/1000 km

—

No lubrication because only the gas mixture passes through the intake port.

—

Little solid lubrication because of minimal combustion residue.

Hydrogen Wear rate 3–70 μ m/1000 km

—

No lubrication because only the gas mixture passes through the intake port.

—

No lubrication because of lack of combustion residue, increased corrosion caused by water vapor.

—

Evaluation: ++ Very good; + Good; o Average; – Poor; — Very poor Figure 7.180 Influence of fuel type on wearing action at the valve and valve seat insert.

Valve

Valve Seat Ring

Inlet

270–370 HBW2.5/187.5 Hardened >48 HRC

220–320 HBW2.5/187.5

Exhaust (hard-faced)

30–50 HRC

30–46 HRC

Figure 7.181 Comparison of hardness for the valve and the valve seat insert.

7.13.2.2 Materials and Properties 7.13.2.2.1 Materials Casting alloys: Various production methods, including die or sand casting and centrifugal casting, are employed to form components from these alloys. The manufactured materials are as follows: •• Cast iron [7-113]: Low-alloyed gray-casting material is used at both the intake and the exhaust ports for engines developing low internal loading. The high share of free graphite in the material ensures good emergency (dry) running properties. The material’s properties can be further improved by heat treatment, for example, to enhance ductility, which is necessary when using titanium valves. Austenitic cast iron is used to harmonize with the coefficients of thermal expansion found in aluminum cylinder heads. Increasing the amount of carbide increases wear resistance in this material. •• Martensitic steel casting [7-113]: These materials are based on tool steels and rust-free martensitic steels. They are generally employed as hardened qualities for intake and exhaust valve seat inserts in utility vehicle engines involving moderate and high loading, at temperatures of up to approximately 600°C. Good corrosion resistance is achieved by adding chrome.

•• Nonferrous alloys [7-113]: This group of materials comprises high-alloy nickel- or cobalt-based alloys. Such alloys are used particularly at the exhaust side in engines where high loading occurs. Characteristic of this group of materials is the high percentage of carbides and the inter-metallic phases. Excellent high-temperature characteristics, capable of handling up to 875°C, are attained. Disadvantageous are the high costs for the materials, their low thermal conductivity, and the difficulties in machining. In high-performance engines (racing and Formula 1) copper-based alloys with added beryllium are used because of their great thermal conductivity. Powdered metal materials: Here a powder mixture is compacted at pressure of up to 900 MPa inside a mold that is close to the final contours. The resulting blanks, the so-called green bodies (powder preforms), are sintered at high temperatures (1000 to 1200°C for ferrous alloys) and then subjected to heat treatment. Mechanical machining—turning and polishing—concludes the production process. Additional manufacturing steps may, however, be required, depending on the type of material used. The goal of modern powdered-metal development is to keep down the number of manufacturing steps in the interest of achieving major cost savings [7-114]. Powdered-metal materials are subdivided into several groups: •• Low-alloyed steels: Low-alloyed steels are used primarily for intake valve seat inserts in gasoline engines. These materials are based on a Fe–Cu–C system. The structure is ferritic/pearlitic in nature, with a share of cementite. Small amounts of nickel or molybdenum are used to improve wear properties. Solid lubricants (such as MnS, Pb, MoS2, CaF2, or graphite) are often used to improve amenability

Internal Combustion Engine Handbook | 191

6606_Book.indb 191

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• Medium-alloyed steels: These materials are generally used as the valve seat inserts for gasoline engines and at both the intake and the exhaust ports in diesel engines. This group of materials is the one most widely used and provides a broad range of variants, of which the three most common groups are worthy of mention. In the martensitic steels, the microstructure is essentially a martensitic tempered structure with finely divided carbides, solid lubricants, and, if appropriate, hard metal phases (intermetallic phases of great hardness and temperature resistance such as Co–Mo–Cr–Si Laves phases and Co– Cr–W–C phases [7-115]). High-speed steels derive their superior wear strength from a martensitic matrix with a fine distribution of specially formulated M6C or MC type carbides, which can be made by alloying elements such as Cr, W, V, Mo, and/or Si. Taking the standard high-speed steel alloys (such as M2, M4, and M35) as the basis and using alloying technology modifications, such as diluting with iron powder, adding solid lubricants, or adding other hardphase elements, finally culminate in the valve seat material. In contrast to the other two materials groups, bainitic steels do not have any tempered structure but instead a thermally more stable bainitic basic structure. The addition of solid lubricants, carbide-forming agents, and hard phases in combination with the fundamental structure produces good wearing properties when hot. Typical alloying elements include Co, Ni, and Mo. The medium-alloyed steel groups can also be purchased as copper-infiltrated qualities. Here, the open volume between the pores in the sintered body is filled with liquid copper during the sintering process. The advantage of this alloy, in addition to improved heat conductivity, is found in better machinability. •• High-alloyed steels: This group includes martensitic and austenitic materials. They are used in engines with higher demands for resistance to high-temperature oxidation and corrosion. Typical alloying elements include Ni, Cr, and Co. Because of the high alloying element content, these materials are very costly when compared with the other materials groups. It is for this reason that dual-layer technology is often used in which the valve seat insert is made of two layers of different materials—a high-alloy material at the valve seat and a low-alloy material facing the port [7-116]. •• Nonferrous alloys: The basic Ni and Co alloys, in contrast to the casting alloys, are seldom encountered in powderedmetal technology. Copper-based materials are of particular interest for racing applications. One objective of modern materials development efforts is to identify substitutes for toxic beryllium as an alloying element. Adding ceramic particles (such as A12O3) has already made it possible to achieve wear values comparable to those in the standard applications [7-119].

7.13.2.2.2 Properties Valve seat insert materials must exhibit certain properties to satisfy the material technology requirements. The key properties are enumerated below: •• Hot hardness: A material’s hardness generally corresponds to its wear resistance. For this reason, hot hardness is used as an indicator of a material’s wear resistance at elevated temperatures. Severe drops in hardness at rising temperatures may point to potential temperature limits for a given material (Figure 7.182). Hot hardness comparison of cobalt, nickel and ferrous alloys

Hardness HRA

to machining by cutting. Overall, the amount of alloying material is less than 5%.

90 80

Co based

70

50

Ni based

FE based

60 0

200

400 600 Temperature in °C

800

Figure 7.182 Hot-hardness comparison [7-117].

•• Structural stability at elevated temperatures: Structural stability at elevated temperatures identifies changes in the material because of the influence of heat. Figure 7.183 summarizes various effects. One must assume that there are diffusion-related changes, in particular, for materials with tempered structures when they are subjected to thermal stress. Temperature

Process

Effect

–190–21°C

Conversion of residual austenite into martensite

Increase in hardness Dimensional changes

250–900°C

Reduction of intrinsic stresses Diffusion processes Precipitation processes

Hardness changes Properties changes Structural changes

Figure 7.183 Effects because of thermal loading.

•• Coefficient of thermal expansion: The coefficients of thermal expansion for valve seat inserts and cylinder head materials are of considerable significance when mounting the inserts in the cylinder head with a press fit. It is beneficial if the materials used for both items exhibit similar coefficients of thermal expansion. If this is not the case, then a reduction in the holding force may occur when the system heats up (which is the case when combining ferrous valve seat inserts and aluminum cylinder heads). This can cause the valve seat to be dislodged from the cylinder head bore and result in damage to, or destruction of, the engine. Figure 7.184 shows typical values for coefficients of thermal expansion.

192 | Internal Combustion Engine Handbook

6606_Book.indb 192

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.13 Valve Seat Rings

also keeps the notching effect of the pores from initiating fatigue, culminating in material breaking away.

Thermal Expansion (10–6 K) Cylinder head Valve seat ring

In contrast to cast valve seat inserts, one must expect a certain volume of pores in powdered-metal products.

Cast iron

 9–11

Aluminum

23–27

Ferrous (martensitic)

 9–13

Ferrous (austenitic)

17–19

Ni based

12–16

FE based

 17–35

Co based

12–14

Fe-based (Cu-infiltrated)

 40–49

Ni based

 16–18

Co based

 14–15

Cu-based

100–200

Thermal Conductivity (W/mk)

Figure 7.184 Coefficients of thermal expansion.

•• Thermal conductivity: To keep the valve temperature within reasonable limits, it is necessary to ensure good transfer of heat from the valve disk, via the valve seat insert, to the cylinder head. This is achieved, in addition to engineering good heat transmission interfaces, by selecting materials with high thermal conductivity. Figure 7.185 depicts the theoretical heat flows at the valve.

Figure 7.186 Thermal conductivity.

•• Resistance to oxidation and corrosion: Because of the extreme operating situation, valve seat inserts must be able to withstand corrosion and oxidation resulting from exposure to the hot exhaust gases. This can be achieved either with the chemical composition of the material or by a carefully defined passivation of the component’s surfaces by preoxidation, for instance.

Theoretical calculations [7-118] have revealed that an increase in conductivity from 20 to 40 W/mK reduces the operating temperature at the valve seat insert by 50°C and that at the valve by 30°C. Measurements in various engines have confirmed this reduction in valve head temperature [7-119]. One method commonly used to achieve these values is to infiltrate the medium-alloyed materials used on the exhaust side with copper. Figure 7.186 summarizes some representative values. When engineering the cylinder head, one must take into account the fact that the increased injection of heat into the aluminum comprising the cylinder head around high-conductivity valve seat inserts causes a loss of strength in the aluminum. Fissuring in the web area is the result of this type of thermal overloading.

•• Wear resistance: The following wear-inducing mechanisms are effective here: Adhesion: Local micro-welds with subsequent failure at the contact points. Material is transferred from one interface surface to the other and pitting will take place. Abrasion: Material removal because of grinding and cutting mechanisms in the microscopic range. Material transfer is found to only a limited extent. Oxidation: Forming brittle, loose oxide layers, which will spall off under load. This is called tribo-oxidation. Corrosion: The formation of reaction phases, such as the nickel-sulfur eutectic with its low melting temperature, can cause high nickel content material to weaken and break away.

•• Density: To keep the stresses on materials as low as possible, materials with higher density are favored because of their higher specific contact area at any given loading level. This

Sliding contact • Q6

• Q6

Stem guide

• Q5

Valve seat

• Q4

• Q2

• Q5

• Q4

• Q3 • Q1

• Q1

• Q2

• Q3

Figure 7.185 Thermal flow at the valve [7-119].

Internal Combustion Engine Handbook | 193

6606_Book.indb 193

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• Machining properties: Good machining properties are an important criterion when evaluating materials for valve seat inserts because the final machining of the valve seat has to be effected after the insert has been mounted, which is because of the close tolerances for the cylinder and the valve seat insert. The nature of the micro-structure, the highest possible density and the addition of solid lubricants can have a positive effect on tool lives. 7.13.2.3 Geometry and Tolerances Valve seat inserts, in general, exhibit a simple ring shape. Special shapes with contoured exterior surfaces are used for components that are cast in place during cylinder head manufacture. These contours are intended to create a positive connection, to keep the valve seat inserts from being dislocated [7-120]. Figure 7.187 shows a typical valve seat insert contour Figure 7.188 summarizes common tolerance values.

Valve seat angle

Subordinate angle

Two engineering options available to reduce wear at the valve seat are found in reducing the valve seat angle and increasing the width of the valve seat. Reducing the valve seat angle or widening the valve seat reduces the loads that are effective parallel to the seating surface, as is depicted in Figure 7.189. Investigations have revealed that reducing longitudinal surface loading results in a reduction of the wear rate. Common values for the valve seat angle and valve seat widths are given in Figure 7.190. •• Installation chamfer: The chamfer positions the valve seat insert and lowers the forces required for pressing it in place, prior to and while mounting in the cylinder head. Turned chamfers are normally a simple sloped area with an angle of from 10° to 45°. When valve seat inserts are formed in a powder metallic production process, the chamfers that are imparted are often rounded, with radii of from 0.4 to 1.4 mm, and with an area sloped by 10°–15° on the outside surface. It may be assumed, in principle, that smaller angles for the sloped areas result in lower assembly forces.

140

Installation chamfer

Inside diameter(ID) Outer diameter

Figure 7.187 Typical valve seat insert contour.

•• Valve seat: The valve seat in the insert is the actual functional area for this component. As a rule, final finishing by milling is carried out only after the component has been mounted in the cylinder head so as to achieve exact congruence of the valve axis and the valve seat insert axis (centerline offset of 0.02 to 0.03 mm in new engines).

Outer diameter

Inside diameter

Height

Installation chamfer

45°

120 100

30°

80 60

20°

40 20 0

0

± 0.013 mm

Orthogonality

0,03 relative to

Surface

Cylinder dimension

Run-out dimension of sloped areas Coaxiality Angle

Surface

Dimension

Parallelism

1 1,5 Seat width in mm

2

2,5

± 0,010 mm bevel side Ra = 1,25 ± 0,1

± 0,15

Ra = 3,2 0,2

± 1°

Ra = 3,2 ± 0,05 0,04

Surface front faces

Ra = 1,6

Tolerance angle incline

± 2°

Tolerance radius

0,5

Figure 7.189 Comparison of surface loading depending on valve seat angle and width.

Da < 45 mm

Da > 45 mm

Surface Seat

Surface load in %

Height

± 0,15 – ± 0,3

Figure 7.188 Tolerance ranges in valve seat engineering.

194 | Internal Combustion Engine Handbook

6606_Book.indb 194

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.13 Valve Seat Rings

Seat Width (mm) Spark-ignition engine

Inlet

Exhaust

Valve Seat Angle

1.2–1.6

1.4–1.8

45°

Diesel engine  Passenger cars

1.6–2.2

1.6–2.2

45°

 Utility vehicles

2.0–3.0

2.0–3.0

20°–45°

Gas engine

1.8–2.5

1.8–2.5

20°–45°

Figure 7.190 Valve seat widths and angles.

In addition, it is necessary to ensure that no burrs are created at the assembly surface during milling. This is prevented by fine grinding of the components. •• Inside diameter: The inside diameters of valve seat inserts are generally not machined. To optimize gas flow patterns, the inner surfaces of intake valve seating rings in certain families of motors are specially shaped to impart Venturi contours, for example. To improve run-in conditions and to achieve constant valve seat widths following final machining of the valve seat insert (in the cylinder head), subordinate angles are often provided at the valve seat area. The normal value for such angles is 30° (Figure 7.191).

A further orientation value for the design of aluminum cylinder head assemblies is calculated as follows: Insert length differential = 0.3% to 0.4% ´ of the diameter of the bore in the cylinder head. The amount of excess length should always be selected to suit the particulars of the application. Transferring heat to the cylinder head requires good contact with the inside diameter of the cylinder head bore, particularly at the face toward the combustion chamber, because it is here that the greatest amount of heat transfer takes place. Figure 7.193 shows the temperature distribution inside a valve seat insert at the exhaust port. When valve seat inserts are made using powdered metal technology, it is necessary to ensure that the ratio of the outside diameter to the wall thickness is in a range of from 10 to 13. This is necessary to ensure sufficient “green-body” stability in the powder blanks before they are sintered. No such limitation is imposed on cast parts. The roughness of the outside surface has an influence on the forces required to press the valve seat insert into the cylinder head.

Machined seat angle

Subordinate angle

•• Wall thickness: More compact designs for modern engines impose demands for thinner walls at the valve seat inserts. This is limited by the mechanical loading on the valve seat insert and by aspects associated with production reliability. The wall thicknesses normally produced in mass production exceed 1.8 mm The ratio of height to wall thickness should be as is shown in Figure 7.192. •• Outer diameter: To achieve a sufficiently tight press fit in the cylinder head, the insert is usually from 0.05 to 0.13 mm shorter than the bore in the cylinder head [7-117]. Figure 7.191 Subordinate angles.

Figure 7.193 Temperature distribution inside a valve seat insert at the exhaust port.See color section page 1071.

Internal Combustion Engine Handbook | 195

6606_Book.indb 195

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Insert Height H

Height/Wall Thickness

5–6 mm

≤2.5

6–9 mm

≤3.0

>9 mm

£4.0

Figure 7.192 H/W ratio.

7.13.2.4 Cylinder Head Geometry and Assembly The geometry of the cylinder head has a significant influence on the functioning of the valve seat inserts. The temperatures inside the insert can be influenced, particularly with appropriate engineering and assembly procedures. Good contact between the insert’s outside surface and the inside of the bore in the cylinder head is critical. Consequently, perfect roundness and an exact 90° angle between the outside surfaces and the bore are important factors, as is the tendency of the cylinder head to warp. When using valve seat insert materials with enhanced thermal conductivity, it is necessary to remember that this will cause increased thermal loading at the web area in the cylinder head. This can lead, in turn, to fissuring in this area, particularly in higher-performance engines. When installing valve seat inserts in the cylinder head at room temperature, there is a danger that—because of the slight differential between the lengths of the insert and the bore—plastic deformation of the cylinder head material with a displacement of material can occur during assembly. To avoid this effect, an angle of < 10° is recommended for the installation chamfer. Pre-installation chilling with liquid nitrogen offers the advantages of reducing slightly the length differential and lowering the insertion forces. Disadvantageous is the fact that the valve seat insert material is more brittle at low temperatures. In addition, exact process design is absolutely necessary because any delays during assembly immediately change the thermal insertion conditions, where the consequences are increased insertion forces and the risk of imprecise seating.

combinations of materials that permit running dry, that is, without additional lubricating oil. Increased abrasive or adhesive wear, particularly at the ends of the valve guides, will result in poorer performance and emission values for the engine. Adhesive wear can, in fact, cause seizure. As at the valve seat inserts, there are various influencing factors that have to be taken into account when engineering and using valve guides. 7.14.1.1 Loads at the Valve Guides The loads encountered inside the valve guide are reactions to forces that the valve stem introduces into the tribologic system represented by the valve guide and the valve itself; these forces tend to tip the valve. They consist of [7-121] •• friction action at the end of the valve (Fq) •• the lateral forces exerted by the valve spring (Ff) •• the standardized eccentric force on the end of the valve (Fn) •• the forces exerted by gases on the valve disk (Fgas). The moments thus generated are neutralized by opposing forces at both ends of the valve guide. Figure 7.194 illustrates this equalization of forces:

∑F = 0 = F + F q



n

N=4

+ Ff + Fgas + ∑ Fvfn

(7.15)

n=1

When running dry, the loading at the ends of the valve guides causes metal-to-metal contact with the valve stem. Oil inside the valve guide forms a hydrodynamic lubricating film as a result of the valve’s reciprocating motion; pressure is developed at the ends of the valve guide. This lubricating film separates the mating surfaces through to the point that the motion is reversed. Subsequently, there is a brief period of direct contact between the surfaces’ solid bodies, which then reverts again from adhesive sliding to sliding action.

Fn

7.14 Valve Guides Valve guides, just like the valves and valve seat inserts, are essential components in the valve train. Consequently, annual demand is comparable to that for the mating components, totaling more than one billion units per year (see Figure 7.177) In terms of the materials, the market is divided into powdered metal, reformed brass and cast iron qualities.

Fq

Ff Fn

Fvf1

Fvf3

Fq

7.14.1 Requirements for Valve Guides

The function of the valve guide is to stabilize the reciprocating valve in such a way that it is always perfectly positioned at the sealing surface inside the valve seat insert. The tribologic system is formed by the valve stem and the valve guide. Lubrication occurs when motor oil seeps through the gap between the valve stem and the valve guide. In some materials, certain alloying additives and/or components in the micro-structure contribute to lubrication. Because of the increasingly stringent exhaust emission laws, it will become more important to reduce oil seepage rates in the future. Required here are

Fvf2

Fvf4

Ff

Fvf1

Fvf3

Fvf2

Fvf4

Fgas

Fgas

Figure 7.194 Forces at the valve and valve guide.

196 | Internal Combustion Engine Handbook

6606_Book.indb 196

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.14 Valve Guides

In principle, the contact between the valve stem and the guide cycles continuously through the friction situations described in the so-called Stribeck curve, depending on the sliding velocity. The following items influence loading inside the valve guide:

situation on the port side of the valve guide can fluctuate, thus influencing the oil seepage rate. Investigations have revealed that gauge pressure of 0.8 bar on the port side can cause the oil to be forced out of the valve guide, resulting in insufficient lubrication with increased wear and the potential for seizure [7-121].

(a) Valve train: The forces occurring at the ends of the valve guide vary, depending on the type of valve train that is used. Consequently, the lateral forces for rocker arm valve trains are as much as five times greater than those found in valve lifter designs. Figure 7.195 shows the typical cycle of lateral forces in a rocker arm valve train. (b) Valve clearance: Dynamic processes in valve lifting induce additional forces (Figure 7.196). Increasing valve clearance by 0.1 mm increases the lateral force by 22% [7-121].

Specially engineered shapes for the valve stem seals (gas lip seals, for example) can eliminate this problem. (d) Valve guide clearance: The valve guide is responsible for the exact positioning of the valve in the seat at the valve seat insert. To ensure that this task is fulfilled, the valve guide bore and the outside diameter of the valve stem have to be sized to match one another, always striving to achieve the smallest possible amount of play at the valve guide. In addition to improved heat transfer, the hazard of the valve’s tipping is reduced. Moreover, this geometric matching of the mating components supports the establishment of the hydrodynamic lubricating film. The lower limits for the difference between the diameters are determined by the divergent coefficients of thermal expansion for the guide and the valve stem. Figure 7.197 provides some basic valves for valve guide clearances.

(c) Valve stem seal: Creating a hydrodynamic lubricating film in the contact area between the valve stem and the valve guide requires both a sufficient quantity of oil and an adequate valve sliding velocity. This is achieved with valve stem seals that allow defined volumes of oil to pass through the stem sealing area. Normal values lie in a range of from 0.007 to 0.1 ccm/10 h. When using turbochargers or engine braking in utility vehicles, the pressure 200

Rotational speed 2000 1/min

Transversal force [N]

150 100 50 0 –50

80

100

120

140

160

180

200

220

260 280 Crankshaft angle [°]

–100 –150

Rotational speed 1000 1/min

–200

Figure 7.195 Lateral forces at a valve guide, at varying speeds [7-121] (engine driven, valve play 0.1 mm, valve guide play 45 μ m, oil temperature 50°C, rocker arm valve train).

200 150

Play 0,2 mm

Force [N]

100 50 0 –50 –100

80

100 120 140 160 Crankshaft angle [°]

200

220

240

Play 0,1 mm

260

Figure 7.196 Lateral forces at a valve guide, at varying valve plays [7-121] (engine driven, speed 1000 rpm, valve guide play 45 μ m, oil temperature 60°C, rocker arm valve train).

Internal Combustion Engine Handbook | 197

6606_Book.indb 197

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Stem Diameter (mm)

Intake (µ m)

Exhaust (µ m)

 6–7

10–40

25–55

 8–9

20–50

35–65

10–12

40–70

55–85

Figure 7.197 Basic valves for valve guide clearance [7-122].

(e) Valve: As the component mating with the valve guide, the valve itself has a critical influence on wear phenomena by two factors. 1. The heat applied via the valve stem: Theoretical calculations assume that some 10%–25% of all the heat impinging on the valve is dissipated through the valve guide. This effect depends on the thermal conductivity of the valve stem material (12–21 W/mK), while the engineering design of the valve is also of critical importance. Hollow valves filled with liquid sodium serve to lower (by between 80 and 150°C) the temperature at the critical curved area at the back of the valve head. Cooling is achieved by the liquid sodium inside the valve transporting heat from the head to the stem area. The higher thermal loading thus imposed upon the guide makes particular demands on the material and system tuning. 2. The material for the stem: Distinction is made here among the following groups of materials: Ferrous alloys: Valve stems are made up primarily of martensitic or austenitic qualities. Surface roughness is R a < 0.4. The surface finish can be improved by chromium-plating or nitriding. Typical thickness values for chromium-plating are from 3 to 15 μm and from 10 to 30 μm where nitriding is employed [7-122]. Post-treatment of the finished surfaces by polishing is indispensable because residues from the production process (chromium nodules or nitride needles) have to be removed completely to prevent increased wear at the valve guides. The target value for surface roughness is Ra < 0.2.

7.14.2 Materials and Properties 7.14.2.1 Materials Powdered-metal materials: This group of materials, which accounts for a continuously rising share of the market, can be used in all types of passenger cars and utility vehicles. •• Ferrous materials: The micro-structures of these steel qualities, containing small amounts of alloying elements Cu, P, and Sn, are generally ferritic or pearlitic. Copper, when used as an alloying element, assumes a variety of tasks. On the one hand, it improves dimensional stability during the sintering process; moreover, it has a positive influence on thermal conductivity and mechanical properties such as hardness and strength. When tin is also present, there are reactions with the copper, including the formation of a bronze phase with a low melting point. This gives rise to liquid phases even at relatively low sintering temperatures, culminating in greater density in the sintered component. Phosphorous, with iron and carbon, forms the Fe–P–C hard phase known in materials used for casting. Solid lubricants—such as MnS, MoS2, graphite, CaF2, and BN—improve dry running properties in case lubrication is interrupted. Powdered-metal valve guides are relatively porous, and this is reflected in a density value of from 6.2 to 7.1 g/cm3. These pores are often filled with oil to provide basic lubrication between the valve stem and the valve guide when engines are first started. The pores can be charged with oil by immersing the component in a heated oil bath. Capillary action and surface tensions cause the oil to enter the open pores in the sintered part. This process is very sensitive to outside influences including oil condition, component cleanliness, temperatures, oil viscosity, etc. Another far better reproducible process involves impregnation with oil. Here the valve guides are first placed in a vacuum chamber to evacuate the air from the pores. The chamber is then flooded with heated oil that enters the pore under ambient pressure. In this way, one can be sure that almost all the open pores are filled with oil.

Nickel-based alloys: This group of materials is used, in particular, wherever exhaust valves are exposed to high thermal and mechanical loading. In general, this group of materials is known as “nimonic” alloys. When compared with the ferrous alloys, there are no particular factors of interest regarding the tribologic system comprising the valve stem and valve guide.

•• Nonferrous materials. In this context, application is restricted to copper-based materials. In addition to special materials such as dispersion-strengthened copper [7-126], various powdered metal brass qualities have been tested. Market introduction has, however, not been achieved; when compared with current materials, neither cost nor functional advantages have been demonstrated.

Lightweight metal alloys: To reduce the masses in motion in the valve train, current research activities are focusing on the use of titanium and aluminum alloys for valves.

Nonferrous metals: Copper-based wrought alloys (Cu–Zn compounds) are often specified for use in valve guides for vehicular engines. These materials are purchased as drawn tubular or bar material, which is then turned to make the valve guides. The micro-structure comprises two main phases,

Non-metallic materials: The types of ceramic materials now in use exhibit good wear-resistance properties. No special adaptive measures are required when such stems are used in conjunction with conventional valve guide materials. The reason is found in the excellent surface quality of the ceramic valves.

•• The cubic, surface-centered α phase: This is characterized by good cold-reforming capability and is thus characteristic for all wrought brass alloys. The values for hardness and tensile strength are relatively low. This phase dominates where the tin content in the alloy is less than 37.5%. When

198 | Internal Combustion Engine Handbook

6606_Book.indb 198

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.14 Valve Guides

used as valve guide material, the α content should be more than 20%, the risk of seizing increases significantly otherwise.

guide can allow exhaust gas components to enter and be deposited in the sliding contact area.

•• The cubic, surface-centered β phase: The presence of this phase permits increases in both hardness and tensile strength. Toughness is reduced. An increase in the share for this phase is attained by raising the tin content from 38% to approximately 46%.

7.14.2.2 Material Properties To satisfy the requirements in the application technology, it is necessary that valve guides exhibit certain key properties, discussed below. •• Wear resistance: The main loads on valve guides occur at the ends, wherein the end toward the port generally exhibits more severe wear than the end toward the camshaft; refer to Figure 7.198. This is because of higher thermal loading at this end. The wear mechanisms effective here are both abrasion and adhesion. The latter, in borderline cases, can cause seizure between the valve guide and the valve stem and, consequently, result in engine failure. Austenitic stem materials show a greater propensity for adhesive wear. When using chromium-plated or nitrided valve surfaces, wear appears mainly in the valve guide. The increased wear at the port end on the exhaust side poses a problem. The increase in the clearance between the valve and its

Cam side Channel side

Figure 7.198 Wear-prone areas in the valve guide.

In extreme cases, this can cause the valve stem to block inside the guide; engine failure will result. •• Density: Nonporous metals such as the reformed nonferrous metals and casting materials, because of their high specific contact area, have the advantage that the loading on the material is kept down at a given load level. This reduces susceptibility to wear. This also allows avoidance of fatigue phenomena and the associated fissuring and propagation of fissuring, which can appear because of the notch effect induced by the pores. When dealing with powdered metal products, one must always assume a certain amount of porosity. When making powdered metal valve guides, there appears, because of the compaction of the powder at each end, a density gradient; the area of greatest porosity is at the center of the valve guide (Figure 7.199). This type of density distribution and pore distribution is a beneficial property for this category of valve guide since the greatest density is found in the area of greatest loading. The center section of the powdered metal guides can, because of greater porosity, take on a larger amount of oil and thus serve as an oil reservoir. Refer to Figure 7.200 for the density values for the various groups of materials used in valve guides. 7,2 7

Density g/cm2

The heterogeneity of Cu–Zn alloys offers the ability to modify their properties to suit the particular application while increasing the materials’ amenability to cutting operations. Adding aluminum boosts strength without any adverse influence on warm-reforming capabilities. At the same time, the slip properties are improved [7-123]. The material used for valve guides is primarily the CuZn40Al2 alloy. Various additions of further alloying elements such as Mn and Si serve to improve wear resistance. In addition to the superior machinability in comparison with other valve guide materials, high thermal conductivity is a further beneficial property of this material. Cast iron/cast steel: Valve guides made of ferrous casting alloys are widely used, particularly in the utility vehicles sector. The micro-structure comprises a ferritic/pearlitic fundamental structure with free graphite elements (at sizes of approximately 4–7 μ m). These act as an “integral” solid lubricant. The share of ferrite here is generally less than 5%. When phosphorous is present, phosphide compounds, individual and finely distributed structural components and distinct networks may be formed. When the demands on the component are more severe, careful addition of alloying elements (Si, P, Cu, Mo, or Mn) can increase wear resistance. To be mentioned here, in particular, is the ternary Fe–P–C compound, which is often found as the hard phase in casting alloys. Cr is of rather lesser significance as an alloying element and is used in special materials chosen where good corrosion resistance at high temperatures is required. Sand-casting is the manufacturing process of choice. The manufacturers indicate that these materials are compatible with all types of fuels. Maximum operating temperature is 600°C.

6,8 6,6 6,4 6,2 6

Figure 7.199 Density distribution inside a powdered metal valve guide.

Internal Combustion Engine Handbook | 199

6606_Book.indb 199

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

consisting of the valve stem and guide, then one sees that there are combinations of materials that can narrow, or in extreme cases even eliminate, a pre-established valve guide clearance because of outside temperature influences causing the valve to seize. This is always the case when

Density (g/cm3) Nonferrous materials (based on CuZn40Al2)

>8.0

Powdered metal materials (ferrous)

6.2–7.0

Casting materials (ferrous)

>7.1

Figure 7.200 Density values.

•• Thermal conductivity: Thermal conductivity is a critical value for exhaust valve guides. On the one hand, a part of the heat in the valve has to be transferred to the cylinder head through the valve guide. Measurements at test beds have shown that the temperature at the valve head can be lowered by up to 8%, depending on the thermal conductivity of the material used for the guide. On the other hand, the exhaust valve guides are exposed to hot exhaust gases. Consequently, good thermal conductivity reduces the thermal loading on the component itself. The temperature at the end of the valve guide toward the camshaft should not exceed 150°C as otherwise the functioning of the valve stem seal is endangered. Figure 7.201 shows the temperature development at exhaust valve guides. Quite apparent here is the divergence of thermal loading at the two ends of the valve guide; the physical processes associated with dissipating heat to the cylinder head take place in the lower half of the valve guide, at the end toward the valve port (positions A–D). Above this area, the various thermal conductance capacities are of lesser significance. Figure 7.202 summarizes some typical values for thermal conductivity.

Inner diameter in mm

600

PM: Cast: Brass:

Temperature [°C]

500

30 W/mK 43 W/mK 85 W/mK

400

lValve stem ≥ lValve guide



(7.16)

where λ = coefficient of thermal expansion. Thermal Conductivity (W/mk) Nonferrous materials (based on CuZn40Al2)

46–100

Powdered metal materials (ferrous)

21–48

Casting materials (ferrous)

38–45

Figure 7.202 Thermal conductivity.

If this relationship is inverted, the end toward the port will dilate, causing an increase in the clearance inside the valve guide. This opens the possibility for exhaust gas contaminants to enter the valve guide and be deposited on the sliding surfaces. The result is that the valve seizes. Any hard particles entering the gap between the stem and its guide promote abrasive wear. Figure 7.203 summarizes some coefficients of thermal expansion. •• Hardness: The requirements for hardness in the valve guides are relatively low. This can be traced back to the fact that the loading on this valve train component is not extremely high. In addition, the polished and (in some cases) coated surfaces on the valve stems do not provide much opportunity for abrasive attack. Figure 7.204 shows the normal hardness ranges for valve guide materials. •• Oil content: Oil content is a characteristic that is found only in valve guides made by powdered metal sintering. This figure indicates the amount of oil (in percent by weight) held in the component’s pores. The characteristic values are at an order of magnitude of from 0.5% to 1.2% by weight. •• Mechanical machining: Final machining of valve guides is undertaken with the guides mounted in the cylinder head, parallel to machining the seat in the valve seat insert. This ensures that the centerline offset between the valve guide and the valve seat insert is kept within certain limits. Values for a new engine lie in a range of from 0.02 to 0.03 mm [7-122].

300 200 100 A

B C

D

E

Thermal Expansion (10–6K) Valve Guides

Figure 7.201 Temperature distribution in valve guides at differing thermal conductivity values [7-124].

•• Thermal expansion: Like valve seat inserts, valve guides are held in the cylinder head by a press fit. Because of the lower temperature level and the larger mating surfaces, the danger of loosening because of differences in thermal expansion is low. If one observes the tribologic system

Valves

Nonferrous materials (based on CuZn40Al2)

18–22

Powdered metal materials (ferrous)

 9–13

Casting materials (ferrous)

 9–11

Ferrous (martensitic)

 9–13

Ferrous (austenitic)

17–19

Ni based

12–16

Figure 7.203 Coefficients of thermal expansion.

200 | Internal Combustion Engine Handbook

6606_Book.indb 200

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.14 Valve Guides

Hardness HB 2.5

Loss of Hardness in % up to 250°C

Nonferrous materials (based on CuZn40Al2)

150–170

approximately 20%

Powdered metal materials (ferrous)

120–200

0%

Casting materials (ferrous)

190–250

0%

Figure 7.204 Hardness ranges for valve guide materials.

The inside diameter of the valve guides is set by reaming. To do this, broaches with from one to six blades made of TiN-coated hard metal qualities are used. Machining tools made of cubic boronitride or polycrystalline diamonds are used only in exceptional cases. Tool life depends on a variety of influencing factors. Narrow tolerances for the centerline offset between the guide and the valve seat insert have a beneficial effect. Burr-free reaming and a homogeneous micro-structure also extend tool lives. Hard phases or martensitic components in the micro-structure have an adverse effect because of their extreme hardness. Small inside diameters for long valve guides should also be avoided, as this generates high torsional torques in the broaching tool. Common values for the inside diameter, as a function of length, are shown in Figure 7.205. 100

Length in mm

90 80 70 60 50 40 30 20

2

4

6 8 10 Inner diameter in mm

12

14

Figure 7.205 Ratios of valve guide inside diameter to length [7-125].

7.14.3 Geometry of the Valve Guide

Valve guides are typically cylindrical; the ends may assume any of a number of shapes, depending on the exact design. At the side toward the port, simple chamfering may be found, serving as an aid in press-fit installation. There is greater variety at the end toward the camshaft, which depends in each case on the type of valve stem seal used in the particular instance. Over and above this, there are versions with a collar at the outside which forms a stop when pressing the valve guide in place (refer to Figure 7.206 for examples). Figure 7.207 provides some standard tolerance values for valve guides. •• Outer diameter: The valve guide outside diameter has to be matched carefully to the bore in the cylinder head as this is responsible for a perfect press fit in the cylinder head. The standard values for the difference in length between the cylinder head bore and the valve guide are from 0.02 to 0.05 mm for cast iron and from 0.04 to 0.08 mm for aluminum cylinder heads [7-125]. The following ratio should be maintained when manufacturing valve guides in a powdered metal process: Length/Outer diameter ≤ 4 (powdered metal valve guide); 6 (cast-iron valve guide)

(7.17)

•• Wall thickness: The minimum wall thickness for powdered metal valve guides is 1.8 mm (this is affected by the flow properties for the specific powder used and restrictions imposed by the compression technology). If stem seal seats are turned into the guide, then the initial wall thickness should be no less than 2.6 mm, because this is reduced by turning operations. The shorter the valve guide, the thicker the wall should be since—because of the shorter lever arm—the reaction forces involved in guiding the valve stem are increased. This induces greater loading at the ends of the valve guides. Figure 7.208 shows standard values for wall thicknesses as a function of the length of cylindrical valve guides. •• Inside diameter: The inside diameter of the unmounted valve guide is generally not machined. •• Length: Using valve guides with the maximum possible installed length is fundamentally beneficial to keep the tipping angle of the valve as small as possible. The length

Inner diameter Di Outer diameter Stem seat DaS

Length L

Outer diameter Da

Figure 7.206 Valve guide contours.

Internal Combustion Engine Handbook | 201

6606_Book.indb 201

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

of the valve guide should be at least 40% of the length of the valve [7-122]. Outer diameter

Inside diameter

Height Installation chamfer

Da

±0.01 mm

Cylindrical shape

0.01

Surface

Ra = 1.6

Di

±0.1 mm

Surface (before machining the cylinder head)

Ra = not machined

Surface (after machining the cylinder head)

Ra = 2.0

Centering on the outside surface (coaxiality)

0.15

Cylindrical shape

0.1

Dimension

±0.25 mm

Surface front faces

Ra = 6.3

Tolerance radii

±0.15 mm to ±0.3 mm

Tolerance angle incline

±1°

Length in mm

90 80 70 60 50 40 30 2.5 3 3.5 Wall thickness in mm

•• The length of the valve guide outside surface should be as long as the bore in the cylinder head so that the ends of the valve guides, exposed to greater loading, are backed by the cylinder head material.

7.15 Oil Pump

100

2

Valve guides are installed by pressing them into the bores in the cylinder head, generally with both the guide and the cylinder head at ambient temperature. The following engineering notes should be observed:

•• The end toward the port should not protrude into the intake or exhaust port. This would have an adverse influence on gas flow, and the end of the valve guide would be subjected to extreme thermal loads. This could, under certain circumstances, cause increased wear; in the event of improper matching with the valve stem material disturbances in valve train operation and even engine failure could ensue.

Figure 7.207 Tolerance ranges for valve guide engineering.

20 1.5

7.14.4 Installing in the Cylinder Head

4

4.5

Qth in Ltr\min at 1000 min-1

Figure 7.208 Wall thicknesses for cylindrical, powdered metal valve guides as a function of the length [7-125].

For the oil supply of combustion engines, positive-displacement pumps of the gerotor, external gear and rotary vane pump principles are commonly used. Although the power density of modern combustion engines continuously increases and the torques are constantly increasing at low speed—mostly because of turbocharging of diesel and SI engines, a direct connection with the pump size is not generally seen. Instead, the number and design of the crankshaft bearings and the number of consumers influence the pump dimensions. Apart from the “family” concept, that is one pump for each engine series, the oil requirement in the so-called hot idle operation of any engine remains the key value for oil pump design. Figure 7.209 (SI engines) and Figure 7.210 (diesel engines) show a nearly directly proportional relationship between engine power and pump size standardized to a speed of 1000 min–1.

Engine power in Kw

Figure 7.209 Pump sizes corrected by the translation, as function of the engine power in SI engines. See color section page 1072.

202 | Internal Combustion Engine Handbook

6606_Book.indb 202

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Qth in Ltr\min at 1000 min-1

7.15 Oil Pump

Figure 7.210 Pump sizes corrected by the translation, as function of the engine power in diesel engines.

Engine power in Kw

Even when variable-delivery pumps are used, it is worthwhile to throttle the engine’s basic need as much as possible. Avoiding large friction surfaces at the pump’s internals immediately causes a reduced consumption of frictional power in the pump. If the drag of an oil pump is measured according to the consumption cycle defined in EC Directive 80/1268/EEG (Neuer Europäischer Fahrzyklus—NEFZ), one finds that the oil pump’s share in the total drag is up to 2.5%. The decisive issues in power consumption are, in addition to the conceptrelated dimensions of the moving parts within the pump, the speed range and the pressure losses up to cylinder crankcase. The better the overall oil management of the motor, the lower the power consumption of the oil pump.

7.15.1 Overview of Oil Pump Systems

There are many types of oil pumps, but not all systems are suitable for use in internal combustion engines. The primary selection criteria are size, costs, and efficiency in a given situation. Also, the ability to use the system in a wide range of applications is important. The only types of pumps found in mass production are rotor-type pumps and dual-gear pumps—the so-called internal and external gear pumps, but also a variation of the sliding-vane pump, that is, the rotary vane pump. 7.15.1.1 Internal Gear Pump As mentioned above, the internal gear pump is a member of the double-gear pump family. These are double gears because two interlocking elements, the inner and outer gears, together execute the rotary motion. As a rule, the system is driven by the inner rotor. The outer rotor is driven by teeth contact. The outer rotor is always positioned eccentric to the inner rotor by one-half of tooth height so that the teeth mesh when engaged. On the opposite side, a seal is formed between the tips of the teeth, depending on the difference in the tooth count between inner rotor and outer rotor, the tips of the teeth slide off each other or pass a filling part of the so-called crescent. This crescent may also be known as sickle. Two principal forms of internal gear pumps with and without crescent are the result.

7.15.1.1.1 Oil pumps without crescent In oil pumps without crescent, gears are commonly used, in which the tooth count of the inner rotor (ZI) is one tooth less than the tooth count of the outer rotor (ZA): ZI = Z A − 1



They are also known as gear pumps. Typical tooth counts lay between 4/5 and 13/14 teeth. Using the schematic pump toothing shown in Figure 7.211, their mode of action can be described as follows. Observing the processes in the gear set in direction of rotation, the teeth first slowly draw apart because of the axis offset caused by rotation. The resulting increase in volume creates a vacuum that is responsible for drawing in the fluid. The displacement process begins as soon as the two circles describing the tips of the teeth make contact, that is, immediately following the tooth “pocket” with the greatest volume in the area where the tips meet to form a seal. The reduction in volume follows continually—equal to the intake process. The gradient at which the volume is increased at the intake side and reduced at the pressure side can be effectively influenced by the selected tooth geometry. The area changes approximately relative to the angle of rotation of a sine function. The volume change as a temporal

2 4

3 1

Figure 7.211 Schematic pump display. See color section page 1072.

Internal Combustion Engine Handbook | 203

6606_Book.indb 203

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

derivation of the surface function times gear set width accordingly follows the cosine function. This has a crucial influence on the change in pressure ratios in the gear set, cavitation and noise behavior which will be discussed below. Prerequisite for the transport of the fluid from the intake space to the pressure chamber is the separation of both areas from each other. They are separated, on the one hand, by contact at the flanks where the teeth are engaged and, on the other hand, by the two tips of the gears passing one by the other in the tip sealing area. In addition to the axial gap of the gears to the housing, the areas of the sealing and the drive web are also axial, which belong to the housing or its cover and are not shown here. This group of oil pumps is sealed by the teeth themselves, wherein there is a seal only along a narrow line at mating teeth. This low sealing capability makes it clear why these pumps are used mainly in the low-pressure range, particularly because there is no compensation for changes in the gap because of thermal effects. Thus, there is always a small gap that, in turn, causes hydraulic losses. On the other hand, a gap such as this is naturally desirable as it helps to reduce friction losses. For the sake of completeness, it should be mentioned that the use of such pump systems is in principle possible for pressure ranges of 30–40 bar. However, the challenges regarding manufacturing tolerances, oil pressure pulsation and acoustics are so great that other systems are preferred for these applications. One of most common tooth geometries of gear pumps is a profile constructed from continuous circular arcs at the outer rotor originating in the work of Myron F. Hill [7-127] and called “generated rotor” (or gerotor); see Figure 7.212.

primarily in Europe. The more or less unlimited selection of the geometry for the sealing and driving flanks makes possible a more compact design when using Duocentric toothing. It also makes it possible to engineer higher teeth, giving better utilization of the pump’s physical size, which can be used to reduce overall pump size. As a rule, the size advantages when compared with the gerotor concept is up to 8% In addition to the tooth designs mentioned above, a cycloid tooth design has recently appeared on the market. This tooth profile is made up of hypocycloid and epicycloid elements, and it is called the Duocentric IC; see Figure 7.214.

Figure 7.213 Duocentric tooth design.

Figure 7.212 Gerotor—toothing.

Another commonly used toothing is the Duocentric® toothing, made of noncontinuous circular arcs (see Figure 7.213). The gerotor is used primarily in the North American market, where it was developed, while Duocentric toothing is found

Figure 7.214 Duocentric-IC tooth design.

204 | Internal Combustion Engine Handbook

6606_Book.indb 204

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

Calculated

Measured

Pulsation Figure 7.215 Measured and calculated pressure progression in the crescent of a Trochocentric pump. See color section page 1072.

By appropriate selection of the profile correction—weighing between two cycloids—the quiet running and the pulsating behavior of the volume flow can be positively influenced. This pump system is found both on the crankshaft and in auxiliary drive trains, for example, in the oil sump. Pumps driven directly by the crankshaft typically exhibit from 8/9 to 13/14 teeth. The selection of tooth counts is mostly determined by the dimension of the crankshaft journal from which the inner rotor is powered. In pumps running in the oil sump, with diameters that are reduced accordingly and thus with lower speeds at the circumference, common values are between 4/5 and 7/8, while 6/7 is typical here.

Two toothing systems are used, one with involute toothing and a second with a profile formed from trochoid toothing. In the first, one finds tooth counts of 19/24; see Figure 7.216. In the case of the Trochocentric® toothing with a trochoid profile, this ratio is 11/13; see Figure 7.217. These systems are used exclusively in pumps driven directly by the crankshaft. Because of their suitability for higher pressures, they are usually found in transmission oil circuits as well.

7.15.1.1.2 Crescent-type oil pumps As the name says, a crescent serves as the sealing element in this type of pump. It extends over several tooth tips and forms a long sealing space. Consequently, this type can also be employed for higher pump pressures. Because this design requires a tooth count difference between the inner rotor and the outer rotor of at least two, it requires more space than pumps without crescent: ZI = ZA − X;  X ≥ 2 The pumps compensate for this disadvantage with very quiet operation and low pressure pulsation. In addition to the long sealing space of this pump type, there is no change in volume in the area of the sealing space, unlike pumps without crescent. Pressure builds in steps over the crescent as shown in Figure 7.215. The number of steps depends on the number of teeth present in the vicinity of the crescent. This prevents large pressure gradients. Furthermore, by optimizing the crescent geometry, it is possible to mostly avoid sharp pressure transitions.

Figure 7.216 Involute toothing.

Internal Combustion Engine Handbook | 205

6606_Book.indb 205

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

measured at the speeds of modern combustion engines. In principle, a cavitation-free operation to 8000 1/min pump speed can be achieved without issues. By implementing additional measures, this range can still be extended. However, it must be noted that power consumption which increases quadratically with speed will be raised disproportionately—a fact that applies to all other pump types discussed here as well. Furthermore, with increasing speed, the flow velocities in the gear set increase, which can also damage the flanks. In this context, the build-up of trapped oil in the tooth fillet ground must be mentioned. These pressures can be notably reduced by using suitable relief geometries. However, this is a spatial problem that is difficult to master with increasing gear widths. With the possibility of radial filling not available in internal gear pumps, this design principle gives special advantages where higher displacement volumes are involved.

Figure 7.217 Trochoid toothing.

7.15.1.2 External Gear Pump For external gear pumps, two or more externally toothed gear are used. Involute toothed gears have prevailed; see Figure 7.218. For cost reasons, they are manufactured as straight tooth gears from sintered steel. Helical toothing can also be considered. With helical toothing, the volume change during the intake process and the volume displacement during the meshing of the teeth progress more steadily. Because of the axial forces in a helical toothing, only small helix angles of up to approximately 9° are recommended. The teeth which are not engaged form with housing the sealing space, similar to a pump with crescent.

Figure 7.218 External gear pump.

For this reason, this pump type can be used for pressure ranges between 30 and 40 bar, even without gap compensation. There is no speed limit for the usability of this pump type,

7.15.1.3 Vane Pumps Vane pumps are members of the moving-pusher systems. The displacement vanes, also called moving-pushers, are mounted so that they can move in radial slots in the rotor. They are guided by an eccentric ring along a circular curve. The displacement cavity formed by two vanes, the rotor ring and the side plates rotates and so moves oil under pressure from the intake area into the pressure chamber. Here, the volume of the cavity changes continuously. Agostino Ramelli, a 16th century engineer developed the first vane pumps. He is considered to be the father of so-called rotary pumps or crank-driven displacement systems. The most famous displacement system is the ancestor of modern vane pumps shown in Figure 7.219.

Figure 7.219 Vane pump by Agostino Ramelli, about 1590 [7-128].

It is assumed that the separately drawn element A represent a vane that is guided on a concentric rib shown in section “a.” The arrangement of the guide rib ensured that the vane was

206 | Internal Combustion Engine Handbook

6606_Book.indb 206

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

always in contact with the cylindrical outer wall. With just a few minor modifications, this design is still used in modern oil pumps (see Figure 7.220).

Figure 7.220 Vane pump.

Another invention by Ramelli inspired a further vane pump principle, a pump with articulated arranged vanes. In 1875, Prof. Franz Reuleaux discussed the crank-driven displacement systems by Ramelli and their systematic modifications in a theoretical essay about kinematics One of these modifications is a pump type filed for patent by Albert Sylvain Toussaint Vrolix in France on February 26, 1943 (Figure 7.221).

Figure 7.221 Pendulum-slider pump, registered by Albert Sylvain Toussaint Vrolix [7-134].

System

This pump too is still used, in a slightly modified design, in modern combustion engines but usually as a volumetric flow adjusting pump. The large friction diameter of the outer follower has energetic disadvantage and is thus now limited to just a few engine types. 7.15.1.4 Benefits and Drawbacks of the Various Pump Systems The unambiguous evaluation of pump characteristics made in the past can no longer be represented on the basis of modern knowledge. Much more than the individual criteria regarding the space, noise, installation location, efficiency and more, specific customer conditions influence the decision for a specific pump design. Regarding the individual criteria, we can only give some general notes here. Figure 7.222 provides a general summary of the achieved values of pumps used in combustion engines and transmission systems. Installation space: If the available space demands a short, diameter-expanded type, gear and vane pumps are first choice, in particular, when the option of adjustability must be taken into consideration. A smart arrangement of the gears and the fact that external gear pumps are capable of operating without laterally arranged nodules, that is, they draw radially, makes their use in the described space a possibility. Noise: There are no clear assessment criteria for the noise behavior of various pump types. As a matter of course, the working principle and the internal pressure build-up are mostly responsible for the noise behavior. A resonance situation with surrounding components, adverse flow paths or severe oil foaming can be named as causes for noise, as is the pump itself, regardless of the technical principle.  haracteristics values for pump systems C Installation location: Regarding their usability as crankshaft pump, external gear pumps are not a suitable concept. All other types up to vane pump types can be used as pumps on crankshafts. All types can be used in the oil sump, without limitations. Volumetric efficiency: Crescent pumps are most advantageous here. Thanks to the sealing gaps along the crescent, they are suitable for applications in high pressure ranges. Volumetrically “tight” pumps have, by their nature, high pressure gradients. This frequently causes noise problems. For this reason, they must be made “non-tight,” which means that other pump designs such as

Maximum Drive Speed (rpm)

Typical Operating Pressure (bar)

Typical Installed Width (mm)

Permissible Operating Temperature (°C)

Kin. Viscosity (mm2/s)

Crescent pump

650–80001

1–40

8–25

−40 to +160

2…

Gear pump with crescent

600–7500 (crankshaft) 350–75001 (auxiliary drive)

1–25

8–14 (crankshaft) 20–32 (auxiliary drive)

−40 to +160

5…

External Gear Pump

350–9000

1–30

16–60

−40 to +160

2…

Vane pump

400–7000

1–13

15–30

−40 to +160

5…

Depending on flow path and gear set width.

1

Figure 7.222 Typical values for oil pumps for internal combustion engines.

Internal Combustion Engine Handbook | 207

6606_Book.indb 207

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

external gear pumps with comparable efficiency and noise behavior are as effective. Mechanical efficiency: The deciding factor for the mechanical power consumption of every pump is friction processing in the toothing of gear pump, but also between moving parts and oil shearing in the gaps between moving parts and tthe housing. The tooth shape and the sensitivity in respect to tooth deviations and position change heavily influence the friction within the toothing. Crescent pumps and external gear pumps with involute toothing have advantages in this context. Regarding the loss shares arising from the oil shearing between moving parts, gear pumps are more beneficial. The large friction diameters of gear pumps and pendulum-slider pumps adversely affect their mechanical friction power consumption. Ultimately, the concept is selected individually according to the customer’s requirements. Volumetric flow control: All pump concepts, except for crescent pumps, can be represented as a flow-variable pump. In crescent pumps, flow control would be only possible by changing the pump speed. Susceptibility to contamination and wear resistance: Every solution can be realized to have long service life by using appropriate materials and/or protective coating. Oil flow pulsation: Under the ideal conditions of a non-foamed oil and in a cavitation-free speed range, pulsation is not a critical issue in any pump type. Special measures in respect to filling tooth spaces, pressure build-up and oil recirculation produce, even under difficult conditions, acceptable results regarding pulsation in all pump designs. A prerequisite is, however, sufficient space. Costs: The complexity of a pump determines its costs. Vane types, and not just the classic ones, have a deciding disadvantage

in the number of components. Crescent pumps are more expensive than gear and external gear pump, because of the expensive machining of the crescents. The additional shaft in an external gear pump essentially determines the cost delta.

7.15.2 Regulating Principles

Because of discrepancies between effective displacement of the pump and the actual oil requirements of the internal combustion engine (see Figure 7.223), it is necessary to integrate some kind of regulation into the system. The purpose is to limit the maximum system pressure in the engine. Generally, control is differentiated in pressure level regulation, multi-stage regulation or constant pressure modification. Depending on the pressure data used for the control, one also differentiates between direct and indirect regulation. The pressure-regulating devices can be adjusted either at the raw oil end, that is, the oil upstream of the filter, or at the clean oil end downstream of the filter. Figure 7.224 summarizes the systematology.

A pressure level

Direct regulation

Raw oil end adjustment

Discrete pressure stages

Constant pressure adjustment

Indirect regulation

Pure oil end adjustment

Raw oil end adjustment

Figure 7.224 Pressure control systematology.

System pressure [bar] Theoretical delivery Eff. delivery Engine consumption System pressure

Engine speed [1/min]

Figure 7.223 Engine consumption curve, system pressure, and theoretical and effective displacements. See color section page 1073.

208 | Internal Combustion Engine Handbook

6606_Book.indb 208

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

7.15.2.1 Direct Regulation Direct regulation is referred to when the regulating device is located inside the oil pump and the pump pressure itself is the leading variable. A schematic representation of this control type is shown in Figure 7.225. Once this regulation system has gone into action, a virtually constant pressure is achieved downstream of the pump, regardless of speed and temperature. Depending on the number of flow restrictions and consumers before the main gallery, the system pressure is reduced by the time the oil reaches the main bearings. These throttle losses in the filter or, cooler/filter module in particular, are proportional to the throughput volume or the square of the flow velocity. Because the oil throughput in the engine usually heavily increases because of rising temperatures and declining oil viscosity, these throttle losses also increase heavily. Any system that uses the pressure signal downstream of the pump for direct regulation will not be able to automatically compensate these losses. To prevent an inadmissibly large pressure drop in the bearing main gallery, the regulating device is set to maximum pressure loss. The result is that, at low speeds and lower temperatures, pressure tends to be too high, causing unnecessary hydraulic losses. The advantage of this system is the simplicity of the design. The preferred configuration is to divert the excess oil into an internal bypass in the pump. This results in somewhat more favorable conditions in the intake, which in turn helps to reduce cavitation. Much more important, however, is the fact that the internal oil return does not contribute to additional foaming in the sump.

M

Direct regulation Figure 7.225 Diagram of a direct pressure regulating system.

7.15.2.2 Indirect Regulation One refers to indirect regulation when the regulation valve itself is located inside the pump, in this case the lead variable is not pump pressure but a pressure tapped elsewhere in the engine (Figure 7.226). As a general rule, this is the pressure at the main gallery. With this regulation system, one can achieve a nearly constant system pressure in the main gallery, regardless of engine speed and temperature, as soon as the

system begins regulating. At low temperatures, this type of regulation can temporarily cause overly high pump pressure because the throttling losses at high viscosities in the lines and the filter are also high.

M

Indirect regulation Figure 7.226 Diagram of an indirect pressure regulating system.

The behavior of this system is particularly critical during cold starts. In this situation, because of the small amount of oil required by the engine and the lag in the system, excessive pressure may be applied to the oil filter. Because the pressure peak usually acts only for some milliseconds, downstream components are rarely adversely affected. To prevent this, one can install a so-called “panic valve” at the oil pump, limiting maximum pressure at that point. Safety valves such as these are usually set to open at between 10 and 13 bar. 7.15.2.3 Regulation at the Raw Oil and Clean Oil Side A disadvantage in both the regulation systems described above is that they are installed upstream of the oil filter and thus are exposed to contaminated oil. The result may be that one or more particles of grime carried in the oil stream can cause the regulating device to seize up. This can bring about one of the following two basic states: The regulator sticks in the closed position; maximum pressure can no longer be regulated causing damage to the filter or dilating the hydraulic valve lifter. The other situation is that the regulator sticks in the open position, the result being that the oil pump no longer draws any oil because the regulation system diverts the flow to the bypass. Considering the components within the pump, these situations occur extremely rarely. In most cases, the regulating valves are loaded with very high compression force and only massive particles would be strong enough to block them. Much more interesting is the issue of regulating energy, in particular, in respect to response times. As described above, the pressure downstream of the pump is usually higher than the main bearing pressure because of the throttle losses up to

Internal Combustion Engine Handbook | 209

6606_Book.indb 209

1/19/16 8:33 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

System pressure [bar] Theoretical delivery Engine consumption System pressure

Eff. delivery Engine consumption, two-stages System pressure two-stage regulation

Engine speed [1/min]

the main bearing gallery. If particular purity requirements can be ignored, the regulation at the raw oil end basically offers an advantage regarding the regulating force. 7.15.2.4 Two- or Multi-Stage Regulation The standard regulation unit design has a response point at which the valve opens, thus stabilizing the pressure level when this value is reached. The resulting system pressure corresponds to the solid line shown in the chart in Figure 7.227. It has been found that it is not necessary in every situation to have a pressure curve like this in the engine. Rather, it is quite possible to use other pressure curves, such as pressure reduced in dependency on speed, at medium and lower engine speeds; see Figure 7.227 (dashed line).

Figure 7.227 Two-stage regulation. See color section page 1073.

By selecting this curve, it is possible to reduce the hydrostatic output at the oil pump and, nonetheless, maintain minimum oil pressure in the engine. This curve can be realized in constant pumps by a corresponding charge on the regulating valve. In flow rate regulating pumps, it is customary to freely select the switching point and to execute the switch using a 2/2 direction valve that is actuated by the engine control unit. This valve is designed so that it always regulates to the high-pressure level in the event of an electrical failure; see Figure 7.228. 7.15.2.5 Stage-Regulating Pumps Because of the requirements regarding delivery volumes in combustion engines, it is no longer possible—because of space limitation—or even no longer desired to cover the required flow

Temperature MCU Rotational speed

M

Stage regulation

Figure 7.228 Diagram of an electric stage regulating system.

210 | Internal Combustion Engine Handbook

6606_Book.indb 210

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

rate with a single-stage oil pump. A possible solution is twoor multi-stage oil pumps. Here, two or more pumping stages are connected in parallel. This parallel configuration makes it possible to achieve at least twice the delivery volume. The advantage of this parallel system is that, when the regulating valve switches when a specific system pressure is exceeded, the second (or all other stages) are switched off. As soon as the system pressure drops below this set regulating pressure, the stage or stages are reactivated. In this way, only the temporary increases in oil demands are met without generating full pump output at all times. Most of the mechanical losses in the system, however, are permanent. Now, only two-stage oil pumps are used for large-displacement combustion engines. Because there are only two states, that is pumping or not pumping, the pressure curve in the engine always exhibits a pressure jump during activation or deactivation of the stage(s) with partially adverse effects on certain engine elements.

7.15.3 Flow Rate Regulating Pumps

The power consumption of an oil pump comprises a hydraulic and a mechanical component. The mechanical portion corresponds to the energy share to be spent to overcome the pump’s internal friction. This is mostly viscous friction. Mechanical output is approximately proportional to the product of viscosity ν , density ρ , axial gap height x, speed n gear set diameter d with the following relationship: Pmech ~ ν ⋅ ρ ⋅ x–1 ⋅ n2 ⋅ d4 The hydraulic share corresponds to the energy added to the fluid volumetric flow Q during transport to a higher pressure level ∆p: Phydr = Q ⋅ ∆p

Power consumption, oil pump [W]

Depending on the operating state and driving cycle, the total amount consumed can be up to 4% of the engine drag. 2.5% savings in power can be attained in the NEFZ cycle. Using a volumetric flow regulating pump offers the option of saving hydraulic power by adjusting the pump’s delivery volume to the actual engine consumption, as shown in Figure 7.229.

In the simplest case, a regulation to a specific pressure is intended. That means that the volumetric flow remains constant after a specific pressure level is attained, regardless of the increasing engine speed. The advantage of such a regulation is mostly limited to the hydraulic power. Even if the surface (3) between the power consumption of a constant pump (1) and a regulating pump with nominally equal size (2) promises enormous gain in respect to power savings, the advantage relative to the consumption cycle according to the valid EU standard is sobering. Because of the selected gear transmission ratio and the ratio between engine and oil pump, the speed range is limited to only a narrow area in which the discrepancy between the power consumption of both pump concepts is insufficiently utilized. For this reason, an early lowering of pressure us useful which enables a further reduction of the power consumption (4a and b). How much the pressure can be lowered, depends on the engine’s wear limits. In principle, two methods are possible: two-stage regulation (4a) and constant regulation (4b) of the pressure. In a two-stage regulation, the regulating pressure within the pump is controlled by two independent effective areas. One effective area is usually permanently connected to the pressure in the main bearing gallery. The other area can be activated or deactivated with a solenoid valve. Because each pressure level is mechanically regulated to two set values in the pump, this system requires only a simple controller. Feedback of the pressure signal is not required. For easier servicing, the electric valve can be installed outside of the oil chamber. However, because of throttle losses in the control lines at higher viscosities, the valve must be installed as close to the pump as possible. Complete freedom in respect to pressure regulation can be achieved by checking the position of the actuators using a proportional valve; see Figure 7.230. Pressure regulation is challenging because the pressure is no longer set mechanically (via compression forces, for example) but electrically using the proportional valve. In the simplest case, the pressure is measured in the main gallery and controlled by the engine control unit. In principle, the regulating strategies mentioned can be applied to all pump

Power consumption, continuous pump Power consumption, regulating pump Power savings Potential at pressure decrease For the consumption cycle relevant speed range

Engine speed [min -1 ]

Figure 7.229 Potential of power savings. See color section page 1074.

Internal Combustion Engine Handbook | 211

6606_Book.indb 211

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Temperature MCU Rotational Speed

M

Continuous regulation

Figure 7.230 Diagram of an electric stage-less regulating system.

design types. What benefits can be expected for the power consumption will be described below, using the power consumption in the NEFZ cycle. 7.15.3.1 Internal Gear Regulating Pump (Volumetric Flow Variable Gear Pump) The first series volumetric flow variable oil pump in a combustion engine is the internal gear regulating pump (IRP) (Figure 7.231) developed by SHW for the VW four-cylinder FSI engines. The throughput volume changes in the dependence of the pump pressure by rotating the eccentric axle. To enable fast regulation, this rotation is executed by an adjustment ring rolling in the toothing of the housing; see Figure 7.232. The toothing following the adjustment ring changes its working stroke according to the position of the eccentric. The actual angle position of the adjustment ring is determined by the balance of forces at the adjustment ring. The principal determining factors are the pump pressure, the compression force and the internal gear set forces.

Figure 7.231 The world’s first volumetric flow regulating pump in series (VW FSI engines).

Position maximal delivery volume

Position minimum delivery volume Outer rotor rotor Outer

Outer rotor Eccentric axle in 0° position Intake space

Pressure spaces

Eccentric axle in 90° position

Intake Intake space space

Pressure spaces

Inside rotor Adjustment ring

Return spring

Inside rotor Adjustment ring

Return spring

Figure 7.232 Volumetric flow variable gear pump.

212 | Internal Combustion Engine Handbook

6606_Book.indb 212

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

Because of volumetric flow regulation as required, the power consumption can be reduced, depending on the operating point, by up to 30%, compared to a constant pump. The comparison of power consumption in a direct-controlled conventional oil pump and an IRP shows that between 0.5% and 0.8% fuel can be saved in the NEFZ cycle. 7.15.3.2 External Gear Volumetric Flow Regulating Pump In regulated external gear pumps, the delivery volume is regulated by the axial shift of the gears against each other. One of the two gears is shifted against the other gear with a fixed location. The shift is made possible by the pistons arranged to the right and left of the shifting gear. The pistons also assume the shaft bearing for the idle wheel. By adding pressure to this regulating unit, the active gear width is reduced. This continuously changes the tooth width involved in active pumping; see Figure 7.233.

Figure 7.234 Adjustable external gear pump with clean oil regulation.

Position maximal delivery volume

Position reduced delivery volume Figure 7.235 Adjustable external gear pump with constant pressure regulation.

Figure 7.233 Volumetric flow variable external gear pump.

In the case of a direct regulation, the axial shift force is immediately provided by the pump. In indirect regulation, the pressure of the main bearing gallery represents the reference variable. A return spring ensures that the gears are always reset in the direction of maximum pumping when the pressure drops. The fuel savings potential of such a regulating pump can be up to 1% in the NEFZ cycle when regulating to a pressure level. This type of pump is used in series since 2006; see Figure 7.234. The maximum savings of hydraulic power is offered by map regulation. In this case, an electric regulating valve assumes control of the pump delivery, allowing the setting of nearly any pump delivery volume. This pump concept was introduced in series in 2007. Figure 7.235 shows the realized series design.

7.15.3.3 Vane Pumps In the vane pump, shifting the central eccentric ring relative to the outside contour causes a change in the geometric volume of the oil pump. The adjustment may be done by turning around a fulcrum (see Figure 7.236) or by a linear shift of the outer ring relative to the inner rotor. Guide rings arranged on the vanes’ inside ensure that the vanes are permanently touching the cylindrical outer wall. In this manner, the spaces, which are limited by two vanes each depending on the adjustment of the eccentric ring, change their volume over the angle of rotation. If the eccentric ring is arranged concentrically to the inner rotor because of the adjustment, the chamber volume will not change and the pump will not deliver. This enables the infinitely variable adjustment of delivery volumes between maximum and “zero” delivery. The resulting savings from this pump concept are comparable to the potential of the external gear pump.

Internal Combustion Engine Handbook | 213

6606_Book.indb 213

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

It consists of a city cycle (urban conditions) lasting 780 s and an overland cycle (extra-urban conditions) lasting 400 s. This cycle, shown in Figure 7.237, also forms the basis for analyzing the potential in power consumption savings. To be able to compare different regulating concepts, it is necessary to create a uniform option for deducing the pump speed from the vehicle speed defined in the EC Directive. Starting with a mean tire diameter that can be uniformly assumed to be 70 cm for all vehicles, the actual engine speed is determined by the translation in the axle differential and transmission. While the gear transmission is initially an unknown, the correct transmission between engine and oil pump is known to the oil pump manufacturer. The comparison of existing transmissions for vehicles with different engines made by different manufacturers shows that the essential differences in transmission is limited to a few characteristics. The main characteristic is the number of gears. Depending on whether it is a five-gear or six-gear transmission, the different number of gear levels in the extraurban driving cycle must be taken into consideration. Within one transmission family, the translations between the various gears spread comparatively little; see Figure 7.238. Only the axle translation and the first gear exhibit a certain spread in a six-gear transmission in particular. With a higher gear, we see a significant reduction. By defining mean values for the individual gears, “standard transmissions” can be derived so that, during the subsequent analysis of the oil pump’s power consumption, it is only necessary to differentiate between five- and six-gear transmissions. Another important parameter when analyzing power consumption is the temperature progression during the cycle. It determines the oil viscosity and, therefore, materially the oil pump’s power consumption. The EC Directive only defines the temperature at the beginning of the measurement. It should be between 20 and 30°C to take the cold phase of the engine operation into account.

Position maximal delivery volume

Position reduced delivery volume

Figure 7.236 Adjustable vane pump.

7.15.4 Consumption Savings in NEFZ Cycle

Since January 1, 1996, fuel consumption in motor vehicles has been determined in the European Union using a standardized driving cycle specified in the EU Directive 80/1268/EEC in the 93/116/EEC version. It is also known as the New European Driving Cycle (NEFZ). In the English-speaking world, the abbreviations NEDC (New European Driving Cycle) or MVEG (Motor Vehicle Emissions Group) are established. The standardized driving cycle lasts overall 1180 s and defines initially only the vehicle speed and the gear selection matching the speed level in manual transmission systems.

NEFZ- Geschwindigkeitsverlauf

NEFZ speed progression 0 100° 100°

140

90° 90° Außerstädtischer Extra-urban Fahrzyklus driving cycle

100

80° 80° 70° 70° 60° 60°

80 StadtfahrStadtfahrStadtfahrStadtfahrUrban Urban Urban Urban driving 1 driving zyklus cycle 1 zykluscycle 2 zykluscycle 3 zykluscycle 4 2 driving 3 driving 4

60

50° 50° 40° 40° 30° 30°

40

20° 20° 20 0

Temperatur [°C] [°C] Temperature

Speed [km/h] Geschwindigkeit [km/h]

120

10° 10° 0

200

400

600

Zeit [s] [sec] time

Geschwindigkeit [km/h]

800

1000

Temperatur [°C]

0° 1200

Figure 7.237 NEDC cycle according to 80/1268/EEC. See color section page 1074.

214 | Internal Combustion Engine Handbook

6606_Book.indb 214

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

Five-speed transmission 5-Gang-Getriebe

Six-speed transmission 6-Gang-Getriebe

4,11

4,02

Axle ratio Gear I Gear II Gear III Gear IV Gear V Gear VI

Ratio [--]

3,45 3,54

2,38

2,03

1,59

1,34

1,17 1,0

1,0

Figure 7.238 Transmissions in various vehicles by different Original Equipment Manufacturers (OEMs); mean values displayed in circles. See color section page 1074.

0,80

0,80

The assumed temperature curves often base, however, only on the engine’s temperature gradient in dependence on its construction and the running resistance. In fact, however, the temperature curve in the oil pan must be analyzed, because only this variable returns the material oil viscosity value for determining the friction power in the pump. Figure 7.239 shows

the temperature in a vehicle’s oil pan during a highway tour with a constant pump and a regulating pump in comparison. Because of the “non-pumping” of oil because of regulation, one can clearly see a discrepancy between the oil sump temperature (red) and the engine temperature (blue) at comparable speed and pressure profile.

Vehicle measurement, continuous pump, highway driving Pressure downstream filter

Temperature downstream filter

Temperature oil sump

Pressure [bar]

Temperature [°C]

KW Rotational speed [rpm]

Engine speed

time [s]

Vehicle measurement, regulating pump, highway driving Pressure downstream filter

Temperature downstream filter

Temperature oil sump

Pressure [bar]

Temperature [°C]

KW Rotational speed [rpm]

Engine speed

time [s]

Figure 7.239 Progression of oil sump temperature during a highway tour with a regulating pump (top) and a constant pump (bottom) of the same nominal dimension. See color section page 1075.

Internal Combustion Engine Handbook | 215

6606_Book.indb 215

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Basis continuous pump Regulation to pressure level Regulation to two pressure stages Map control Figure 7.240 Achievable fuel savings in NEFZ cycle. See color section page 1075.

This can result in a reduction of power savings. For the sake of completeness, it should be mentioned that the oil sump temperature during urban driving significantly differs from the engine temperature, although not to the extent shown. The reduction in fuel consumption achievable under these conditions is summarized in the graphic below (Figure 7.240).

7.15.5 Engineering Principles

At this point, one should first discuss the theoretical basics so that later the relationships among the dimensions for the gear sets used in sump and crankshaft pumps are clear. In the ideal situation, there is neither space nor cost specifications at the beginning of a project. There are engine oil requirements and pressure curves for various temperatures. These may be calculated or drawn upon measurements made with comparable engines. It is on the basis of these measured and calculated key values that the theoretical delivery volume for the oil pump is figured, taking the volumetric efficiency into account If it has not been determined whether the oil pump is to be located in the sump or at the crankshaft, then a set of gears should be dimensioned for each variation. Theoretical gear set design (gear pump) If the amount of oil required by the engine is known roughly, the pump is designed for the so-called hot idle operation of the engine. This means that the oil pump has to be matched to this situation so that sufficient oil pressure is available during operation. The size of the gears is ascertained first during engineering. One calculates backward, from the known or assumed delivery volume, to the theoretical pumping surface and the required root and outside diameters for the inner rotor. To do this, it is necessary to know the gear width, which is taken from the amount of installation space available. If the installation space has not been defined, one may use values based on experience. If the aforementioned dimensions are known, it is possible to determine the outside diameter, taking appropriate wall thicknesses into account. The formulas given below are employed in these calculations: qth = 1000 ⋅

Qeff hvol ⋅ n



where qth = t heoretical pump delivery volume per revolution (cm3) Qeff = effective delivery volume, taken from the engine consumption curve (dm3/min)

η vol = volumetric efficiency n

= engine speed at Qeff (min–1).

Once this value is known, then the theoretical delivery surface can be determined by the gear width: A=



qth RB

where A = delivery surface of the gear pair (cm 2) RB

= gear width (cm).

Knowing the delivery surface, it is possible to use the following equation:

(

2 A = da1 − d 2f 1



) p4



where da1 = Outside diameter of the inner rotor (cm) df1

= Root diameter of the inner rotor (cm),

while this formula d0 = m ⋅ z where d0 = pitch circle m = modulus z = number of teeth (inner and outer rotor) is used to calculate the dimensions for the toothing. The geometric dimensions for the pairs of gears for the crankshaft and the pump in the sump are now available. Using the gear dimensions as the basis, one can calculate the depths required for the kidney-shaped ports used to fill and empty the pair of gears. In the next step, the intake speeds and the critical circumferential speeds of the pair of gears are determined in conjunction with engine and pump speeds. Ascertaining the gear width in this way permits an estimate of the space required for the oil pumps. Once the designs for both pumps—at the crankshaft and in the sump—have been determined, it is necessary to decide at which location the pump is to be installed. The following

216 | Internal Combustion Engine Handbook

6606_Book.indb 216

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

selection criteria are available, from a technical viewpoint, when making this decision: •• installation space •• drive power and output •• noise and pulsation. The cost factor has to be considered from the commercial point of view. The following sections provide details on the variants available for crankshaft and sump pumps. 7.15.5.1 Crankshaft Pump Crankshaft pumps are used today, above all, by the automobile industries in North America and in Japan. In Europe, they are used primarily by companies that are influenced by parent firms in North America. The structure of the crankshaft pumps is usually as follows:

collar on the inner rotor. The housing also has a bearing bore to accept the inner rotor. The disadvantages are higher costs and somewhat more friction. The benefit of this design is that the inner rotor no longer needs to follow the movements of the crankshaft. The outer gear clearance and the distance between the tips of the teeth can be less than that in the version attached directly to the crankshaft. Here, the offset between the pump and the crankshaft centerline and the crankshaft motion has to be taken up by outer gear play and play between the tips of the teeth so that there are no clashes. To ensure the most precise alignment of the pump on the crankshaft centerline, the crankshaft pumps are normally centered on the engine using centering pins, bushings, or tabs.

•• die-cast housing with shaft seal pressed in place and integral regulation valve •• die cast or steel cover •• inner rotor centered directly on the crankshaft •• outer rotor driven by the inner rotor. Internal gear pumps in the Duocentric, DuoIC, or gerotor designs are normally used. Crankshaft pumps are mostly employed for economic reasons. Because every engine requires a front cover to take the shaft seal, it is logical to integrate the oil pump into this cover. Because of the crankshaft diameter and the required sealing space between the cover and the crankshaft bore in the housing, on the one hand, and the root diameter for the inner rotor, on the other hand, a certain root diameter in the inner rotor results automatically. This is the determinant factor in the geometric dimensions. As already described, the selection of the numbers of teeth for the inside toothing gives a theoretical delivery surface at the appropriate root diameter. Selecting a suitable width for the gear, assuming a suitable volumetric efficiency, gives the dimensions for the crankshaft mounted oil pump. In practice, crankshaft diameters lie between 35 mm (in small three- to four-cylinder engines) and approximately 50 mm (V6 and V8 engines). It is quite conceivable that, at crankshaft diameters greater than 40 mm, the outer gear will have a very large outside diameter (Figure 7.241). In practice, it is often necessary to strike compromises in crankshaft pumps because the overall length of the engine must kept down. This means that there is often an insufficient cross-section available for oil intake. Moreover, for cost reasons, a steel cover is used to enclose the gears. This is normally a stamped component 4 to 5 mm wide. This cover, however, makes it impossible to fill the oil pump very easily from the side at the cover if the gear sets are large and wide. Consequently, intake is possible only on the casing side, and this will result in the filling, cavitation, and noise problems at medium and high speeds. Normally the inner rotors are attached to the crankshaft. It is less often that designs are chosen that exhibit a bearing

Figure 7.241 Crankshaft pump.

Power is normally transferred to the inner rotor via two flats, hex heads, or inside gearing with a variety of tooth geometries. Special designs such as polygonal transfer journals are also found. The pickup tube is attached with a threaded intake flange and gasket. The pressure outlet is directly realized between the oil pump to the engine block, with a seal or O ring. Because the crankshaft pump often serves as the terminating cover, there is a gasket between the engine block and the oil pump. In this case, the shaft seal is mounted in the oil pump case. The oil pan is often flange-mounted to the bottom of the oil pump, which means another area requiring sealing. 7.15.5.2 Sump Pump Sump pumps are primarily used by German car makers. The goal is to achieve power savings at somewhat higher costs. It is also possible to shorten engines in this way. Sump pumps normally consist of the following components: •• die-cast housing with bearing bore for the drive shaft and integrated regulation or cold-start valve •• die-cast cover with bearing bore for the drive shaft •• inner rotor, pressed onto the drive shaft •• outer rotor, driven by the inner rotor. Sump pumps are usually driven directly by the crankshaft, via a chain and sprocket. Multi-stage pumps are also often found in the oil sump. These comprise one or more pumping stages and may also

Internal Combustion Engine Handbook | 217

6606_Book.indb 217

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• oil pump integrated into the engine block and driven via a chain or spur gears.

include one or more scavenger stages. Such an arrangement is not possible on the crankshaft.

7.15.5.3 Key Oil Pump Values Taken from Practice Figure 7.243 shows key oil pump values found in practice. Compared here are four-, six-, and eight-cylinder engines. The four-cylinder, all-aluminum engine with the crankshaft pump has the most oil-sing points; consequently, a large oil pump is used. The four-cylinder engine with a sump pump, the six-cylinder engine and the eight-cylinder engine have been on the market for some time now. 7.15.5.4 Comparison Between Crankshaft and Sump Pumps Every pump system offers advantages and disadvantages and will be selected depending on the specific application. Experience has shown that compromises have to be reached on the following items:

Figure 7.242 Sump pump in 3-D explosion representation.

Because these pumps are located in the oil sump, the long pickup tube with sieve is often eliminated. In the ideal design, this can be integrated directly into the case or the cover, as a die-cast component. Because the oil pumps are in the sump, no additional sealing is required to retain any leaking oil. The oil pumps are mounted on the engine block by centering pins or bushings to ensure that the chain is properly aligned. The oil is transferred between a cast bore in the housing and a machined bore in the engine block. Internal gear pumps in the Duocentric, DuoIC, or gerotor designs are used. External gear pumps are also used occasionally. A subset of the sump pumps comprises oil pumps that are driven by auxiliary drives. The following drive types and locations are common:

•• build size (engine height and length)

•• oil pump driven via a shaft from the distributor shaft

Costs: The sump pump is normally more expensive. This is primarily because of the additional costs for the chain and drive

•• costs •• drive power and delivery •• pulsation and noise. Installation space: If a crankshaft pump is to be used, it is normally limited by the first engine bearing and the path for the timing chain or belt. Normally, attention is paid to attaining the shortest possible engine length. Given this objective, there is little room for a wide set of gears and good filling properties. Unlike crankshaft pumps, sump pumps can usually be very wide in design. In most engine designs, there is enough space to install a sump pump between the engine block, the outline described by the conrod, the crankshaft webs, and the oil pan.

•• oil pump integrated into the engine block and driven via a chain or spur gears

Designation

FourCylinder Crankshaft

FourCylinder Sump

Six-Cylinder Crankshaft

Six-Cylinder Sump

EightCylinder Crankshaft

EightCylinder Sump

Number of teeth

8/9

6/7

9/10

6/7

9/10

6/7

Theoretical delivery volume referenced to engine speed (cm3) (incl. step-down for sump pumps)

20.6

10.3

17.0

12.8

15.72

21.3

Outer diameter (mm)

84.5

58.2

91

58.2

90.0

65.2

Gear width (mm)

14

20

10.8

25

10.7

31.2

Maximum pump speed (min–1)

6800

7000

7000

4500

5800

4450

Output at 6000 min–1* engine speed (W)

2560

830

2330

970

2100

1830

Hydraulic valve clearance compensation

X

X

X

X

Camshaft shifter

X

Key values for engine oil consumers

Turbocharger

X

Timing chain

X

Balance shafts

X

All-aluminum engine

X

X

X X

*Power values at: 80°C oil temperature; 5 bar oil pressure, 5W30 Shell Helix Ultra oil.

Figure 7.243 Typical oil pump values.

218 | Internal Combustion Engine Handbook

6606_Book.indb 218

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

12

Test conditions Medium 5W30 Temperature: 100 °C Viscosity: 12cST Remark: Pumps were driven on an engine consumption curve.

Pump pressure [bar]

9

6

3

0 500

750

1000

1500

2000

2500

3000

3500

4000

4500

5000

5500

6000

6500

Engine speed [min-1] Sump pump

Crankshaft pump

Figure 7.244 Pump pressures for crankshaft and sump pumps.

1500

Drive power [Watt]

1200

900

Versuchsbedingungen: Test conditions Medium Medium 5W30 Temp Temperature: 100 °C Viskosit Viscosity: 12cST Remark: Pumps were driven on an engine consumption curve.

600

300

0 500

750

1000

1500

2000

2500

3000

3500

4000

4500

5000

5500

6000

6500

Engine speed [min-1] Sump pump

Crankshaft pump

pinion. If sump pump engineering is carried out with careful attention to costs, it can come very close to the costs for the crankshaft pump. Technical comparison: A comparison between a crankshaft and a sump pump is to be made below (Figure 7.246). Both pumps are designed for a throughput of 5.2 L/min at 750 min–1 engine speed, 1.5 bar, and 120°C. This means that both pumps can deliver the same flow volume to the engine when the hot engine is idling. Both pumps were run on a component test bed, at conditions relevant to engine operation and at a motor oil consumption curve for oil temperature at 100°C. Similar pressure levels (Figure 7.244) were developed by both versions. Drive power and delivery: The basic assumption for a pump with low drive power requirements is an ideal design of the gear set (ratio of diameters to

Figure 7.245 Drive power of crankshaft and sump pumps.

width). As is seen in Figure 7.245, the drive power required for the crankshaft pump is significantly greater than that for the sump pump. This is essentially for two reasons. The technical data for pumps are as follows:

Designation

Crankshaft Pump

Sump Pump

Tooth count

9/10

6/7

Theoretical delivery volume referenced to engine speed (cm3) (including step-down for sump pumps)

8.1

8.1

Gear width (mm)

8.1

20

Outer diameter (mm)

72

58.2

Root diameter (mm)

46.1

29.4

Shaft diameter (mm)

35

16

Ratio

1

2 (step-down)

Figure 7.246 Technical data.

Internal Combustion Engine Handbook | 219

6606_Book.indb 219

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

First, with the wider design for the sump pump, the outside diameter can be reduced in comparison with the crankshaft pump, the advantage being that friction at the pair of gears is lower. Second, lowering the step-down ratio for the sump pump can further reduce the amount of friction in this type of pump. Stepping down can be damaging when the hot engine is idling. At hot idle conditions, a certain delivery volume is required to maintain oil pressure in the engine. Because of the temperature, however, the oil is quite thin, resulting in relatively high losses through the leakage gap. If there is a high step-down ratio, the percentage leakage losses are high in comparison to the volume of oil delivered to the engine. Here one refers to poor volumetric efficiency at lower speeds and higher temperatures (Figure 7.247). If, however, the step-down ratio and the clearances are selected properly, then the sump pump has a comparable

delivery volume in the lower speed range. Thanks to stepping down, the oil pump is filled well at high engine speeds, and the oil delivery volume does not drop off so severely as is the case with the crankshaft pump (Figure 7.248). The result is better cavitation behavior.

7.15.6 Cavitation and Noise Emission

If a noise complaint results in a measurement of the airborne noise under the oil pan, the source is often found quickly but is not often the true cause. The noise emission of a gear pump depends on the transmission paths which need to be determined in the first instance. Depending on whether it is a transmission of structural noise, a vibration of the oil pan or the cylinder crankcase because of intake or pressure-side oil pressure pulsation, the remedies to be taken can differ widely. Interestingly enough, the emitted airborne noise is the source of the complaint only in very rare cases.

100 90 80

Volumetric efficiency [%]

70 60 50 40

Test conditions Versuchsbedingungen: Medium: 5W30 Medium Temp Temperature 80 °C Viskosität: 18cSt Viscosity Druck: 5bar XX

30 20 10 0 0

1000

2000

3000

4000

5000

6000

Engine speed [min-1]

Crankshaft pump

Sump pump

Figure 7.247 Volumetric efficiencies of crankshaft and sump pumps.

60

Test conditions Versuchsbedingungen: Medium: 5W30 Medium Temperatur: Temperature 80°C Viskosität: 12cSt Viscosity Druck: 5bar Pressure

50

Delivery [l/min]

40

30

20

10

0 0

1000

2000

3000

4000

5000

6000

Engine speed [min-1]

Crankshaft pump

Sump pump

Figure 7.248 Delivery volumes of crankshaft and sump pumps.

220 | Internal Combustion Engine Handbook

6606_Book.indb 220

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

These speed-dependent influencing factors determine noise excitation:

is particularly unpleasant as it turns on and off the noise when the resonance speed range is traveled; see Figure 7.250.

•• mechanical characteristics of the toothing •• cavitation processes

Ground gear set 90°C oil temp.

•• changes in flow speed because of periodic fluctuations in delivery volume •• sudden pressure equalization when cavities at differing pressures meet •• vibrations of pressure control elements •• resonances with oil pan and crankcase structure. Oil viscosity also influences the noise behavior of an oil pump. Because of its damping property at low temperatures, it can positively affect noise emission. However, high viscosity and low volumes consumed by the engine at low temperatures may also cause undesired effects adversely affecting the noise. In the lower speed range, mechanical characteristics of the toothing significantly determine noise. Deviations in toothing because of the manufacturing process, changing tooth spring stiffness, and load-induced disturbances during engagement are visible in the acoustic spectrum with harmonics up to many multiples of the basic frequency. Inaccuracies in machining the teeth on the pinion and the gear with inside teeth appear in addition with a so-called repeat rate, referred to as the “hunting tooth frequency.” This frequency can be determined using the following equation: where fGMF is the tooth pitch and z1 and z2 represent the numbers of teeth at the gears, while na is the number of potential installation positions:

fHT =

fGMF ⋅ na z1 ⋅ z2



The mechanical noises in a toothed system, except the disturbances in engagement for outside teeth as a factor of load, are receding ever further into the background in view of the high quality found in sintered or erosion-machined gear sets. In the involute gears of an external gear pump, selecting angular toothing and proper profile corrections allows a reduction in the noise level. In addition, gears manufactured in a sintering process have another decisive advantage. Because the teeth are manufactured in a regular process such as milling or grinding a toothing, they have stochastically distributed toothing errors tending more toward an unobtrusive noise than a noticeable tonal component. For this reason, sintered gears are often superior to ground units in respect to their noise behavior. Figure 7.249 documents the clear manifestation of the frequency of tooth engagement and its harmonic in a ground toothing compared to the behavior of sintered gears. The strength of the mechanically produced toothing is its low tendency to speed modulation manifesting in the form of out-of-band emissions in the spectrum. A further anomaly that can be prominent in the lower speed range in particular, is the elevation of the excitation caused by resonances with oil pan, cylinder crankcase or other elements of the vehicle structure. This type of noise excitation

Sintered gear set 90°C oil temp.

Figure 7.249 Differences in the structural noise between a sintered (top) and a ground (bottom) toothing. See color section page 1076.

Microphone oil pan

Figure 7.250 Resonance with oil pan near idle speed.

With increasing speed and rising pressure level, the second and third frequency of tooth engagement in addition to the normal frequency becomes visible. This is caused by a further source of noise is—the alternating pressure (pulsation) in the liquid, which is because of the discontinuous oil pumping in the individual cavities between the teeth. When the pumped

Internal Combustion Engine Handbook | 221

6606_Book.indb 221

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

volume in the cavity meets the high pressure side, the volume is first compressed by the greater liquid pressure at the outlet side. The reduction in the size of the cavities during continuing transport ensures that they are emptied. The pulsation in pressure thus arising causes fluctuating speeds for the liquid particles, which is superimposed on smooth flow. In external gear pumps, the noise of oil squeezing in the tooth gaps is often quite noticeable. Corresponding to the number of gaps in both gears, the second of the frequency of tooth engagement is often clearly apparent. The discontinuous opening of the tooth gaps at the intake end also causes heterodyning of the intake flow with alternating velocity of the fluid particles which may cause an excitation of the oil pan floor.

Kavitatonsgebiet Cavity area

Figure 7.251 Speed and frequency analysis of the oil pressure pulsation of a gear pump. See color section page 1076.

With increasing speed, differences in the velocity between the fluid in the housing and the gear set and fast volume changes in the tooth gaps result in a local pressure drop. If the pressure drops below the critical value of 0.4 bar absolute, the gas dissolved in the oil separates as bubbles. This newly formed gas/fluid mixture increases the emitted noise; see Figure 7.252.

If the pressure in the cavity between the teeth again rises because of the rapid meeting of the pumping cavities with the outlet, the vapor bubbles implode suddenly. This process is accompanied by very high, localized pressure peaks that contribute to considerable noise generation and pulsation. This typical event is also called cavitation. The presence of cavitation “seeds” in the form of the air bubbles previously mentioned is important for this process. The start of the increase in noise as a result of cavitation is determined not only by the flow speed, which is proportional to engine speed, but also by the number of cavitation seeds in the liquid. For this reason, one may observe, with increasing oil foaming, an increase in the number of cavitation bubbles at otherwise constant pressure and constant flow velocity. At high air content, cavitation starts even at low speeds. The associated early rise in the noise level then continues gradually. If, on the other hand, the air content in the oil is very low, bubbles start to form only at very high circumferential speeds. In this case, the noise level rises suddenly (Figure 7.251). Subsequent Figure 7.253 shows the result of the order analysis of a gear pump. The rise of the acoustic pressure level toward higher speeds is very apparent. At approximately 4000 rpm, the levels rise steeply, up to the third order. This effect is traceable to the cavitation of the bubbles previously described, which appear more frequently with the increase of the speed. A further phenomenon—which appears in the pump examined here at a speed range of between 2000 and 4000 rpm—is because of torsion resonance in the drive train. A good prediction of the acoustic pressure level, Figure 7.254 and Figure 7.255, is provided by the equation below, taken from Verein Deutscher Ingenieure (VDI—Association of German Engineers) Guideline 3743, applicable to the cavitation-free range. Measurements have confirmed their validity for external gear pumps.



⎛ P⎞ LWA = 78 + 11⋅ l g ⎜ ⎟ ± 3 [ dB ] ⎝ P0 ⎠

One achieves a better approximation for internal gear pumps with a modified form of the previous equation.



⎛ P⎞ LWA = 78 + 17 ⋅ l g ⎜ ⎟ ± 5 [ dB ] ⎝ P0 ⎠

In both cases, the calculated noise levels closely match the measured ones up to the beginning of the cavitation which is identified by the sudden increase in pulsation.

7.15.7 Calculation

Figure 7.252 Bubble formation in a gear pump, recorded at 500 frames/s.

7.15.7.1 Numerical Simulation of Flow CFD Formulation of the motion equations for multi-dimensional flow results in partial differential equations that are dependent on both time and location. The complex geometry of the flow, because of the small amount of space available, renders a unified solution to this equation impossible. It is thus helpful to simplify the spaces used for information so that they can be broken down into volumes that can be calculated more

222 | Internal Combustion Engine Handbook

6606_Book.indb 222

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.15 Oil Pump

Pressure pulsation frequency multi Analyze W orking : Input : spekt r um : FFT Analyzer

[P a]

6 0k

4 0k

2 0k

0 6k 4k 2k 2k

3k [H z]

4k

5k

6k

[RP M ] (N ominal V alues)

25

110

20

100

15

90

10

80

5

70

Figure 7.253 Order analysis as gear pump speed rises.

Sound pressure [dB(A)]

Pressure [bar]

1k

60

0 0

1000

2000

3000

4000

5000

6000

7000

8000

-1

Rotational Speed [min ]

AZ z 9 Sound pressure

Figure 7.254 Surfaceanalyzed noise level of an external gear pump.

VDI 3743

25

110

20

100

15

90

10

80

5

70

0

Sound pressure [dB(A)]

Pressure pulsation [bar]

AZ z9 Pulsation

60

0

1000

2000

3000

4000

5000

6000

7000

Pump speed [min -1]

DuoIC z8 Pulsation

VDI 3743

DuoIC z8 Sound pressure

Figure 7.255 Surface-analyzed noise level of an internal gear pump.

Internal Combustion Engine Handbook | 223

6606_Book.indb 223

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

simply. With the appropriate transfer and peripheral conditions, such complex flow channels can be depicted as the sum of the individual, simple spatial elements. This idea is the basis for modern flow simulation (CFD—computational fluid dynamics). Today, not only the maintenance equations for the mass, pulse energy, and species are calculated, but also the heat transport in laminar and turbulent flows. With this method, the flow path from the intake point onward and within the housing can be checked during the design phase already. Pressure losses promoting cavitation formation can be identified early and, under certain circumstances, eliminated. In addition to a stationary flow analysis, the movement of valves and the changes in cavities in the gear set can be also represented today. With the assistance of so-called moving grid techniques, it is possible to simulate the movement of the flow cavities as well. Calculating the behavior of regulation pistons and optimizing flow in the valve area and reverse flow into the pumping channel can contribute to both increasing the efficiency of the pump and at the same time to delivering important information on the tendency of valves to oscillate. Knowledge of velocity and pressure distribution in the cavities is an important prerequisite not only in respect to cavitation and pulsation but also for issues of wear. While the measurement of the global pressure in the gear set does not present a problem, only CFD analysis can provide information about local pressure and velocity gradients. The advanced technology of deformable grids enables the observation of the variables of the working medium in the working chambers. Figure 7.256 depicts an example of the pressure distribution in the cells of a vane pump. Figure 7.257 displays a snapshot of the calculated velocity distribution in the gear set of an external gear pump.

Figure 7.256 Calculated flow within the pump housing of a vane pump. See color section page 1076.

Corresponding to the gear speed and the oil properties at one temperature, the calculated values represent the velocities in an isothermic state change. If the energy equation is resolved in parallel, one obtains additional information about local heat formation and, therefore, density and viscosity

changes and so also obtains crucial information about the bearing capacity of the lubrication gap.

Figure 7.257 Calculated distribution of flow velocities in the gear set of an external gear pump. See color section page 1076.

The rotation of the gear sets, their intake and discharge behavior influences the flow in the area of the crescent-shaped area in particular. Today, these processes can be also simulated very well. To avoid numerical singularities that may arise from the compression of theoretical volumes to “zero,” the toothing profile of one of the gears is given an equidistant offset. The gap flow created which was originally considered to be a problem has proven to be not such a problem after all. 7.15.7.2 One-Dimensional Simulation of Flow Grids In addition to the classical multi-dimensional simulation of flow, knowledge-based programs used to calculate branched flow grids have been available for some years now. These programs are based on the flow string theory. They are capable of describing static and dynamic flow events in compressible gases and non-compressible fluids. The particular strength of these programs is found in their modular structure, while the individual modules—such as those for oil pumps, valves, bearings, camshaft shifter systems, manifolds, and many more—are knowledge-based. Equations tailored individually to the particular component, usually backed up by test results, reflect the flow properties for these components. Often the database can also be expanded with one’s own experience or test results. Using the modules mentioned above, it is possible to construct and examine very complex multi-strand grids such as the oil management concept for an internal combustion engine. These procedures are often used in an initial projection of oil requirements at various engine operating states.

7.16 Camshaft The internal combustion engine is a machine that works intermittently. A fresh fuel charge flows through an open intake port and into the cylinder where it is compressed and

224 | Internal Combustion Engine Handbook

6606_Book.indb 224

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

ignited; it expands and passes through the open exhaust port into the exhaust system. Cam-actuated valves are normally used in four-cycle engines, less often in two-cycle engines, to open and close the ports. In Wankel and two-cycle engines, the piston itself normally takes care of opening and closing the ports. Other potential embodiments such as rotating or reciprocating sleeves are no longer used in series production.

7.16.1 Camshaft Function

The primary function of the camshaft is to open and close the intake and exhaust valves so that gases can be exchanged; these actions are synchronized with the position of the piston and thus with the crankshaft. Normally the valves are opened by transferring force from the cam to the cam follower, to other actuation elements where required, and ultimately to the valve, opening (or lifting) the valve against the force of the valve spring. During the closing cycle, the pre-tensioned valve spring closes the valve. When the follower is in contact with the cam’s base circle (with the cam exerting no lift), the valve spring keeps the valve closed against any gas pressure in the port (turbocharger pressure or exhaust gas counter-pressure). During engineering, it is important to pay attention to the dynamics of all the peripheral conditions. Desmodromic systems employed to increase potential engine speed (both the opening and closing phases are cam driven) are rarely used in mass production because of reduced valve train masses in multi-valve engines and because improved valve springs have brought about an improvement in performance. In the four-cycle engine, the camshaft is driven by the crankshaft and rotates at half the crankshaft speed. The valve timing for each individual valve is determined by the geometry and the phase rotation angle of the individual cams, normally separate for intake and exhaust valves and for the cylinders that are located along one or more camshafts. In multi-valve engines, it is possible to actuate several valves using a single cam with the intervention of linkages or forked levers. In special designs, the valves of multiple cylinders or the intake and exhaust valves are activated by the same cam. In addition to the movements of the intake and exhaust valves required to control gas flow, the camshaft can also be used to generate the additional valve movements required for engine braking systems used in medium- and heavy-duty utility vehicles. Here existing or additional cam lobes are employed so that engine drag is increased during overrun or coast down; the exhaust valve might, for example, be opened briefly around dead center in the compression stroke. A further function of the camshaft, in addition to supplying power to auxiliary units (such as vacuum, hydraulic, fuel, or injection pumps), is actuating individual injection pumps in the engine block (pump-line nozzle) or pump nozzles in the cylinder head. This is used mostly in utility vehicle applications. Here, in addition to the cams that actuate the load change valves at the cylinders, further cams are provided to generate the stroke motion in the injection pump(s). Because of the

additional loading encountered here, the cams usually have to be considerably more stable in design. Torque, power output, fuel consumption, and pollutant emissions are influenced decisively by valve timing. The high specific power desired by the customer, smooth torque development, low fuel consumption and pollutant emissions all across the speed range are difficult to achieve with conventional valve trains (see also the sections on camshaft shifting systems and variable valve actuation). In every application the valve stroke length, velocity and acceleration are the products of compromises between the fastest possible opening and closing for the individual valves and the forces and surface pressures created thereby. The friction and friction losses at the camshaft and the valve train, as a whole, are also important criteria in engineering.

7.16.2 Valve Train Configurations

When using overhead valves (OHVs) the camshaft is located in the engine block, with the lift motion transferred to the valve by tappets or cam followers, push rods and rocker arms. The configuration used for this type of drive train is usually simpler, but the stiffness is markedly lower than in systems with an OHC or DOHCs. In the latter designs, the camshaft or camshafts are located in the cylinder head and driven off the crankshaft by gears, chains or belts (and, in a few cases, toothed chains). The valves are actuated by rocker arms, cam followers or valve lifters. The various types of valve trains used in passenger cars and utility vehicles and their application ranges are shown in Figure 7.258. The materials listed here for the cams and cam followers are discussed later. When the cam lift stroke is transferred to the cam follower (rocker arm, tappet or valve lifter), one may differentiate between sliding contact and rolling contact. Current development trends are toward rolling contact to reduce drive losses and increase the tolerable loading. Another trend of simple valve lifter drives is toward sliding contact (single-art cam follower without hydraulic clearance compensation) to reduce costs. In addition to reduced friction losses (which means greater engine efficiency), the improved tribologic characteristics can also reduce wear. Where rolling contact is used, the tolerable surface pressure between the cam and the cam follower is considerably greater than for sliding contact. In the same comparison, Hertzian pressure rises because of the transition from sliding to rolling contact and the curved radii. Materials with adequate rolling fatigue strength have to be selected when engineering for rolling contact; hardened steel (such as anti-friction bearing steel) is normally used. Two variants in the camshaft bearing concept are “open bearings” and “tunnel bearings.” In the open bearing concept, the bearing races are part of the camshaft; split bearings have to be used to support the camshaft. In tunnel bearings, the camshaft has bearing races with a diameter greater than the maximum cam height. The camshaft can thus be slid completely into solid bearing races in the cylinder head or the engine block (Figure 7.259).

Internal Combustion Engine Handbook | 225

6606_Book.indb 225

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Trend

Variants

Cam follower (cam contact)

Cam material Rolling contact Sliding contact

OHV push rod

OHV Rocker arm

OHV Cam follower

OHV Pushrod

Uncommon, for simple V engines and utility vehicle applications

Uncommon, constant

Standard

Common

Sliding contact rolling contact with/without HVA

Sliding contact rolling contact with/without HVA

Sliding contact rolling contact with/without HVA

Sliding contact

Steel cast iron (GJL,GJS)

(Rolling contact) Steel (Sliding contact) cast iron (GJL,GJS)

Steel Cast iron SHG (GJL,GJS)

(Rolling contact) Steel (Sliding contact) cast iron (GJL,GJS)

Steel Cast iron SHG (GJL,GJS)

with/without HVA

(Rolling contact) (Sliding contact) Steel

Steel, P/M Cast iron SHG (GJL,GJS)

(Rolling contact) (Sliding contact)

Cast iron SHG (GJL,GJS)

Figure 7.258 Valve train configurations for passenger car, motorcycle, and utility vehicle engines.

7.16.3 Structure of a Camshaft Open bearing in cylinder head Direct bearing on camshaft stem

Tunnel bearing cylinder head/cylinder head Additional bearing rings

Figure 7.259 Camshaft bearing variants.

The basic camshaft design is shown in Figure 7.260. The main component is the cylindrical shaft (either hollow or solid), upon which the individual valve actuation cams are located. As was previously mentioned, additional cams for the injection system may also be included. The actuation forces are backed at camshaft bearings, most of which are axial bearings that stabilize the camshaft along the longitudinal direction. The crankshaft is driven by a drive sprocket that is attached either permanently or detachably to the drive flange at the end of the camshaft, or camshaft shifter. The camshaft shifter can be solidly connected to the camshaft tube to eliminate the drive flange. As an alternative to this arrangement, the second camshaft in DOHC engines may be driven by the first camshaft. In this case, the first camshaft is fitted with an additional driving wheel (usually a sprocket or gear). Auxiliary units are driven with an additional driving flange or takeoff at the free end of the camshaft or, for example, by an eccentric or lift profile at some point along the camshaft. A trigger wheel (generating one or more pulses per revolution) may also be mounted on the camshaft to ascertain the angular position of a camshaft, in particular when camshaft shifters are used. The cam comprises one section with a constant radius (base circle) and the lifting area (run-up and run-down ramps, cam flank and cam nose). The difference between the base

226 | Internal Combustion Engine Handbook

6606_Book.indb 226

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

Impulse generator ring (phase position of camshaft)

Camshaft stem (oil supply through camshaft) Axial bearing (between cams) Injection pump cam

Output element (e.g., vacuum pump)

Camshaft tube with partial section (red)

Valve cams

Cam tip Hexagon (service and assembly)

Cam lift

Drive flange (e.g., camshaft adjuster)

Cam flank Base circle Figure 7.260 Design of a camshaft with partial section (red).

circle and the highest point on the cam represents the cam lift stroke, which is selected to be proportional to the desired kinematic valve stroke. In systems with mechanical clearance adjustment, faults in the cam’s base circle (deviations of the base circle from constant radius) have no effect on operational properties. A system with hydraulic valve play adjustment, by contrast, responds to every change in the base circle. Where there is a fault opposite the direction of movement, the hydraulic valve lifter (HVA) compensates for this fault as valve play; in this case the valve stroke increases. If there is an error in the cam base circle in the lift direction, then the valve is already opened in the base circle segment because of the associated rise in force. This “pumping up” can, in extreme cases, result in complete loss of combustion chamber compression, burn-out of valves or valve seat inserts and subsequent engine failure.

7.16.4 Technologies and Materials

Camshafts made of cast iron are very widely used and differ in terms of the micro-structure and hardness. Figure 7.261 provides an overview of the technologies and materials used Assembled camshafts become more popular. Assembled camshafts are made up of individual components (tube, cams, drive flange, etc.) that have been assembled. The materials can thus be matched exactly to the particular requirements. When demands are extreme, camshafts forged from steel or machined from solid material (bar material) are used. 7.16.4.1 Cast Camshaft A camshaft made of cast iron with nodular (GJS) or laminar (GJL) graphite is often the ideal tribologic match for sliding contact and low-load rolling contact in many applications. With

proper alloying and closely defined hardening of the cams, tolerable pressure levels of well over 1000 MPa can be attained. In the case of chilled cast iron (SHG), the cam area is cooled quickly following casting to create a wear-resistant carbide structure (ledeburite) with great hardness and good tribologic compatibility. A gray casting with good machining properties is available for use in the core area and the camshaft bearing points (Figure 7.262). A special version of the cast camshaft is the so-called trilobe camshaft of the VarioCam Plus valve actuation system. Figure 7.263 shows the structure of the cams with large cam stroke (outer cams) and small cam stroke (inner cam). In combination with a switching tappet, the valve stroke change and thus, a variable valve train is shown. 7.16.4.2 Assembled Camshaft Serving as the basis for an assembled camshaft is a tube to which individual cams are attached by thermally inducted shrink fit, friction press fit, interior high-pressure forming or a comparable joining process. It is possible to distinguish between camshafts in which the tube and all the attached components are present as finished parts when they are attached and require no further machining and those processes in which the camshaft following assembly is available as a rough component (either in whole or in part), which has to be ground like conventional (unitized) camshafts. The advantage of these assembled camshafts is, in addition to the savings in weight, the design degree of freedom for individual camshaft components and the flexibility in material selection. For example, the material for cam, tube, bearing, drive and takeoff elements can be selected independent of each other and for the specific application optimized in respect to costs, properties and manufacturing technology.

Internal Combustion Engine Handbook | 227

6606_Book.indb 227

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Technology

Cast camshaft

Cam materials

In Serie für:

Cast iron with nodular graphite (GJS) inductive hardened

Passenger cars

Cast iron with flaked graphite (GJL) remelt hardened (WIG)

Passenger cars

Cast iron with nodular graphite (GJS) chilled casting

Passenger cars / Trucks

Cast iron with flaked graphite (GJL) chilled casting

Passenger cars / Trucks

Steel

Passenger cars / Trucks

Sinter materials

Passenger cars

Sinter materials (precision cams)

Passenger cars

Assembled camshaft

In development Passenger cars / Trucks

Cast Forged camshaft

Steel

Rod material machined

Steel

Camshaft

Trucks

Figure 7.261 Camshaft technologies and materials.

Cross-section

Camshaft: "gray" set area

180° irradiation 360° structure with "aligned" carbides

Flaked graphite

Modular graphite Camshaft:

Longitudinal section not etched

etched

Figure 7.262 Chilled cast iron camshaft in cross section.

Steel or sintered materials (P/M steel) are used as the material for cams, cast cams are in development. Steel cams are normally forged as a rough part; the inner bore is then machined, and the cam is mounted on the tube. To attain the required materials properties, the cam can be hardened and tempered before or after attachment.

Using sintered material at rolling contact points makes it possible, because the cam geometry can be sintered more exactly than the required manufacturing tolerances, to build a camshaft that need not be further worked once the inside bore has been machined and the cam has been mounted on a tube with final geometry.

228 | Internal Combustion Engine Handbook

6606_Book.indb 228

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

for actuating injection pumps (PD/PLD) or when long service life must be ensured (stationary or marine applications). When there are high demands in terms of torsional and/or tensile strength, steel shafts also have to be used for sliding contact.

Large cam lift Small cam lift

Figure 7.263 TriLobe camshaft with stroke changeover.

Figure 7.264 shows some examples of cam materials for assembled camshafts. A high-alloy, liquid-phase sintered, powdered-metal steel was developed for use as a sintered material for sliding contact. Assembled camshafts with chilled-cast iron cams are in development. With this concept, the advantages of the cast material in the sliding contact, such as cam followers, can be combined with the advantages of the assembled camshafts such as low weight. The chilled-cast iron cams can be manufactured in individual casting processes or as bar material. After the cams in the bore are machined, they are joined, similar to steel or sintered cams. 7.16.4.3 Steel Camshaft Used for almost all applications with rolling contact in utility and stationary vehicles and in many passenger cars are forged steel camshafts or steel camshafts that are machined from solid material (Figure 7.265). Steel camshafts are used especially when camshafts have to absorb high forces via additional cams

Cast cams

Figure 7.265 Segmented camshaft for stationary engine: a replaceable cam/bearing element for each cylinder.

With the high tolerable pressure levels and the good mechanical properties of the material, these camshafts can be used for maximum demands, provided that correct tribologic mating materials are used. 7.16.4.4 Special Camshaft Designs The assembled camshaft can make a crucial contribution in meeting the varied requirements of modern combustion engines on the valve actuation. To the well-known advantages compared to cast camshafts (less weight, material selection), the integration of auxiliary functions and the optimization of the camshaft regarding the application circumstances can be added. To reduce valve drive friction and wear caused by insufficient lubrication or continuously increasing loads, camshafts with

Precision sintered cams Sintered cams

Structure Technology Material Hardening process

Particularity

Particularity, Croning

Double pressed and sintered

Liquid phase sintering

100Cr6 Hardened and tempered

Chilled casting

Carburized Sinter hardened and tempered

Hardness Application Particularity

Rolling contact (Sliding contact)

(Sliding contact) (Sliding contact) (Sliding contact) Rolling contact Rolling contact Rolling contact Cam profile Suitable for high Sliding cup Sliding cup grinding contact pressure coating coating not required in rolling contact not required not required

Figure 7.264 Cam materials for assembled camshafts.

Internal Combustion Engine Handbook | 229

6606_Book.indb 229

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

coated functional surfaces can be used. Coatings, such as are used in cam followers, are worth considering. The design of these camshaft types is now under development. The camshaft on roller bearings (low-friction camshaft—LFC); see Figure 7.266, allows, with minimal changes, a reduction of losses in the combustion engine by using rolling bearings instead of the traditional plain bearings. Thanks to the closed rolling bearings installed on the camshaft, bearing friction can be halved and pressure oil supply at the bearing points, as required for plain bearings, can be omitted. The rolling bearings are lubricated by oil mist in the cylinder head. The direct integration of the LFC in the simplified cylinder head (shroud) is possible.

freezing and a significantly simplified cylinder head shroud with reduced height. The use of variable camshafts has positive effects on consumption, emissions and torque characteristics because a compromise between high torque in lower and medium speed ranges and the required nominal engine output is not necessary. Thanks to partially or fully variable valve actuation by function integration into the camshaft, the actuating time and/or the valve stroke function can be selected according to the engine operating point in the map. With the MAHLE CamInCam (see Figure 7.268), the function of two camshafts in the space of just one camshaft has been integrated.

Figure 7.266 Camshaft on rolling bearings.

In a camshaft with integrated oil mist separation, the blow-by gas is fed through the camshaft where a swirl generator rotates with the camshaft (Figure 7.267). Centrifugal force carries the heavy oil shares in the blow-by gas to the outside and forms a wall film in the interior of the camshaft tube. This film is separated from the cleaned blow-by gas at the camshaft end pointing away from the camshaft drive side. In comparison with conventional, separately installed separation systems, this active separation system is characterized by increased oil separation rates, its immunity from

Figure 7.268 MAHLE CamInCam.

This camshaft comprises an outer camshaft on which, on the one hand, cams are solidly connected with the camshaft tube via a thermally induced shrink-fit, just like a traditional camshaft. The other cams are attached to the camshaft tube on which they can rotate, and are connected with the second shaft in the interior via pin connection. By using a simple, a dual-effect or two camshaft shifters, both shafts, and thus

Cleaned blow-by

Dip tube separator Drallerzeuger Swirl generator Einlass Inlet blow-by Blow-By Dichtring Sealing ring Oil film generation Erzeugen eines Ölfilms

Spritzschutz Splash protection

Oil return

Figure 7.267 Camshaft with integrated oil mist separator.

230 | Internal Combustion Engine Handbook

6606_Book.indb 230

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

adjacent cam pairs, can be adjusted in their angles independent of each other (for example, the inlet and outlet time of the corresponding valve lifts). Compared to engines with two camshafts (DOHC), it is possible to save further weight and space in engines with space for only one camshaft (SOHC and OHV). The free space can be used to improve aerodynamics and passive pedestrian protection systems. With the use of a CamInCam, it is possible, in engines designed with two camshafts, to continuously optimize the opening time of the valves for a favorable gas exchange cycle in respect to the present operating cycle (Miller/Atkinson cycle) by relatively adjusting the inlet and/or outlet. By using an intermediate lever in combination with a modified CamInCam, a mechanically fully variable valve traincan be provided for throttle-less load control (Variable Lift and Duration—VLD; see Figure 7.269). For this purpose, cams with a special shape are attached to the camshaft tube which only affects the opening or closing of the valves. The general structure of the camshaft in this fully variable valve train system is the same as the CamInCam. The integration of a stroke changeover in the camshaft has been realized in the partially variable AVS valve train system (Audi Valvelift System) (see Figure 7.270). The camshaft comprises a basic shaft with rolled-on involute toothing and spring-loaded interlocks for the cam segments which are movable in an axial direction on the shaft and featuring longitudinal toothing at the interior. The toothing of the shaft and the cam segments interlink and transfer the torque introduced from the crankshaft via the timing chain over the different cam contours of the cam segments to the cam follower. Another special version of camshafts is a camshaft made from injection-molded plastic (see Figure 7.271) which are used in portable engine-driven devices such as lawnmowers, leaf blowers, high-pressure cleaners and similar equipment.

Figure 7.269 Fully variable valve train VLD.

The significantly lower mechanical and thermal stresses in comparison to passenger cars, utility and motorcycle engines and the moderate requirements on service life at low weight allow the use of plastic materials as they are used in chain tensioning systems.

Figure 7.270 Exploded view of an Audi Valvelift system camshaft [7-146].

Internal Combustion Engine Handbook | 231

6606_Book.indb 231

1/19/16 8:34 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.16.5 Reduction of Mass

Figure 7.271 Plastic camshaft for portable engine-driven devices.

7.16.4.5 Materials Properties and Recommended Matches Figure 7.272 shows, for example, the spreads for torsion and tensile strengths for various cast materials. Various potential matches for rolling and sliding contact and the tolerable Hertzian pressures in each case are shown in Figure 7.273 and in the summary of trends in Figure 7.258. Starting with the simplest gray-cast camshaft, with cast tappets as the cam followers for sliding contact, it is possible to cover the entire range with the pairs of materials depicted, through to high-load rolling contact with cams and rollers made of roller bearing steel (100Cr6).

Similar to the situation for the vehicle as a whole or for the overall valve train, the camshaft as an individual component is subject to the necessity of reducing masses. On the one hand, the engine’s static mass is minimized, while, on the other hand, the moving (rotating) masses have great influence on the dynamics of the total system. At the same time, it is always necessary to reach a compromise among technical feasibility (minimum wall thickness, etc.), costs (materials, machining steps, etc.), and functioning (cam width, diameter of the base circle, torsion stiffness etc.). One possible option is hollow drilling or cylindrical hollow casting for camshafts (20% reduction in mass). When using hollow casting techniques with a graduated inner contour (profiled cavity), the mass is reduced even further. Figure 7.274 shows some examples of mass reductions in comparison with camshafts made of solid material and examples of camshafts with hollow cylinders and profiles. The assembled camshaft today presents the greatest potential for reducing masses. The steel tube’s wall thickness can be reduced further than the wall thickness in the casting process. Integrating the camshaft bearing into the camshaft itself (tube diameter = inner race diameter) permits additional savings in masses. An important design criterion for such shafts is the joint between the cams and the tube, with its influence on the moment, which can be transferred.

Torsion resistance and Tensile strength [MPa]

Torsion resistance

Tensile strength

Samples from camshafts

Specification values: - Separately cast tensile samples - DIN 50.125 Ø 6 x 30 mm

Material Figure 7.272 Strength values for various casting materials.

232 | Internal Combustion Engine Handbook

6606_Book.indb 232

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

Cams (C): Cam follower (CF): C: P/M CF: Steel

Steel Steel

Rolling contact

C: SHG CF: Steel C: Steel CF: Hard metal

Sliding contact

C: SHG, Steel CF: Steel (nitrided) C: SHG, Steel CF: SHG (alloyed, hardened) C: SHG, Steel CF: SHG (low-alloyed, nitrocarburized)

Hertzian pressure [MPa]

Mass/cylinder [g]

Figure 7.273 Pairs of materials and Hertzian pressures.

Solid material Cylindrically hollow Profile hollow Assembled camshaft

Cylindrically hollow (stepped)

Cylinder spacing [mm]

7.16.6 Factors Influencing Camshaft Loading

Valve drive kinematics are the primary determinant for camshaft loading. The peripheral geometric conditions such as the step-down ratio or cam profile (e.g., high acceleration rates) are decisive here in particular. Moreover, the camshaft is loaded by the valve train masses in motion and the total forces exerted by the valve springs and exhaust gas counter-pressure. An integrated engine braking system can impose further and usually very significant loading on the camshaft (five to ten times the forces encountered during normal changes of gas charges). Figure 7.275 shows some of the influencing factors for camshaft loading.

Profile hollow

Figure 7.274 Reducing masses in camshafts. See color section page 1077.

The contact forces created between the cam and the camshaft induce both torsional and flexural moments in the camshaft which, with the drive moment for auxiliary units, give the total torsional and flexural loads for the camshaft. In addition to the contact forces and the subsequent moments, the geometrical surfaces and the elasticity modules determine the Hertzian pressures and deformations. Geometrical surfaces of the cams and cam followers include the form progression in the movement plane, the contact surface (point, line, or surface contact), the contact width and the crowning across the contact width.

Internal Combustion Engine Handbook | 233

6606_Book.indb 233

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Engine braking operation (truck engines) (e.g., "jake" brake)

Ventilfederkräfte (Abgasgegendruck) Massenkräfte (bewegte Bauteile): Ventil Federteller+Befestigungskeile Ventilbrücke Kipphebel Stoßstange Nockenfolger Zylinderdruck bei “ Auslass öffnet“

Kontaktkraft zwischen Nocken- Nockenfolger (Drehmoment am Nocken) Torsions- und Biegebelastung der Nockenwelle

„Scharfe“ Nockenprofile (hohe Beschleunigungen) Nockenwellenantrieb; Antriebsmoment für Hilfs- und Nebenaggregate

Figure 7.275 Factors influencing camshaft loading.

7.16.7 Design of Cam Profiles

High engine torque in low and medium speed range and high-rated engine output at the same time make contradictory demands on the valve stroke curve. In valve trains with fixed timing, the valve stroke curve is a compromise for optimal filling over the entire engine map area. This conflict of goals can be partially or fully resolved by using partially or fully variable valve actuation systems in which different, adjusted valve stroke curves are achieved for each engine operating point in the map area. The valve stroke curve demanded from the thermodynamic calculations of the engine manufacturer and the peripheral geometric conditions such as valve diameter, valve stroke and valve clearance to the piston at TDC are most important on one hand, while the demands in terms of functioning and manufacturing (such as jerk-free transitions in the entire valve stroke cycle or thermal loading of the exhaust valve while opening) are, on the other hand, the most important parameters. This specified and targeted advance for the valve stroke, depending on the type of valve train and its kinematics, is recalculated to form a cam profile matched to the cam follower. If mechanical valve clearance adjustment is implemented, there is always some play in the total system between the cam and the valve before the valve opens to offset any thermal expansion and mechanical setting during operation. This play causes inconsistency at the start of the stroke and thus always creates a sudden load. During the closing cycle there is also

a “bump” because the valve contacts the valve seat before the cam stroke is completed. To limit the seating velocities and sudden accelerations for the valve train components involved, it is necessary to provide the appropriate opening and closing ramps. Variances in the valve stroke occur in systems with mechanical valve clearance compensation, the extent depending on wear and temperature; valve overlap also varies (phase during which the intake and exhaust valves are both open). In valve trains incorporating hydraulic valve lifters (HVA), this valve clearance is not present and these ramps can be manufactured much flatter (Figure 7.276); the valve stroke and overlap are nearly constant with HVA. An important criterion for design is Hertzian pressure. This indicator value describes the compressive load on the mating components. Using the maximum tolerable Hertzian pressure allows us to preselect potential materials for cams and cam followers. The dynamics calculation usually shows, in comparison with the basic kinematic design, more realistic values for the location and size of maximum pressure values (Figure 7.277). When a roller is used as cam follower (roller tappet, roller lever, roller cam follower), there are often concave radii in the cam flanks to produce the required valve stroke curve in respect to valve stroke height, valve opening time and valve acceleration. The calculated concave radii of the cams must always be larger than the contact radius of the cam

234 | Internal Combustion Engine Handbook

6606_Book.indb 234

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

Cam angle Theoretical valve stroke Theoretical valve speed Theoretical valve acceleration

Cam angle Theoretical valve stroke Kinematic pressure Dynamic pressure

follower to ensure constant valve stroke curves. Here, it is necessary to consider manufacturing limitations in reference to grinding. It may be necessary under certain circumstances to accept deviations from the specified valve stroke curve. When using sintered cams, the outer contours of which require no additional machining, any concave radius can be realized (at least in principle). When using an assembled camshaft, it is necessary, depending on the system, to pay attention to the moment that is to be transferred or can be transferred as a decisive magnitude. During engineering, one must ensure that the maximum dynamic moments can be transferred with the required degree of confidence.

Figure 7.276 Valve stroke, speed, and acceleration plotted against the cam angle for a roller cam follower valve train with hydraulic valve lifters. See color section page 1077.

Figure 7.277 Theoretical valve stroke and Hertzian pressure (kinematic and dynamic) for a roller cam follower drive train with hydraulic valve lifters. See color section page 1077.

7.16.8 Kinematics Calculation

In the kinematic (quasi-static) calculation, the moving masses in the individual valve train are reduced to one single mass and one spring (the valve spring). A targeted motion (corresponding to the progression of the valve stroke) is imposed upon this individual mass. The mass and spring forces are considered in this way; additional outside forces such as gas forces coming into play when the exhaust valve is opened can be taken into account. The most important results of kinematic calculations include the hydro-dynamically effective speed for sliding contact, roller speed for rolling contact, contact forces and/or Hertzian pressures between the cam and its follower (as well as bearing

Internal Combustion Engine Handbook | 235

6606_Book.indb 235

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

loads for the valve train components), loading, and relative motion of the driving element at the end of the valve shaft or a valve link (e.g., valve finger radius, elephant foot, etc.). The hydro-dynamically effective speed (total speed, lubrication index) is a measure of the cohesion of the lubricating film between the components in contact (Figure 7.278). In the case of sliding contact, there are usually two “zero intersections” (change of sign) in this curve during each cam revolution. Because, at that particular moment, the load-bearing capacity of the lubricating film collapses, the risk of wear can be reduced by suitable design. When there is rolling contact with roller bearings (such as a needle bearing in the case of a roller cam follower), it is possible to forecast the service life (taking into account various loading populations).

stiffness, also depicts the natural frequencies for the system being observed. The output depicts the motions of the individual components and the forces and pressures effective upon them. One sees in Figure 7.279 that the force between the cam and roller deviates distinctly from the progress determined kinematically, which is the result of the oscillations superimposed on the targeted motion. Particularly in valve trains with hydraulic clearance compensation, loss of contact force can result in grave problems (pumping up the hydraulic valve lifters, resulting in wear or failure of valve train components). A dynamic analysis of the valve train can identify critical components at the engineering stage (long before parts are developed for measurements and engine operation) and thus shorten the development process considerably.

7.16.10 Camshaft Shifter Systems

7.16.9 Dynamics Calculations

Calculations of the dynamics supply a far more accurate image of real system behavior than does the relatively simple kinematics model. Accordingly, greater effort is required for modeling. Multi-body simulation is the tool used for dynamics calculations. Common to all such programs is that the mechanical systems being assessed are broken down into individual masses and they are then coupled one with another by means of spring and damping elements corresponding to the stiffness of the components and their damping properties. In addition to integrating hydraulic subsystems (hydraulic valve lifters) into the simulation, it is also possible to use the results from FEM calculations, for example, force- or path-dependent stiffnesses for components. The degree of detail for the dynamics calculation is virtually as desired and is limited only by the ratio of benefit to effort. With all these elements and peripheral conditions, there arises a model capable of oscillation that, in addition to the

Cam angle Theoretical valve stroke Cam contour radius Hydrodynamically-effective speed

To comply with future exhaust gas regulations and to reduce fuel consumption, elements that influence valve timing are used more often in gasoline engines. The camshaft shifter is one such device. It enables continuous change in the timing for a camshaft, across a wide angular range. This makes a change possible in valve overlap in DOHC engines and thus influences the residual gas content in the combustion chamber. In addition, it is possible, above all at idle and full throttle, to tune timing for maximum comfort and/or maximum torque and highest performance. Camshaft shifters have been used in vehicles since the mid-1980s, initially as two-state shifters with simple controls but today more often as continuously adjustable systems operating under closed-loop regulation. In DOHC engines, camshaft shifters are used mostly on the intake shaft; typical adjustment angles lie between 40 and 60 crankshaft degrees. There are, however, also shifters in mass production, used on the exhaust side, preferably in turbocharged engines. Both degrees of freedom may be

Figure 7.278 Cam contour, theoretical valve stroke, and hydrodynamically effective speed plotted against the cam angle at contact between cam and flat tappet. See color section page 1078.

236 | Internal Combustion Engine Handbook

6606_Book.indb 236

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.16 Camshaft

Cam angle Theoretical valve stroke Kinematic contact force Dynamic contact force, n = 6000 1/min

combined where there are maximum demands regarding performance and exhaust gas quality. In some DOHC engines camshaft shifters are used for dethrottling, that is, reducing consumption by closing the intake valve late. In this concept, however, neither an increase in performance nor an improvement of comfort at idle can be achieved since the valve overlap is not changed. Continuous camshaft shifting operates in a closed-loop regulation circuit and is hydraulically powered in all cases nowadays. In the engine management system, the required set-point angle for timing adjustment is taken from an engine map dependent on load and speed. This is compared with the actual, measured angle. Deviations between the set-point and actual angles are evaluated with a regulation algorithm and cause a change in the electrical power applied to the control valve. Thus, the valve diverts oil into the chamber at the valve shifter in a fashion corresponding to the desired adjustment direction, while oil is allowed to escape from the opposite chamber. The angular position of the camshaft changes in accordance with the degree of fill at the oil chambers in the shifting unit. Sensors scan the trigger wheels at the camshaft and crankshaft; the actual value is calculated on the basis of these signals. This regulation process runs continuously, at high frequency, and thus leads to good response characteristics when there are rapid changes in the set-point angle, giving high angular accuracy in maintaining the set-point angle. The system generally uses the motor oil circuit as its power supply; systems with a separate high pressure supply are also found in sports engines. The following components are needed to implement camshaft shifting: •• The hydraulic shifter unit, mounted on the drive end of the camshaft. This component sets the adjustment angle in response to the alternating filling of two oil chambers. Low

Figure 7.279 Theoretical valve stroke, kinematic contact force, and dynamic contact force plotted against the cam angle for a rocker arm valve train with hydraulic valve lifters. See color section page 1078.

leak rates and sufficiently large piston surfaces ensure good stiffness under load. The shifter unit is built in various styles—with a linear piston and helical toothing or with a rotary piston. •• The regulation valve, built into the cylinder head or an attached component, should be located near the point at which oil is transferred to the camshaft. This valve is controlled electrically, usually with a pulse-width modulated signal; it regulates the flow of oil into and out of the chambers in the shifter unit. A high flow rate during adjustment phases and precise regulation capacities to fix the angle are the most important features of the valve. •• The regulation circuit for continuous adjustment comprises suitable software and a power output stage in the engine management unit as well as trigger wheels and sensors at the crankshaft and camshaft. Components already present in the engine can be used for this purpose, although the trigger wheel at the camshaft has to be modified. The overall system for continuous camshaft shifting and the components described above are shown in Figure 7.280. Two concepts for the hydraulic shifter unit have become commonplace. A brief review of their basic design is provided below. The camshaft shifter with helical toothing consists of these main functional components: the drive sprocket (joined with the crankshaft), adjustment piston, and output hub (bolted to the camshaft). These components are joined one with another in pairs, via helical inner toothing, so that an axial shift of the adjustment cylinder causes the drive hub to rotate in relationship to the drive wheel. The transfer of the torque using inner toothing is very robust. The design shown in Figure 7.281 is completely sealed, for use in toothed belt drives.

Internal Combustion Engine Handbook | 237

6606_Book.indb 237

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Camshaft timing devices

Camshaft timing diagram

Camshaft position sensor trigger wheel

Crankshaft sensor trigger wheel

Control valve

Engine control Chamber connected with engine oil pressure

EMS

Chamber unloaded/oil return

10 40

30

0

20

Figure 7.280 Continuous camshaft shifting.

When the engine starts, the spring shown in the illustration keeps the shifting piston in its home position. Both chambers are filled with oil during regulated operation; good sealing between the two chambers provides good stiffness under load. Quick responses demanded by the engine are achieved with engine oil at a pressure of approximately 1.5 bar. Shown in Figure 7.282 is the slewing motor or vane shifter in a version for chain drives. This version of the camshaft shifter is more compact and economical than the version with helical toothing; it comprises only the drive gear and the output hub. Rotary torque is transferred during operations by the oil fill in the chambers. Only during engine starting does a locking element normally ensure a fixed mechanical link between the drive and output elements. This locking element is unlatched hydraulically once the camshaft shifter has filled with oil. The locked end position here is, as a rule, “late” timing when adjusting the intake camshaft and the “early” timing setting when adjusting at the exhaust camshaft. The regulation valve comprises a hydraulic section and a solenoid. The hydraulic slider is located in a bore with connections for oil supply, actuator chambers for the camshaft shifter, and oil return. A spring moves the slider toward the home position. When power is applied to the solenoid, the slide is shifted against the force of the spring. This changes the flow of oil into and

out of the two chambers; in the so-called regulated position all the oil ports are largely closed. This achieves stiff holding of the adjustment piston in the camshaft shifter. In accordance with the accepted parameters of the particular application, the regulation valve either is integrated directly into the cylinder head or is attached by an intermediate housing. The regulation valve is connected electrically to the engine management unit.

7.17 Chain Drive The primary function of the camshaft is to ensure that the valves open and close at the correct times. In modern overhead valve engines this is done by power transmission from the crankshaft. In most cases, toothed (synchronous) belts, or roller toothed (silent) or bushed roller chains [7-151], [7-152] of various weights are used. The selection of the design depends on the engine maker‘s design philosophy. The most important criteria in the choice of drive concept are the costs, the amount of space taken up, maintainability, service life, and noise generation. Figure 7.283 shows a comparative evaluation of timing chain and synchronous belt.

238 | Internal Combustion Engine Handbook

6606_Book.indb 238

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.17 Chain Drive

Camshaft position in regulating position

Basic position

Chamber unloaded/oil return Chamber connected with engine oil pressure

Figure 7.281 Slewing motor or vane shifter.

Camshaft position in regulating position

30°

Corresponds to 60° Crankshaft angle

Basic position

Chamber connected with engine oil pressure Chamber unloaded/ oil return

Figure 7.282 Camshaft shifter with helical toothing.

Internal Combustion Engine Handbook | 239

6606_Book.indb 239

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Timing Chain

Synchronous Belt

Installation space

+

O

Service life

++

O

Costs

O

+

Maintainability

O

O

Noise generation

O

O

In the course of decades of experience, certain dimensions for the roller and bushed roller chains used in timing drives have proven to be particularly suitable.

7.17.1 Chain Designs

Among the standard chains, one differentiates between roller and bushed roller chains. In addition, there are both simplex and duplex chains (Figure 7.285). A special version of the chain is the toothed chain also called a silent chain (Figure 7.286).

Legend: ++ very good + good O adequate

Figure 7.283 Comparative evaluation of timing chains and belts.

In modern engines, these power transmission systems often serve not only the camshaft, but other components such as the oil pump, water pump, and fuel injection pump as well. Examples of possible arrangements are shown in Figure 7.284. DOHC diesel 4V

DOHC-4V

Split chain drive with two hydraulic tensioners

SI engine with single-strand chain and oil pump drive

Section A-A

Figure 7.285 Chain designs.

A

A

Figure 7.286 Silent chain. Figure 7.284 Timing chain drive.

Because neither the camshaft nor the crankshaft runs entirely smoothly and the power required by the injection pump is subject to severe, periodic fluctuations, this drive system is exposed to very complex dynamic loading [7-153], [7-154]. Oil pump drive Mass balancing drive 3.5

Control chain for SI engines

4.76

5.72

8 9.525

˘ 7

˘ 5

˘ 6.35

11.7 9.8

7 mm8 mmBushed roller Bushed roller chain chain

14

Traditional toothed chains used in timing drives always have a pin joint. By combining the advantages of a bush chain and a toothed chain, a new chain type for drives was developed, the bushed tooth chain (Figure 7.287). This chain variant is recommended for every application focused on low wear, and good acoustic and dynamic behavior [7-155]. Control chain for diesel engines 5.72 5.72

min. 5.72

9.525

9.525 9.525

˘ 6.35

˘ 6.35 ˘ 6.35 10.46 15.5

max. 23.7 3,8'' 3,8'' Roller chain Roller chain

10.5 23.8

3,8'' 3,8'' Bushed roller Bushed roller chain chain

Figure 7.285 Chain designs.

240 | Internal Combustion Engine Handbook

6606_Book.indb 240

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.17 Chain Drive

sprocket diameter using a simplex chain requires a number of teeth < 18, it is advisable to use a multiple chain with the same or smaller pitch.

7.17.2 Typical Chain Values

Three essential factors characterize a chain’s suitability for use as a timing chain: •• breaking strength •• endurance (Figure 7.288) •• wear resistance. Figure 7.287 Bushed tooth chain.

In toothed chains, the plates are shaped so as to enable direct force transfer between the chain and the sprocket while, in roller and bushed roller chains the interface with the sprocket takes place at the pivot joint via pins, bushes or rollers. Silent chains can be made of any conceivable width without any fundamental change in design. Guide plates are provided to keep the chain from wandering off the sprocket; they may be located either at the center or on either of the outside edges. The rollers, rotating over the bushes in a roller chain, encounter a small amount of friction when rolling along the sprocket’s teeth. Thus, the contact point at the circumference changes continuously. The lubricant between rollers and bushes contributes to noise and impact damping. In a bushed roller chain, by contrast, the fixed bushes always mate with sprocket teeth at the same point. Thus, perfect lubrication for such drives is particularly important. At the same pitch and failure strength, a bushed roller chain exhibits a larger joint surface than corresponding roller chains. A larger joint surface causes lower pressures at this surface area and thus less wear. Bushed roller chains have proven their value, particularly for heavily loaded camshaft drives in high-rpm diesel engines. Whenever the transfer of a given torque at a certain maximum

One cause that might be responsible for failure is exceeding the static or dynamic breaking load. Particularly in timing drives, one does not encounter uniform loading. Pulsating loading on the chain results from fluctuating torques at the camshaft and the injection pump (in diesel engines, for example), non-uniform camshaft rotation, and pulsating longitudinal chain forces caused by the polygonal effect. Here the chain’s fatigue strength must never be exceeded because the number of load alterations during an engine’s service life is in all cases greater than 108. In today’s engines with their precise timing, minimal stretching because of wear, at from 0.2% to 0.5% of chain length at up to 250,000 km in service, can be attained. A chain timing system with its mass, stiffness and damping represents a system capable of oscillation, having several degrees of freedom. In response to excitation by the camshaft, crankshaft, injection pump, etc., this can cause resonance effects that result in extreme loading of the timing drive system. Engineering measures make it possible to increase stiffness in the chain while retaining its specific mass (Figure 7.289). This shifts resonance points toward the higher frequencies.

7.17.3 Sprockets

The shapes of the teeth in sprockets intended for use with roller chains, bushed roller chains, and silent chains are standardized (DIN 8196). Proper tooth profile is just as important to

10000

Load peak in N

Bushed roller chain, diesel applications

Bushed roller chain prechamber, diesel engines

5000

Roller chain, high-performance control drives for SI engines Roller chain, oil pump chain

0 10000

100000

1000000

10000000

Charge change number

100000000

Figure 7.288 Fatigue strengths for roller and bushed roller chains.

Internal Combustion Engine Handbook | 241

6606_Book.indb 241

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

2400 2100

1000

2800

800

1500 1200

600

900

400

600

200 0

Specific rigidity (kN)

Specific mass (g/m)

1200

300 Toothed belt

Singlestrand roller chain

Doublestrand roller chain

Singlestrand bushed roller chain

Doublestrand bushed roller chain

reliable operation of the timing system as, for example, the chains’ wear resistance. Usually sprockets with the widest tooth gap are used. This makes possible uninterrupted engagement and disengagement of the chain even at higher chain speeds, because of the short teeth and the wider gap between teeth. Depending on the amount of space available and the particulars of the application, pulleys or sprockets with one or two rows of teeth may be used (Figure 7.290).The selection of the materials depends on the timing drive system parameters, the operating conditions, and the amount of power to be transferred. Carbon steel, alloyed steels and sintered materials are used for the sprockets. The materials used for precision punched sprockets include C 10 or 16MnCr5 for sprockets made with cutting processes, and D 11 for sintering processes, with the heat treatment suitable for each particular material.

Singlestrand bushed roller chain

Doublestrand bushed roller chain

0 New chain generation

Figure 7.289 Typical chain values: stiffness and mass.

Sintered sprocket

7.17.4 Chain Guide Elements

With the introduction of continuous-action tensioning and guide elements that are matched exactly to the particular engine, the drive can be optimized to such an extent that its service life equals that of the engine, without any special care being required beyond the prescribed engine maintenance. The chain tensioner (Figure 7.291) assumes a number of functions in the timing drive. First, the timing chain is preloaded (along the slack span) to a defined value under all operating conditions, even where stretching because of wear has occurred. A damping element, using either friction or viscous damping, reduces oscillations to an acceptable amount. Simple rails made of plastic or metal are used as guide elements. They usually have a plastic surface and are either flat or curved to fit the chain’s path (Figure 7.292). The newer versions of these rails are usually injection-molded plastic. As regards the tensioner rails, a slip-promoting covering made of PA 46 is injected or clipped onto a backing element made of PA 66 with 50% glass fiber content for reinforcement

Machined sprocket

Precision-stamped sprocket

Figure 7.290 Chain designs.

242 | Internal Combustion Engine Handbook

6606_Book.indb 242

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.18 Belt Drives

Hydraulic chain tensioner Chain tensioner screwed-in

Tensioning unit

Chain tensioner, flanged

Figure 7.291 Chain tensioner.

purposes. The slip rails are usually manufactured as a unitized component

Tensioning rails 2 components plastic

to advantages found in the simplicity of the drive concept, flexibility in belt guidance, low friction, and cost advantages when compared with other drive systems. Moreover, auxiliary units such as oil or water pumps can be integrated into the drive concept. 7.18.1.1 Synchronous Belt Drive  ynchronous belt design S The synchronous belt is a bonded system made of three components (Figure 7.293): •• polyamide fabric •• rubber blend •• tensile member, usually an endless-wound glass cord.

Nylon fabric

Rubber blend

Tensile member

Sliding rails 1 component plastic

Figure 7.292 Guide elements.

7.18 Belt Drives This section provides an overview of the demands and functions of today’s belt drives in internal combustion engines, synchronous belt drives used to drive the camshafts, and Micro-V belt drives used to run auxiliary components.

7.18.1 Belt Drives Used to Drive Camshafts

Synchronous belt camshaft drives today hold a 50% market share in European engines. This can be traced essentially

Tensile member

Figure 7.293 Synchronous belt design.

The facing fabric is made of high-strength polyamide and is coated to reduce wear. It protects the rubber teeth and the cord in the area of belt web against wear. The rubber blend is a high-strength polymer. Polychloropene (CR) was used in early versions. Because of stringent requirements in terms of dynamic strength and resistance to temperature and aging, HNBR (hydrogenated nitrile rubber) materials are used exclusively today.

Internal Combustion Engine Handbook | 243

6606_Book.indb 243

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

The cords in the tensile member are made of glass fiber—a material distinguished by its great tensile strength and flexibility in bending. Consequently, it is particularly well-suited for camshaft drives in which the crankshaft sprockets are small in diameter. The manufacturing process is such that the strands in the tensile member are twisted, clockwise and counterclockwise, in pairs, to achieve largely neutral running properties for the belt.  ynchronous belt profiles S There has been a very versatile evolution in the profiles used for synchronous belts because they were initially employed as timing belts. A wide variety of profiles are in use today. The various profiles and their properties are discussed below. The first camshaft drive belts were based on the classical Power Grip design with its trapezoidal teeth, at that time already in widespread use in industrial applications. In response to increasing demands for power transmission, ratcheting resistance, and quiet running, curvilinear profiles (Power Grip HTD/High Torque Drive) were developed. When compared with the trapezoidal shape, the forces are introduced more smoothly to the tooth with the rounder profiles, and this in turn reduces the possibility of tension peaks (Figure 7.294). Rounded profiles are used exclusively these days.

Loadin

Toothed gear

Loadin

Power Grip

®

HTD toothed belt

g

oth

Belt to

Split 9.525 mm C-Tooth

Split 9.525 mm B-Tooth

1.90

2.30

PowerGrip ® Power Function Profile

Split 9.525 mm CF-Tooth

Split 9.525 mm BF-Tooth

2.20

2.80

PowerGrip ® HTD Profile Split 9.525 mm

Split 8.00 mm 3.20

3.60

PowerGrip

®

HTD 2 Profile

Split 9.525 mm

Split 8.00 mm 3.10

3.50

Figure 7.295 Tooth profiles. g

Belt tooth

® ® Power Power Grip Grip Trapezoid Trapezzahnriemen toothed belt

PowerGrip ® Trapezoid profile

Toothed gear

Figure 7.294 Development of synchronous belt profiles.

In the first generation of synchronous timing belts—with the trapezoidal teeth—there were two different tooth shapes, the smaller “C tooth” for gasoline engines and the larger “B tooth” for diesel engines, each with a pitch of 9.525 mm (Figure 7.295). This differentiation is no longer made in the newly-developed HTD tooth profiles. When the HTD profile was introduced to the market, it was necessary to take into account the fact that some car makers continued to use the existing trapezoidal tooth sprockets. To suit these applications, the profiles were optimized in regard to the radius at the root, flank shape, and tooth height (power function profile) so that they could be used with the existing trapezoidal sprockets. The associated sprockets, type ZA (C or CF tooth) and type B (B or BF tooth) are defined in ISO 9011.

HTD stands for the “high torque drive,” which was developed and patented by Gates. This curvilinear profile represented a considerable improvement in noise reduction, in power transmission, and, in turn, in terms of service life. With the introduction of the succeeding HTD 2 generation, the existing advantages of HTD profiles were further enhanced. Here the radii at the root and the flank angles are once again enlarged. Unique sprocket profiles are used for both types of profiles. The exact data for the profiles are available from Gates. Two pitch values are used for the two profiles:9.525 mm and 8.00 mm. The smaller pitch has benefits in regard to noise and, because of the smaller sprocket diameter, permits a more compact design. Both of the above-mentioned profiles can also be used in a double-sided synchronous belt (Figure 7.296). Double-sided synchronous belts are used, for example, to drive balancer shafts. Twin Power(R) Profile

Figure 7.296 Double-sided synchronous belt.

244 | Internal Combustion Engine Handbook

6606_Book.indb 244

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

B

D

PLD/2

W

7.18 Belt Drives

P

P D W B PLD/2

= = = = =

PD = Split × Number of teeth

OD = PD – PLD

Outer diameter OD

Effective diameter PD

PLD/2

Characteristic values—synchronous belt and sprocket The most important values for the synchronous belt are shown in Figure 7.297. The height of the tooth and depth of the backing together give the overall thickness of the belt. The pitch line distance (PLD—the distance from the root of the tooth to the center of the tensile member) depends on the belt design, the thickness of the fabric, and the diameter of the tensile cords. The width of the synchronous belt is selected in accordance with the alternating dynamic loading; in internal combustion engines, it normally lies between 15 and 25 mm and, in isolated applications, is as much as 30 mm. The profile of the sprocket has to be selected to match the diameter. The effective diameter is the product of the number of teeth and the pitch; the outside diameter of the sprocket is reduced by a value corresponding to the pitch line distance (Figure 7.298).

PLD/2 = Pitch line distance (half distance of the outer diameter to the pitch line)

Figure 7.298 Sprocket characteristic values.

7.18.1.2 Synchronous Belt Drive System The most important demand placed on the toothed belt system is synchronizing the camshaft over the engine’s entire lifetime. This is an important criterion for maintaining emission values even after extended periods in service. With proper selection of the materials for the belt, the use of an automatic tensioning system, and the use of optimized system dynamics, stretch in the synchronous belt can be kept to less than 0.1% of belt length. In four-cylinder engines, this represents a timing deviation of from 1 to 1.5 crankshaft degrees. This deviation is usually low enough that it does not require correction.

Split Depth of tooth Depth of backing Width Pitch line distance

Figure 7.297 Synchronous belt characteristic values.

The usual requirements in engine building continue to apply, regarding engine life (now 240,000 to 300,000 km), temperatures of approximately 120°C, the smallest possible build size, and minimum weight. Bothersome noises generated by the belt drive are not acceptable.  esign criteria D Complex synchronous drives are engineered with computer support using in-house created programs. A survey of the most important parameters considered in the design and some general design criteria is provided here. Important input data include the arrangement of the components, that is, the drive configuration, torque development at the components, and the dynamic circumferential forces calculated from them, along with the data for the belt itself. With these data at hand, it is possible to calculate and optimize not only the system geometry such as span lengths and wrap angles, but also the belt’s lifetime in reference to various failure modes. The dynamic forces and oscillations are used to calculate the other components in the system in the same way, such as the design of the reversing pulleys and the idler pulleys. Given below are a few general design criteria that must be observed in synchronous belt systems to engineer a functional system that will achieve the 240,000 km lifetime required today:  ecommended minimum wrap angle R Crankshaft150° Crankshaft/injection pump Auxiliary unit sprocket Tensioning pulley (smooth or toothed) min. 30° and better Deflection pulley (smooth or toothed)

100° 90° >70° 30°

 eriodic tooth engagement P Periodic tooth engagement means that a given tooth always engages with the same sprocket groove. This is to be avoided so as to preclude irregular belt wear and the belt damage that it may cause. The appearance of periodicity is calculated as follows: X.nnn = Number of teeth at the belt ÷ Number of teeth at the sprocket Here the following values for X.nnn are to be avoided: X.nnn = X.0, X.5 (must in all cases be avoided) X.nnn = X.25, X.333, X.666, X.75 (ought to be avoided)

Internal Combustion Engine Handbook | 245

6606_Book.indb 245

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

 inimum diameters for sprockets and deflector pulleys M 17 teeth Pitch 9.525 mm Pitch 8.00 mm

18 teeth

Smooth deflector pulleys

Æ 50 mm

crankshaft sprocket. In this case, the crankshaft damper often serves as the forward flange. The rear flange is attached to, or integrated into, the crankshaft sprocket. Additional flanges may be required in complex, multi-valve trains, depending on the number of sprockets and deflector pulleys. In these cases, it is advisable to locate the flanges at sprockets and not at deflector pulleys. In general, it is important to ensure that sprockets with flanges are aligned exactly with the other pulleys and sprockets to avoid deflecting the belt from its prescribed path. Sprockets and pulleys with just a single flange or without a flange are made wider than the belt itself to ensure that the belt runs steadily on the sprocket or pulley. The width of the sprockets and the geometric design of the axial guide flanges are depicted in Figure 7.299.

 olerances for sprockets and deflector pulleys T Run-out/Lateral run-out: ∅ 50 to 100 mm ± 0.1 mm ∅ > 100 mm ± 0.001 mm per mm ∅ Outside circumference conicity: ≦0.001 mm per mm pulley width

Parallel alignment of bore and toothing: ≦0.001 mm per mm pulley width Surface roughness: Ra ≦ 1.6 µm

Pitch error < 100 mm ∅ ± 0.03 mm groove/groove/ 0.10 mm through 90°

Belt tensioning systems

100 to 180 mm Æ ± 0.03 mm groove/groove/0.13 mm through 90° Axial guidance A synchronous belt has to be guided on at least one sprocket by flanges to keep the belt from wandering out of alignment. As a rule, guide flanges for the belt are located at the

 ixed tensioning pulleys F In the past, tensioning pulleys were always fixed. Eccentrically mounted deflection pulleys were most often used (Figure 7.300). Preload was set mechanically on the line and was checked with suitable measuring instruments (span frequency measurement). One disadvantage of fixed tensioning pulleys is the increase in tension that results from the greater expansion

8–25°°

bf

b f¢

Eccentric tensioning pulley

Mechanical compact tensioner

˘ Disk

b = Nominal belt width bf ¢¢= b + (1.5 x positive tolerance of belt width) bf ¢ = b + 1.75 mm bf = b + 3.50 mm

˘ Disk + 0.38

> 2.4

b f¢¢

>1.0

Figure 7.299 Sprocket width and belt guidance.

Hydraulic tensioner

Figure 7.300 Belt tensioning systems.

246 | Internal Combustion Engine Handbook

6606_Book.indb 246

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.18 Belt Drives

of the engine in comparison to belts when the engine heats up. Another problem is that they cannot compensate for the loss of belt tension through the service life because of stretch and wear.  utomatic tensioning pulleys A Because of the drawbacks associated with fixed tensioning pulleys and because of the increased dynamic forces in camshaft drives, accompanied at the same time by increased expectations regarding length of service, automatic tensioning idlers are used to a greater extent. This technology compensates both for the temperature-related rise in tension and for belt stretch. It also keeps constant the high tension required for dependable operation at high engine dynamics. The most widely-used is the mechanical, friction-damped, compact-design tensioner. Hydraulic tensioning pulleys are used in some applications where very high dynamic forces are found in the belt drive system. With their asymmetrical damping, they exhibit very good damping properties even at low preload values. 7.18.1.3 Synchronous Belt Dynamics Optimizing system dynamics is an important step along the road to synchronous belt drives lasting across the engine life because forces and loads can be minimized and at the same time monitored. Here, it is important to ensure that all the components in the system reach their targeted lifetime under these conditions. Dynamic loading on the drive, rotational oscillations, dynamic forces, and oscillations along the spans are optimized as a whole.

To do this, many parameters are optimized to minimize dynamic loading on the system. These parameters include the tensioner response characteristics, preload and damping, belt values, belt stiffness and damping, the belt profile, and the moments of inertia for the sprockets at the camshaft. Figure 7.301 shows two important values for the dynamics in the synchronous belt drive—the alternating load at the crankshaft and the rotary oscillations at the camshaft. System resonance, here at 4000 rpm if possible, is reduced to a minimum with optimized system design and has to be monitored over the service life of this drive. At the same time, the loads on other system components such as deflection and tensioning pulleys are also minimized. 7.18.1.4 Oval Sprocket Technology Today, belt drives for combustion engines with nonuniform transmission are state-of-the-art and relevant patent documents originated in the 1990s. Besides high-pressure injection pumps for diesel engines, the oscillating torques of the camshafts caused by valve actuation are, in medium and higher speed ranges in particular, the main exciting source for dynamic effects and thus primarily influence the service life and the function of such a system. The oval sprocket counteracts these dynamic influences, that is, they are mostly compensated by a load-synchronous change of partial belt lengths. The ovality and phase position, specifically selected for the application, of such a sprocket in a synchronous belt timing drive on the crankshaft for example, allows the minimization of the effects on the drive. The effect of a drive equipped with

Crankshaft alternating load

2000

[N]

1000

0

-1000 -2000

1000

2000

3000

4000

5000

6000

Rotational speed [1/min] Rotational oscillations NW

[°]

1,5 1

0,5

0

1000

2000

3000

4000

Rotational speed [1/min]

5000

6000 Figure 7.301 System resonance.

Internal Combustion Engine Handbook | 247

6606_Book.indb 247

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• constant engine performance at reduced exhaust gas emissions and less fuel consumption thanks to increased consistency of the timing over the lifespan

oval sprockets depends mostly on the change in dynamics over all the operating states; in an ideal case (constant exterior dynamics), the oscillation of the components can be fully eliminated while the dynamics of the belt forces are kept to a level matching the external dynamics of the components. Empirical experience shows that the belt forces can be reduced by up to 45% and the vibration amplitudes of the camshafts by up to 50%. Figure 7.302 and Figure 7.303 present a corresponding example.

•• reduced friction losses •• less noise development because of a reduced power level.

Figure 7.303 Reduction of vibration amplitudes of the camshaft, compared to a round crankshaft sprocket. See color section page 1079.

Combinations of non-round sprockets on crankshaft and camshaft for example, or on crankshaft and injection pump promise further optimization potential in respect to the advantages mentioned above.

Figure 7.302 Reduction of belt forces, compared to a round crankshaft sprocket. See color section page 1079.

7.18.1.5 Application Examples Depicted in Figure 7.304 are typical application examples for two engines. In both cases, the water pump is integrated into the drive system. In many diesel engines, the injection pumps (distributor injection pump or common-rail pump) are integrated into the primary belt drive.

The use of such systems results, among other things, in these advantages and design-related options: •• reduced belt width, use of a less expensive belt design and/ or increase in system lifespan

Camshafts

CM2

CM1

Camshafts Injection pump

CM1 IP

Tensioning pulley Water pump

IDR

+

Deflection pulley

+ Tensioning pulley

W_P Water pump

W_P CRK

CRK Crankshaft Inline SI engine

Crankshaft Inline diesel engine

Figure 7.304 Application examples.

248 | Internal Combustion Engine Handbook

6606_Book.indb 248

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.18 Belt Drives

7.18.2 Toothed V-Belt Drive to Power Auxiliary Units

Auxiliary units were driven in the past with simple V-belts. The increased complexity triggered by owners’ increased demands in terms of comfort, integrating the alternator, water pump, power steering pump, and air conditioning compressor into this drive system have made this feature state-of-the-art. The complexity of the drives is further increased by additional units such as the fan and mechanical turbochargers or pumps used for secondary air injection. Today the auxiliary units are driven in a serpentine configuration with multi-rib V-belts (Micro-V belts).The major benefits that the Micro-V belt offers when compared with V-belt drives are greater power transmission and reduced installation space in complex drives.

PK profile, 1270 mm reference length. When components that draw a great deal of power—such as the alternator, power steering pump, or air conditioning compressor—are driven with the back of the belt, the belt can also be designed as a double-sided Micro-V belt, with ribs on both sides (Figure 7.306).

Double-sided Micro-V ® belt

7.18.2.1 Drive Element, Micro-V belt  tructure of the Micro-V belt S The Micro-V belt is a bonded system made of three components (Figure 7.305): •• fiber-reinforced rubber blend •• tensile cords •• overcord or rubber backing. The tensile cords transmit drive power from the crankshaft to the auxiliary units, absorb dynamic loads at low stretch, and provide good resistance to alternating flexure. Micro-V ® belt

Figure 7.306 Double-sided Micro-V belt.

 haracteristic values of the Micro-V belt and sprocket C Figure 7.307 shows the most important values for the Micro-V belt. The belt width is calculated by multiplying the number of ribs by 3.56 mm (PK profile). Depending on the design, the belt height is between 4.3 and 5.3 mm. Width (Number of ribs x 3.56 mm) Height 4.3 - 5.3 mm, Designdependent

Cover fabric or rubber coating Tensile cords

Reference circumference 300 mm

Rubber blend Figure 7.305 Structure of the Micro-V belt.

The cords are made of nylon, polyester, or aramid; the widely differing moduli of elasticity for the tensile cords enable optimized tuning of system dynamics. The rubber forms the V ribs and transfers the drive forces from the pulley into the tensile cords. Chloroprene or EPDM is used as the material; fiber material is added to the rubber blend to stiffen the product. The overcord can either use a backing fabric or be made through rubberizing. During the manufacturing process, the cords in the tensile member are twisted, clockwise and counterclockwise in pairs, to achieve largely neutral running properties for the belt. The Micro-V belt is manufactured in a vulcanization process. The V ribs are either molded from the very outset or are cut into the belt after vulcanization. In double-sided belts, this grinding process is carried out on both sides.  icro-V belt profile M It is the PK profile (as per ISO standard) that is normally used for automotive applications. The groove spacing is 3.56 mm. The designation for the belt, such as 6 PK 1270, means six ribs,

Length measurement

Measurement force 100 N per rib

Figure 7.307 Characteristic values for the double-sided Micro-V belt.

The reference belt length is determined on a two-pulley test bed at a defined preload (ISO 2790). The reference circumference of the pulleys used here is 300 mm. The standardized profile used for the pulleys is shown in Figure 7.308. The outside diameter of the flanges is one dimension used to describe the pulley. More important for the design and determination of belt length, however, is the pulley diameter across the test balls (2.5 mm diam.). The pulleys are made of either steel or plastic. 7.18.2.2 Auxiliary Component Drive System The most important demand on any auxiliary unit drive system is slip-free drive for all auxiliary units, at all loading states, for the length of the engine’s working life. In modern engines with full drives, it is thus possible, using the Micro-V belts in a five- or six-rib design, to transfer maximum torques

Internal Combustion Engine Handbook | 249

6606_Book.indb 249

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

a

dB

dB

Effective diameter Outer diameter

A Do Do

Profile outer diameter

dB

Sphere diameter (2,5 mm)

DoB

DoB Diameter over sphere

a

Groove angle

A

Distance outer diameter to effective diameter

of up to 30 Nm and maximum power of from 15 to 20 kW with all the auxiliary units running at full load. The ambient temperatures at 80 to 100°C on average are somewhat lower than in a synchronous belt drive. It is important to avoid, in particular, noises such as the well-known belt squeal caused in cold and damp weather by slippage between the belt and the pulley. This is achieved with optimum system design in regard to the geometry and dynamics. It is also necessary to avoid belt noises caused by misalignment of pulleys, doing so right from the engineering stage. For auxiliary units, too, 240,000 km is taken today as the desired life expectancy in current engineering development work. Design criteria The auxiliary unit drives are engineered with computer support using in-house created programs. A survey of the most important parameters considered in the design and some general design criteria is provided here. Important input data include the arrangement of the components, that is, the drive configuration, torque development at the components and the moments of inertia of the components and the data for the belt itself. With these data at hand, it is possible to calculate and optimize not only the system geometry such as span lengths and wrap angles, the system’s natural frequencies, the slippage limit values, but also the belt’s lifetime. Given below are a few general design criteria that must be observed in Micro-V belt systems to engineer a functional system that will achieve the 240,000 km lifetime required today.  ecommended minimum wrap angle R Crankshaft 

150°

Alternator 

120°

Power steering pump, A/C compressor 

90°

Tensioning pulley 

60°

Figure 7.308 Characteristic values for the Micro-V pulley.

Minimum diameters for pulleys and deflector pulleys In practice, the smallest pulley is often found at the alternator, which is needed to achieve the high rotation speeds required there. Typical alternator pulleys have a diameter of from 50 to 56 mm. Belt fatigue rises exponentially when small pulleys are used; this has to be taken into account when engineering the belt. It is advisable to use diameters of no less than 70 mm for deflection pulleys.  elt tensioning systems B Belt tensioning in auxiliary unit drives is normally handled today with automatic tensioning pulleys. The tensioning pulleys ensure constant tension throughout the service life and compensate for belt stretch and belt wear. The design of the tensioning pulleys is determined essentially by the available installation area (Figure 7-309). In long-arm tensioners, the spring-and-damping system lies in the same plane as the belt drive system; where Z-type tensioners are used, the tensioner housing is recessed into the area behind the belt drive. Preload is generated by a leg spring; the tensioner is friction damped at the same time. The preloads for 6 PK belts normally lie in a range of from 250 to 400 N, the exact value depending on the system’s dynamics.

 lignment error/run-in angle A To avoid unacceptable belt wear and noise, the belt’s run-in angle into the grooved pulleys should not exceed 1°.

Figure 7.309 Automatic belt tensioning systems.

 ystem natural frequency S The system’s natural frequency should not be in the engine’s idle range (second engine order).

7.18.2.3 Application Examples Figure 7.310 shows a typical Micro-V belt drive. In many drive concepts, the power steering pump and the air conditioning

Long-arm tensioner

Z-type tensioner

250 | Internal Combustion Engine Handbook

6606_Book.indb 250

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

compressor have already been integrated into the standard belt drive design. Particularly when the drive configurations are complex, additional deflector pulleys are required to ensure the required wrap angle at all driven units and guarantee slip-free operation. Power steering pump

Alternator

Deflection pulley Water pump

Tensioning pulley

Climate compressor Crankshaft

Figure 7.310 Example of auxiliary component drive system.

7.18.2.4 Belt-Driven Starter–Alternator (RSG/Start–Stop System) An automatic start-stop system can reduce emissions but also fuel consumption and wear. Independent studies confirm fuel reduction between 4% and 25% for these systems, depending on the driving cycle. The starter-alternator revs the engine to a high crankshaft starting speed. This ensures a very fast, reduced noise and fuel-saving engine start. The belt-driven start-stop system also offer the option of using stored electrical power for acceleration and to convert braking energy to electric energy that is returned to the battery. A conventional starter can be usually eliminated with these systems, which has advantages relative to costs and weight. Modified starters are now offered for start-stop systems, however, they have no benefits in respect to noise or weight. A special high-performance V-ribbed belt and special belt tensioning systems had to be developed for RSG applications

(Figure 7.311). New requirements for the belt properties were specified because the combustion engine is started now using the starter-alternator integrated in the belt drive. Thanks to optimized adhesion of the belt components, the development of a rubber compound with higher load-bearing capacity and improved tensile cords, the new high-load Micro-V belt is capable of transferring moments of 70 Nm and more under all operating conditions and with a lifespan of up to one million startups. This capacity represents a breakthrough in belt technology and allows the inexpensive integration of a starter-alternator system without drastic changes to the auxiliary unit drive or the engine. RSG systems require the same amount of space as the conventional belt drives of modern engines while starter-alternators (KSG) installed on the crankshaft need space between the engine and the transmission. Furthermore, belt-driven starter-alternator systems have a lower weight and are significantly less expensive. For belt tensioning systems in start-stop systems, the functionality has been expanded to provide, even during engine start, sufficient belt force. After the engine has started, load and slack span alternate. Concepts for tensioning systems are in series production, which feature hydraulically-damped tensioners, doubletensioning systems or an asymmetrically-damped tensioner which is placed (by the alternator) in the tight span of the belt drive. The decision for the tensioning system depends mainly on the individual functional demands (such as start-stop, torque assistance) and the space available. Because of the low-noise starting properties and the reduced fuel consumption, the RSG system is very much accepted by the customers. The target market for RSG systems are all SI and diesel engines used in passenger cars and small utility vehicles. The belt drive system does not differ if the on-board power system is 12 or 42 V.

7.19 Bearings in Combustion Engines The shafts found in multi-cylinder reciprocating engines—the crankshaft, valve train, and balancers—generally run in plain

Figure 7.311 Belt tensioning systems for starter–alternator drives. See color section page 1079.

Internal Combustion Engine Handbook | 251

6606_Book.indb 251

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

(sliding or friction) bearings. The reasons for selecting this type include their great ability to withstand shock and their damping properties, easy division for assembly around the crankshaft or camshaft, low space requirements, insensitivity to grime, and, last but not least, the low costs incurred when compared with rolling bearings. The fundamental disadvantage of plain bearings compared to rolling bearings is the higher friction level and the resulting greater oil requirements. Rolling bearings are employed in engines in some cases wherever the advantages of the plain bearing are not fully exploited: At the crankshaft for small, single-cylinder engines, at the bearings for the sprocket drive, and, to an increasing extent, at the valve train (roller tappets).

7.19.1 Principles 7.19.1.1 Radial Bearings Constant loading The lubricant is drawn into a plain, radial bearing by adhesion, filling the lubrication gap between the surfaces that move relative onto the other; this causes a buildup of pressure that keeps external forces in balance and that keeps the mating components—journal and bearing—separated by an oil film (Figure 7.312). The dimensionless Sommerfeld number describes the interrelationships in a cylindrical radial bearing: p ⋅y 2 SoD = = f ( b/d, e ) . h⋅ w



(7.25)

angular velocity

ψ   –

relative bearing clearance, s/d

η N/m2 ⋅ s

dynamic viscosity

ε   – relative eccentricity of the journal centerline within the bearing clearance. Every load and velocity value corresponds to a certain eccentric equilibrium situation for the journal in the bearing:

ε = 0 → SoD = 0;  ε = 1 → SoD = ∞  ynamic loading D A characteristic feature for the bearings used in engines is loading, which alternates periodically in both magnitude and direction; this results, for example, from the ignition and inertial forces at the crankshaft and from the pulsating loads resulting from the camshaft’s actuating the valves. The change in force causes an imbalance that causes the shaft’s centerline to shift in the radial and circumferential directions. This eccentricity rises with rising loads; resistance to the displacement of the lubricant damps the radial motion. The high shock resistance of the plain bearing is the result. The resultant additional bearing capacity is defined by the Sommerfeld number for lubricant displacement: SoV =

p ⋅yn = f ( b/d, e ) . η ⋅ ( ∂e/∂t )

(7.26)

The overall force at the bearing results from vectoral addition of both effects (Figure 7.313).  riction F If a continuous and complete separation of the sliding surfaces were to be achieved by the oil film, then no bearing

The terms in the above formula are as follows: p N/m2

ω s–1

specific bearing loading F/(b ⋅ d)

b

DB

P

MB

dZ MZ

e

e0 =½ Spiel

MZ(MB bei n = ¥ )

Pmax

h0

n

Pmax

MZ

MZ bei n = 0 Gümbel circle

Figure 7.312 Hydrodynamic pressure buildup as a result of rotation.

252 | Internal Combustion Engine Handbook

6606_Book.indb 252

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

b

vL

v

W

h

MW

ML

e

d

dL

b

FV

dW

h0

FD

ax

m

pm

ax

pmax

pD

max

pV

Bearing width [mm] Bearing diameter [mm] Shaft diameter [mm] Excentricity Bearing load [N] Bearing load portions caused by rotation and squeeze effect [N] Lubricant film thickness [mm] h Minimum lubricant film h0 thickness [mm] pmax Maximum lubricant film pressure [MPa] pDmax Maximum lubricant film pVmax pressure caused by rotation and squeeze effect [MPa] Angular position of h 0, b relative to F0 d Angular postion of h 0 g Angular position of bearing load F vL Angular velocity of bearing Angular velocity of shaft vW dL – dW j Relative clearance dL

dL dW e F Fp, F V

(

)

Figure 7.313 Hydrodynamic pressure buildup as a result of rotation and displacement.

material would be required; the bearing would run entirely in accordance with hydrodynamic principles. Friction, in this case, is determined only by the oil’s shear strength and is very low, on an order of magnitude of μ = 0.002 to 0.005. In real-world operations, however, there is contact between the mating surfaces because the bearing cannot form a sufficient hydrodynamic lubricating film for every operational state. This “mixed friction” situation is associated with far greater friction levels, increasing by as much as a factor of ten. The familiar, generalized Stribeck curve describes the interactions (Figure 7.314).

Coefficient of friction µ

Mixed friction

Fluid or viscous friction Static friction Breakaway point (transition speed)

0

nü

Rotational speed n

Figure 7.314 Stribeck curve.

The system becomes thermally unstable if the friction energy thus generated cannot be dissipated. The probability that a thermally-unstable situation is reached in a plain bearing,

that is, the susceptibility of the bearing to malfunctions, is dependent on the energy density in the bearing system (load and velocity). Following dynamic loading, the shaft centerline periodically describes within the bearing a certain displacement path (see also Figure 7.317 below) with the smallest lubrication gap changing in size and location. The results are, on the one hand, that a far higher degree of direct material contact can be handled and that the dimensions of the bearing can be far smaller than one that is under constant loading; On the other hand, every area is subject to pulsation loading and the material’s endurance becomes an issue. 7.19.1.2 Axial Bearing Axial bearings are used to stabilize the shafts longitudinally and absorb the axial thrust generated by helical toothing and by any angular positioning. Higher loads may occur briefly, emanating from the clutch or resulting from shock triggered by acceleration. Axial bearings may be engineered as thrust washers or combined with a radial bearing to form a so-called locating bearing. These bearings are simple, plane surfaces made of bearing metal. They work in the mixed lubrication sector; that is, no hydrodynamic pressure is established. It is important that the wetting of the surfaces with lubricating oil is assured. Overheating is generally the reason for axial bearing failure; failure resulting from overloading because of shock or vibration is unlikely.

Internal Combustion Engine Handbook | 253

6606_Book.indb 253

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.19.2 Calculating and Dimensioning Engine Bearings

Z

A bearing is dimensioned in several steps during engine design work. The major dimensions, diameter and width, are determined primarily on the basis of the design parameters for the engine and mating components. Once bearing loading has been calculated, it becomes possible, during the concept phase, to use specific bearing loading (F/b ⋅ d) as a rough reference value. Because of the great influence exerted by load characteristics, the ratio of width to diameter, bearing clearance, oil viscosity, and engineering details, exact calculations for bearing dimensioning have to be made as early as possible. The primary results of calculation work are the selection of the appropriate type of bearing for the application and establishing the bearing dimensions, in conjunction with the acceptable boundary values.

0

180 90 450 0,75

lateral force of piston

Stroke

cylinder pressure

0,25

0,25

0,50

0,75

360

7.19.2.1 Loading The loading on engine bearings changes cyclically. The forces effective at the crankshaft are illustrated by a representative example in Figure 7.315. These forces are made up of cylinder pressure and the reciprocating and rotating inertial forces. Figure 7.316 uses polar coordinates to show the progress of bearing forces—in both magnitude and vector—at the camshaft bearing in a diesel engine over a complete operating cycle and running at maximum torque. At higher speeds with lower loading, the peaks triggered by ignition decline and the ellipse representing inertial forces increases.

+pgas

0,50

630 270

540

Figure 7.316 Polar diagram of the forces on the conrod bearing of a diesel engine.

When designing the crankshaft system, the bearing loads are normally calculated with the stiffness and oscillation situation in the crankshaft, taking account of elastic deformations. Thus for the main bearings above all(statistically indeterminate bearing), one can ascertain more exactly the distribution of loading across the individual bearing points. Having calculated cyclical loading in this way, one may then calculate the

+pmasse

rod force

cylinder pressure inertia force of piston rotating inertia force of conrod

radial inertia force of conrod reaction of the conrod pin force

TDC force onto conrod pins

Force onto neighbouring bearing pins

rotating inertia force of crank throw

tangential force onto crankshaft

BDC

conrod force including translational inertia force of conrod

Figure 7.315 Forces effective at the crank gear of an engine.

254 | Internal Combustion Engine Handbook

6606_Book.indb 254

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

•• smallest lubrication gap

hydrodynamic pressures that are generated and the widths of lubrication gaps. The most commonly used method here is to calculate the bearing journal displacement path.

•• maximum lubricating film pressure. Additional information is obtained, and this includes the oil throughput rate, hydrodynamic friction, and the resultant oil heating. The period through which the smallest lubrication gap remains in a certain area provides information on the concentration of friction energy and thus on the amount of wear to be expected. The calculation of the displacement path is suitable particularly for parameter studies in an early stage of motor engineering, for example, to determine the ideal layout for the balancer in view of the crankshaft bearings and/or the influence of design parameters such as the ratio of width to diameter or bearing play. Calculations for loading and the displacement path are often integrated.

7.19.2.2 Bearing Journal Displacement Path The displacement path that the journal executes in each full cycle, shown in Figure 7.317, can be calculated using relatively simple means. The results are strongly influenced by the nature of the model (method after Holland-Lang or the mobility method after Booker), by the peripheral conditions for the pressure curve, and by the assumptions of oil viscosity. Thus, it is possible to compare the results delivered by different programs only if these assumptions are identical. The acceptable boundary values, determined by applying experience from practical operations and drawing on test results for the calculated data, apply only for comparable calculation models. The path is iterated across the full cycle through to convergence in steps of a few degrees of crankshaft angle. Calculations are carried out separately for each loading situation. As a rule, the values are ascertained for operation at nominal load and at maximum torque with low engine speed. The most important results from the calculation are as follows:

7.19.2.3 Elasto-Hydrodynamic Calculation Elasto-hydrodynamic calculation is a more precise method, one developed in recent years, to calculate engine bearings. Here the distribution of the lubrication film in the bearing is calculated locally, taking elastic deformations into account (Figure 7.318).

Displacement path Fixed shell

Fixed pin

S

Z 120 0 180 240 540 60 600

0

480 660

60

300

0,75 0,50 0,25

0,75 0,50 0,25

0,25 0,50 0,75

0,25 0,50 0,75

300 120 360

180

660

420 360 540

600 240

Figure 7.317 Displacement path of a bearing journal (seen relatively from bearing and journal).

Crush

Bearing width

420

480

Oil groove

Bearing circumference

Pin width

Oil drilling

Pin circumference

Figure 7.318 Housing and bearing model (developed) used to calculate the elasto-hydrodynamic lubricating film.

Internal Combustion Engine Handbook | 255

6606_Book.indb 255

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

With the numerical solution for the Reynolds differential equation, one can take into account the stiffness of the bearing’s environment and the influence of local geometry characteristics for the bearing and the journal. In addition, this calculation method makes it possible to examine and verify values, such as the degree of fill at the gap, which are taken as givens in the global observation. Figure 7.319 shows the results for a conrod bearing with a slightly asymmetrical impingement of load, shortly after ignition. This method requires far more detailed data and significantly greater calculation effort than calculating the displacement path. Consequently, it makes sense to employ this at an advanced stage in engineering and to examine local influences. An estimate of the service life supported by cumulative damage models may follow if the load population and the necessary materials data are known. As a rule, service life and operational reliability are verified today by field testing and accompanying component testing. 7.19.2.4 Main Dimensions: Diameter and Width The bearing diameter and width are defined within narrow limits by the engine design and the dynamic forces at the shafts. It is possible to influence specific bearing loading within these limits, and this may be decisive for the selection of the bearing design. The ratio of width to diameter is usually from 0.25 to 0.35. At the same specific loading F/(W * D) a relatively smaller diameter and greater width causes a larger lubrication gap, lower peak pressures, and smaller friction losses. Because of the low circumferential velocity, sensitivity to contact with foreign objects, and malfunction falls, this situation is one that should be targeted. The minimum journal diameter required for sufficient crankshaft stiffness imposes limits on optimization. 7.19.2.5 Oil Feed Geometry Additional information on the distribution of lubricating and cooling oil is provided in Chapter 9. Here, only the features that affect the bearing directly are described. The establishment of the hydrodynamic lubrication film is greatly influenced by the grooves and bores required for lubricating oil supply—at the main bearings, for example. An annular groove in the main bearings is ideal for continuous

m

cu

c ir

g

Be ar in

Be ar in

g

cir

cu

m fe re nc e

fe re nc e

[d eg ]

[d eg ]

27 27 27 24 0 24 0 24 0 21 0 21 0 21 0 0 18 18 0 18 0 15 0 15 0 15 0 12 0 12 0 12 0 90 0 90 0 90 0 1e+000 60 60 60 1200 1e+001 3 30 30 0 600 0 0 0 1e+002 1 –3 –3 –3 0 0 0 0 0 –6 –6 –6 Bea 22 0 Bea 22 0 Bea 22 0 – 0 ring –9 0 –9 0 ring ring 0 wid –22 90 wid –22 0 –22 t w h th [ idth [mm mm [mm ] ] ]

Figure 7.319 Results of the elastohydrodynamic calculation.

supply to the conrod bearings; this does, however, in otherwise identical conditions, reduce the smallest lubrication gap to approximately 30%. This is compensated, in part, by better oil delivery to the bearing point so that the load-bearing capacity declines to only about one-half. Thus, one must attempt to achieve sufficient oil supply with bores and grooved sections in those areas around the bearing that exhibit lower loading and large lubricating gap widths. The displacement path described above provides information on the most favorable location for grooves (bearings) and bores (shafts). In passenger car engines, a semicircular groove in the upper shell of the main bearing and a bore in the crankshaft, exiting at the conrod journal approximately 45° before the apex in the direction of rotation, have been established as standard. To avoid inconstancies in oil flow, which could cause oil starvation and cavitation, it is often necessary to eliminate coarse inconstancies in the oil feed geometry. This is done by rounding bores and with tapered run-out at the grooves. When planning the lubricating oil supply, it is necessary to pay attention not only to sufficient delivery but also to adequate cross sections in drain channels. This is true particularly for thrust bearings where continuous, radial grooves in the running surface ensure both wetting of the axial bearing surface and a slight restriction of flow out of the radial bearing. Grooves are often required in the bearing housing for oil distribution, and here it is important that there be no hollows behind the bearing shells in zones that are subjected to loading because the shells could bend in response to lubricating film pressure and breaks in the bearing metal could occur. 7.19.2.6 Precision Dimensions The actual bearing engineering work concentrates not only on the selection of the bearing type, but also on precision dimensioning: •• tight seating, overhang •• bearing clearance •• progress of bearing thickness around the circumference, gap at the bearing shell ends •• surface properties, shapes, and positioning tolerances at the ends.

256 | Internal Combustion Engine Handbook

6606_Book.indb 256

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

SN pS

SN

pS

W

d

sL

sL

+d sL

pr s G pr ·m t pr

pr pS Sn W

radial pressure [MPa] bolt force [N] crush height [mm] bearing wall thickness [mm] d angel of section cutted out m fricton coefficient sL circumferential pressure in the bearing [MPa] sG circumferential pressure in the housing [MPa] t tangential stress caused by friction [MPa]

 ight seating, overhang T The bearing force has to be transferred to the housing. To do so, it is necessary that the bearing be seated firmly in the housing, to reliably suppress the relative motions resulting from the pulsating load. This tight seating in the radial bearing is affected by the excess diameter and the so-called “crush height,” Sn, at the ends of the bearing shells (Figure 7.320). The tabs or pins normally used to ensure correct positioning of the bearing are not suitable for fixing the bearing. The limits are, on the one hand, sufficient radial pressure (see Figure 7.320) and, on the other hand, tangential strain that can be tolerated by the bearing shell without great plastic deformation. All standard bearing metals are overtaxed at the low bearing thicknesses prevailing today; a steel backing shell is required to provide sufficiently tight seating. An engine bearing thus is made of a composite material incorporating steel and the actual bearing metal, with or without any additional coating, depending on the composition and particulars of the employment. Only in isolated cases, such as in large wrist pin bushings, can a single, solid material be used. The low-alloy steels required in bearing manufacture have a maximum compression yield point of 360 N/mm 2; this dictates a lower limit for bearing thickness, at approximately 2.5% of the diameter. Of particular significance is temperature development in aluminum housings. Because of the differing degrees of thermal expansion for steel and aluminum, there is a reduction as temperatures rise; this can even lead to a loss of the preload. At lower outside temperatures, by contrast, the strength limits for the bearing shell and/or the bore may be exceeded. In such cases the area immediately around the bearing is stiffened by sintered or cast steel components that are cast in situ. The global models for press fit calculations are no longer adequate when engineering bearing locations for composite designs; local strains and deformations have to be ascertained using finite element methods.

Figure 7.320 Excess diameter and installation strains.

 earing play B Bearing play is the most important, user-definable magnitude in bearing design. A smaller amount of clearance, nominally, creates greater hydrodynamic load-bearing capacity and—because of the greater damping to counter displacement—better acoustic conditions. In contrast, at larger bearing play, the lubricating oil throughput rises excessively (more than with the square of the clearance); the bearing becomes more tolerant of deformations and disturbances. Thus, one sets the value for minimum play to be as small as possible while still ensuring operational reliability. The maximum play results from the manufacturing tolerances for bearing wall thickness (6–12 μ m) and for the adjoining components and can become unacceptably high for small engines where D < 60 mm. Classification of bearing thickness is often a more favorable method than more exact manufacturing to limit the tolerance in play. As for the press fit, mastering bearing play for aluminum housings is difficult. Across the operating temperature range there is an unacceptable amount of change from, for instance, 15 μ m at 30°C to as much as 120 μ m at 130°C (at 50 mm diameter). Limiting maximum play requires more exact classification in which the bore, shaft, and play are associated one with another.  all thickness, clearance at ends W An undisturbed, perfectly cylindrical bore is ideal for the bearing function. The strains resulting from bearing installation and inertial forces usually, however, result in a bore that is not a true circle; this is compensated by a continuous change in bearing shell thickness, from the center to the ends. When bearings are split into two semicircular shells, a gap some millimeters in length and approximately 5–15 μ m in depth equalizes the differing thicknesses for the shells. Figure 7.321 shows typical values. Also essential to uninterrupted bearing function is the correct design of the bore and journals in regard to alignment,

Internal Combustion Engine Handbook | 257

6606_Book.indb 257

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Bearing Location

Operating Conditions

Design Parameters

Movement Type

Load Type

U m/s

Pmax N/mm2

ψ min %

B/D ––

Wristpin bushing

slewing

Pulsating load from the cylinder pressure, oscillating masses

 2–3

70–130

0.8

<1.0

Conrod bearing

Nonuniform rotating, ~ n

Pulsating load from the wristpin force and rotating masses

10–20

50–100

0.5

0.28–0.35

10

Crankshaft bearing

rotating, n

Pulsating load from adjacent conrod bearings

12–22

40–75

0.8

0.25–0.32

 8

Axial bearing

sliding

Thrust, coupling force, impact load

15–24

<2 permanent <5 brief <12 impact

––

–––

––

Rocker arm bearing

slewing, >0

pulsating

60–90

0.7

0.5–0.8

9

Camshaft bearing

rotating, n/2

pulsating

20–50

Mass balancing

rotating, 2n

planetary

20–40

1.2

0.3–0.4

Gearing drive, sprockets, auxiliary assemblies

rotating

uniform

pr min N/mm2

Crankshaft drive:  9

Valve train: 8 >10

dictated by the engineering design

Figure 7.321 Characteristic vales and typical reference values for the most important bearing locations.

roundness, crowning, waviness, and surface roughness. Reference is made here to the applicable engineering guidelines and standards. Acceptable boundary values are applied when selecting the type of bearing, which is done in consideration of loading and other peripheral conditions. The loading limits and the characteristics for use in the normal bearing designs are described in greater detail at Section 7.19.4. It is important that simultaneous development be carried out by the bearing maker right from the engine’s draft design stage.

one may categorize those that are most important for use in internal combustion engines in three groups of bearing metals and two groups of overlays (Figure 7.322).

Bearing metals

Anti-friction metal, cast (DIN-ISO 4381, SAE 12 - 17) PbSb14Sn9Cu SnSb8Cu4, SnSb12Cu5

Aluminum alloys, Roll-bonded

7.19.3 Bearing Materials

In addition to its primary function, which is transferring load during relative motions, the bearing has the additional important task of concentrating any disturbances in the system upon itself and to protect adjacent components such as the crankshaft for as long as possible. Because of increasing loads, increase of the mixed friction portion during the cycle and the legally-restricted use of materials such as lead, this procedure becomes more and more difficult and requires new approaches. With the familiar standard materials comprising a harder matrix with good thermal conductivity (CuSn and AlCu, for example) with inclusions of soft and immiscible phases (mainly Pb and Sn), good emergency operating behavior can be attained by lubrication from the components with low melting points. In the newer and more homogeneous materials, damages will arise only from major faults because of special, tribologically active alloy elements. Thin run-in layers support the bearing run-in by quickly adapting in the run-in and preventing seizing during engine start. Every good bearing material is a compromise between the contradictory requirements for strength and good tribologic properties. The best composition takes account of the weighting for the particular application. In spite of the multitude of different but, in some cases, very similar materials made by various bearing manufacturers,

(SAE 770 - 788)

AlSn40Cu, AlSn25CuMn, AISn20Cu, AlSn6Cu AlSn12Si4, AlSn10NiMn AlZn4,5SiPb

Leaded bronze, cast, sintered (DIN 1716; DIN-ISO 4382,4383; SAE 790 -798) CuPb30 CuPb25Sn4, CuPb20Sn2 CuPb15Sn7, CuPb10Sn10

Lead-free copper materials, cast CuSn 5 zu 1, CuZn20, CuAl8, ... Layers

Anti-friction metal, Galvanically (SAE 19, ...) PbSn8, PbSn10Cu2, PbSn16Cu3, Pbln9 SuSb12Cu, SuSb7, SuCu4

Synthetic layers PAlMoS2C

Al alloy, Sputtered AlSn20Cu Figure 7.322 Main bearing alloys for composite bearings.

258 | Internal Combustion Engine Handbook

6606_Book.indb 258

1/19/16 8:35 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

AIZn5SiPb Hardness HB 55–65

AIZn5SiPb Hardness HB 55–65

AIZn5SiPb Hardness HB 55–65

AIZn5SiPb Hardness HB 55–65

AIZn5SiPb Hardness HB 55–65

AlSn6Cu Hardness HB 36–45

AlSn20Cu Hardness HB 34–38

AlSn6Cu Hardness HB 36–45

AlSn20Cu Hardness HB 34–38

AlSn6Cu Hardness HB 36–45

AlSn20Cu Hardness HB 34–38

AlSn6Cu Hardness HB 36–45

AlSn20Cu Hardness HB 34–38

AlSn6Cu Hardness HB 36–45

AlSn20Cu Hardness HB 34–38

AlSn25CuMn Hardness HB 43–55

AlSn25CuMn Hardness HB 43–55

AlSn25CuMn Hardness HB 43–55

AlSn25CuMn Hardness HB 43–55

AlSn25CuMn Hardness HB 43–55

Figure 7.323 Comparison of micro-structures for Al bearing alloys.

The exact definitions, tolerances for the material composition, and mechanical properties are listed in [7-158] and in the above-mentioned standards. 7.19.3.1 Bearing Metals  hite alloys W Steel and white-alloy metals are now found only (on rare occasions) in passenger car engine designs, in bearings that are subjected only to low loads (camshaft bearings, sprockets). The SnSb8Cu and PbSn8 alloys have superb running properties, but their long-term strength is insufficient to handle the pulsating loads occurring in the drive train in modern engines. The composite material incorporating steel is manufactured in stationary sand casting or centrifugal casting for thickwalled bearings and in strip casting for thin-walled bearings of smaller dimensions.  luminum alloys (Figure 7.323) A Alloys based on aluminum have proven their effectiveness as main bearings and camshaft bearings across a broad range of applications.When used as a two-material bearing without an overlay, they represent a very economical solution for moderate loads; as a three-material bearing and grooved bearing, they are in direct competition with leaded bronze compounds. Aluminum alloys are not suited, as per today’s standards, for heavily loaded bushings where a slewing motion is encountered, for example, in the small conrod end and the rocker arm; neither do they provide a satisfactory basis for sputter bearings. AlSn alloys are the ones most frequently used. Upward of approximately 15% tin content, these alloys exhibit good slip characteristics; their excellent corrosion resistance makes it possible above all to use them in gas-fired engines and large four-cycle engines fired with heavy oil. Both AlSiSn materials and AlPb alloys are used in the English-speaking regions and in Japan. AlZn4 and 5SiPb are used when dealing with heavy loads such as those found in conrod bearings. This material does not have an embedded soft phase and thus is suitable for use as the substrate for three-material or grooved bearings only when an overlay is applied. The manufacture of aluminum bearing alloys is affected in a continuous or semi-continuous casting process; the process

windows are limited by the formation of separations (liquidation) in the soft phase and by the appearance of fissures in the hard phase. The stronger the matrix and the higher the tin content, the narrower the processing window. The method used most widely today is horizontal extrusion casting, which is non-critical for AlSn materials but which, however, cannot produce any higher-strength micro-structures. A somewhat more homogeneous structure can be achieved with vertical extrusion casting although the process is more sensitive to interference because the cooling conditions are more difficult to control. Belt casting, the newest technological development, permits a broader bandwidth in the process and, beyond that, the combination of a high share of matrix-strengthening elements and higher soft phase content. Because the ingot referred to here—in contrast to the other two processes—is actually a belt that runs simultaneously, the chilling and solidification parameters are better tuned to suit the particular material composition. The use of this casting procedure enables bearing metals such as AlSn25CuMn mentioned above. After casting, the strips are rolled out in several steps and heat treated; AlSn alloys are then joined with a thin aluminum bonding layer and, depending on the thickness of the finished bearing, are wound into coils or stored as strips. The amalgamation with the steel is made by roll bonding, which is essentially a friction welding process (Figure 7.324). The surfaces of the two strips are cleaned and activated; they are heated and rolled together, then they are rolled down by 20%–35%. The finished strip is then coiled up again. In smaller batches, plating is more economical in strips that are several meters in length; the process is essentially the same. The newer AlSn alloys are also roll bonded with alloyed intermediate layers such as AlZn so that their higher strength can also be utilized in the composite.  opper alloys C The copper-based materials used for bearings are many and varied. CuPbSn type alloys are used almost exclusively for composite materials. Other alloys such as CuAl or CuZn are used as solid materials only in special cases. Leaded bronze comprises a fixed CuSn matrix in which the lead is embedded. From 1% to 10% tin and from 10% to 30% lead is alloyed in. The higher the tin content, the stronger

Internal Combustion Engine Handbook | 259

6606_Book.indb 259

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Unwinder

Dressing roller

Tape welding machine

Al-strip Unwinder

Tape welding machine Dressing roller

Per bath

Brushing machine

Plating rolling machine

Induction furnace

Shears

Belt-grinding machine

Winder

Per bath

Figure 7.324 Production of steel-aluminum composite material (from [7-158]).

St belt

the material. The higher the lead content, the better the slip properties. Two groups are formed: •• CuPb(18–23)Sn(1–3) for higher slip speeds as found in conrod bearings and main bearings •• CuPb(10–15)Sn(7–10) for rocking movements as found in rocker arms and wristpin bushings. Homogeneous materials such as brass or bronzes are more prevalent in the area of lead-free copper-based materials: •• CuSn-based bronzes with good deformation properties •• CuZn-based brass alloys with good corrosion resistance. In rotating applications, leaded bronzes are suitable only with an additional electroplated or sputtered overlay. Wristpin and rocker arm bushings may be used with or without an overlay, the choice depending on their size. A major disadvantage of leaded bronze is lead’s sensitivity to corrosive attack by sulfur and chlorine compounds.

Consequently, aluminum alloys are given preference when running with heavy oil and in gas-fired engines. The bronze/steel composite material is made by casting or sintering. Strip casting is a suitable process for composite material of up to approximately 6 mm thick; centrifugal casting is used for thicker bearings. In the strip casting process used most widely for passenger car bearings, the edges of the pretreated lead strip are bent upward and the molten metal is cast into the “trough” thus formed. After cooling, the surface is milled down and the edges are trimmed. Stretching the strip slightly during these last two steps ensures stable steel strength. An optional, subsequent rolling step boosts the strength of both the steel and the bronze in heavy-duty bearings (sputter bearings). The strip is coiled up again for intermediate storage (Figure 7.325).

Brushing machine

Unwinder

Winder

Roughing machine

Tape welding machine Profiling machine Annealing path Melting furnace Pouring box

Cooling path

+

Circular shears Finishing machine Rolling mill

Shears

+

Figure 7.325 Production of lead–bronze composite material in strip-casting process (from [7-158]).

260 | Internal Combustion Engine Handbook

6606_Book.indb 260

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

CuPb20Sn Strip casting

CuPb20Sn Centrifugal casting

When sintering, the sheet metal strip is pretreated, and then bronze powder is spread over it. The sintering process proper (sintering and rolling) is carried out in two steps to achieve a structure with only a very few, very small, pores. The micro-structures differ markedly (Figure 7.326) and the strength of cast bronze is, without having to take any additional steps, greater than that of sintered bronze. 7.19.3.2 Overlays Overlays have to be applied to all higher-strength bearing materials to achieve running properties of adequate quality and insensitivity to disturbances. Basically there are three fundamentally different types of coatings: •• white-alloy metals deposited electrochemically •• sprayed or printed polymer-based synthetic layers •• AlSn alloys applied with the PVD (physical vapor deposition or sputter) process. Surface modifications such as zinc phosphating are found in certain application niches but have not made a broad breakthrough. An intermediate layer is required to ensure good bonding with the substrate and/or to suppress diffusion effects; nickel or NiSn is normally used for this purpose. Nickel is not a material that offers good slip properties; consequently, the thickness of this layer should be considerably less than the surface roughness. Common are from 1 to 3 μm, as otherwise larger, contiguous Ni areas appear on the running surface, and the bearing responds aggressively to disturbances where the overlay is worn.  lectroplated overlays E These overlays are, from the alloy technology viewpoint, similar to the cast white-alloy metals but exhibit less hardness and a finer structure since they are precipitated out at temperatures below the melting point, sort of in a “frozen” state (see Figure 7.331 below, three-material bearing). They are very insensitive to mixed lubrication but also wear very quickly because of their low hardness level, from 14 to 22 HV. The most widespread is the PbSn(8–18)Cu(0–8) system, where the amount of tin reduces corrosion sensitivity and

CuPb20Sn sintered

Figure 7.326 Matrix of the CuPb20Sn2 alloy from various manufacturing processes.

the copper increases durability. Tin content in excess of 16% leads to faster diffusion and thus to long-term instability, while more than 6% copper can cause brittleness so that the strength-enhancing effect is negated. Sn-based overlays with Sb or Cu shares have become another standard. They also represent an approach for lead-free threematerial bearings. Bi overlays have also attained a certain importance in this context. These layers are applied in galvanic baths with the application of current. This is done in a four-stage process encompassing pretreatment, applying and activating the intermediate layer, precipitating the overlay, and using subsequent heat treatment to stabilize the structure and to induce a sufficient diffusion bond. Overlay thickness is limited for several reasons: •• Durability drops rapidly with increasing thickness. •• The geometry of the lubrication gap must not change unacceptably as a result of wear. •• A concentration of electrical voltages causes the layers to be thicker at corners and edges. For economic reasons, also, mass electroplating is to be targeted if at all possible. As a rule, overlays of from 15 to 35 μ m thick are applied in mass electroplating processes; where thicker layers are necessary—in large bearings, for instance—they have to be reworked retroactively.  ynthetic layers S They are further development of solid lubricants on the basis of new polymer groups with the necessary aging stability. The overlay is applied with spraying or printing techniques and attains its strength from the polymerization process during a controlled heat treatment. The tribologically active filler materials are mostly graphite and MoS2. The layers have effects because of the reduction in friction with a defined wear in the event of contact with solid objects, thus significantly reducing energy input and preventing bearing failure. A cross section of such a layer is displayed in Figure 7.327. These layers are also used as run-in layers with thicknesses between 6 and 10 μ m.

Internal Combustion Engine Handbook | 261

6606_Book.indb 261

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

AlSn20 Extrusion casting, plated

Figure 7.327 SYNTHEK cross section.

 puttered overlay S A development that has made considerable advances in mass production only in the last decade is the use of the sputter process to deposit AlSn layers on plain bearings. Sputtering (cathode ionization) is a coating process in which a working gas (argon) is ionized in a high vacuum. An electrical field accelerates the ions to the cathode, the “target,” and atoms are dislodged from the target by the impact of the atoms. These atoms condense on the bearing running surface and form the slip-promoting film (Figure 7.328). Atomic deposition creates a strong structure with an extremely fine distribution of the soft phase, which, in spite of high hardness at approximately 90 HB, gives good running properties (Figure 7.329).

Bearing shell Al

Bearing shell

Al

Plasma

e Al+

e–

Al

Target

Al Al Sn Al Al Al Sn Al

Target

Al Sn Al

Lattice atom

Figure 7.329 Comparison of structures for roll bonded and sputtered AlSn20 layers.

Today AlSn20Cu is used almost exclusively as the sputter layer for heavy-duty bearings, but the process is, in principle, quite flexible and enables deposition of a very much broader range of alloys than is possible with conventional electrochemical processes. The only major drawback is the high cost of the coating.

7.19.4 Bearing Versions—Structure, LoadHandling Capacity, Application

Al

Plasma

Sputtered AlSn20 alloy

Primary impact

Figure 7.328 Sputtering process, schematic.

A further advantage of the process is the increase in bonding strength achieved by pre-cleaning the substrate by “sputter etching” under vacuum, producing a highly-active surface.

For cost reasons, one strives to satisfy the requirements for the particular application with the simplest possible bearing design. But contradictory demands for strength, tight seating, and good running properties ultimately lead to a “division of labor” and to a multilayer structure for the bearing. The use properties of the bearings and, above all, their dynamic load-bearing capacities are influenced not only by the selection of materials but also with the structure and thickness of the layers and other engineering measures. Thus, there are, beyond classical multilayer concepts, newer types that optimize the bearing’s utility with a closely-defined sequence of layers and/or design of the running surface. The fundamental advantages and disadvantages have already been mentioned in the discussion of the materials. Figure 7.330 provides a survey of the types of bearings most commonly used for a particular application range. 7.19.4.1 Solid Bearings Solid material is used primarily in large, industrial engines in the form of hard bronzes for thick-walled bushings and AlSn6 for thrust washers (axial bearings). The advantage is simple manufacture and, in thrust rings, the additional option of enabling use at both ends with proper engineering. In passenger car engines, the slow-running camshafts are borne directly in the aluminum cylinder head. Although these

262 | Internal Combustion Engine Handbook

6606_Book.indb 262

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

Backing Shell

Bearing Metal

Running Layer

Maximal pacross

Main Application

Solid bearings

None

CuPb15Sn7 AlSn6

None

60

Wristpin bushings thrust washers, camshaft bearings

Two-material bearing

Steel

CuPb10Sn10 CuPb15Sn7 AlSn6 AlSn20 SnSb12Cu

None

120

Wristpin bushings, rocker arm bushings

45 40 20

Thrust washers, camshaft bearings main bearings, conrod bearings camshaft bearings

CuPb10Sn10 CuPb20Sn2

90 55

Large wristpin bushings Conrod bearings, main bearings

65 70 65 70 50

Conrod bearings

Three-material bearing

Steel

AlZn4,5

PbSnCu PbSn16Cu PbSn10 Cu PbSn10 SuSb7 ceramic SYNTHEC SnCu4 SYNTHEC PbSn16Cu2

Grooved bearings™

Steel

CuPb20Sn2 AlZn4,5

SnSb7 PbSn16Cu2

50 55

Main bearings, large engines Main bearings, conrod bearings

Sputter bearing

Steel

CuPb20Sn2 CuPb10Sn10 CuSn5Zn CzSn5Zn

AlSn20

>100

Conrod bearing

AlSn20 + SYNTHEC

>120 >130

CuZn5Zn

Figure 7.330 Most important bearing versions and their application.

alloys are not bearing metals, proper functioning is reliable because of the low energy density in the bearings. 7.19.4.2 Two-Material Bearing Here there are two essentially-different application areas: •• Suitable for use at wristpin and rocker arm bearings are rolled bushings made of leaded bronzes, because the high specific loading of up to 120 N/mm2 requires endurance strength and the disadvantage of low running capacity, because of the low sliding speed, is of little significance. When there is an insufficient oil supply, these bushings tend to liberate lead from the material and cause oil carbonization (“burn traces”). •• Bearings based on AlSn, because of their excellent ratio of performance to cost, are the preferred solution for moderately loaded applications involving rotational movement and thus primarily the main and conrod bearings in gasoline engines and industrial diesels Their wear is low but there are limits to their adaptability. The low wear also harbors a risk: The appearance of the bearings changes hardly at all.; consequently, evaluating their condition with a visual evaluation is difficult. This makes necessary adequate statistical confidence for service life during the testing stage. The continuously rising loads found in new developments and advances in engines have resulted in the development of two-material bearings using higher strength AlSn alloys. The above applies to them in principle, but these bearings, because of the smaller lubrication gap and the greater energy density, are at greater risk of friction and wear damage. The increase of strength achieved by reducing the tin content is thus not an option, which helps meet development objectives. And the bonding layer made of pure aluminum can become a weak point.

7.19.4.3 Three-Material Bearing Three-material bearings with an overlay applied by electroplating and, above all, on a leaded bronze basis are the types used predominantly for crankshaft bearings. They represent a fully mature technology, are available worldwide, and offer a good ratio of cost-to-benefit. They are distinguished by good adaptability and are tolerant of grime and error for as long as the soft overlay is present. In larger engines, three-material bearings based on aluminum are also used. Three-material bearings are suited with some limitations where high loading situations are encountered, above all in the conrod bearings for modern direct-injection engines (both gasoline and diesel). Their weak point is faster wear at the overlay as loading increases. Corrosion resistance, too, which becomes more important at longer oil change intervals, is not high. Wear at the overlay, from 15 to 30 μ m thick, has, in and of itself, only an insignificant effect on the bearing function; exposure of the substrate, however, leads to a drastic increase in sensitivity to disturbances. The classical three-material bearing with a PbSnCu overlay is therefore supplanted to an even greater extent by higher-strength, two-material aluminum bearings in the lower load range and by the true high-performance concepts—grooved bearings for industrial engines and SYNTHEC or sputter bearings for passenger car and utility vehicle engines. 7.19.4.4 Miba Grooved Bearing The grooved bearings developed by Miba almost 20 years ago and shown in Figure 7.332 delay the degradation of the running layer with a special geometry for the surface. The overlay is embedded in very fine grooves in the running direction; between them are lands made of the harder bearing material. The ratio of materials at the running surface is approximately 75% overlay to 25% bearing metal. With this geometry, it is

Internal Combustion Engine Handbook | 263

6606_Book.indb 263

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Sputter bearing Steel support cup Lead-bronze or high-strength aluminum alloy as intermediate layer or without intermediate layer Running surface: AISn20 Sputter Sputter surface: SYNTHETIC Run-in layer

CuSn5Zn Bronze

AlSn20

Grooved bearing Steel support cup Depending on the application, various aluminum or lead-bronze alloys are used for intermediate layers and various galvanic running surfaces.

CuPb20Sn

AlZn4,5

Three-material bearing Steel support cup Depending on the application, various aluminum or lead-bronze alloys are used for intermediate layers and various galvanic running surfaces; SYNTHETIC running surface offered as alternative.

CuPb20Sn

CuSn5Zn-Synthec

Two-material bearing Steel support cup Depending on the application, various aluminum or lead-bronze alloys are used.

CuPb20Sn

Figure 7.331 Material structure (examples).

AlSn20Cu

Running surface structure

Running layer approx. 75 %

200 µm

Al alloy approx. 25 % Nickel dam max. 5 %

View of running direction

Section vertical to the running direction

Figure 7.332 Miba grooved bearing.

264 | Internal Combustion Engine Handbook

6606_Book.indb 264

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

7.19.4.5 Sputter Bearing The bearings that can stand the most extreme loading and are produced in large numbers today are three-material, leaded bronze bearings with a sputter overlay. Because of their greater load-handling ability, up to more than 100 N/mm2, and with good running properties at the same time, they are installed in engines with high power density and used for passenger cars, utility vehicles and drives for fast ships. Guideline values for application limits and costs of bearing types are shown in Figure 7.333(a)–(e). Today, hardly any other type of bearing can be considered, particularly conrod bearings

possible to continue to determine the tribologic properties by selecting the overlay material but to protect that layer against wear with the harder lands. Thus the good running properties are retained for a much longer time than in threematerial bearings. The grooved bearing today finds its primary application in diesel engines with greater specific power and is used to drive locomotives and ships; in the passenger car and utility vehicle engine segment, it has been supplanted in recent years, to an increasing extent by the sputter bearing, because of the continuously-rising loads.

Permissible limit values from calculation Sputter bearing - lead-free with run-in layer

Smallest lubrication gap (µm)

Sputter bearing Synthetic bearing

Grooved bearing, Pb-based

Three-material bearing, Sn-based Three-material bearing, Pb-based Two-material bearing

Anti-friction metal bearing

Figure 7.333 Guideline values for bearing types. (a) Permissible limit values from calculation.

Lubricating film peak pressure (bar)

Bearing capacity in MPa

Load-bearing capacity limits (MPa) by types

Antifriction metal bearing

Twomaterial bearing

Threematerial bearing, Pb-based

Threematerial bearing, Sn-based

Grooved bearing

Synthetic bearing

Sputter bearing

Sputter bearing lead-free with run-in layer

Figure 7.333 (b) Load-bearing capacity limits (MPa) by types  (continues)

Internal Combustion Engine Handbook | 265

6606_Book.indb 265

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Peripheral speed in m/s

Maximum permissible peripheral speed (m/s)

Antifriction metal bearing

Twomaterial bearing

Threematerial bearing, Pb-based

Threematerial bearing, Sn-based

Grooved bearing

Synthetic bearing

Sputter bearing

Sputter bearing lead-free with run-in layer

Figure 7.333 (c) Maximum permissible peripheral speed (m/s)

Wear rate of major bearing designs (at identical load)

Antifriction metal bearing

Twomaterial bearing

Threematerial bearing, Pb-based

Threematerial bearing, Sn-based

Grooved bearing

Synthetic bearing

Sputter bearing

Sputter bearing lead-free with run-in layer

Figure 7.333 (d) Wear rate of major bearing designs (at identical load).

Emergency run capacity (without consideration of resistance)

Antifriction metal bearing

Twomaterial bearing

Threematerial bearing, Pb-based

Threematerial bearing, Sn-based

Grooved bearing

Synthetic bearing

Sputter bearing

Sputter bearing lead-free with run-in layer

Figure 7.333 (e) Emergency run capacity (without consideration of resistance.

266 | Internal Combustion Engine Handbook

6606_Book.indb 266

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.19 Bearings in Combustion Engines

in direct-injection diesel engines for passenger cars. Some of them may be given a SYNTHEC run-in layer for bearings in utility vehicles with extreme loads and sputter bearings with larger bearing diameters. The only major drawback to the sputter bearing is its price. Because of the complex vacuum coating process, a sputter bearing is five to eight times as expensive as a three-material bearing. Thus, in the conrod and main bearings, a sputter shell on the side subjected to heavy loads is combined with a three-material or grooved bearing shell on the side with less loading. This combination offers the additional advantage that tiny particles of grime cannot become embedded in the soft overlay. The application limits and costs for the various bearing designs are shown in Figure 7.333.

7.19.5 Bearing Failure 7.19.5.1 Progress of Damage Bearing damage (Figure 7.334) in the narrower sense is always an interference in the geometry of the slip space to an extent that precludes stable operation of the bearing location. The results are great friction and, with such friction, local overheating and destruction of the bearing and adjoining components, right through to complete failure of the engine. In internal combustion engines where, in contrast to mechanical engineering bearings, the sizes of the loads and the direction change cyclically, the course of damage depends on the location, time, and loading level, and thus is to be seen statistically. This can cause total damage to one bearing while, in the adjacent bearing, almost no damage at all can be found.

Total damage

Disturbances can be covered for a period of time and even corrected if the cause is eliminated (e.g., excess temperature, oil shortage, etc.); an error in the geometry may be attenuated by wear or adaptation. Because of the serious subsequent damage that can follow bearing failure, even events that, seen on their own, do not cause a failure are deemed to be bearing damage. Events such as these are seen as early warning signals for impending bearing damage and thus are very important when diagnosing the status of the system. 7.19.5.2 Types of Bearing Damages DIN-ISO standard 7146 and trade literature published by plain bearing manufacturers describe the most frequent types of bearing damage; consequently, only a brief overview is given here. This depiction follows the organization used in DIN-ISO 7146. Damage to the running surface  oreign objects, grime F Foreign objects that are swept into the bearing with the lubricating oil continue to represent the most frequent cause of bearing failure, particularly in main bearings and in spite of great efforts to maintain cleanliness during assembly, as well as in operation. The problem found in such events, in addition to the permanent disturbance, which reduces service life, is the scoring or embedding process itself. While this is happening, extremely high friction is generated locally.  road-surface wear B At high loads, with the wrong oil (too thin) or the selection of an unsuitable bearing design, premature wear can appear in the zone where the narrowed lubrication gap is found. As a rule, wear is not a problem in normal operations; threematerial bearings do, however, become more susceptible to disturbances when the protective overlay is no longer present.  dge collar, local overloading, overheating E Deficiencies in the geometry, localized contact points because of elastic deformations and minor assembly errors can be attenuated by localized wear at the soft layer. This process, however, leads to an increased degree of mixed lubrication, corresponding to a local increase in temperature and, in an extreme situation, to instability and damage.

Figure 7.334 Total failure of a leaded bronze bearing.

 atigue fracture F The bearing material has to exhibit sufficient durability so that the pulsating loads can be reliably transferred throughout the required service life. If this is not the case, then fine fissures appear and later particles will spall off. The hazard potential represented by fatigue fracture is dependent on the thickness of the layer affected: Spalling at the running layers seldom leads directly to bearing failure. Fractures in the bearing material, about ten times as thick, have an enduring adverse effect on slip gap geometry.

Internal Combustion Engine Handbook | 267

6606_Book.indb 267

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

 avitation C Cavitation is the result of vapor bubbles in lubricating oil which arise when the lubrication oil pressure at some points falls below the vapor pressure. These bubbles implode when they again enter an area of higher pressure. The pressure surge thus created tears particles out of the bearing surface and, in serious cases, right through the bearing metal and into the steel in the backing shell. Cavitation is quite often a design problem (groove shape, bearing play, etc.). In addition to changes in the geometry of the oil flow, its prevalence can also be reduced by measures that raise the oil pressure in the system.  orrosion C Of the materials commonly used in bearing technology, the lead in the electroplated overlay and in the leaded bronze is most often affected by reactions with sulfur and chlorine. In those cases where corrosion is to be anticipated during operation, for example, where industrial engines are run on heavy oil or landfill gas, an increase in the tin contact in the CuPb materials or the use of AlSn instead of CuPbSn is necessary Damage at the back of the bearing I nsufficiently tight seating The second important functional surface of the radial bearing is the outside diameter. Sufficient friction is necessary to transfer the force. The tight seating of the bearing in the housing bore is achieved by sufficient overage of the diameters and bearing halves by excess length, the so-called “crush height.” Because of elastic deformations resulting from the operating forces, there is thrust loading at the interface between the bearing and the housing; insufficiently-tight seating can result in relative movements between the bearing and the housing. The consequences are material displacement, fretting corrosion, material transfer (pitting), and, in serious cases, shell movement. These relative movements can be suppressed by greater crush height. The limit is imposed by the tangential stresses in the steel shell, which may not exceed the creep limit. Increasing the operating speeds for existing engines thus frequently necessitates engineering modifications.  ssembly errors A In addition to the operating loads and geometric deficiencies, errors in assembling the bearings are often the reason behind serious bearing damage. Thus, bearings should be designed in such a way that incorrect positioning, interchanging and the like can be positively avoided.

7.19.6 Prospects

Rapid developments in engine technology, which are further accelerated with the introduction of direct-injection engines, are flanked by component development and, in some cases, made possible by such developments. The major driving forces behind new developments in bearings include

•• loading capacity (higher ignition pressures, mean pressures, service periods) •• costs (heavy-duty, multilayer bearings are expensive) •• Environmental aspects (lead, cleaning, manufacturing processes). Even if all the requirements are covered today from the technical side, there are combinations of loading capacity, running properties, and manufacturing costs that are not ideal. The goals of new developments are above all an improvement in the ratio of costs-to-utility. For economic reasons the use of bearings without an overlay is also targeted at higher loading levels so that most material developments in recent years are aimed at increasing strength with the least possible limitations in regard to running capacities. Because of the changes in technology, developments are moving in very different directions: •• Improving the load-bearing capacities of two-material bearings so that they can replace three-material bearings in some applications. This is done by developing new aluminum alloys in conjunction with advanced casting technology. Newer developments are, for example, AlSn10NiMn, AlSn12Si4, and AlSn25CuCoZr—the first, because of the significant reduction of the tin content, is more suited for use in smaller (passenger car) engines at moderate loads. •• Increasing the wear resistance and durability of electroplated overlays by new material combinations within the Sn system; sometimes hardening by means of ceramic particles but also completely new systems on bismuth basis. •• New synthetic overlays based on polymers with embedded solid lubricants. •• New PVD layer combinations to increase temperature resistance properties. Several new developments of this type are close to introduction for mass production. They will supplant conventional bearings in areas relevant to the environment such as the passenger car industry and will provide alternatives for longer service life and higher loads, taking into account the tribologic necessities.

7.20 Intake Systems The air intake systems in modern internal combustion engines serve a number of functions in addition to routing and filtering the combustion air. The demands placed on intake systems continue to rise with increasing engine complexity. Two major trends are emerging.  ystem competence S The entire air routing configuration is seen as a system extending from the intake opening to the cylinder head; it is engineered and manufactured by the supplier and delivered ready for installation. This presumes that the supplier will fully understand the system, going beyond the air supply

268 | Internal Combustion Engine Handbook

6606_Book.indb 268

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.20 Intake Systems

External air pipe – Water/snow separator – Warm air addition

Air filter – Approach flow filter element – Adjustable intake cross-sections – Add-on components (ECU, actuators) Clean air pipe – Approach flow HFM – Turbocharger, mech. charger – Intercooling

Intake manifold – Supercharging – Introduction EGR/TE/blow-by – Charge movement – Vacuum receiver – Add-on components (throttle valve, fuel bar, actuators, sensors)

Figure 7.335 Air flow system for an internal combustion engine (schematic).

system proper and including the exhaust system in view of the exchange of gases in the cylinder.  odularization M A second trend is the increasing modularity of the intake system. A modular design makes good sense because the air supply system is spread out all around the engine and, simply because of its size, lends itself to the attachment of discrete components. These components are not necessary components of the air management system proper. One example is locating the motor management circuitry inside the air filter; passing air is used to cool the electronics. Modularization requires, in addition to an understanding of the system as a whole, increased competency in manufacturing and integration. Figure 7.335 schematically shows the air path in a fourcylinder engine with the main functions and some of the attached components. The thermodynamic situation along the air path is explained below along with the proximity to the fields of acoustics and filtration.

7.20.1 Thermodynamics in Air Intake Systems

The thermodynamics in the air supply system depend on the combustion process (gasoline or diesel) and on the charging principle. The external air piping and air filter are similar in all the variants. The systems differ markedly, however, downstream from the air filter, depending on the charging principle; see Figure 7.336. Turbocharged engines are fitted with an elaborate clean-air section with a compressor and after-cooler, while the intake manifold is simple in design, as it serves primarily to distribute air to the cylinders Naturally aspirated engines, by contrast, have a simple clean-air section but in most cases complex, active intake manifolds to improve cylinder fill.

External air

Air filter

Clean air

Intake manifold

Figure 7.336 Air routing in naturally aspirated engines (bottom) and turbocharged engines (top).

 xternal air pipe E The external air path, that is, the section of the intake system between the intake opening and the air filter, not only guides the air but also is used for the addition of warm air and the elimination of dirt. Blending in warm air influences the engine’s operating properties, particularly in the cold starting phase. This function wil grow in significance in the future as more stringent limit values are adopted in exhaust gas legislation. Drying the filter element and melting snow are further reasons for adding warm air. Fuel consumption can be favorably influenced by intelligent temperature regulation for the intake air. Warm air is drawn in through a second take-up point near the exhaust manifold; it is activated by flaps. The flaps are actuated by thermostat elements or by vacuum or electrical actuators. A suitable external air routing system also separates coarse particles (droplets, snow, dirt) with minimal pressure loss by diversions. This preliminary separation helps to keep down the amount of grime collected at the air filter and protects the filter element against moisture. Particle separation and

Internal Combustion Engine Handbook | 269

6606_Book.indb 269

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.337 Air flow in filter elements: uneven with high pressure drop (left)—close to optimal (right).

pressure loss at diversion points are determined in advance with the help of CFD (computational fluid dynamics). Air filter What might colloquially be referred to as the “air filter” or “air cleaner” comprises the filter element, the air cleaner body, and the cover. In addition to its acoustic effect, the body serves to optimize the airflow path and air distribution around the filter proper. The ideal here is the most uniform distribution possible. Air velocity perpendicular to the filter element has to be homogeneous over the entire filter surface area. When the influx of the air is not uniform, there is greater pressure loss at the filter element, and engine efficiency is degraded. The dirt and dust trapping capacity of the filter material is also optimized with homogeneous airflow. Three-dimensional CFD techniques are employed at a very early stage to engineer the airflow in air filter bodies (Figure 7.337). This makes it possible to determine the ideal geometry very early and with a minimum of effort involving physical testing. For example, in the system shown in Figure 7.337, the geometry could be reduced by 30% at identical pressure loss and identical dust trapping capacity. It was possible to attain air distribution that was within 3% of the ideal test bed value.  lean air pipe C The air impinging on the mass flow meter (MM) that measures the intake air on the clean-air side is analyzed for new intake systems, using CFD simulation, to achieve uniform flow. In view of more stringent emission limits, reliable functioning of this meter, in all operating states and for the life of the vehicle, is specified. Gradual degradation of the meter resulting from deposits on the sensor [oil droplets from the crankcase or from the exhaust gas return system (EGR)] can also be reduced dramatically by applying CFD simulation to the air flow path. The gas pulsations generated by the engine become more intensive downstream on the clean-air side. If thermodynamics and acoustics are not seen as a whole, then this has to be done at the latest in the clean-air runner since both disciplines

exert an effect on air routing. In the area associated with the clean-air runner, one finds acoustic components (shunt resonators, λ /4 tubes) that also have an influence on gas exchange in the cylinders. Today, simulation tools are used for such components. The airflow rate and noise at the inlet tube are thus calculated at a very early point in the design phase. The effort required for modeling can be significantly reduced since a single calculation model delivers both results. Supercharged and turbocharged engines have a longer airflow path than naturally aspirated engines. In engines with a turbocharger, the intake air passes from the forward module and through the air filter to the compressor located near the exhaust manifold. The compressed air is then returned to the forward module, where the after-cooler is located. Finally, the clean-air runner terminates at the intake manifold at the engine. I ntake manifolds Engines with mechanical superchargers or exhaust-driven turbochargers require intake manifolds and runners to distribute the combustion air to the cylinders. The aim here is to achieve short intake runners with little pressure loss and good uniformity of distribution to the cylinders. Naturally aspirated engines use the wave effects initiated by the piston to compress intake air. The procedure known as “resonance tube charging” is described in Figure 7.338. When the intake valve is opened, the piston, as it moves downward, creates a vacuum wave that moves opposite the direction of airflow, away from the combustion chamber and along the resonance tube. The vacuum wave is reflected at the collector, because of a change in cross section. The pressure wave moving back toward the combustion chamber can be utilized to improve cylinder filling, provided that it arrives before the intake valve closes. Ideal tube length, at constant speed of sound a, is inversely proportional to engine speed n. To achieve good cylinder fill over a broad engine speed range, all vehicle classes are seeing an increasing use of intake manifolds that can alternate between short and long resonance tubes. A typical response curve for an active intake manifold with two resonance tube lengths is shown in Figure 7.339.

270 | Internal Combustion Engine Handbook

6606_Book.indb 270

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.20 Intake Systems

Vacuum wave

Intake manifold

Reflection Inlet valve

Overpressure wave

Optimal pipe length [m] » 10 · OT

a [ms–1 ] n [min–1 ]

OT UT

Figure 7.338 Principle of ram tube charging.

300

Torque [Nm]

280 260 240 220 200 1000

Torque tube Power tube 2000

3000 4000 5000 Engine speed [1/min]

6000

Figure 7.339 Torque progression of a six-cylinder engine with controlled intake manifold.

7.20.2 Acoustics

Sound is understood to be mechanical oscillations and waves in an elastic medium. Chapter 7.20.1 used Figure 7.338 to illustrate how piston movement, after the intake valve was opened, triggered a vacuum wave moving against the gas flow

Gap increasing

Lines of identical sealing technology

Potential for increase in air expenditure

better

With increasing intake system complexity, the increase in the airflow rate depends more and more on the quality of the

manufacturing and materials. Figure 7.340 uses gasket quality to illustrate the sensitivity of airflow rate to leakage. A gasket that permits an increase in the airflow rate in a two-stage active manifold can, in three- and four-stage manifolds, quite conceivably lead to a reduction of the airflow rate. In addition to leakage through the switching elements, there is a series of other variables that influence the change of gases in the intake manifold. Figure 7.341 provides a survey of potential sources of loss. This makes it necessary for suppliers of modern intake systems to define the entire intake system in both thermodynamic and mechanical terms at a very early point in development work. This requires linking and networking all the CAE tools right from the outset of the development project.

rigid

two-stage

three-stage

stepless

Figure 7.340 Influence of sealing technology on airflow rate.

Internal Combustion Engine Handbook | 271

6606_Book.indb 271

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Switching elements tightness

Joint tolerances

Mat eri

tig ung

λa EGR introduction and EGR even distribution

Ge

Manufacturing tolerances

F

l

er

a

Wall thickness distribution, cross-section progression

Surface roughness

om

n

etr y / p a c k a g

e

Inflow collector

Resonance tube curvature and progression

Figure 7.341 Influencing factors on the airflow are in controlled intake manifolds.

direction. These pressure fluctuations are propagated as sound through the air filter intake opening (intake noise). Moreover, the pulsation inside the components induces wall vibration (structural noise), which is then again propagated as airborne noise. Those in the vicinity do not always perceive this sound as pleasant, which is why restrictions have been imposed; every vehicle must meet these limits (see also Chapter 27).  Legislation The noise problem of a motor vehicle can be divided in two areas: •• vehicle interior noise •• pass-by noise. While the reduction of interior noise goes hand in hand with passengers’ increasing comfort expectations and thus is a factor in the brand image, a legislated limit value applies to noises generated by vehicles accelerating past a given point. The procedure used to measure this value is illustrated in Figure 7.342. Since October 1, 1995, passenger cars may not exceed a value of 74 dB(A) in accelerated pass-by testing.

• • • •

Measurement acc. to ISO R362 Measured in 2nd/3rd gear Sound level limit 74 dB(A) Recorded in vehicle title

 oise generation N In an internal combustion engine, the pistons, with their reciprocal motion, create fluctuations in air pressure (air pulsations) and airborne noise results from them. The pistons thus act as an air-pulsation noise source. Disturbances in airflow along the intake system can also act as aerodynamic sources, contributing to intake noise. This noise is emitted primarily through the intake opening and thus passes directly into the environment. A second portion of the pulsation energy inside the intake system incites structural noise oscillations in the elastic structure. These are then transmitted from the exterior surfaces to the surrounding air or through attachment points to the body. This situation is illustrated schematically in Figure 7.343.

Microphone

7,5 m

Accelerated passing (passenger car)

The overall acoustic level is the sum of just a few individual noise sources.Regarding pass-by noise, these are engine noise, intake and exhaust noise, tire noise, and wind noise.

10 m

10 m

A Start of acceleration

B End of acceleration Microphone

Figure 7.342 Passenger car noise: EC Directive 92/97/EC.

272 | Internal Combustion Engine Handbook

6606_Book.indb 272

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.20 Intake Systems

Radiation

Muzzle noise

Resonator Clean air pipe

Air filter

throttle valve

Airflow meter

External air pipe Intake opening

Cylinders Solid-borne noise transmission

Sound excitation

Intake manifolds 1 2 3 4

Abstrahlung Collector

Effects on muffle noise Air inlet, external air pipe

 coustic elements of tube systems A A variety of acoustic principles can be used to attenuate intake noises; see Figure 7.345. The most important damper design is constructed in principle like a so-called series resonator. This is a system taking the form of a Helmholtz resonator, in which a damping chamber is connected to a section of tubing. A resonator such as this functions, in principle, like a sprung mass system in which the spring is represented by the compressible air in the chamber, the mass, on the other hand, by the air pulses in the pipe. Depending on its dimensions, a resonance frequency f0 can be calculated at which a resonator such as this amplifies the sound introduced. The following formula is used to calculate the frequency: f0 =

c Aw 2 ⋅ p lacoust ⋅V

(7.27)

where Aw is the mean cross section of the resonator throat lacoust the effective acoustic length of the throat and V the chamber volume. Conversely, frequencies are dampened from f0 ⋅ 2. This relationship must be utilized with the damping filter. To attain best possible damping, f 0 must be as low as possible, that is, far below the frequencies occurring during operation.

Effects on charge change

 ptimization measures O The objective of the measures undertaken to optimize intake noise is consistent acoustical development wherein the noise is to be reduced right from the draft stage. The work carried out to optimize noise is subdivided into primary efforts and secondary efforts. Primary efforts: These exert an influence on the sound source. The noise created by airborne sound means a reduction in the alternating pressures, while the noises excited by structural sound require a reduction in the exciting forces and a change in structural noise behavior and in propagation (admittance and degree of propagation). Secondary efforts: These retroactively reduce the airborne sound generated and reduce noise emissions with mufflers and/or encapsulation. The pulsed-air noise source in the intake system is the engine; any influence on this source often conflicts with the objectives set in the thermodynamics analysis. That is why one employs secondary measures such as damping filters and shunt resonators to reduce intake noise. The effect of acoustic corrections on the throat noise and on the gas exchange is shown in the example in Figure 7.344.

Figure 7.343 Noise sources of an intake system.

Air filter

Clean air pipe

Collector

Resonance tubes

Figure 7.344 Effect on throat noise and the gas exchange.

Internal Combustion Engine Handbook | 273

6606_Book.indb 273

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Type

Designs

Absorption damper Reflection damper Resonance damper Helmholtz resonator Branched Helmholtz resonator (λ/4-Pipe)

Application Broadband, suitable for medium and higher frequencies Relatively broadband, suitable for low and medium frequencies Narrow band, suitable for low and medium frequencies Narrow band, substantial damping, suitable for medium frequencies above Helmholtz resonance Narrow band, suitable for medium frequencies

Figure 7.345 Types of acoustic dampers and their application.

This can be achieved by increasing the air filter volume, reducing the intake cross section or extending the intake manifold length. Because of the usually limited space available, the housing volume cannot be increased at will. Even a very much reduced intake cross section has undesired side effects as it throttles the intake airflow. Increased pressure loss always means a loss of engine performance, which is why, in practice, the pressure loss in the intake tube is kept in bounds by designing a diffuser like intake opening, similar to a Venturi tube. Lengthening the intake tube also runs into system-imposed limits, while a corrective measure such as this also harbors the hazard of tube resonances that can counteract damping at certain frequencies. That is why exact tuning of the entire system is needed to identify the ideal compromise between expense and profitability.

In addition to the finite element method, 1-D calculation programs based on the transfer matrix method or the finite differences method have become established here. The latter offers the advantage that, in addition to the acoustic values, thermodynamic values can also be calculated. The calculation results can be validated at simple component test beds as soon as initial samples are available. Final optimization with parts close to the mass production components is then carried out on the engine acoustic test bed or in the vehicle. In addition to the pure reduction of the acoustic pressure level, the quality of the noise is playing an even more important part in development.Here “dummy head” recordings are used to log the noises that are then evaluated subjectively by test persons in comparison listening tests. These tools are shown in Figure 7.346.

 coustic measurement and simulation tools A Many tools are available for use in designing an intake system; the simulation tools, in particular, have grown in significance in recent years because they can be used to predict acoustic properties even in a very early development stage.

 uture systems F In addition to passive measures, adaptive measures are employed more and more in intake systems. Active manifolds are used primarily to increase the airflow rate. But in the air routing systems, too, these components can be employed to optimize acoustic properties.

Tool

Option

1D-FDMComputer simulation

Optimization of the entire intake system

3D-FEM/BENComputer simulation

Detailed examination for noise irradiation

Pulsation test stand

Realistic irradiation measurement without engine

Engine acoustic test stand

Realistic examination at the engine

Chassis dynamometer for passenger cars

Total vehicle measurement, transfer path analysis

Artificial head measurements

Psycho-acoustic assessment sound design

Speaker sonication test

Damping spectrum and component damping

Figure 7.346 Acoustic measurement and simulation tools.

274 | Internal Combustion Engine Handbook

6606_Book.indb 274

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

Lower speed range

Medium speed range Upper speed range

Figure 7.347 Stepped actuated intake manifold.

Active noise control on the basis of wave interference 180°

Loudspeaker

Engine

By superimposing (= addition) the original sound wave with anti-noise (phase shift 180°), the noise is canceled. Figure 7.348 Active noise control.

By using a single-stage, active manifold in the lower speed range, when the engine does not require its full volumetric flow, a smaller intake cross section can be used to achieve low frequency tuning of the Helmholtz resonator. Figure 7.347 shows such a configuration as an example. The advent of electronics in the intake system has paved the way for entirely different system designs, such as the use of inverted or dephased sound to cancel out noise. If the noise arriving from the engine is countered by a wave of the same amplitude but 180° out of phase, then these two waves cancel each other out. This principle is also known as active noise control and is depicted in Figure 7.348.

7.21 Sealing Systems Many different types of seals and gaskets, and almost as many different materials, are found in internal combustion engines. One normally becomes aware of these inconspicuous engineering elements only when they fail. In such cases, however, functioning of the entire system is endangered. The immense importance of seals and gaskets is clear right from the early stages of engine development. A lack of properlyfunctioning seals makes it virtually impossible to undertake component testing.

Modern sealing systems are extremely reliable. Great development effort has been devoted to devising solutions that ensure long and dependable service life even under critical conditions such as aggressive media, high pressures, and extreme temperatures. This section is intended to provide the reader with an overview of the various types of seals, their uses, and basic information on how they function.

7.21.1 Cylinder Head Sealing Systems

The head gasket is becoming more important in modern engines. In addition to sealing off the combustion chamber, the cooling system, and the oil passages, the head gasket also serves to transmit forces between the cylinder head and the engine block. Thus, it exerts considerable influence on force distribution within the entire assembly system and the associated deformations in elastic components. More stringent requirements for fuel consumption and emissions have given rise to engine designs with optimized weights and, particularly in diesel engines, higher ignition pressures. The use of aluminum and the reduction of wall thicknesses in castings are reasons to anticipate further reductions in component stiffness. To further reduce cylinder warping, which is detrimental to exhaust gas composition, engineers

Internal Combustion Engine Handbook | 275

6606_Book.indb 275

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

are striving to reduce bolt forces. These efforts result in a considerably greater load on the head gasket in the form of dynamic fluctuations at the sealing gap. The combustion chamber seal must be able to ensure the minimum sealing force, permanently and at all operating states. This causes very high demands on the durability of the sealing system selected for use here. 7.21.1.1 Ferrolastic-Elastomer Head Gaskets The head gasket made of asbestos-free ferrolastic elastomers (Figure 7.349) is the system most widely used after the conversion to materials containing no asbestos at the end of the 1980s. The structure consists of a notched metal substrate with elastomers rolled onto both sides. The sealing effect is distributed over the entire surface area, and that requires high bolt forces. The disadvantages of this system are found in the relatively low elastic resilience. Great dynamic fluctuations in sealing gap width or changes in pressure because of thermal effects cannot be compensated and can be neutralized, only in part, by greater bolt forces.

Figure 7.350 Metal-elastomer head gasket.

Metal combustion chamber bead for gas-sealing

Gas

Engine block

FPM elastomer for cooling water and oil sealing

Cooling water or oil

Figure 7.351 Section through a metal-elastomer gasket.

Figure 7.349 Head gasket made from ferrolastic composite material.

Engines with high thermal loads, narrow web widths, and wide oscillations in the sealing gap mark this system’s limits, triggering the development of higher-performance systems. 7.21.1.2 Metal-Elastomer Head Gaskets Metal-elastomer head gaskets (Figure 7.350) are now used primarily in heavy-duty utility vehicle engines. The principle behind this design (Figure 7.351) is distinguished by the separation of functions (separate sealing for the combustion chamber and the liquid circuits) and the system’s immense potential in each case. Not only are bead concepts with purely plastic properties used to seal the combustion chamber, but elastic systems as well. The passageways for liquids sealed with elastomer sealing lips exhibit great adaptability and elastic resilience. Selecting a suitable elastomer material ensures suitable aging resistance when exposed to fuel, coolant, and oil. Depending on the overall concept for the gasket, the elastomer lips may be injected onto the end of the sealing plate or on the surface. As an alternative, so-called inserts, i.e., metal substrates with a sealing lip vulcanized in place, may be used. To avoid component warping and to achieve closely defined introduction of pressure into adjacent components, support elements may optionally be provided at the outer edge of the gasket.

Because elastomer elements require only low sealing forces relative to the bolt force, the entire bolt force can be used for sealing the combustion chamber and component support, if necessary. The available bolt force is thus very efficiently utilized; component warping or the number of bolts used can be reduced. 7.21.1.3 Layered-Metal Head Gaskets Metaloflex Multi-layer steel gaskets have been used as head gaskets (Figure 7.352) in mass production since 1992. Particularly in modern diesel engines and in high-performance gasoline engines, intense effort is required to devise a solution suitable for mass

Figure 7.352 Multi-layer steel head gasket.

276 | Internal Combustion Engine Handbook

6606_Book.indb 276

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

production when using the elastomer seals employed up to that time. For the developer, an essential advantage of the layered-metal head gasket is that the gasket dimensions can be optimally adapted for the engine’s technical requirements, thus avoiding the time- and cost-intensive iteration steps. The metal head gasket is composed of one or more layers, depending on the application.  unction F The sealing function of the layered-metal head gasket is essentially dependent on the beads in the spring steel layers. The deformation characteristics permit plastic adaptation to component stiffness, on the one hand, and, on the other hand, great resilience to compensate for dynamic fluctuations in the sealing gap and for thermally induced component deformations. With the use of half-beads in the liquid sealing areas and full beads at the combustion chambers, the compression levels along a given line, required for sealing in each case, are achieved (Figure 7.353).

“micro bead” coined in serpentine pattern creates a swelling which can substitute the welded stopper at nearly identical stiffness (Figure 7.354). The reason: The multitude of windings created by the serpentine geometry increases the stiffness of the stopper avoiding settling during engine operation and unwanted elasticity. Such an elastic stopper would cause an increase in sealing lip fluctuations under ignition pressure in the engine and have negative effects on system durability.

Figure 7.354 The serpentine stopper achieves a stiffness nearly identical to the one of a laser-welded stopper.

Figure 7.353 3-D section through a metal-elastomer gasket.

The stopper induces an elastic preload in the components around the edge of the combustion chamber. In this way, the fluctuations in the sealing gap resulting from the gas forces are reduced, while at the same time unacceptable deformation of the full beads is prevented. Standard stopper thicknesses are in the range of 100 to 150 μ m. An intermediate layer may be inserted to achieve the required installation thickness or to accommodate differing thickness adjustments for diesel engines; this intermediate layer has no influence on the sealing function. Previously, laser-welded and crimped stopper layers have been used very successfully in series production while coined stoppers are now used for newer generations. In addition to the permanent assurance of the stopper effect, this concept allows the integration of additional functions into the gasket. The integration of a coined stopper in an existing sealing layer achieves very economical solutions. As a rule, one differentiates between stopper coinage in the spring-steel layers and those in the carrier plate. The manufacturing processes for honeycomb and serpentine stoppers allow nearly any geometrical profiling in respect to stopper width and also stopper thickness. Beyond the traditional stopper surface, the designer now has the option to create additional supports on almost any location on the gasket. Application examples  erpentine stopper in spring steel layers S With the serpentine stopper, the surface for the stopper geometrically defined by the engine is optimally utilized. A

 oneycomb stopper in the substrate H The coined stopper in the carrier plate is designed as a honeycomb geometry (Figure 7.355). The cavities coined on both sides as truncated pyramids generate bell mouths on each opposite side which are calibrated in a second work step, that is, they are coined to the required stopper thickness. This process ensures significant strain hardening in the ductile basic material creating a stopper structure with very high mechanical strength. The stiffness of these stoppers is comparable with that of welded stoppers.

Figure 7.355 A honeycomb stopper is used in nearly all head gaskets with carrier plate.

The calibration process described can be used with planar or profiled tools allowing the production of topographical stoppers in very economical process.  ariable stopper thickness V Proper stopper design makes it possible to exert a closely defined influence on sealing pressure distribution and thus, the sealing gap fluctuation. The gasket is normally between 0.10 and 0.15 mm thicker in the area around the stopper, the exact amount depending on engine stiffness. This causes an increase in pressure and elastic preload in the sealing system (Figure

Internal Combustion Engine Handbook | 277

6606_Book.indb 277

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.356). Coined stoppers in particular enable a topographical stopper design required for almost any engine component. The height profiling (Figure 7.357) can be defined variably, both for each cylinder as well as for other areas on the gasket.

increased. Supports in these areas (Figure 7.358) are wellknown and are used in many engines to optimize distortions of bearing races. They limit the flexure of the cylinder head to a minimum (Figure 7.359) and reduce the bending loads in the camshaft. Another positive effect is the reduction of running noises due to increased bearing play.

hoch/high

niedrig/low

Figure 7.356 Comparison of distribution of sealing pressure: Left: a stopper with constant thickness. Right: the optimized stopper with variable thickness. See color section page 1079. Figure 7.358 Carrier plate with face honeycomb support elements.

Figure 7.357 3-D view of a head gasket with height-profiled stopper.

The topographical stopper allows compensation for inhomogeneous component stiffnesses. Areas with low stiffnesses can be prestressed, and thus the application of a uniform compressive load is ensured. In this way, the available bolt force can be exactly distributed and optimally utilized over the required areas. Because the head gasket is a central element of the complete mounting system between cylinder head and engine block, it can be used to precisely control the sealing pressure distribution and thus the force transfer into the engine components. By using additional support elements, flexures and sealing pressure distribution can be optimized in the cylinder head and the block. Thanks to the special coining technology, the supporting elements are possible at nearly any location of the gasket and in combination with conventional stoppers. This is possible not only with serpentine but also honeycomb stoppers. With the gasket design, the deformations of the engine components are influenced and limited in a focused manner, taking the bolt force into account. For example, tensions in the engine components can be reduced, thus effectively preventing cracks in cylinder heads. Due to the high, locally introduced bolt forces, the cylinder head distortions are usually considerably

 ulti-function layer design (Figure 7.360) M Excessive sealing gap fluctuations cause dynamic overloading of the beads; the full beads at the combustion chamber are in particular danger. Relaxation—a reduction in bead force and resilience—occurs and fissures may even appear in the beads. The functional layers in the Metaloflex head gaskets that are provided with beads use their resilience to compensate for the sealing gap fluctuations occurring in the engine. With the use of multiple functional layers, the overall amplitude can be distributed to the individual layers and thus reduced to an acceptable level. The total resilience of the gasket rises with the number of functional layers used. In this way, it is possible to ensure function and durability, even at low bolt forces and high peak pressures.  artial elastomer coating (Figure 7.361) P Due to partial coating, only the surface areas of the head gasket which are relevant for sealing are coated. This makes it possible to omit the coating on the sealing surfaces that extend into the coolant or the motor oil; thus, there is no coating there that could peel off under critical conditions. Further advantages of this process are that, with the special application procedure, both the thickness of the coating layer and the coating medium can be selected to suit the application. The coating requirements in the combustion chamber and liquid areas, which differ in part, can thus be properly met. For coolant and oil sealing, a thicker layer and softer elastomer are beneficial if, for example, the mating surface is rough or porous. At the same time, thinner layers are necessary to contain the ignition pressure at the combustion chamber. These conflicting goals can be resolved by selective coating.

278 | Internal Combustion Engine Handbook

6606_Book.indb 278

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

Engine longitudinal axis

Engine longitudinal axis

Cylinder head flexure

Deformed cylinder head (without support elements) Figure 7.359 Cylinder head deformation with and without face support elements.

Figure 7.360 3-D view of a head gasket with multi-function layer stopper. Figure 7.362 3-D view of a head gasket with double stopper: folded bead.

Figure 7.361 Metal layer head gasket with partial coating.

 ouble stopper design for gasket designs D A modified gasket design is required in many cases where a separate cylinder liner is used. To avoid plastic deformations and to keep the liners from being shifted downward, the necessary sealing and preload forces have to be introduced into the gasket system in a defined manner (Figure 7.362 and Figure 7.363).

Figure 7.363 3-D view of a head gasket with double stopper: honeycomb design.

Internal Combustion Engine Handbook | 279

6606_Book.indb 279

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Force application at the liner is closely defined with the use of a so-called double stopper. In this configuration, one stopper is formed by a folded bead around the edge of the combustion chamber, while a second stopper is formed behind that bead by overlapping two sheets of metal. To achieve ideal response during operations, the stopper force acting on the liner must not cause plastic depression of the liner. By coining varying stopper thicknesses, the distribution of the pressure to the two stoppers can be regulated individually. Thus, for example, the stopper at the outside may be thicker by 20 μ m; consequently, the larger share of the preload is directed not to the liner but rather to the outside area of the cylinder tube. This stratagem ensures the required preload on the components while at the same time avoiding any displacement of the liner. The partition of the stopper enables a significant reduction in cylinder distortion in many situations.  topper-less design S In gasoline engines, and particularly where aluminum engine blocks are used, it is possible under certain circumstances to do without the stopper. In this way the elastic deformations of components caused by the head gasket are reduced dramatically. In addition to reducing cylinder deformation, deformations in the area around the valve seats can also be significantly reduced. The implementation of this concept does, however, require that the bead geometries be matched exactly to the details of the mating components. In standard gaskets with stoppers, the deformationof the full beads are determined by the thickness of the stopper. Protecting the beads in this way creates ideal conditions with respect to durability and resilience. Without stoppers (Figure 7.364) the deformation of the beads depends largely on the stiffness of the components. This means that, depending on the stiffness of the cylinder head and the engine block, the beads are deformed to a greater or lesser extent. Attaining sufficient sealing pressures while achieving ideal durability requires individual adaptation to the conditions prevailing in the engine.

Figure 7.364 3-D view of a head gasket without stopper.

I ntegrated extra functions Integrating a high-sensitivity sensor system directly into the head gasket provides for even more dependable monitoring of processes in the engine: integrated sealing gap sensors (Figure 7.365). The sensor system uses the enormous pressures created by combustion inside the cylinder.These pressures cause relative movement between the engine block and the cylinder head.

Figure 7.365 Head gasket with integrated sealing gap.

The sensor registers this movement and is thus able to detect at an early date irregularities in the engine, such as misfiring or other ignition problems. The measurement of coolant and component temperatures inside the engine is becoming even more significant because, in conjunction with cooling regulated according to the engine map, for instance, the values registered at the measurement points that were previously used are hardly representative. Particularly in operating ranges where there is little coolant flow or none at all, the temperature, of necessity, has to be measured at critical points in the engine. 7.21.1.4 Prospects for the Future The requirements of future engine designs for the head gasket are characterized essentially by higher peak pressures, greater thermal loading, reduced component stiffness, and new materials. With its modular construction, the layered-metal head gasket offers every option for individual adaptation to the specific conditions prevailing in the engine. The engineering freedom offered by this system permits influencing component mounting and strains and the distribution of pressures in the engine. In this way, the available bolt force can be utilized efficiently while, at the same time, minimizing component deformations. The advanced Metaloflex layered metal head gasket will continue to represent a reliable, durable, and economical sealing concept.® The metal-elastomer technology will also be the predominant head gasket design in the heavy utility vehicles sector. Separating the combustion chamber and liquid sealing functions enables ideal adaptation of the gasket, particularly in engines with “wet” cylinder liners.

7.21.2 Special Seals 7.21.2.1 Flat Seal Functionality Flat seals are highly effective, cost-favorable seals both for a number of liquid media and for gases. A broad range of pressure

280 | Internal Combustion Engine Handbook

6606_Book.indb 280

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

and temperature loads can be managed. The requirements on the flange surfaces at the components being sealed are low; surfaces machined with the milling head are sufficient. To achieve positive sealing for static, flat seals, sufficient surface pressure is guaranteed at all operating states: Influencing parameters such as operating media, fluctuations in temperature and operating pressure, engineering elements (such as bolts and sealing surfaces), the location of the gasket within the assembly, and the seal’s long-term influence on the sealing assembly have to be taken into account during engineering. Thus, the following requirements apply for the sealing element: •• adaptation to component surfaces (micro-structure—roughness/macro-structure—not plane) •• pressure resistance (setting behavior) under the influence of heat and/or operating media •• tightness across the entire surface of the seal •• cross-sectional tightness in the seal material •• mechanical stability (tensile strength) •• elastic resilience properties •• temperature resistance Consequently, the ideal seal is an elastic rubberized metal with great strength and resistance to media and temperatures. 7.21.2.2 Elastomer Seals Elastomer seals (Figure 7.366) are employed in a broad spectrum of applications. They are made up of a composite material comprising fibers, fillers, and binders (Figure 7.367). Since the end of the 1980s, rubber-asbestos elastomer seals have been replaced almost completely by asbestos-free qualities. In high-quality elastomer seals, aramid fibers have largely been substituted for asbestos fibers. This material has superb mechanical and thermal properties. Cellulose and mineral fibers are used for economical gasket materials used in less critical areas.

Figure 7.366 Elastomer seals.

Figure 7.367 Elastomer seals—composite structure.

The multitude of material qualities available, such as EWP sealing materials, makes it possible to select a suitable sealing material for almost every application. Elastomer sealing material is available in thicknesses between 0.20 mm and beyond 2.5 mm. Using the material thickness, the seal can be customized in respect of adaptability, mechanical stability and setting behavior. The performance capacities of the elastomer seal can be further improved by applying additional elastomer layers along a line. In these areas the prescribed preload force on the surface (low sealing pressure) is reduced to narrow line-like areas (high sealing pressure). Elastomer seals are cut on modern CNC water jet machines. Gaskets are cut without conventional tools when this technology is employed. The limits for the use of asbestos-free elastomer seals are found in areas that are subjected to severe thermal loads. 7.21.2.3 Metal-Elastomer Seals Metal-elastomer seals (Figure 7.368) differ from the elastomer seals described in the previous section in that they have a metal insert at the center of the material (Figure 7.369). They are used primarily in automotive applications and are found in the coolant, oil, fuel, and exhaust areas.

Figure 7.368 Metal-elastomer seals.

Internal Combustion Engine Handbook | 281

6606_Book.indb 281

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

of up to 150°C, they are comparable with composite materials (Section 7.21.2.2). In exhaust system seals, graphite and mica materials, with their great resistance to high temperatures, are used. As was described in the previous section on elastomer seals, the performance capacity of the metal-elastomer seals can be further boosted by an additional elastomer coating applied along a line. The seal quality over broader surface areas in particular can be improved significantly in this way. 7.21.2.4 Special Metaloseal Gaskets The term Metaloseal® is derived from the words “metal sealing.”®â The fundamental structure for metal seals (Figure 7.372) is based on a metal substrate that is coated with an elastomer, usually on both sides.

Figure 7.369 Structure of a metal-elastomer seal.

The metal insert (substrate plate) is normally sheet steel that is toothed, perforated, or glued to a smooth surface. The metal insert provides a number of benefits: •• high tensile strength •• mechanical ruggedness •• high-dimensional accuracy of the seal •• procedural benefits (coil manufacturing) •• reduced costs due the reduction of the fiber content •• application of various sealing materials on the carrier plate. Because the required tensile strength is attained by the carrier plate, as shown in Figure 7.370, it is possible to optimize, in a targeted manner, other specific properties of the sealing materials. The specific properties of the materials listed in Figure 7.370 are determined primarily by the composition of the sealing surface. Figure 7.371 indicates the most important adjustment parameters in the selection of the sealing layer. Materials

Optimized Properties

Application Example

FW 522

Pressure resistance, crosssectional tightness, medium resistance

Head gasket

FW 715

Adaptability, cross-sectional tightness

Oil pan

FW 520

Temperature resistance up to 450°C

Exhaust manifold

FW 501

Adaptability, temperature resistance up to 500°C

Exhaust gas recirculation

FW 610

Temperature resistance up to 800°C

Turbocharger

Figure 7.372 Oil filter attachment seal with fins.

One of the major advantages is found in the fact that any of a variety of metals can be combined with varying elastomer compounds to suit the application at hand. Thanks to the beads that are formed in addition, the substrate material’s properties can be matched perfectly to the sealing system (Figure 7.373). As was already described in Section 7.21.2.1, the demands made on the sealing element can be satisfied only by metal gaskets that are coated and mechanically modified.

Figure 7.370 Summary of metal-elastomer materials.

The compound used for the sealing layers is determined most strongly by the thermal requirements. In a temperature range Pressure Resistance

Adaptability

Interior Tightness

Resilience Properties

Temperature Resistance

Filler content

↑

↓

→

↓

↑

Fiber content

↓

↑

↓

↑

→

Elastomer portion

↓

↓

↑

↑

↓

Impregnating agent content

↓

→

↑

↓

↓

Compression

↑

↓

↑

↓

→

Figure 7.371 Material parameters and their effect on functionality.

282 | Internal Combustion Engine Handbook

6606_Book.indb 282

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

Substrate

Coating

Bead

Figure 7.373 Structure of metallic seals.

 ubstrate materials S The choice of substrate materials has a direct influence on sealing properties. Ideal adaptability of the seal to the flange surfaces (macro-sealing) can be achieved with two parameters: substrate material properties and bead geometry. Figure 7.374 provides a survey of various substrate materials. Substrate Materials

Application Conditions

Cold-rolled strip

Standard design

Spring steel

Dynamic sealing gap fluctuations, high pressures

Stainless steel

Aggressive media, corrosion protection, increased protection against frictional wear

Temperature-resistant steels

Exhaust area or at temperatures between 400°C and 1,050°C

Aluminum

Preventive for avoiding contact corrosion in magnesium, aluminum or gray-cast housings

Figure 7.374 Metaloseal substrate materials.

Standard material thicknesses are between 0.20 and 0.30 mm. In special cases, thicker or multi-layer seals may be used. This enables the achievement of the best possible macro-sealing properties for virtually every application, through the selection of suitable materials and bead geometries.  oating C The elastomer is selected primarily for the media to be sealed off and the prevailing operating temperature. One of the most important tasks in coating is compensating for surface imperfections. Thus, the medium being sealed is kept from leaking across the surface area. The thickness at which each coating layer is applied can vary between 5 and 100 μ m (on either side), depending on the particular situation. Figure 7.375 shows a few application examples for the various elastomer materials. Elastomer Materials

Application Conditions

NBR

Coolant, oil, air, limited for fuel

FPM

Fuel

EPDM

Brake fluid, hydraulic oil

Temperature-resistant coating

Exhaust area for flange temperatures <1,000°C

Graphite coating

As slip coating, to offset high relative movements of the components

Functioning of a metal seal In the past, it was necessary, in some cases, when using conventional elastomer materials, to resort to engineering tricks in order to achieve a reliable seal. Thus, elastomer gaskets required exactly defined bolt torque in order to achieve sufficient surface pressure while at the same time avoiding excess pressure on the material, which would inevitably damage the elastomer and result in leaks. In addition, there is always a conflict of goals among the various sealing properties when selecting the elastomer material (see Figure 7.371, Section 7.21.2.3). This is where the advantages of metallic seals become apparent. A bead pressed into the substrate reduces the surface pressure to a line-shaped pressure. Thus, at the same bolt forces, higher surface pressure values can be achieved or, conversely, the same surface pressures can be achieved at lower bolt force. With the use of a metallic substrate, all of the physical properties of a metal can be exploited. In addition, one creates a further scale that can be adjusted as required: the bead force. Bead force is influenced both by a certain ratio of the bead’s height to width and by the shape of the bead itself—half or full bead—and can be modified individually to suit each application point. When the component is first tightened down, the elastomer coating is pressed into the surface by the force at the bead and closes any imperfections present there. In addition, the degree to which the seal adapts to the flatness of the component is determined by the bead. The bead functions, in the classical sense, like a spring that develops the required sealing force in response to deformation. Figure 7.376 describes the interrelationship between the component’s demands on the seal and the capabilities that the various functional elements have to exert an influence. Requirements of the Seal

Functional Element

Adaptability to component roughness

Adaptable elastomer coating

Adaptability to component unevenness

Beads

Cross-sectional tightness of the seal

Non-porous elastomer coating

Pressure resistance (setting behavior)

Metallic substrate, thin elastomer coating

Mechanical stability

Metallic substrate

Elastic resilience properties

Substrate material such as spring steel, bead

Temperature resistance

Substrate and coating material

Figure 7.376 Each of the various functional elements in the Metaloseal gasket responds to a specific requirement.

 pplication conditions A Because of the multitude of options for combining metallic substrate materials and various elastomer compounds, almost all of the application points in the engine can be covered. Naturally, every sealing system has to be analyzed, and the corresponding sealing properties, such as material structure and bead geometry, have to be defined. Figure 7.377 provides an overview of the wide range of applications for metallic gaskets.

Figure 7.375 Application conditions for various elastomer materials.

Internal Combustion Engine Handbook | 283

6606_Book.indb 283

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Criteria

Application

Temperature

−40°C–1,050°C

Pressure

Up to 350 bar

Media

Coolant, oil, exhaust gas, brake fluid, hydraulic oil, air, fuel, bio-diesel, urea solution

Surface parameters Roughness

Rmax ≤ 25 μ m

Unevenness

≤0.30 mm

Figure 7.377 Application range of metallic seals.

7.21.2.5 Prospects for the Future In the course of increasing requirements for exhaust gas emissions to Euro 5 and Euro 6 in passenger cars and trucks, the demands on sealing systems have also grown. This results in new and innovative products. In engine development, there is a trend towards downsizing and providing engines with exhaust-driven turbochargers. At temperatures above 1,000°C in the exhaust area, specially coated metallic seals are required which exhibit sufficient mechanical spring behavior to cope with even these temperatures. Nickel-based alloys are increasingly used as a sealing material in this context. Metallic sealing rings (V- or C-rings) from nickel-based alloys present here a technical solution (Figure 7.378). By using special manufacturingprocesses, extremely low leak rates at high temperatures are possible. High-temperature ranges need special attention because lowest leak rates usually require additional coating of the sealing systems. Additional exhaust gas after-treatment systems such as SCR (selective catalytic reduction) make high demands relative to corrosion on the sealing systems. The use of bio-fuels are makes it necessary to prevent damage to seals. By using metallic substrate materials, seals can assume additional functions such as the integration of oil splash plates or sensors for efficient engine management. Furthermore, pre-assembly solutions such as holding clips and centering elements offer additional advantages for the installation of the component.

Figure 7.378 Metal seals from nickel-based alloys.

7.21.3 Elastomer Sealing Systems

Greater performance at less weight, reduced fuel consumption, and lower emission levels—these central demands of the power plant engineer mean greater demands on sealing systems. Thus, engine components and attached assemblies are more frequently manufactured from plastic for weight and functional reasons. Reduced component stiffness (lower Young’s modulus) in comparison to the aluminum and magnesium previously used are the consequences. When parts are clamped, greater deformation is encountered, and the sealing system has to compensate for this. The elastomer-based sealing systems are superb in satisfying these exacting requirements. On the one hand, the sealing pressure required by elastomers is very low and, on the other hand, their superior elastic properties enable tolerance compensation over a broad range. Because of elastomer materials’ ability to resist extreme temperatures, these are used exclusively for containing liquids and gases. A metallic structure is used in elastomer gaskets to seal off the combustion chamber. Depending on the medium to be sealed and the prevailing temperatures, the elastomer materials are selected according to the specification profile. Figure 7.379 gives an overview of the elastomer compounds available and typical applications.

Elastomer material Abbreviation (ISO 1629) FPM MVQ MFQ ACM AEM EPDM ECO

Chemical Engine application Coolant Fuel designation + Fluoroelastomer + o Silicone rubber – o Fluorosilicone rubber – – Polyacrylate rubber – – Ethylene-acrylate rubber – + Ethylene propylene diene – rubber – Epicholorohydrin rubber +

HNBR

Hydrogenated nitrile rubber

+ well suited / o suited / – unsuitable

o

o

Thermal applications

Application examples

–20 bis +230 °C –50 bis +200 °C –70 bis +180 °C –30 bis +150 °C –35 bis +160 °C –50 bis +130 °C

CHG, Intake area CHG, Special applications CHG, Special applications Oil pan, cylinder head shroud Oil pan, cylinder head shroud Water pump

+

–40 bis +120 °C

+

–30 bis +150 °C

Special applications in the fuel area Special applications

Oil + o + + + –

CHG = Cylinder head gasket

Figure 7.379 Elastomer materials.

284 | Internal Combustion Engine Handbook

6606_Book.indb 284

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

7.21.3.1 Elastomer Seals Elastomer seals (Figure 7.380) have no substrate. To prevent overloading the elastomer profile, these seals are, for instance, installed in a groove in the component. These components are always designed so as to eliminate any external deformation. The height-to-width ratio is characteristic of the design of this seal. The cross-section is considerably thicker (higher) along the direction of the compression forces than it is wide. At compression of 20% to 30%, this gives a very broad working range for the seal, and also enables the sealing of plastic components which are subject to severe deformation. This type of gasket is used, in particular, in combination with valve covers, intake manifolds, or water flanges made of plastic. When sealing camshaft bearings and other threedimensional passages in components, the elastomer gasket is the only option for certain management of the sealing point.

With special cross-sections calculated using the finite element method (FEM), the gasket’s profile is matched to the specific properties of the component being sealed. A rectangular gasket cross-section is used only very rarely because the deformation properties are very limited. The T-section is the preferred sealing profile for acoustic purposes. In combination with specially designed decoupling elements for the bolts, this design is used for valve cover seals that integrate acoustic decoupling. Because the components being sealed are pressed together by elastomer elements (see Figure 7.381), this system can no longer be calculated with the engineering methods used in the past. To make these systems more functionally reliable, an analysis of the complete mounting system—comprising the seal, the decoupling element, bolt, and bushing—by FE calculations is unavoidable (see Figure 7.387 and Figure 7.389 below in Section 7.21.4.1). Demands made on acoustically decoupled systems include •• structural noise decoupling •• positive bolting of components •• sealing •• pre-assembly of individual parts. The interplay of FE calculations, laboratory simulation, and material development work is the basis for tailor-made, acoustically decoupled sealing systems.

Figure 7.380 Intake manifold gasket.

In these areas, a special shape of the elastomer gasket can ensure optimal sealing.

7.21.3.2 Metal-Elastomer Gaskets Because some components, for geometric or functional reasons, cannot use pure elastomer seals (they require a groove in the component), the metal-elastomer seal was developed (Figure 7.382). In this type of seal, the elastomer is vulcanized directly to an aluminum or steel substrate. The thickness of the elastomer is coordinated with that of the substrate but is always considerably thinner than in solid elastomer seals. Here, like the pure elastomer seals, the elastomer is not installed in the path of primary force flow. No groove is required in the component because the substrate, made of aluminum or steel, is at the main force transfer point (Figure 7.383).

Bushing (forcetransmitting element)

Decoupling element

Elastomer seal Screw (with centering tip), movable mounted

Cylinder head shroud

Elastic shroud mounting (captively pre-assembled) Elastomer

Figure 7.381 Example for an uncoupled cylinder head shroud system.

Internal Combustion Engine Handbook | 285

6606_Book.indb 285

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

developing other components and are offering them, together with the seals, as pre-assembled, multi-functional systems. These modules, ready for immediate installation, replace the previous individual parts more and more. Here, every conceivable combination of sealing system and component (made of aluminum, magnesium, steel, or plastic) is possible. Lightweight designs are indispensable for reducing fuel consumption when dealing with rising engine performances. Plastics offer decisive advantages here and replace, to a greater extent, the materials used for engine components in the past. The expertise and the system proficiency arising from sealing technology, and from elastomer processing in particular, form the basis for developing innovative plastic modules. These are employed especially in the following areas:

Figure 7.382 Metal-elastomer engine block gasket. Metal support

AEM elastomer for oil sealing

•• valve shrouds (Figure 7.384) •• engine compartment capsules •• oil separators •• coolant flanges •• intake manifolds. Depending on the requirements imposed on the plastic components, PA 6 is used for parts where appearance is important and PA 6.6 for components that have to introduce or transfer forces.

Bolt

Component

Fluid flow

Figure 7.383 Section through a metal-elastomer gasket.

Design freedom in engine development is increased considerably with this concept by integrating supplementary functions into the substrate. In addition, this system is distinguished by great functional reliability and economy. Functions that are normally integrated in practice are •• calibration of fluid flows •• exhaust gas recirculation •• assembly aids •• pre-assembly using clamps •• cable grommets. The use of two-component injection machines makes it possible to vulcanize two different elastomers on a single substrate. The advantage is that the most suitable elastomer can be used for each of the media to be contained. Indispensable to this process are reliable coupling agents to ensure a good bond to the metal. Metal-elastomer head gaskets made of metal substrates to which elastomer profiles are vulcanized are described in Section 7.21.1.2. This type of gasket is used in commercial vehicles and in the large engines found in ships and locomotives. 7.21.3.3 Modules Important to achieving a properly functioning sealing system is not to see the system in isolation but rather to observe the complex interaction of all the individual systems involved. Consequently, seal and gasket manufacturers are now also

Figure 7.384 Valve shroud module with integrated gasket and oil separation.

286 | Internal Combustion Engine Handbook

6606_Book.indb 286

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

Initial attempts to replace PA 6.6 with PA 6 are in development. To achieve the required strength and processing properties, one blends glass fibers and in some cases mineral fillers into the basic resin. Elastomer sealing systems are employed in modules with integral sealing functions, because these can be ideally tuned to the medium to be contained and the requirements for component stiffness. Numerous functions can be integrated into modules, both very efficiently and economically, due to plastics processing properties. Major advantages are also found—as previously mentioned—in the amount of weight reduction and the manufacturing technology because plastic components make it possible to completely eliminate post-injection work such as deburring, tapping threads, or finishing surfaces. Compared to aluminum, thermoplastic material has the advantage of allowing the integration of components by welding. Examples for the multi-functionality of modules include •• acoustic decoupling of the component •• integrating the blow-by gas exit from the crankcase •• integrating oil separator systems into a cylinder head shroud •• integrating valves to regulate crankcase pressure •• integrating cable passages from the cylinder head •• providing a pre-assembled, complete system. In order to ensure reliable functioning of the module over the engine’s entire service life, exhaustive tests of functions and geometries are conducted during the development phase. In addition, simulation tests are worked out that allow for depicting the loading conditions that occur during vehicle operation, making it possible to reduce testing times. When developing these tests, the experience and results drawn from practice are always taken into account. The interplay of FE calculations, simulations, and engine testing makes it possible to prepare for mass production, in the shortest period of time, plastic modules that satisfy all the requirements for loading capacity and service life.

7.21.4 Development Methods

Engine tests continue to be a major factor in gasket testing. These tests, carried out on an engine on a test bed and driven by an external power source, are expensive and time-consuming. Because the trend is toward shorter development cycles, calculations for the sealing systems and laboratory testing under conditions that closely replicate those in the engine are moving further into the foreground. This is intended to

enable fundamental assessments of the functioning of the gasket design even before actual testing in an engine, thus reducing to an absolute minimum the number of costly engine tests required. Preliminary examinations of seals without real-world engine components provide in-depth insights into the functional capacities of the product. The finite element method is used as a calculation tool. This term describes the mathematical algorithm used to translate for computer processing a physical phenomenon at a section of the component being analyzed. A finite element model is the depiction of a geometry by a sufficient number of discrete elements. 7.21.4.1 FE Analysis The task for the person conducting the analysis is to identify the phenomena required to describe the problem at hand and to enter them into the calculation software. FE calculations are used to optimize the components, both in the engineering phase and in the subsequent test phase. This preliminary selection allows one to reduce the number of prototypes required. Many of the engineer’s problems can be converted directly into an FE calculation model right in the CAD program and, provided with the appropriate material behavior data and operating conditions, can be forwarded to an FE analysis program for calculation. The basis for this approach is linear calculation, using small component deformations, elastic material laws, and unequivocal mounting and loading. A further special field application is found for component calculations whenever one of the fundamentals for the linear approach is violated. Nonlinearity of a calculation problem (see Figure 7.385) arises as a rule with major deformations of a component under load wherein, for example, the length of the lever arm used for chucking is shortened and a smaller flexural moment is created than what is defined in the basic dimension. If there are also path limitations for component deformation, then these are described as nonlinear contact conditions. The behavior of most technical materials is also linear only in a very limited range; there they adhere to „Hooke’s Law“, which links tension and elongation, expressed as “Young’s modulus.” Optimization strategies lead to weight reduction or better utilization of the material in the spirit of a uniform strain level push to the boundaries of this range. If one departs from this linear range, then plastic deformations typically appear at metals, creep elongation at plastics, or stress-induced relaxation processes.Nonlinear responses in the tension elongation function are always found in rubber materials. There, time factors—that is, how quickly the load

Flexure beam representations Load Deformation

Linear, little deformation

Load

Load

Deformation

Deformation

Geometrically non-linear, Non-linear contact with other the load application point bodies limits the deformation travels through large deformations

Figure 7.385 Bending bar linear, nonlinear, and with contact.

Internal Combustion Engine Handbook | 287

6606_Book.indb 287

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Bead with stopper, radial cut 1.5 layer or 3.5 layer

Bead force/amplitude [%]

a)

b)

140 130 120 110 100 90 80 70 60 50 40

Maximum permissible amplitude Line load of the bead on stopper length Line load of the bead at 10 μm rebound travel

80

90

100

110

120

is applied and the effective time—play an essential part in the deformation response of a given body.

Force [N/mm]

 roduct calculations P Preliminary calculation and optimization of component properties require both detailed knowledge of material responses and a good understanding of the manufacturing path followed from the semi-finished product to the part ready for shipment. At a full bead in a layered metal gasket (see Figure 7.386, above), several re-forming steps are carried out prior to final assembly in the engine. All the steps are “memorized” by structural changes in the metal and determine the bead’s properties, spring characteristic, and tolerable sealing gap amplitude. With suitable tool dimensions the spring element can be designed for constant width at high force with appropriately smaller permissible sealing gap amplitudes or for great sealing gap amplitudes at lower forces (see Figure 7.386, below). The tuning required for the bead depends on the stiffness of the engine components and the ignition force. Elastomer sections are frequently used at the engine to seal covers and shrouds, intake manifolds, and caps. They are characterized by great adaptability at the sealing surfaces 5 4,5 4 3,5 3 2,5 2 1,5 1 0,5 0

Bead height [%]

Figure 7.386 Sealing gap amplitude and line load of a bead as as function of the bead height.

and, at the same time, low preload force. A T-section (see Figure 7.387), installed between the valve cover and the cylinder head, is used to contain the oil splashed by the valve train. The vertical compression of the section generates the sealing pressure at the base of the groove and at the dual sealing lip towards the cylinder head. The section is designed for acoustically decoupled systems and has two blocks at the side that prevent direct contact between the cover and the head. The tension-induced relaxation of the elastomer material reduces the sealing force of the compressed profile over time; this has to be taken into account during design work.  alculating the component system C The cylinder head gasket forms the link joining the engine block and the cylinder head and, working in conjunction with the head bolts, forms the sealing system. To analyze the sealing system, one requires—in addition to the geometric descriptions of the component in the form of an FE model, the material properties, and the sealing characteristics—information on the temperature distribution in the components and the ignition pressure in the combustion chamber. An engine runs under a wide range of load conditions and always has

Deformation, measurement 0.01 mm/s Relaxed 30 min measurement Deformation, FEM with 2 mm/s Relaxed 30 min FEM

0

0,5

1

1,5

2

2,5

Deformation [mm]

Figure 7.387 Section through a t-profile in groove Calculating the force–deformation curve.

288 | Internal Combustion Engine Handbook

6606_Book.indb 288

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.21 Sealing Systems

250 Engine cold

Normal tension [N/mm2]

200

Engine warm (full load op.) Engine warm and gas force

150 100 50 0

Web 0

Bolt 45

Inlet 90

Bolt 135

Web 180

Bolt 225

Exhaust 270

Bolt 315

Web 360

Figure 7.388 Force distribution around the combustion chamber with rigid stopper.

Line along the bore [degree]

to be tight as regards gases and liquids. Extreme operational situations for the cylinder head gasket appear at full throttle with maximum coolant temperature and at cold start. Thanks to the bolt preload, the gasket is compressed to the height of the stoppers at the combustion chamber and, in other areas, locally to the thickness of the metal. The stopper acts like a wedge at the combustion chamber and places the components under elastic preload. The pressure on the stopper at the edge of the combustion chamber has to be greater than zero in order to ensure positive sealing in all operating states. Figure 7.388 shows a raised area on the exhaust side when subjected to ignition pressure; it has to be corrected by adjusting the stopper height in order to protect the combustion chamber bead against high sealing gap amplitudes. When the pressures at the stopper are too great, material overloads can occur at aluminum components, for example, causing damage to the component. The high temperature of the components at the combustion chamber further limits load-handling capacities. Acoustic decoupling of a component interrupts mechanical transmission paths by elastic mounting, between elastomer elements (Figure 7.389). Impinging on a valve cover are, on

Compression seal nominal 2,21 mm Decoupling element nominal 0,99 mm Total tolerance compression 2,10 mm Tolerance range +1,45/–0,75 mm

Decoupling system shroud 1800

Relaxed components Seal length 130 mm

1600

the one hand, the sealing forces between the cylinder head and the valve cover and, on the other hand, the forces at the bearing point, i.e., the decoupling element. The decoupling system (see Figure 7.381, Section 7.21.3.1) comprising the cover, gasket, and several bearing points is preloaded with bolts and spacer bushings. If the deformation characteristics for the gasket and the decoupling element are known, then one can determine the working point at a given preload. Because all the components exhibit certain manufacturing tolerances, the actual preload in a system deviates from the design value. Calculations for the sealing profile ascertain the smallest permissible deformation at sure sealing; this is then specified as the minimum sealing pressure. In this way, the lowest seal compression force required for operations can be ascertained. The system’s maximum preload force is limited by the loadbearing capacities of the decoupling elements; tolerable stress levels in the elastomer may not be exceeded. Within these limits, the system is operationally reliable and can be fixed for a working range by fine-tuning the preload and tolerance situation. The objective is to work with the lowest possible forces and thus to minimize deformations of the components.

Compression 4.65 mm/Force 1650 N

Force [N]

1400 1200 1000

Decoupl. element relaxed Seal relaxed Force max. Force normal Force min.

800 600

Compression 3.20 mm/Force 385 N

400

Compression 2.45 mm/Force 240 N

200 0

Minimum sealing pressure at seal, compression 1.4 mm/Force 194 N

0

1

2

3

Decoupling element

4

5

6

7

Verpressung [mm]

8

9 Seal

10 Figure 7.389 Decoupling system with work space from component tolerances.

Internal Combustion Engine Handbook | 289

6606_Book.indb 289

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.21.4.2 Simulation in the Laboratory—Testing Functions and Service Life Real stresses are simulated under lab conditions depending on the seal version by conducting, for example, tests to determine media and temperature resistance, durability, adaptability, setting behavior and sealing effects. Commonly used laboratory testing procedures employ servo-hydraulic testing equipment to perform hydraulic combustion pressure simulation to test head gaskets, shakers, and temperature controlled chambers for assembly tests. Hot gas generators are used to test items associated with the exhaust system.  ervo-hydraulic testing equipment S Servo-hydraulic testing equipment is employed for quasi-static, thermal quasi-static-and dynamic testing. Quasi-static tests, which can also be conducted using electromechanical testing equipment, provide insights into the compression and resilience properties of seals and gaskets. Thermal quasi-static tests are used to examine the durability and creep characteristics of sealing materials when subjected to pressure and temperature. Dynamic tests used to preselect and examine the seal design are of significance, particularly for layered metal gaskets (Figure 7.390). The gasket area surrounding the combustion chamber is chucked between metal flanges and is loaded repeatedly for a prescribed number of cycles (e.g., 107) at a given frequency, at constant force amplitudes, or, preferably, at constant distance amplitudes. The objective is to determine the maximum permissible oscillation amplitude compatible with gasket durability. The clamping flanges can be formed to exhibit a defined surface quality (roughness, porosity) so that compression tests can be conducted to determine the minimum sealing pressure necessary to achieve a satisfactory seal.

Figure 7.390 Head gasket test at the servo-hydraulic test stand.

 ydraulic simulation of internal pressure H Operating on the basis of the tests conducted at the servohydraulic test stand, one uses dynamic internal pressure simulation (Figure 7.391) to test the sealing system as a whole, under conditions closely approximating reality.

Figure 7.391 Dynamic internal pressure simulation with original engine components.

For this purpose, the head gasket is installed between the original engine components (engine block, cylinder head). The individual combustion chambers are then “fired” hydraulically, in the normal firing order, using fast-acting servo valves. Temperature cycles are run through, superimposed on the application of internal pressure, using a media circuit connected with the engine’s water jacket. The interplay between component stiffness and gasket design is evaluated by measuring the dynamic variations in gap width that occur. Weak points in the components can be identified at an early stage in development; optimization work for the seal design can thus be conducted before the start of engine testing proper.  ervice life testing S These test procedures are used to examine the long-term properties of seals, seal materials, and modules.These are, in the main, tests of elastomer materials and plastics (modules). Exceptions are examinations of the setting behavior and pressure resistance of elastomer sealing materials. In normal operations, elastomer seals and plastics are subject to aging over time, which does not occur during the pressure tests, brief thermal shock tests, and thermal conditioning procedures that are normally employed. To ensure full functioning of the module over its entire service life, simulation tests have been devised that both take account of the loading conditions found during vehicle operation and allow for reasonable testing periods. To do this, it is necessary to include temperature, media, and pressure loading in a testing program. This is done by connecting external media circuits (oil and coolant) to the test specimen and/or by exposure in a temperature chamber. With these tests, which imitate the engine operating states for temperature, one can simulate within a period of 2000 h, the loading corresponding to about ten years of vehicle operation. If the influences of oscillations and vibrations are to be assessed, a test of this type may also be conducted using an appropriate vibration generator.

290 | Internal Combustion Engine Handbook

6606_Book.indb 290

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.22 Threaded Connectors at the Engine

 ibratory testing systems V Engine components and modules are subjected during operation to mechanical vibration loads due to the influences of the road surface and direct vibration induced by the engine. Similar dynamic loads can be imposed upon the component examined using so-called “shaker” units. Both hydraulic and electrodynamic shakers are used, the latter being more common. A combination of a sliding-top table and a vibration loading along the vertical axis makes it possible to test horizontal vibration loading also. Mechanical systems make it possible, where required, to impose vibration loads along several axes. Acceleration sensors are used to register component vibrations at the test specimen so that testing can be carried out specifically in the critical oscillation resonance range. In this way, fatigue phenomena at the test specimen can be examined with a considerable “time-lapse” effect.  ot gas simulation H The thermal loading at the components, and thus at the sealing points in the exhaust system, can be simulated with hot gas generators. These deliver defined exhaust gas flows at constant temperature, burning heating oil, diesel fuel, or natural gas to do so. To achieve great component deformation, such as is found at the exhaust manifold during engine operation, the specimen is subjected to a thermal shock series in which hot gas and cold ambient air are passed through it alternately. The sealing function can be examined by pressure tests at room temperature (before and after the test series). This is, however, not a significant restriction for evaluating the gasket because it is particularly at low temperatures that the loss of bolt force due to thermal expansion in the mounting system comes fully to bear. When it is necessary to take account of dynamic influences, as well, the hot gas generator can be combined with a vibratory testing system. Either electrodynamic shakers or servo-hydraulic systems may be used, depending on the task at hand and the design of the specimen.

7.22 Threaded Connectors at the Engine 7.22.1 High-Strength Threaded Connectors

The basic modern engine contains between 250 and 320 threaded connections, which use from 80 to 160 different types of screws and bolts. The number of threaded connectors depends primarily on the engine configuration (e.g., four-cylinder inline or V-6 engine) and less on the combustion system (diesel or gasoline engine). Engines developed in Japan, when compared with European designs, have about 15% more threaded connectors per engine and, at the same time, fewer different screw designs. The size and number of bolts and screws increases with the displacement and number of cylinders. Mass production among European car makers, in particular, has been heavily automated in the final assembly area since 1983. The front-runner here was VW with its “Hall 54” at the

Wolfsburg assembly plant for the production of the GOLF III, which had just gone into production [7-192]. To accomplish this, it was necessary to design screws and bolts suitable for automatic feed, installation, and tightening. Engine construction involves high-precision component manufacturing; the manufacturing tolerances for the basic units (e.g., cylinder block and head) are very close and the positioning accuracy for operating equipment and robots is better than 0.5 mm. In fully automated assembly lines, the connector elements are moved by feed systems to the installation point; the bolts are screwed in and torqued down by a single or multiple power driver at an automated bolting station, necessary if only to absorb the reaction torque. Full automation does not make sense if many different engines are built on the same assembly line. With the further development of electrical control systems and ergonomic designs, hand-held power screwdrivers with integrated electronics (torque and rotation angle sensors) are used even more to monitor or control the tightening phase [7-194]. This lowers the investment and maintenance costs for the assembly line and increases flexibility, moving toward “joint production systems.”

7.22.2 Quality Requirements

If defects occurring while installing thread connectors are not detected, then there will be disturbances in the production process. One may count on malfunctions in the assemblies delivered to the customer. The screw or bolt is normally to blame for the disturbance, although, in addition to screw quality, the tolerances and properties of the components being joined and the threading in the nut as well as quality in assembly operations can have just as much influence on the connection. Consequently, high-quality screws and bolts have to be used in automated systems. It is for this reason that reputed manufacturers not only make spot checks during manufacture but also often conduct a full test at the end of the manufacturing process, using automatic testing equipment. Thus, a full account is taken of the quality expectations held by screw and bolt users, with their “zero defects” targets. In practice, it is possible to achieve a reject ratio of less than 50 ppm, referenced to the major features examined, at screw sizes up to M14; up to this size automated quality control can be implemented without any technical problems. The most modern automatic machinery can process, depending on the scope and nature of testing, between 100 pieces per minute (mechanical testing) and 300 pieces per minute (optical testing). At larger dimensions, fully automated testing and the associated handling is often made uneconomical by the screw weight and size with the result that visual checks are made, usually combined with another step (such as hanging the parts on racks for surface finishing or when packing the parts). In conjunction with manufacturing using reliable processes, in which only random errors (occasional defects referenced to annual production volumes, at long intervals

Internal Combustion Engine Handbook | 291

6606_Book.indb 291

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• cylinder head bolt

and at low rates) and no individual defective parts appear, defect rates of less than 50 ppm are achieved as a rule and otherwise less than 300 ppm. This degree of process reliability has been achieved in recent decades due in no small part to consistent introduction of Technical Specification ISO/TS 16949: QM systems: Specific requirements achieved with the application of ISO 9001 [7-195] in the operations. Thus, the defect rate in manufacturing could be reduced from 1,000 ppm to 300 ppm without undertaking any further efforts. To avoid mixing parts later and to satisfy the demand for freedom from foreign parts, this test is made immediately before packing. The products are filled in special containers (small load carriers) or in clear plastic bags and then sealed. Another option, even though seldom used, is to have the screws and bolts tested at the user’s site. A design proposal for screws amenable to assembly is shown in Figure 7.392. Experience has shown that there are difficulties if the bolts that are installed are drawn from a mixture made by different manufacturers unless exact specifications have been imposed in regard to the material, the offset strength, and friction values. It is often necessary to set up the system anew following a change of suppliers [7-196], [7-197], [7-198].

•• main bearing cap bolt •• conrod bolt •• belt pulley bolt •• flywheel bolt. In addition, the following threaded connections can be problematic. They need not be characterized as critical from the applications technology viewpoint but may be among the major applications in the engine: •• camshaft bearing cap bolt •• oil pan fixing bolt, valve cover fixing bolt. Screw connections for subassemblies and flange mounting points are not discussed further at this point, with the exception of threaded connections for magnesium components. High-strength screws upwards of M6 in size are used in most of these cases, and these are largely either standard designs or close to standard designs. 7.22.3.1 Cylinder Head Bolt The function of the cylinder head bolts is to make an operationally reliable connection for the complete system, comprising the cylinder head, cylinder head gasket, and engine block, over long-term operations, taking the maximum possible ignition forces into account. The primary goals are uniform, low component loading and tight seals against combustion gases, lubricants, and coolant.

7.22.3 Threaded Connectors

At the engine there are generally five critical threaded connection areas; these are explained below: dc dpr 120°±1°

S

Internal torx

1,5d

a

c

l

30

6g 6h for heavy corrosion protection

la

(X) A

b

°

15°

lges

Rounded thread run-out

dr

Shank-loaded bolt: dc ‡ lges + 2mm (mm)

35°+10°

Head-loaded bolt: dc > lges + 2mm (mm)

lz

Thread tolerance position

d

dc

T (Z) B

dw dr

dw

B

0 to 1°30 '

R

0,2+0,1

ds

k

at , ng ati on e S xag ic A he on tc ga n i r nte Ce lar l o c Transition rounded or conic

90° ± 1° Head support selectable

kpr

Pressing contour

dz

(Y)

90° ± 2°

Figure 7.392 Design proposal for screws amenable to assembly [7-192].

292 | Internal Combustion Engine Handbook

6606_Book.indb 292

1/19/16 8:36 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.22 Threaded Connectors at the Engine

While in the past, cylinder head bolts had to be retightened once or even twice to compensate for gasket setting, a cylinder head configuration requiring no retorqing is state of the art today. This has been made possible by using waisted-shank bolts or waisted-thread bolts with great elasticity, closer tolerances for tensile strength and friction properties, cylinder head gaskets that resist setting (e.g., all-metal gaskets), and a tightening process with low scatter in the values for preload force. Rotation-angle controlled (turn-of-the-nut) tightening to beyond the elastic limit has established itself as the most common torquing process. Lightweight engineering is promoted more vigorously, and the resulting reduction in component stiffness at the engine block and cylinder head is normally compensated for by reducing the maximum screw strength. The minimum required screw force can be maintained only with a drastic reduction in the tolerances for tensile strength and friction values. When designing the cylinder head bolting constellation, it is necessary to understand the influence of temperature. It is conceivable that, while the engine is heating up, the cylinder head bolts heat up more slowly than the cylinder head and engine block that they join. There may be a considerable rise in the preload force if components such as aluminum, with higher coefficients of thermal expansion, are used for the latter. Considering this aspect, too, the use of waisted-shank bolts or waisted-thread screws (Figure 7.393) is advantageous since, by virtue of the lower rise of the spring characteristics, the increase in screw loading is significantly less [7-192], [7-200].

around € 190 /to (previously this figure was around € 80 / to). Owing to these two factors economical manufacturing has not yet been implemented. The ever-increasing pressure to reduce costs has been responded to in two areas when optimizing the cylinder head bolt: •• Using waisted-thread screws as a compromise between sufficient elasticity and reduced manufacturing costs in comparison to waisted-shank bolts requiring a significantly more complex manufacturing process. •• Replacing the washer even in aluminum cylinder heads by integrating its function into the screw head, in the form of a bolt with a flanged head. To avoid seizure during screw assembly, it is necessary to impose narrow limits on the geometry of the contact surface under the head and to select manufacturing technology that adheres to those limits. This includes surface treatment with extremely low variation in the friction values and excellent adhesion to the substrate material as is found, for example, in the thin-layer phosphatizing process with quasi-amorphous crystal formation. 7.22.3.2 Main Bearing Cap Bolt The main bearing cap bolts connect the main bearing caps with the engine block at the crankshaft bearings. As a rule, two such bolts are used for each main bearing cap; these are usually fully threaded, collared bolts and may be used with washers. Figure 7.394 shows the installation situation for such a connection and the associated force flows. Where lk = clamping length, l′ k = plate thickness, FB= operating force. lk FB

FB lk′

Figure 7.393 Waisted-shaft or waisted-thread screws for cylinder head bolts (KAMAX Company).

The expansion properties of steel can essentially only be influenced by alloying with nickel. Consequently, the latest developments provide for cylinder head bolts made of austenitic materials whose coefficients of thermal expansion are similar to those of aluminum. An as yet unsolved problem is the high degree of tool wear resulting from this material’s great strength. Furthermore, the alloying premiums for steel have risen sharply since the beginning of 2004 and are now trading

Figure 7.394 Installation situation and force flow at the main bearing cap bolt.

The critical problem when designing this configuration is the tight installation space available for the bolt head in most instances. Very close attention must be paid to maintaining the permissible surface pressure for the rear of the bolt head and its mating surface. Every main bearing cap bolt is installed

Internal Combustion Engine Handbook | 293

6606_Book.indb 293

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

twice: the first time for machining the bearing shell seat to press-fit dimensions and then again after assembling the crankshaft and positioning the bearing caps. In the second assembly cycle, seizure may occur at the threads if the bolt exhibits damage such as impact dents at the tip or start of the threading. This is avoided preventively in screw engineering with ideal tip design and in manufacturing with the shortest possible drop heights (maximum 300 mm). The design of the tip is understood to include chamfering the start of the screw shaft before rolling the threads to ensure that the threads do not break out during rolling. At the start of the thread there appear only dull thread teeth that are not inclined to dent in response to impact. To increase the engine block stiffness, so-called ladder frames are being used more frequently in engines. They are used to interconnect individual main bearing caps. In this way, the lower section of the engine can be stiffened to avoid twisting and warping. Usually the bearing caps are cast in place in the ladder frame made of aluminum alloy. In this case the main bearing bolts are used to fix the complete unit in place. Tightening processes using the offset limit or the rotation angle as the lead variable have become the most common assembly techniques. 7.22.3.3 Conrod Bolt The conrod bolt represents a typical case for a threaded connection subject to high dynamic loading. The range of sizes in passenger car engines is from M 7 to M 9, for utility vehicle engines from M 11 × 1.5, M 12 × 1.25, M 14 × 1.5 up to M 16 × 1.5. To achieve correct dimensioning of the conrod bolt, one draws on data from the predecessor engine or for engines of similar design and size. Concerning the bolt for the large conrod eye, the operational loading on the bearing case due to the physical forces acting on the crankshaft system (masses and gas forces) are known. Not known at the outset, however, are the operational loads by size, direction, and location, referenced to the bolt centerline in the parting plane and introduced into the individual threaded connection; this information is needed to ascertain the deformations and loads for the bolt. The professional literature [7-201], [7-202] mentions various analytical procedures used to calculate the axial force FA, the lateral force FQ (calculated magnitude derived from the friction value in the parting plane), and the eccentric distance a for the axial force from the screw centerline, dependent on the design parameters of the conrod bearing case. If these values are available, then it is possible, using software programs such as “VDI Schraubenberechnung,” “SR1” or “KABOLT” (screw calculations as per VDI 2230: [7-203], [7-204], [7-205], [7-206]), to determine the preload value required to prevent partial liftoff and lateral shift of the connected components, and thus, to ascertain the appropriate thread dimensions and the strength class for the bolt. The determined values are used to designate the specifications for bolt tightening. Once the design calculations for the conrod joint have been concluded, pulser tests are used for the entire connecting

rod to demonstrate durability. Subsequently, the calculated and laboratory results are verified with testing in the field. The calculation parameters for a conrod bolt connection are shown in Figure 7.395. The example refers to a four-cylinder gasoline engine with displacement of 1996 cm3. The conrod bolt design is based primarily on the loading and on the assembly of the conrod. Depending on whether or not a nut is used, the bolts are equipped with heads shaped to accommodate torque transmission or with antirotation devices. The two halves of the conrod are centered with knurling, a fitting bushing, or a separate spline. Large conrods in utility vehicles often use interlocking areas following the tongueand-groove principle. The use of a sintered conrod makes good sense when a particular model is manufactured in medium-range numbers. While in conventional manufacture, the large conrod eye is cut away after machining in order to mount the conrod bearing shells. In recent years, “cracking” has established itself in large-volume manufacturing of sintered, cast and forged conrods. Here the conrod end is separated from the conrod shaft in a device that applies a defined, external load to areas laid out to promote fracture. The advantage, in addition to eliminating the cutting work, is that the two halves of the conrod are selfcentering. Then the fracture surfaces (postassembly) can be used to permit turning out the bearing shell seats. That is why a cracked conrod does not require an exact fit for the shaft at the conrod bolt. Here the screw diameter may exhibit a tolerance of 0.1 mm. Each conrod is assembled twice after cutting. The first time is in preparation for machining the seats for the bearing shell. Here the preload force used for assembly must be similar to that found later during operation, so that similar deformations are induced in the conrod bearing housing. It is for this reason that the bolts are tightened to just below the offset limit under torque and rotation angle or under direct offset limit control. The conrod is disassembled after machining (to insert the bearing shells) and is then mounted on the crankshaft. Here a rotation angle controlled tightening process is used, which tightens the bolt into the range beyond the elastic limit; alternately, tightening under offset limit control is employed. If one decides in favor of the rotation angle as the control magnitude, then it is necessary to conduct extensive laboratory trials in advance in order to formulate specifications for tightening. When using the 0.2% offset limit as the lead magnitude, it is sufficient, in a few tightening trials, to define the so-called “window.” Particularly because the conrod bolt, because of the manufacturing process for the conrod, has to be assembled twice and tightened into the offset limit range, one must ask which screws are particularly suitable for tightening beyond the 0.2% offset limit [7-211]. When dimensioning threaded connections, it is necessary to remember that the threaded section, in the event of overloading due to static tensile forces, breaks at its weakest point. This is normally the case in the nonengaged threaded section or in the waisted-shank area. In the multiwaisted bolts

294 | Internal Combustion Engine Handbook

6606_Book.indb 294

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.22 Threaded Connectors at the Engine

Characteristics: Linear calculation set (may be calculated non-linear: Parting line surface rectangular, force engagement eccentric) Expanding bolt thread tempered, no washer Phosphated surface Torque tightening

Friction value Tightening factor

KABOLT definition: Other bolt with cylinder shank MB x 1 x 43 – 11.9 (Rp0,2 = 1035 N/mm2)

0

S α

d1

l1

a¢

d2

l2

Bolt section:

FA

lf

3

lrG

lK

d4

7,75 d1 3,0 l1 6,5 d2 l2 14,0 8,3 d3 6 l3 7,35 d4 3,55 l4

l4

n · lK

rm

d3

l3

2

Nr. 1 Diameter Length Nr. 2 Diameter Length Nr. 3 Diameter Length Nr. 4 Diameter Length Freely selected length

FA 2r 0

d

Loads and characteristics Operating force Operating force Req. residual clamping force Extension limit Tightening factor Friction, thread Friction under head Flange material

FAo FAu

= 10600 N = 0N

FKerf Rp0,2 aA mG mk

= = = = = =

– 1035 N/mm2 1,18 l 0,10 l 0,14 l C 45

Geometry: Bolt diameter Thread pitch Head support diam. Friction diameter Compressed bush Clamping length Force introduction factor Force introduction factor

d p dw dkm DA lK n nred

= = = = = = = =

8,00 1,00 12,70 10,60 13,18 28,55 0,50 0,45

mm mm mm mm mm mm l l

Bore diameter Bore chamfer diameter Eccentricity, bolt Eccentricity, force Length Länge Breite Conrod bore hole gauge Radius conrod bearing eye Distance of both bolt centers Clamping angle

dh DF s a cT cB b

= = = = = = =

8,50 9,20 ± 0,48 2,70 13,20 12,70 33,10

mm mm mm mm mm mm mm

l l1 l 2 rm

= = = =

132,00 31,50 100,50 27,00

mm mm mm mm

2r a

= 62,00 mm = 23°

Figure 7.395 Conrod bolt connection ratios [7-204].

Internal Combustion Engine Handbook | 295

6606_Book.indb 295

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

recently developed, the failure is also in a waisted area. The conrod bolts shown in Figure 7.396 are particularly suited for tightening into the range beyond the elastic limit.

Figure 7.396 Conrod bolts for tightening beyond the elastic limit [7-211].

When using bolts with a shaft (similar to DIN EN 24014), there should be at least six nonengaged turns in order to distribute plastic elongation over a larger area and, thus, to avoid the hazard of premature narrowing. The best tightening properties in the range beyond the elastic limit are demonstrated by waisted-shank bolts and screws that are threaded right up to the head (similar to DIN/EN 24017). The measured flexibility places the multiwaisted bolts between the waisted-shank bolt and the screw that is threaded along its full length. The durability of threaded connections is determined exclusively by the magnitude of local stress concentrat1ions. In bolt materials the fracture strength of the notched area compared to the smooth rod should, as a rule, be greater than 1, indicating a material with sufficient ductility. Permissible in high-strength bolts are durability values in the pulsating tensile range of σ A = +/−55 N/mm2 [7-207]. Screw durability is increased if the threads are rolled after annealing. The additional dynamic forces resulting from dynamic operational forces (which are absorbed by the screw) are lowered (in connection with eccentric loading) as the preload force level gets higher. This, too, favors a tightening process that goes beyond the elastic limit. 7.22.3.4 Belt Pulley Bolt The belt pulley is secured with a bolt at its center. Often mounted on the crankshaft in addition to the belt pulley are a gear for the oil pump drive and possibly the vibration damper. The inside bore of the belt pulley is mounted on the crankshaft end journal. The large bore diameter in the belt pulley makes it necessary to create a positive connection between the bolt and the pulley with a large washer or a large-diameter bolt collar. Often an M 12 bolt is fitted with a washer or collar diameter of up to 38 mm (in gasoline engines) or an M 18 bolt with a collar diameter of up to 65 mm (in a diesel engine with 2.5

liter displacement, for instance). The pulley is press-mounted separately on the crankshaft journal or is pulled onto the crankshaft by the screws at a previously defined tightening torque. In utility vehicle engines, sizes of up to M 24 × 1.5 are used and the washer is positioned just before assembly. In large utility vehicle engines, the pulley is seated on the vibration damper and, passing through oversize bores, is bolted directly to the crankshaft with six or eight screws or lug bolts (e.g., M 10). In the past, belt pulley bolts were tightened down with only a torque wrench. Today, the rotation angle technique has become more or less standard. Tightening takes place through a snug-tight fit until all mating planes are seated firmly one on the other. The bolt is then turned down further, the amount based on a measurement of the rotation angle. Extremely high ultimate tightening torques are achieved in this way. When using an M 12 × 1.5—10.9 bolt, torques of up to 260 Nm can be achieved, while the ultimate torque calculated theoretically lies between 120 and 150 Nm. The great spread in the ultimate torqueresults from the large head contact area, which causes “seizure” if there is even the slightest misalignment. A tightening technique based on the 0.2% offset limit cannot be applied for the pulley bolts when there is an extremely large screw head collar diameter or where several components have to be bolted together with the result that there are many mating planes between the parts to be joined. Manufacturing inaccuracies and unavoidable grime results in greater setting and a connection that is so flexible that the 0.2% offset limit point is sensed not only for the bolt but also for the connection. 7.22.3.5 Flywheel Bolt Because of the engineering design there is a relatively small pitch circle at the crankshaft. During assembly it is necessary to ensure that there is sufficient clearance between the bolts to accommodate the tightening tool. The bolts are all tightened simultaneously, using a multispindle tool and using the 0.2% offset limit as the control variable. This is also done because shorter clamping lengths (e.g., 7 mm) are present. Because of the narrow clearance between the crankshaft journal and the flywheel, the bolt heads are not as high as the standard heads. To be sure that the required torque can be applied safely and positively, a twelve-pointed (bihexagonal) head or a hexalobar head or, if necessary, an inside, multispline socket is used at the head. When oil is supplied by runners inside the crankshaft, the bolts used to seal against oil leaks are provided with a microencapsulated sealing adhesive or with an all-round nylon coating. Some engine manufacturers still tighten down flywheel bolts under torque control, and then snug them down manually. In the dual-mass flywheels that are used more frequently today, the module is delivered to the vehicle manufacturer complete with the bolts and is then assembled as a unit. The bolts are tightened down with a multispindle power driver, through bores in the clutch plate spring and the clutch disk.

296 | Internal Combustion Engine Handbook

6606_Book.indb 296

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.22 Threaded Connectors at the Engine

7.22.3.6 Camshaft Bearing Cap Bolt This threaded connection usually uses collared screws that are similar to the standardized styles, in sizes of M 6, M 7, and M 8 for passenger car engines and M 10 and M 12 for utility vehicle engines. Because during torque-controlled tightening, there is the hazard that differing clamping forces could be imposed on the camshaft bearings, the rotation-angle-controlled technique is used to a greater extent to achieve defined preload values. A special technique used by some passenger car makers is to screw grub screws into the cylinder head; the bearing cap is then positioned and fixed with nuts. 7.22.3.7 Oil Pan Attaching Screw The oil pan is also secured to the engine block with collared screws similar to the standardized styles (Figure 7.397; outside or inside socket wrench application, M 6, M 7, M 8 screws). To achieve complete freedom from leaks, the surface pressure must be uniform across the entire oil pan gasket. This is achieved with the smallest possible screw diameter that thus exhibits appropriate flexibility and with a suitably large collar diameter or a washer, where the screw forces are introduced uniformly. In addition, many screws are needed so that when forces are introduced there are large overlaps in the “pressure cone” in the area of the seal. In spite of the great demands for tightness, this connection point is considered to be trouble free. The screws are, usually only tightened under torque control (using a multispindle power screwdriver unit). To prevent canting, tightening is started at the middle of the engine block, continuing outward from there.

Ribs on the lower surface of the engine block reduce noise propagation. The oil pan itself is a source of high noise emissions because of its large surface area and low weight. Here the solution involving structural noise decoupling by the mounting screws used for the oil pan is a viable option. Widespread use has not been implemented, however, due to the costs involved. In order to cut costs, the market is exhibiting a trend towards preassembling connector elements in system assemblies such as oil pans, valve covers, timing belt covers, etc. Here leading screw manufacturers are working on economical and spacesaving preassembly solutions such as, for example, using self-tapping screws, which are preassembled by forming an inside thread in component to be secured. Another solution, which at the same time could serve the interest of acoustic decoupling, employs a plastic bushing that is slipped over a special screw and is then premounted together with the component (Figure 7.398).

1 2 3 4 5

Figure 7.398 Oil pan screw with inside hex lobes and collar (KAMAX Company).

7.22.4 Threaded Connections in Magnesium Components Structural noise-coupling fixing elements before and after assembly (right): 1 Connecting element 2 Decoupling element 3 Housing cover 4 Seal 5 Housing

Figure 7.397 Oil pan screw with structural acoustic decoupling element (depiction before [left] and after [right] assembly) (KAMAX Company).

The trend toward lightweight construction in automotive engineering, prevailing for years now, requires not only optimization of the components made from proven materials such as steel, aluminum, or plastic, but also the use of alternate materials such as magnesium. The advantage of magnesium is its relative stiffness even where cavity walls are thin. It is comparable to the density of plastic. In engine design, this material is now used only in secondary assemblies such as the cylinder head shroud for encapsulated engines or for the air filter body intake tube. Here magnesium replaces plastic. In the engine block itself, the thermal loads

Internal Combustion Engine Handbook | 297

6606_Book.indb 297

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

are too high for connection points as a whole. Thus, steel screws in conjunction with magnesium can be used only at room temperature due to the setting and relaxation properties. Therefore in the engine component area heat-treated aluminum screws are used, made from AL 6056 in conjunction with the die-cast magnesium alloy AZ 91, AS 21 up to 120°C (maximum temperature: 150°C), for example. When using magnesium components, contact corrosion properties in conjunction with steel or aluminum screws have to be taken into account [7-208], [7-209].

7.22.5 Screw Tightening Process

When selecting the assembly and tightening technique, one must remember that passenger car engines are manufactured in large numbers; utility vehicle engines, by contrast, are built in short production runs or even individually [7-210]. 7.22.5.1 Torque-Controlled Tightening Torque-controlled tightening is normally used only for secondary applications (minimum preload force need not be exactly defined). It is employed only in demanding applications (such as mounting the belt pulley) in automated assembly lines. It continues to be used in service work. The problem is that the preload value applied under torque control has to be selected so that in the worst case (i.e., smaller actual coefficient of friction than what was estimated when establishing the torque level) the 0.2% offset limit is not exceeded as otherwise the screw would be stretched. Preload is the force that is present in the threaded connection after the completion of the assembly. At

Figure 7.399 Aggregate assembly using hand-held power screwdrivers (Image: Atlas Copco Tools) with integrated torque and rotation angle sensors (Atlas-Lopco).

a very high actual coefficient of friction (higher than what had been assumed), the preload value is very low. Consequently, the properties of the screw cannot be fully exploited with this technique. Screw and bolt manufacturers and the automotive industry have agreed upon the coefficients of friction to be expected. These lie between μ total = 0.08 and 0.14. They are a component in the quality agreement in each case and are spot checked for each batch of screws at a friction value test device [7-211]. A special form of torque-controlled tightening is the combination with “snugging down”; once the tightening phase is completed, the connection is retightened with a torque wrench. This technique is used at the remaining manual assembly stations in mass production and for all critical connector elements in short production runs. When manual connection is used, a torque-controlled pneumatic screwdriver is employed to tighten the screw or bolt down to the specified moment, then snugged down using a measuring torque wrench and usually marked with paint. The torque required to restart rotation is the snugging moment (also called retightening torque). Experience has shown that snugging usually goes beyond the adjustment value for the wrench with the result that an indirect, rotation-anglecontrolled tightening process is often used. 7.22.5.2 Rotation-Angle-Controlled Tightening When tightening the nut using the rotation angle as the lead variable, through to the 0.2% offset limit, the preload value is on average from 25–30% higher than for torque-controlled tightening. While in torque-controlled tightening the preload force varies by about ±25% (practically to the same extent as the friction), the preload force where the rotation angle or 0.2% offset limit is used as the lead variable varies by only about ±10%. When tightening using rotation control, the spread in preload is dependent on friction only in the range up to the snugtight torque. The snug-tight torque is the moment that has to be applied until, by tightening the connection, all the mating surfaces are seated solidly one against another due to elastic and plastic deformation. The spread results primarily from the differing 0.2% offset limits for the bolts, provided that the required repetition accuracy when approaching the set angle is achieved. This is the case in today’s pulse transducers. Beyond that, we see from the progress of the curves above the 0.2% offset limit that angular scatter has only a subordinate influence on assembly preload (Figure 7.400). Torque monitoring is used to ensure quality in the connection. We see that when tightening under torque control, the minimum preload force Fm lies between 48 and 57 kN. When working with yield point (0.2% offset limit) control, this value is between 67 and 85 kN while rotation angle control yields between 77 and 94 kN. Consequently, tightening under torque control gives the greatest spread in preload force at the smallest preload level. The preload force level when using turn-of-thenut tightening is on average about 10% greater than that for the yield point technique. The area around the Rp0.2 points represents the window for tightening under yield point control. The switch-off point

298 | Internal Combustion Engine Handbook

6606_Book.indb 298

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

∆FM = 17 kN u-Breakaway

Rp0,2-Point

∆FM = 19 kN MA-Breakaway Joining moment

0 10

mG

=m

K

0

,1 =0

50

mG

= mK

Nm

,14

=0

100

130

50

Rp0,2 = 1100 N/mm2 Rp0,2 = 940 N/mm2

umax at tightening under torque

kN 50 40 30 20 10 0

mG = mK = 0,10 mG = mK = 0,14 mG = mK = 0,10 mG = mK = 0,14

100

100

MA max at tightening under torque

Assembly pre-tensioning force Fm

7.23 Exhaust Manifold

0

Breakaway torque MA

50

grd

100

150

Tightening angle u

Figure 7.400 Tightening curve for a screw as per DIN EN ISO 24014—M 12 × 1.5 × 70—10.9 for control by torque (left) and control with rotation angle and 0.2% offset limit (right), illustrating the influences of thread and head friction as well as screw strength.

for the power driver has to lie within this area so that the threaded connection is registered as “OK” and can receive the paint marking, if that is specified. When dimensioning threaded connections, it is necessary to remember that the threaded section, in the event of overloading due to static tensile forces, breaks at its weakest point. This is normally the case in the nonengaged threaded section or in the waisted-shank area. Using the turn-of-the-nut process (as a process that goes beyond the elongation limit) is not critical in screws and bolts where the shank length is greater than 2 × d or there are more than 10 turns of nonengaged threading. In that case, even tolerances as great as 20° are acceptable when specifying the rotation angle used for tightening. In a threaded connector with a pitch of P = 1.5 mm, turning the screw by 30° beyond the 0.2% offset limit induces plastic elongation of about 0.125 mm. Referenced to the 60 mm effective clamping length, this represents permanent deformation of 0.21%. This value is not critical. Conversely, when using short screws (<2 × d shank length) the angle of rotation has to be specified so precisely, that the switch-off is close to the offset limit , particularly because today screws are often tightened into the offset limit range. The rule of thumb is that in these cases, referenced to the grip length, a maximum of 1% permanent deformation is acceptable. It is necessary to note here, however, that if the screw is tightened several times, then the head contact surface and the engaged threads can be damaged and thus tend to scuff and possibly seize. The required preload force cannot be attained in this event. A further advantage of rotation-controlled tightening is its reproducibility even when using simple tools; consequently, it is a favorite technique for initial tightening on the assembly line and for service work. 7.22.5.3 Offset-Limit-Controlled Tightening When compared with the rotation-controlled technique, this process offers the advantage that it always approaches the real 0.2% offset limit for the particular screw being installed. This process is used only to a very limited extent and in those situations where greater setting effects are expected during

and shortly after tightening. The permanent elongation of the bolt each time it is tightened lies between 0.1% and 0.2% (the exact amount depending on the sensitivity of the power driver system) and thus below the yield point. Unacceptable permanent elongation of the screw or bolt beyond the offset limit is hardly possible. In comparison with tightening under rotation angle control, the mean preload force value is 4%–7% lower. Quality assurance for the connection is affected by monitoring the window. This window specifies the power driver’s switch-off point (defined by specifying maximum and minimum angle and torque values) within the tensile yield range of the bolt.

7.23 Exhaust Manifold Economical cast manifolds were the standard in vehicle engineering for many years. Only in sportier vehicles, in the interest of optimizing torque and performance, were single-walled tube-runner manifolds used. They enabled individualized runner lengths, diameters, and configurations. Combustion at full throttle was largely substoichiometric with the result that exhaust temperatures were relatively low. In the mid 1980s legislators in Europe imposed pollutant emission limits, which made it necessary to equip the vehicles with catalyticconverters. As emission laws became more stringent, exhaust pollutants following a cold start had to be reduced further and more quickly. One of the options for rapid reduction was found by reducing the exhaust manifold’s thermal mass (or capacitance). In the cast iron version the mass for a four-cylinder manifold, at 4–8 kg, is quite high. If the exhaust manifold’s thermal mass is low, then the heat in the exhaust can bring the catalytic converter up to the so-called light-off temperature more quickly. The light-off temperature is defined as the exhaust temperature at which half of the pollutants are converted. Options for reducing the mass are presented in Sections 7.23.2 to 7.23.4. Figure 7.401 shows the influence of manifold design on the

Internal Combustion Engine Handbook | 299

6606_Book.indb 299

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

500

Exhaust gas temperature [°C]

100

Air gap insulated Manifold 2 mm Cast manifold Driving speed

80

400

60

300 40

200

20

100 0

Speed [km/h]

600

0

25

50

0 75 100 125 150 175 200 225 250 275 300 325 350 375 400 Cycle time [s]

temperature ahead of the catalytic converter when using the MVEG cycle. A further aspect that has had a negative effect on traditional exhaust gas manifold design is the increase in exhaust temperatures, resulting from the increase in the power density and operating with a stoichiometric fuel-air mix across wide areas of the engine map. Whereas in the early 80s we found exhaust temperatures of 850°C in gasoline engines and 650°C in diesel engines, these levels today have risen to beyond 1000°C in gasoline engines and as much as 850°C in diesel models. Especially in gasoline engines, this fact has a significant influence on the selection of the casting material. Earlier cast manifolds using silicon-molybdenum (SiMo) alloys reached their application limits at exhaust temperatures of up to 900°C. Higher-quality gray casting qualities containing 20–36% nickel can be used up to about 1000°C. To handle even higher exhaust temperatures, it is necessary to resort to nickel- or cobalt-based alloys like those that are also used in turbine engineering. Because cast manifolds are generally operated in the endurance strength because of a typical wall thickness of 4–6 mm (by comparison tube manifolds are h 1.0–1.8 mm), the changes in microstructure occurring at these temperatures and the inadequate thermal strength result in plastic deformation [7-217]. During the cooldown phase microfissures appear, and these lead to manifold failure in the long run. Neither have extensive studies on the development of new manifold casting materials resulted in a sufficiently improved service life [7-212]. One solution is to assemble the manifold from sheet steel or steel tube, which have a higher rupture life due to their design. Thus, there are examples in which cast SiMo manifolds were tested over 250 hours, while assembled manifolds for the same engine were tested under identical conditions for up to 500 hours. For maximum exhaust temperatures, diesel engines offer a better operating environment for cast materials. In response to new legislation, however, there are trends toward replacing, cast components with those made of sheet metal.

Figure 7.401 Influence of manifold design on temperature ahead of the catalytic converter.

Further arguments in favor of substituting sheet material for castings are efforts to reduce overall vehicle weight and ultimately to also reduce the great tendency for a cast manifold to heat up after the engine is shut down (Figure 7.402). The installation situation permits a very compact design using cast manifolds while sheet metal manifolds tend to take up more space, because of optimized runner lengths and minimum bending radii that have to be observed. When the various manifold designs are heated and cooled, we find that cast materials, in comparison to tubing and sheet metal, involve a high degree of thermal lag. An assembled design with air gap insulation lies between casting and tubing in regard to this factor. The need for heat shielding is determined primarily by the component’s surface temperature, the postheating properties, and the proximity of nearby components. Because the energy transmitted in irradiated area rises with the fourth power of the surface temperature, it makes good sense to shield cast and tube manifolds that can reach surface temperatures of up to 800°C. One very good alternative are Air Gap Insulated (AGI) manifolds; here the tubing carrying the exhaust gas is separated from the supporting structure by an air gap. These manifolds, which, by their very nature, incorporate their own heat shield and exhibit maximum surface temperatures of 550–600°C, generally do not require any additional shielding.

7.23.1 Manifold Development Process

The essential steps in manifold development are listed below: •• customer query for a desired manifold concept •• customer specification of the available installation space (may also include the geometry for the draft concept as well as that for the cylinder head flange, exhaust flange geometry, space available for power driver use, surrounding engine compartment geometries, etc.) •• specification of loading data (engine type and performance, vibration induction by the engine and/or road, exhaust gas temperature)

300 | Internal Combustion Engine Handbook

6606_Book.indb 300

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.23 Exhaust Manifold

Air temperature in 30 mm distance [°C]

250 225 200 175 150 125

Air gap insulated

100

Cast manifold

75

Manifold 1 mm

50

Manifold 2 mm

25 0

0

60

120

180

240

300

360

Time after engine off [s]

420

480

540

600

Figure 7.402 Soaking behavior of various manifold designs.

•• definition of the emission standard (EURO 4 or EURO 5 or some other norm)

Good acoustical properties due to high material damping properties

•• development of a detailed concept and the design created using CAE tools including, for example, heat transfer calculations, calculations for flow mechanics, or FEM calculations [7-214], [7-215]

Economical (€ 15–18 for SiMo casting).

•• construction of samples with tooling similar to that to be used in mass production •• certification testing at either the developer’s or the customer’s site •• customer’s production approval for the development •• construction of mass-production tooling •• test with mass-production components to verify the design •• production launch. As a rule, the overall period between the inquiry and production launch is about two years. Development work today is carried out in only a 14-month period; eight months are consumed in pure development time, and the remaining six months are required to build mass-production tooling and set up the manufacturing lines.

 isadvantages: D •• High weight •• The maximum permissible exhaust gas temperatures for cast material are limited. •• If, because of the extreme temperatures, the use of nickel alloys is necessary, then the price will rise up to € 100, depending on the Ni content typical price: € 35–40. •• Cast manifolds operate in a temperature range that can affect service life (bad for endurance, considering the engines’ higher performance densities and resultant higher temperatures). •• High surface temperatures (heat shielding required). •• More critical than tube-runner manifolds regarding emissions following cold start because of the manifold’s high thermal masses.

7.23.2 Manifolds as Individual Components 7.23.2.1 Cast Manifold  ypical materials: T Nodular gray casting (GGG), SiMo gray casting: Nodular gray casting using silicon-molybdenum (GGG-SiMo), SiMo gray casting with vermicular graphite, austenitic cast iron (GGV-SiMo) [7-217] Wall thicknesses: 7–8 mm for GGG manifolds

2.25–4 mm for chilled casting.

 dvantages: A Compact design Wide latitude in designing the shapes

Figure 7.403 Cast manifold for four-cylinder gasoline engine.

Internal Combustion Engine Handbook | 301

6606_Book.indb 301

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

•• Severe post-heating properties because of great thermal mass. •• Any desired or optimized runner lengths can be implemented to only a limited extent with cast material (performance optimization is limited). 7.23.2.2 Tube-Runner Manifolds

wall, from 1.8–2.0 mm. This problem can be countered by reducing wall thickness to a typical value of 1.2 mm. Selected designs now have a wall thickness of 0.8–1.0 mm [7-216]. •• Problematic acoustical properties due to low damping by the material. Additional measures may be required. •• Greater costs (€ 20–35). 7.23.2.3 Single-Wall, Half-Shell Manifold

Figure 7.405 Half-shell manifold for three-cylinder (diesel) engine.

Typical materials: Austenitic steels such as the type 1.4301, 1.4828, or 1.4841 alloys Figure 7.404 Lightweight tube type manifold for a four-cylinder gasoline engine.

Typical materials: Austenitic steels such as the type 1.4301, 1.4828, or 1.4841 alloys Ferritic steels such as the 1.4509 alloy or newly developed ferritic steels containing up to 14% chrome along with titanium and niobium stabilization (examples being SUS 425 Ti, LR 429 EX, and F 14 Nb) Wall thicknesses:  1.0–1.8 mm  dvantages: A •• Performance-optimized design is possible •• Low weight •• Standard steels that are readily available can tolerate high exhaust temperatures. •• Low postheating properties.  isadvantages: D •• More compact designs are possible but should not be implemented in four-cylinder engines because of performance considerations. Designs such as this are developed in some cases today to replace an existing cast design with a tube system occupying the same space. This, however, involved major problems in reaching the required durability levels, in addition to many other disadvantages. •• High surface temperatures (heat shielding required). •• When compared with the cast manifold, a tube type manifold is favorable for emissions at start-up. The situation can, nonetheless, remain critical if the manifold’s thermal mass is still relatively high due to choosing an excessively thick

Ferritic steels such as the 1.4509 alloy or newly developed ferritic steels containing up to 14% chrome along with titanium and niobium stabilization (examples being SUS 425 Ti, LR 429 EX, and F 14 Nb) Wall thicknesses: 1.5–1.8 mm  dvantages: A •• Economical (€ 15–20) •• Low weight •• Standard steels that are readily available can tolerate high exhaust temperatures. •• Low post-heating properties.  isadvantages: D •• Only very short runner lengths can be realized in a fourcylinder engine; the geometry of such a manifold is then typically very limited. •• The shape involves a great deal of cutting loss. •• Very long welding seams are required. •• High surface temperatures (heat shielding required). •• Critical acoustic properties (additional efforts may be necessary under certain circumstances, in the form of double-wall shells). 7.23.2.4 Manifolds with Air Gap Insulation (AGI Manifold) Separation of functions: Inside, there are lightweight components carrying the exhaust gases; outside are the load-bearing elements with greater material thickness. These internal components are decoupled by floating seats. In this way, it is easy to achieve durability in such a manifold.

302 | Internal Combustion Engine Handbook

6606_Book.indb 302

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.23 Exhaust Manifold

7.23.3 The Manifold as a Submodule 7.23.3.1 Integrated Manifold and Catalytic Converter

Figure 7.406 Jacketed manifold with air gap insulation for V6 gasoline engine.

Typical materials for the interior components carrying exhaust gas: Austenitic steels such as the type 1.4301, 1.4828, or 1.4841 alloys Typical materials for the load-bearing outside components: Austenitic steels such as the type 1.4301 Ferritic steels such as the type 1.4509 Wall thicknesses: Interior components carrying exhaust gas 1.0 mm; load-bearing outside components 1.5 mm  dvantages: A •• Relatively low weight and compact design. •• A design with optimized performance can be devised within a defined degree of latitude. •• Standard steels that are readily available can tolerate high exhaust temperatures. •• No high surface temperatures (thus nearby components can be positioned relatively close to the AGI manifold without further protective measures).

Figure 7.407 Catalytic converter near the engine, with welded, cast manifold (six-cylinder boxer engine).

Because a catalytic converter near the engine can be joined with the manifold using techniques such as welding or flanging, all the alternatives depicted at Section 7.23.2 are available for use in the manifold section. 7.23.3.2 Integrated Manifold and Turbocharger The manifold and turbocharger module, shown in Figure 7.408 is employed for both gasoline and diesel engines.

•• Low post-heating properties. •• Suitable for emission-optimized systems. The interior components carrying the exhaust gas have only a low thermal capacitance with the result that energy losses through to the catalytic converter are low; the outer components, with greater thermal mass, accept thermal energy only after the catalytic converter has reached full operating temperatures. •• If a performance reduction contingent on a driving cycle occurs in the start phase of the engine, then the surrounding support structure acts as an “insulation” to prevent rapid cooling of the catalytic converter. •• A concept with a water-cooled outside jacket is even possible [7-213]. •• Good acoustic properties can be attained with moderate effort.  isadvantages: D •• High costs (€ 40–66). •• In some cases, it is necessary to use high-pressure, internal reforming to achieve the complex geometries required while still taking up the least possible space; that means high costs and long lead times for the tools. •• Runners cannot be of any desired length.

Figure 7.408 Cast manifold with integrated cast turbocharger (diesel engine).

Compared with an assembly made up of individual components this module eliminates the masses of the flanges on the components while at the same time simplifying assembly. A clear disadvantage of this modular design is that the entire system has to be replaced even if just one of the components fails. Great costs are involved in unnecessary replacement of the turbocharger.

Internal Combustion Engine Handbook | 303

6606_Book.indb 303

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

If in this area, too, for the reasons already discussed, one opts for other types of manifold, then it is necessary to provide additional support at the engine block for the heavy turbocharger unit. Turbocharger housings made of sheet metal, which can be used to further reduce thermal capacitance and weight are also under development now.

7.23.4 Manifold Components

Components such as connector nipples for the exhaust-gas recirculation or the runners for secondary air supply, which until recently were contained in the manifold or the welded intake flanges, are more frequently being integrated into the engine block itself. Flange concepts for tube manifolds are shown in Figure 7.409. Used here are flange designs ranging from complex, heavy cast flanges with integrated secondary air feed through to very simple, lightweight, deep-drawn flanges made of sheet metal. In some cases, deep-drawn flanges exhibit, while carrying out the same functions, up to a 50% reduction in mass when contrasted with a comparable cast flange. By raising the edge of the deep-drawn flange, for instance, one can achieve the same stiffness characteristics as in the cast flange. The seal is achieved with greater surface pressure induced by beads stamped into the metal around the entry openings. Typically the thermal loads placed on the intake flange are low because of its contact with the relatively cool cylinder head. Consequently, economical, easily annealed standard steels such as the ST52-3 alloy can be used. Because of the higher temperatures at the exhaust ports, flanges of a similar design have to be made of higher quality ferritic or austenitic steels [7-214].

7.24 Coolant Pumps for Combustion Engines 7.24.1 Requirements, Models, and Constructive Design

The coolant pump guarantees coolant circulation in the coolant circuit according to requirements under all of the operating

Cast flange with secondary air intake

conditions and operating states of the combustion engine. High reliability, low operating energy input, and a cavitation-free operation should be ensured at the same time as delivering the smallest installation space and low costs. Single-stage, rotary centrifugal pumps are now used in the majority of passenger car coolant circuits. Their speed and consequently according to the similarity relations for coolant pumps, the pump volumetric flow, is coupled to the engine speed via the direct drive from the engine crankshaft via a step-down ratio. The increasing demand for coolant volumetric flows, which are partially or completely independent of the engine speed can be achieved in future with switchable or electrically driven coolant pumps. Due to the various installation conditions and different requirement profiles of the coolant pumps dictated by the engineering design, different types and manufacturing concepts are being used. The constructive design of a passenger car coolant pump essentially consists, as seen in Figure 7.410 and Figure 7.411, of a bearing case with bearing insert, sliding ring seal, impeller and boss (belt pulley). Leakage tank with cap and with mounted pumps the housing with spiral port. The coolant feed should occur in a flow-optimized manner symmetrical to the axes, however, for pump flow design reasons it can also be executed on the reverse side via the pump casing and the sliding ring seal. Figure 7.412 shows a mounted pump with feed and spiral port integrated in the pump housing. With coolant pumps for combustion engines, in regard to their attachment and/or their constructive design in mounted pumps (Figure 7.410 and Figure 7.412) and unit pumps (Figure 7.411), a distinction is made between different impeller types, the intake-side or pressure delivery-side arrangement of the sliding ring seal as well as the design of the drive. With unit pumps, pump construction parts such as the spiral port and the feed are located in the engine housing. The coolant pump can be driven by a V-belt, synchronous belt or a poly V-belt. If the location of the coolant pump is contingent of the engine design outside of the belt level(s), it can also be driven using a rigid shaft or toothed gears. Coolant pump housings are made from gray cast iron or aluminum, but will be manufactured increasingly from

Support flange concept 2 stamped flanges from 6 mm

Cast flange with secondary air intake at the cylinder head

Finned flange

Stamped flange 8 mm sheet metal

Deep-drawn flange 3 mm sheet metal

Figure 7.409 Flange concepts for tube manifolds.

304 | Internal Combustion Engine Handbook

6606_Book.indb 304

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.24 Coolant Pumps for Combustion Engines

Sliding ring seal Sliding ring seal

Bearing insert Impeller

Leakage tank with cap

Spiral casing

Figure 7.410 Passenger car coolant pump (mounted pump) manufactured by GPM.

Sliding ring seal

Impeller

Toothed belt pulley

Bearing insert

Bearing housing Leakage tank with cap Figure 7.411 Passenger car coolant pump (unit pump) manufactured by GPM.

Internal Combustion Engine Handbook | 305

6606_Book.indb 305

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Figure 7.413 Open impeller.

Figure 7.412 Passenger car coolant pump (mounted pump) manufactured by GPM with housing-integrated feed and spiral port.

plastic in future. Impellers are made predominantly from temperature-resistant plastics, sheet metal and aluminum for passenger cars, whereas gray cast iron is favored in the manufacture of utility vehicles.

7.24.2 Impeller and Spiral Port

Mechanical energy is fed via the drive disc to the pump shaft and converted into pressure and speed energy in the impeller. With radial impellers the coolant enters the impeller axially and is fed radially to the impeller circulation in the impeller port as a result of the centrifugal force. The flow is delayed once the coolant exits the impeller in the spiral port. Passenger car coolant pumps usually employ radial impellers, but mixed-flow or axial impellers can also be used for small impeller diameters and high pump speeds. Radial impellers are manufactured as open (Figure 7.413) or enclosed (Figure 7.414) types. When compared with an open design, an enclosed impeller is more hard-wearing against production-contingent crack size dispersions in regard to characteristic curve behavior and cavitation susceptibility. However, impeller manufacturing tool costs are higher than those for the open design due to the additional cover disk. The constructive design of the impeller and spiral port is performed at the customer’s request after presenting the design value to GPM—specific computing program. In addition to the design data, the respective installation and application conditions are also taken into account. Impeller inflow in the axial direction should be without interruption or swirl as much as possible.

Figure 7.414 Enclosed impeller.

Any obstructions or pre-swirl must therefore be taken into account in the calculation. The spiral port (Figure 7.415) or spiral casing can be construed as a guide port with a guide vane. The spiral corresponds to the angled section of the guide vane.

Figure 7.415 Spiral port (simple helix).

306 | Internal Combustion Engine Handbook

6606_Book.indb 306

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.24 Coolant Pumps for Combustion Engines

Spiral casings are usually used with single-stage pumps. They can be situated in either the engine block (unit pumps) or in the pump housing (mounted pumps). The cross section of the spiral port increases in the impeller’s direction of rotation. The same flow state (i.e., constant circuit velocity) must prevail in each of the two parallel spiral circuits to ensure an homogeneous, axial-symmetric flow. For engine cylinder series, which are arranged in a V-shape, the coolant can be fed to the cylinder banks via a twin spiral port (Figure 7.416). Twin spiral ports can be implemented for both mounted pumps (pump-internal) and unit pumps (engine block internal).

lips run against a specially hardened shaft sleeve, no measurable wear occurs during the service life.

Counter ring

Sliding ring Spring collar Pressure spring Sealant

Sleeve

Cup Housing

Holder

Bellows

Figure 7.416 Twin spiral port.

7.24.3 Coolant-Side Sealing

Coolant-side sealing to the pump bearing is delivered by a sliding ring seal (radial sealing system) [Figure 7.417(a)] or alternatively with a radial shaft seal, such as a VR seal (axial sealing system) [Figure 7.417(b)]. Sealing can be arranged on either the intake or pressure delivery side of the impeller. Friction heat is easily dissipated with the sealing arranged on the intake side due to flushing with the entire pump volumetric flow. A sufficient rinse volumetric flow via the impeller reverse side must be ensured to achieve heat removal if the sliding ring seals are arranged on the pressure delivery side. The sliding ring seal friction pairs are lubricated and cooled by the coolant. Consequently a slight leakage of steam and/or fluid may escape into the atmosphere. Engineering measures within the coolant pump housing can prevent this so-called cosmetic leakage escaping into the atmosphere. The leakage, which remains in the housing as a result, evaporates thanks to the generated engine heat (see Figure 7.410 and Figure 7.411). The VR seal (radial sealing system) represents an economical alternative to the axial sliding ring seal. The sealing principle requires no leakages over the sealing gap. Because the sealing

Support cap

Pressure disk

Water

Bushing Inner sealing sleeve

Outer sealing sleeve Grease chamber

Figure 7.417 (a) Sliding ring seal (axial sealing system). (b) VR seal (radial sealing system).

Special advantages include the small installation space as well as potential use at high speeds (10,000 rpm). Another advantage of the radial sealing system is that noise cannot arise under normal circumstances.

Internal Combustion Engine Handbook | 307

6606_Book.indb 307

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7.24.4 Map and Similarity Relations of the Coolant Pump



The dependency of the pump differential pressure/pump delivery head from the pumped coolant volumetric flow and the pump speed are shown in the coolant pump map. Figure 7.418 shows a typical map of a passenger car coolant pump.

Poperating energy input = Md ⋅ 2 ⋅ p ⋅ n



h=

Peffective output Poperating energy input

(7.31) (7.32)

where g = gravitational acceleration (m/s2) Md = torque (Nm)

Pump speed [min -1 ]

n = speed (s–1) Peffective output = effective output (W) Poperating energy input = operating energy input (W)

Delivery head [m]

∆pTotal = total pressure differential according to Bernoulli’s equation (N/m2) V̇ = volumetric flow (m3/s)

η = efficiency (–) ρ = fluid density (kg/m3)

Volumetric flow rate [l/min] Figure 7.418 Map of a passenger car coolant pump.

The coolant pump map should be depicted as in Figure 7.418 using the parameter ‘delivery head’ and subsequently be independent of the temperature and composition of the coolant, i.e. the map applies for all coolant temperatures and antifreeze percentages. In contrast the characteristic curve course when using the pump parameter ‘differential pressure’ would only apply to the fluid state prevailing at the time of determining the map. The course or gradient of the pump characteristic curve can be influenced by the shape and number of the impeller vanes. H applies for the delivery head: ∆p H = total r⋅g

=

(7.28)

pD − pS c −c + hD − hS + r⋅ g 2⋅g 2 D

2 S

(7.29)

where cD = flow velocity on the pump pressure delivery side (m/s) cS = flow velocity on the pump intake side (m/s)

(7.34)



n12 ∆p1 H 1 Md1 = = = n22 ∆p2 H 2 Md2

(7.35)



n13 Poperating energy input 1 = n23 Poperating energy input 2

where Md = torque (Nm)

H = delivery head (m) hD = geodetic head of the pump pressure delivery side (m) hS = geodetic head of the pump intake side (m) pD = local pressure on the pump pressure delivery side (m/s) ps = local pressure on the pump intake side (m/s). D denotes the measuring point on the pump reverse side, S denotes the measuring point on the pump intake side. Effective output, operating energy input and coolant pump efficiency are calculated from Peffective output = ∆ptotal ⋅ V = H ⋅ r ⋅ g ⋅ V

(7.33)



n1 V1 = n2 V2

H = delivery head (m)

g = gravitational acceleration (m/s2)



Figure 7.419 depicts the dependency of the pump operating energy input on the engine speed and pump volumetric flow. Depending on the volumetric flow requirements, pump efficiency and coolant circuit resistance, it is common to have pump operating energy inputs for engine speeds between 500 W and 3.5 kW. Here it is important to ensure that the pump operating energy input is able to change at the same engine speed dependent on the coolant circuit operating state (for instance heater open or closed, thermostat open, regulating or closed) due to the associated different coolant circuit overall resistances. Using the similarity relations for centrifugal pumps, it is possible to convert the parameters of a predetermined coolant pump characteristic curve or one determined using measuring technology to other pump speeds. The following applies:

n = speed (s–1) Peffective output = effective output (W) Poperating energy input = operating energy input (W) ∆pTotal = total pressure differential according to Bernoulli’s equation (N/m2). Equations (7.33)–(7.35) show that the linear volumetric flow, the pump differential pressure/delivery head as well as the quadratic pump torque and pump effective output rise in third power with higher pump speeds.

(7.30)

308 | Internal Combustion Engine Handbook

6606_Book.indb 308

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.24 Coolant Pumps for Combustion Engines

Drive power [W]

Drive power [W]

Pump speed [min -1 ]

Engine speed [min -1 ]

The operating energy input rises similar to the pump effective output, but is, however, also dependent on the efficiency course. The similarity relations according to (7.33)–(7.35) apply unrestrictedly for a pump map as per Figure 7.418. This means that the pump parameters can be converted to modified speeds. However, when operating the coolant pump in the coolant circuit, the similarity relations only apply for constant coolant circuit flow resistance. Because the coolant circuit resistance at least increases in the lower engine speed area due to a notcompletely, hydraulically rough developed flow, the similarity relations can only be used to a limited extent with the operation of the coolant pump in the coolant circuit for converting pump parameters, i.e. after reaching an hydraulically rough flow. Figure 7.420 schematically shows the characteristic course of a coolant circuit overall resistance by marking the point, from which a hydraulically rough flow is present and the similarity relations can be applied. The same characteristic of the flow resistance course as is shown in Figure 7.420 and is demonstrated by the individual coolant circuit components such as the engine, radiator, heater, thermostat, and oil heat exchanger.

Hydr. rough flow

Figure 7.419 Coolant pump operating energy input dependent on speed and coolant volumetric flow.

Volumetric flow rate [l/min]

An increase in coolant resistance as per Figure 7.420 with small volumetric flows/low speeds is responsible for pumping a disproportionately low amount of coolant from the coolant pump in the lower engine speed range. These fluid-mechanical circumstances must be taken into account by the manufacturer when determining the design values of the pump. The intersection of the pump characteristic curve with the coolant circuit resistance provides the coolant pump working points for the respective pump speeds. The working points are shifted accordingly with a modified coolant circuit resistance caused for instance by blocking the heating or regulating the thermostat. Figure 7.421 schematically depicts the various coolant circuit resistances and working points (AP) for idle and nominal speeds with an open or closed heater in the coolant pump map. The coolant pump is usually designed for a design value predetermined by the engine manufacturer (pump differential pressure/delivery head and volumetric flow) with pump nominal speeds or another pump speed between idle speed and controlled speed, and for a certain coolant circulation state, such as thermostat open or heater closed. This design value should match the resulting working point from the operation of the coolant pump in the coolant circuit. With a different working point, other coolant circuit volumetric flows and thereby partial volumetric flows, which are too big or too small are pumped via the coolant circuit elements than where specified in the design value for the coolant pump. This can have a disadvantageous effect on heat transport and heat dissipation as well as on the compressive load of the coolant elements.

7.24.5 Cavitation n

n Pump ~ Engine;

Figure 7.420 Overall coolant circuit flow resistance.

Cavitation is the name given to the formation and implosion of vapor bubbles in flowing liquids. Cavitation bubbles occur when at a certain point in the coolant circuit, the coolant pressure drops below that of the vapor. The intake side of

Internal Combustion Engine Handbook | 309

6606_Book.indb 309

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Circuit drag Heater closed

Circuit drag Heater open

AP heater closed Delivery head

AP auf open APHz heater Nominal speed

AP heater closed APAPHzheater auf open

Idle speed

Figure 7.421 Coolant circuit resistance and working points in the coolant pump map.

Volumetric flow rate

the coolant pump is particularly prone to cavitation because of its low pressure level. The delivery head of the coolant pump and subsequently the coolant circuit volumetric flow are reduced depending on the size of the vapor pressure lower deviation. Furthermore, the sudden implosion of the vapor bubble in the higher pressure areas, e.g. in the swirl port or in the engine block, can lead to material removal. Pressure drops occur locally at sharp edges and diversions as well as when entering into the vane port with high flow velocities (i.e. large coolant volumetric flows at high pump speeds). Due to the difficult task of calculating and measuring these isolated pressure decreases in the rotating vane port, cavitation experiments are at least conducted for the design value with design speed on a specially designed coolant pump test bench. The pressure on the pump intake side is lowered as far as necessary until a specific delivery head decrease, for instance 3%, is achieved. The so-called NPSH(net positive suction head) value is calculated from the determined intake pressure and the prevailing measurement conditions (fluid temperature and composition) during test bench experiments: NPSH 97 % = where

pS 97 % − pVapor r⋅ g

c2 + S 2⋅g

(7.36)

cS = flow velocity on the pump intake side (m/s) g = gravitational acceleration (m/s2) NPSH97% = NPSH value with a 3% delivery head decrease (m) pVapor = vapor pressure of the fluid (N/m2) pS 97% = static pressure on the pump intake side with a 3% delivery head decrease (N/m2)

ρ = fluid density (kg/m3).

Because this NPSH value, like the delivery head, is independent of the fluid composition and temperature, it can be used to calculate the required pressure on the pump intake side and if necessary the required pressure in the control tank or its connection point in the coolant circuit to prevent cavitation for an operating state of the coolant circuit (e.g. 40% antifreeze content, coolant temperature on the pump intake side = 108°C heater closed) to be taken into account:

pS stat erf = ( NPSH ⋅ r ⋅ g ) + pVapor − r

cS2 2

(7.37)

where cs = flow velocity on the pump intake side (m/s) g = gravitational acceleration (m/s2) NPSH = net positive suction head—absolute energy head minus vapor pressure head = net energy head (m) pVapor = vapor pressure of the fluid (N/m2) pS req stat = required static pressure on the pump intake side to prevent cavitation (N/m2) pS 97% = static pressure on the pump intake side with a 3% delivery head decrease (N/m2)

ρ = fluid density (kg/m3). It is important to remember when operating the coolant pump in the coolant circuit, that cavitation bubbles not only appear at very high temperatures, but also at the medium temperature range due to the generally high pressure reduction at the bypass thermometer plate and the still very low pressure in the control tank during this engine operating state. A cavitation map can be determined to get statements on the cavitation behavior of the coolant pump even for speeds and volumetric flows, which differ from the pump design value. The cavitation value NPSH grows as the pump speed and volumetric flow increase (Figure 7.422).

310 | Internal Combustion Engine Handbook

6606_Book.indb 310

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

NPSH

7.24 Coolant Pumps for Combustion Engines

Volumetric flow rate Figure 7.422 Cavitation map of a coolant pump.

7.24.6 Electric Coolant Pump and Switchable Mechanical Coolant Pump

The demand for coolant volumetric flows independent from the engine speed can be achieved with the use of speed-regulated coolant pumps. Thus achieving low volumetric flows in the engine partial load area for increasing the coolant and oil temperatures to achieve an engine friction decrease as well as high volumetric flows from engine idle speed to achieve heater support, with a large injection of heat into the cooling system after full-throttle journey and with high engine torques in the lower engine speed range. Furthermore, the coolant pump can be switched off during the engine’s warm-up phase to reach the desired operating temperature sooner. On account of the coolant-side saturation of the heater and radiator, the coolant volumetric flow is limited in the upper engine speed range, whereby a significant pump operating energy input is possible as per Figure 7.419. These requirements are met by the electric main water pump with coolant distributor function (Figure 7.423).

pump operating energy input and thereby the installation size of the electric motor, as well as the generator load, can be kept small. For instance, this can be achieved by reducing the main flow resistance and optimizing the coolant circuit network. Electro main water pumps have until now only been used in individual engine series. The limited availability of electric power within 12-V on-board power supplies is an obstacle to further expansion, but particularly the high system costs for electronic DC motors. Using an electro main water pump also means extensive changes in the coolant circuit design. The positive effects of a volumetric flow variability within the scope of thermal management without significant changes to the engine or to the coolant circuit, can also largely by brought about by economical mechanically driven pumps that have a switch-off function. Coolant volumetric flow switch-off allows for significantly faster engine heating, particularly in the cold start phase. This results in lower fuel consumption and reduced emissions. Potential fuel savings in the NEDC (New European Driving Cycle) as a result of volumetric flow switch-off range between 0.5 to 3% according to manufacturer’s data. Various solutions for switching off the volumetric flow have come to light and it is primarily the following concepts that are derived from these solutions: •• directional control valves connected in series to the pump •• coolant pumps with switchable magnetic coupling •• coolant pumps with integrated split ring slide valve (Figure 7.424).

Figure 7.423 Electric coolant distributor pump manufactured by GPM.

The electric coolant pump shown in Figure 7.423 achieves high pump efficiency by employing an axial impeller and therefore requires a relatively small pump operating energy input. Thanks to the integration of circuit control elements such the thermostat, the coolant distributor pump can adopt additional coolant circuit functions economically using spacesaving solutions. With flow-optimized coolant circuits, the

Figure 7.424 Switchable mechanical coolant pump manufactured by GPM.

Internal Combustion Engine Handbook | 311

6606_Book.indb 311

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

The split ring slide valve’s roles as a control element in centrifugal pumps has already been examined in-depth by Pfleiderer. It concerns an impeller-concentric ring slide valve, which can be moved axially over the impeller by suitable actuators. It is actuated by vacuum actuators or even directly with hydraulic pressure. Compared with other concepts, the coolant pump with split ring slide valve has a number of significant advantages: •• Unlike clutch pumps, the slide valve allows actual zero delivery when the engine is running as even thermal siphon flows are easily suppressed. •• In essence, the slider pump requires less installation space than a constant pump and weights only slightly more. •• The pump operating energy input can be considerably reduced in the off-state.

wall in the form of slots, the peculiarities described below result for the drive trains in two-stroke engines. The slots in the cylinder wall make it more difficult to achieve defined lubrication of the tribologic pair of piston and cylinder. To ensure adequate lubrication and to avoid unacceptably high oil consumption, engineers must exercise great care when selecting the mating materials in regard to minimum lubricating oil requirements, metered lubricating oil feed, and/or sufficient oil stripper effect by the piston rings. To prevent the piston rings (and the ends in particular) from entering the exhaust, scavenging, and intake slots due to spring action, it is necessary to observe maximum slot widths (expressed as the ratio between slot width and cylinder diameter). This is explained in detail in [7-223], [7-224].

•• The slider pump turns out to be very economical, particularly when compared to clutch pumps. •• Engine cooling is ensured as a fail-safe function thanks to the spring return mechanism.

7.25 Control Mechanisms for TwoStroke Engines Characteristic of the principle behind the two-stroke cycle is that, in contrast to the four-stroke engine, one complete working cycle is executed per crankshaft revolution; the expulsion of the burned charge from the cylinder and the introduction of fresh fuel and combustion air into the cylinder (scavenging process) takes place at crankshaft angles around bottom dead center (BDC). The requirement here is that, with a suitable design of the mechanism controlling the change of gases, there is minimum mixing of fresh gas and exhaust gas (high scavenging efficiency) with a low required scavenging pressure gradient (low work expenditure for changing the charge), all this within the smallest possible crankshaft angle range around BDC (limited restriction on the useful piston stroke). There are several different scavenging processes available for the change of charges in two-stroke engines; these are explained in greater detail in Section 10.3 (see also [7-222] and [7-223]). Their use requires a far different design for the drive components than what is found in four-stroke engines. Because the working cycle for the two-stroke engine transpires at the same frequency as crankshaft rotation, it is possible, in contrast to the four-stroke engine, to use the piston itself to control the gas flows. Loop scavenging is used particularly in small engines and those running at high speeds; this principle is shown in Figure 7.425. Here the piston controls the discharge of the exhaust from the cylinder through the exhaust slot(s), the inflow of fresh gas via the scavenging ports, and, when using the crankcase scavenging pump concept, the inlet of the fresh fuel-air mix into the crankcase as well. Because of the arrangement of the exhaust, intake, scavenging, and/or transfer passages at the cylinder, which penetrate the cylinder

Figure 7.425 Sectional view of a modern, two-cycle spark ignition engine with loop scavenging, crankcase scavenging pump, reed valves at the intake system, and flat spool exhaust control.

In addition, the slots, normally rectangular in shape, have to be rounded at the corners at the upper and lower ends, and the transitions from the cylinder to the channel walls have to be rounded. Piston ring rotation in the piston grooves, accompanied by the hazard that the ends of the rings enter the slots in the cylinder walls under spring pressure, is prevented where required by pins pressed into the ring grooves. The fact that firing is twice as often as in four-stroke engines and, above all, the piston controls fresh gas and exhaust flow, results in far higher thermal loads on the piston and cylinder in slot-controlled two-stroke engines when compared with four-stroke designs. This is discussed in [7-225]. This loading is seen as the essential cause for the limited service life often found in high-performance two-stroke engines. The situation is made all the more difficult where the incoming air or mix passes through the crank chamber (crank chamber scavenging pump). This largely eliminates effective cooling of the piston with splashed oil, a technique commonly used in higher-performance four-stroke engines. Among the strategies

312 | Internal Combustion Engine Handbook

6606_Book.indb 312

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.25 Control Mechanisms for Two-Stroke Engines

available to reduce the thermal load on the piston, piston rings, and wristpin boss are the following: Limiting individual cylinder volumes; careful designing of cylinder cooling (using water cooling if possible), particularly in the area around the exhaust slots; designing to reduce cylinder warping, which would make it more difficult to dissipate heat from the piston, through the piston rings, and to the cylinder walls; selecting a timing concept that prevents additional heating of the piston and fresh gas by exhaust blowback into the scavenging ports; selecting a scavenging process in which the exhaust flowing out of the cylinder is kept from coming into contact with any large surface area at the piston. In modern loop scavenging cylinders for high-speed twostroke engines, the fresh gas is generally introduced through between four and seven scavenging or transfer passages (in a mirror symmetrical arrangement to the exhaust channel), sweeping the wall at a shallow angle in the direction of the wall opposite the exhaust slot. This causes a rising stream of fresh gas to be formed along the cylinder wall. Near the cylinder head it reverses direction and forces the exhaust gas out of the cylinder. The transfer passages are located at the side of the cylinder and are tapered slightly along the direction of flow. This requires far more space between cylinders in multicylinder engines of this design when compared with similar four-stroke engines. The discontinuities in cylinder wall stiffness caused by the charge exchange runners results in more indirect force flow between the cylinder head and the

crankshaft. Consequently, the highly asymmetrical thermal loading on the piston and cylinder due to the exhaust slots make it necessary to very carefully design the drive assembly and its cooling. It should be noted here that various strategies are used particularly in modern, two-cycle spark ignition engines to increase the fresh gas fill efficiency, to influence fuel and air blending, and to avoid negative influences of gas pulsation in the intake and exhaust sections. Depending on the concept employed, these may involve rotary intake valves, reed valves (oneway valves), bypass reed controls, oscillation chambers, and, on the exhaust side, control spools or cylindrical valves. This may increase the complexity of the drive system considerably. When using uniflow scavenging with exhaust valves, a concept employed particularly in diesel engines, the fresh gas enters the cylinder through scavenging ports under piston control while the exhaust gas flows out through several valves located in the cylinder head; their opening is synchronized with crankshaft rotation frequency. To achieve good scavenging efficiency, it is necessary that the intake runners or slots generally not impart any particular directional effect (aside from a slight tangential orientation to support gas blending); with the result that the volume of the intake plenum located upstream from the scavenging ports, as shown in Figure 7.426, in many cases adjoins the outside diameter of the cylinder sleeve (see also [7-226]).

Figure 7.426 Longitudinal and cross sections through a uniflow scavenged four-cylinder, two-cycle diesel engine made by Krupp [7-226].

Internal Combustion Engine Handbook | 313

6606_Book.indb 313

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

Because the scavenging ports have to be covered by the piston skirt at TDC, long pistons are required particularly in long-stroke engines, resulting in a relatively large overall height for the engine. In contrast to loop scavenging, uniflow scavenging with exhaust valves causes somewhat less and more symmetrical thermal loading at the piston and cylinder. By contrast, the doubled actuation frequency for the exhaust valves (in comparison to four-stroke engines) and the high thermal loading on the cylinder head in fast-running engines places great demands on the design of cylinder head cooling and the kinematics of the valve train. In the design with four exhaust valves, often selected for high-speed engines, one objective in development is to achieve a shallow contour for the runners (small runner surface to be cooled, low exhaust heat losses where an exhaust turbocharger is used) so that the exhaust gas flow at the individual valves is hindered as little as possible. Aside from this, intensive cooling is necessary, particularly in the area around the injection nozzle to avoid carbonization problems. In order to exchange the charges, within the limited crankshaft arc available for this purpose, with the smallest possible amount of work, one must select a suitable valve train concept and implement valve train kinematics with minimal pressure loss as gases flow through the valves.

Figure 7.427 shows the solution used in a 1.0-liter, two-cycle diesel engine being built by AVL now. In this engine, the four exhaust valves at each cylinder are activated by roller cam followers at two overhead camshafts. Figure 7.428 shows an alternate exhaust runner version to this concept.

Figure 7.428 Illustration of the exhaust runner configuration and the valve train for a uniflow-scavenged, two-cycle diesel engine for passenger cars.

Figure 7.427 Longitudinal and cross sections through a uniflow scavenged, two-cycle diesel engine made by AVL for passenger cars [7-227].

314 | Internal Combustion Engine Handbook

6606_Book.indb 314

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.25 Control Mechanisms for Two-Stroke Engines

Bibliography

7-1. Zima, S. 1999. Kurbeltriebe, Konstruktion, Berechnung und Erprobung von den Anfängen bis heute, 2. Auflage, Braunschweig. Vieweg: Wiesbaden. 7-2. Junker, H. and Ißler, W. 1999. Kolben für hochbelastete Diesel-Motoren mit Direkteinspritzung, 8. Aachener Kolloquium Fahrzeug- und Motorentechnik, Aachen. 7-3. Röhrle, M. 1994. Kolben für Verbrennungsmotoren. Verlag moderne Industrie AG: Landsberg. 7-4. Kemnitz, P., Maier, O., and Klein, R. Monotherm, a New Forged Steel Piston Design for Highly Loaded Diesel Engines, 2000. SAE 2000-01-0924. 7-5. Buschbeck, R., Ottliczky, E., Hanke, W., and Weimar, H.-J. 2010. Innovative Kolbensystemlösungen für Verbrennungsmotoren, MTZ Extra 03. Wiesbaden. 7-6. Künstscher. 1995. Kraftfahrzeug Motoren. Verlag Technik: Berlin. § 3.2.2. Kräfte an der Pleuelstange.

7-27. Hoppe, S., Münchow, F., and Esser, J., “DLC-Hochleistungsbeschichtungen für Kolbenringe, Potenzial und Einsatzerfahrungen,” VDI-Bericht 1994:143–156, 2008. 7-28. Martínez, D.L., Valverde, M., Rabuté, R., and Ferrarese, A., “Kolbenringpaket für reibungsoptimierte Motoren,” MTZ 07.–08. 2010, Wiesbaden, Juli–August 2010. 7-29. Mittler, R., “Detaillierte 3D- Ringpaketanalyse,” MTZ 07.–08. 2010, Wiesbaden, Juli–August 2010. 7-30. Herbst-Dederichs, C. and Münchow, F., “Hybrid-Laufbuchsen aus Grauguss und Aluminium,” MTZ 65(10):820–822, 2004. 7-31. Martin, T. and Weber, R., “Compacted Vermicular Cast Iron (GJV) für den V6-Dieselmotor von Audi,” MTZ 65(10): 824–832, 2004. 7-32. Schwaderlapp, M., Bick, W., Duesmann, M., and Kauth, J., “200 bar Spitzendruck—Leichtbaulösungen für zukünftige Dieselmotorblöcke,” MTZ 65(2):84ff, 2004. 7-33. Zeitschrift Aluminium, Ausgabe11/2003

7-7. Vogel Fachbuch, Motorschäden, Greuter/Zima, § 6.2.3. Pleuelstangen.

7-34. N.N. BMW Presse-Information vom 23.06.2004.

7-8. Fisher, S. 1983. Berechnungsbeispiel einer Pleuellagerdeckelverschraubung, VDI-Berichte Nr. 478.

7-35. Flores, G. 1992. Grundlagen und Anwendung des Honens. VulkanVerlag: Essen.

7-9. VDI 2230: Systematische Berechnung hochbeanspruchter Schraubenverbindungen, 1986. Beuth-Verlag: Berlin und Köln.

7-36. Abeln, T. and Klink, U. 2001. Laserstrukturieren zur Verbesserung der tribologischen Eigenschaften von Oberflächen. Tagungsband: Stuttgarter Lasertage.

7-10. Thomala, W. 1986. Erläuterung zur Richtlinie VDI 2230 Blatt 1(1986), RIBE-Blauheft Nr. 40. 7-11. Ohrnberger, V., Hähnel, M., and Aalen, “Bruchtrennen von Pleueln erlangt Serienreife,” Werkstatt und Betrieb 125:3, 1992. 7-12. Adlof, W.W., “Bruchgetrennte Pleuelstangen aus Stahl,” Schmiede-Journal 1996. 7-13. Herlan, T. Optimierungs und Innovationspotential stahlgeschmiedeter Pleuel, VDI, Schwelm, 1996 oder 1997. 7-14. Moldenhauer, F., “Verbesserungen bei bruchtrennfähigen Pleuelstangen durch neuen mikrolegierten Stahl,” MTZ Motortechnische Zeitschrift 61:4, 2000.

7-37. Robota, A. and Zwein, F., “Einfluss der Zylinderlaufflächentopografie auf den Ölverbrauch und die Partikelemissionen eines DI-Dieselmotors,” MTZ 60(4), 1999. 7-38. Herbst, L. and Lindner, H., “Verschleiß- und schmierstoffmindernde Bearbeitung von Zylinderoberflächen mit Excimerlaser im Motorenbau,” VDMA-Nachrichten 12-2002. 7-39. Mahle GmbH (Hrsg.). 2009. Zylinderkomponenten. Eigenschaften, Anwendungen, Werkstoffe, Vieweg + Teubner, 1. Aufl. 7-40. MTZ 10/2008: Pkw-Kunststoffölwanne zur Reduzierung von Kosten, Gewicht und CO2.

7-15. Weber, M., “Comparison of Advanced Procedures and Economics for Production of Connecting Rods,” Powder Metall. Int. 25(3):125–129, 1993.

7-41. ATZ online: Neues Kunststoff-Ölwannenmodul spart 1,1 Kilogramm Gewicht.

7-16. Richter K., Hoffmann E., Rüsselsheim, Lipp, K., Sonsino, C.M., and Darmstadt, “Single-Sintered Con Rods—An illusion?,” Metal Powder Report 49(5):38–45, 1994.

7-42. ATZ online: Crash-Simulationen am gealterten Kunststoff

7-17. Skoglund, P., Bengtsson, S., Bergkvist, A., Sherborne, J., and Gregory, M., “Performance of High Density P/M Connecting Rods,” Powdered Metal Applications (SP-1535).

innovations-report 22.10.2008: DuPont liefert Zytel Polyamid für erstes in Serien-Pkw eingesetztes Kunststoff-Ölwannenmodul.

7-18. Spangenberg, S., Kemnitz, P., Kopf, E., and Repgen, B., “Massenreduzierung an Bauteilen des Kurbeltriebs, Pleuel in Fokus,” MTZ 24–261, 2006. 7-19. Depp, J.C., Ilia, E., and Hähnel, M., “Neue hochfeste Werkstoffe für sintergeschmiedete Pleuelstangen,” MTZ 292–298, 2005. 7-20. N. N.: Kolbenringhandbuch der Federal-Mogul Burscheid GmbH, Neuauflage 2008 (Online-Ausgabe) 7-21. Esser, J., “Einfluss von Ölabstreifringen auf den Ölverbrauch,” MTZ 63:7/8, 2002. 7-22. Mierbach, A., “Radialdruck und Spannbandform eines Kolbenringes,” MTZ 55:2, 1994. 7-23. Herbst-Dederichs, C. and Münchow, F. Modern Piston Ring Coatings and Liner Technology for EGR Applications. SAE Paper 2002-01-0489, 2002. 7-24. Ishaq, R. and Grunow, F., “Wege zur Optimierung des Reibsystems Kolbenring und Ringnut,” MTZ 60:9, 1999. 7-25. Esser, J., Hoppe, S., Linde, R., and Münchow, F., “Kompressionskolbenringe in Otto- und Dieselmotoren,” MTZ 66:7–8, 2005. 7-26. Esser, J., Linde, R., and Münchow, F., “Diamantbeschichtete Laufflächenschicht für Kompressionsringe,” MTZ 65:7–8, 2004.

innovations-report 11.09.2007: Mann + Hummel plant im Jahr 2009 Serienstart von Kunststoffölwanne für Pkw

7-43. Hohnstein, T., Gleiter, U., Glaser, S., and Fritz, T. Erste Serienanwendung von Steinschlagoptimierten Kunststoff-Ölwannen. MTZ 01.2010, Wiesbaden. 7-44. Berg, W. 1981. Aufwand und Probleme für Gesetzgeber und Automobilindustrie bei der Kontrolle der Schadstoffemissionen von Personenkraftwagen mit Otto- und Dieselmotoren, Dissertation TU Braunschweig. 7-45. Gruden, D. 2003. Volume Editor: Traffic and Environment, The Handbook of Environmental Chemistry, Volume 3 Part I. Springer-Verlag: Berlin, Heidelberg, New York. 7-46. Meinig, U., Spies, K.-H., and Heinemann, J. 1997. Canister Purge Flow Influence on EGO- Sensor Signal and Exhaust Gas Emissions (Purgeopt.). Warrendale, PA: SAE- paper 970029. 7-47. N.N. California Code of Regulation (CRR), Title 13, Section 1988.2 7-48. Brunsmann, I. 2002. “Ölabscheidung in Entlüftungssystemen.” in Tumbrink, M. (Hrsg.) und 12 Mitautoren: Filtersysteme im Automobil—Innovative Lösungsansätze für die Automobilindustrie. expert-Verlag: Renningen. 7-49. Verhoefen, U., “Filtration technischer Gase—Teil 1,” pneumatic Digest 16:3–4, 1982. Jahrgang.

Internal Combustion Engine Handbook | 315

6606_Book.indb 315

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7-50. Brauer, H. 1971. Grundlagen der Einphasen- und Mehrphasenströmung. Frankfurt a. M.: Verlag Sauerländer und Aarau.

7-74. Klos R. Aluminium Gusslegierungen. Die Bibliothek der Technik, Nr. 116, Verlag Moderne Industrie, 1995.

7-51. Krause, W. 1995. Ölabscheidung in der Kurbelgehäuseentlüftung. Dissertation Uni Kaiserslautern.

7-75. N.N. Ansicht eines Mitsubishi-Galant-Motors. Prospekt der Fa. Mitsubishi Japan, 2001.

7-52. Löffler, F. 1988. Staubabscheiden, Lehrreihe Verfahrenstechnik. Georg Thieme Verlag: Stuttgart.

7-76. Krebs, R., Böhme, J., Dornhöfer, R., Wurms, R., Friedmann, K., Helbig, J., and Hatz, W., “Der neue Audi 2,0T FSI Motor—Der erste direkteinspritzende Turbo-Ottomotor bei Audi. 15,” Wiener Motorensymposium, 2004.

7-53. Sauter, H., Brodesser, K., and Brüggemann, D., “Hocheffizientes Ölabscheidesystem für die Kurbelgehäuseentlüftung,” MTZ 64(3):180–184, 2003.

7-77. Klaus, B. Die Valvetronic der 2. Generation im neuen BMW-ReihenSechszylindermotor. VDI-Tagung in Stuttgart, 15. und 16. September 2004.

7-54. Ahlborn, St., Blomerius, H., and Schumann, H., “ Neue Wege in der Reinigung von Kurbelgehäuseentlüftungsgasen,” MTZ 60:454–459, 1999.

7-78. Jägerbauer E., Fröhlich K., and Fischer H., “Der neue 6,0 1-Zwölfzylindermotor von BMW,” MTZ Motorentechnische Zeitschrift 64:7/8, 2003.

7-55. Sauter, H. 2004. Analysen und Lösungsansätze für die Entwicklung innovativer Kurbelgehäuseentlüftungen. Dissertation Uni Bayreuth.

7-79. Hannibal, W., Flierl, R., Groenewold, C., and Paulov, M. Konzepte des variablen Ventiltriebs im Vergleich mit Hinblick auf das Potenzial zur Kraftstoffverbrauchssenkung. Vortrag Haus der Technik, 3. März 2009, Essen.

7-56. Hannibal, W. and Meyer, K. Patentrecherche und Überblick zu variablen Ventilsteuerungen. Vortrag Haus der Technik, März 2000. 7-57. Flierl, R., Hofmann, R., Landerl, C., Melcher, T., and Steyer, H., “Der neue BMW-Vierzylinder-Ottomotor mit VALVETRONIC,” MTZ Motortechnische Zeitschrift 62:6, 2001. 7-58. Hannibal, W. and Lukas, F., “Rechnergestützte Auslegung des Audi-Fünfventil Zylinderkopfkonzeptes,” MTZ Motortechnische Zeitschrift 55:12, 1994. 7-59. Dong, X., “Öffnungsquerschnitt von Ventilen,” MTZ Motortechnische Zeitschrift 46:6, 1985. 7-60. Schäfer, F., Barte, S., and Bulla M., “Geometrische Zusammenhänge an Zylinderköpfen,” MTZ Motortechnische Zeitschrift 58:7/8, 1997. 7-61. Eidenböck, T., Ratzberger, R., Stastny, J., and Stütz, W., “Zylinderkopf in Vierventiltechnik für den BMW DI-Dieselmotor,” MTZ Motortechnische Zeitschrift 59:6, 1998. 7-62. Krappel, A., Riedl, W., Schmidt-Troje, D., and Schopp, J., “Der neue BMW Sechszylindermotor in neuer Hubraumstaffelung und innovativer Leichtbauweise,” MTZ Motortechnische Zeitschrift 56:6, 1995. 7-63. N.N. Becker CAD CAM CAST. Broschüre der Fa. Becker GmbH CAD CAM CAST, Steffenberg-Quotshausen, 2001.

7-80. Adolf, W., “Wer an Leichtbau denkt, kommt an einer Stahlkurbelwelle nicht vorbei,” Schmiede-Journal, März 94:13–16 7-81. Heck, K., et al., “Herstellen von Endmaßen-Gusskurbelwellen— Innovative giesstechnologische Entwicklung,” Konstruieren und Gießen 23(3):4–12, 1998. 7-82. IMC Consultants, IMC, Bericht für Georg Fischer/DISA Analysis of alternative strategies designed to increase market share of the magnesium converter, May 1998. 7-83. Nickel, F. ISAD der integrierte Starter—Alternator—Dämpfer, Tagung Motor und Umwelt 98, AVL Graz, S. 175–182. 7-84. Becker, E. and Hornung, K. Projekt 78274 Kurbelwellenfertigung im Masken- oder Grünsandverfahren, Georg Fischer F&E Berichte Aug. 85, Sept. 85. 7-85. Gusseisen mit Kugelgraphit—ein duktiler Werkstoff von Georg Fischer. Georg Fischer, Technisches Merkblatt über Herstellung und Eigenschaften von GJS, 08/93. 7-86. Engel, U., et al. “Influence of Micro Surface Structures of Nodular Cast Iron Crankshafts on Plain Bearing Wear,” SAE Technical Paper Series, 880097, Detroit Michigan USA, March 4, 1988.

7-64. Hannibal, W. “Begleitende Entwicklung der Audi-FünfventilTechnologie mittels Rechnereinsatz,” Wiener Motoren Symposium, 1995.

7-87. Hornung, K. and Mahnig, F. Beanspruchungsgerechte Gestaltung und anwendungsbezogene Eigenschaften von Gussteilen, Georg Fischer, Technisches Merkblatt, 8/87.

7-65. Seifert, H., “20 Jahre erfolgreiche Entwicklung des Programmsystems PROMO,” MTZ Motortechnische Zeitschrift 51:11, 1990.

7-88. Heuler, P., et al., “Steigerung der Schwingfestigkeit von Bauteilen aus Gusseisen mit Kugelgraphit,” ATZ 94(5):270–281, 1992.

7-66. Dirschmid, W. and Schober, M., “Computersimulation in der Ventiltriebsauslegung,” MTZ Motortechnische Zeitschrift 57:4, 1996.

7-89. Fuchsbauer, B. Untersuchungen zur Schwingfestigkeitsoptimierung bauteilähnlicher Proben unterschiedlicher Größe durch Festwalzen. Diss. TH Darmstadt, 1983.

7-67. Nefischer, P., Blumenschein, S., Keber, A., and Seli, B., “Verkürzter Entwicklungsablauf beim neuen Achtzylinder-Dieselmotor von BMW,” MTZ Motortechnische Zeitschrift 60:10, 1999. 7-68. Scheeren, H.W., Koreneef, A., and Fuchs, H. Herstellung von Zylinderköpfen für hochbeanspruchte Diesel- und Ottomotoren. Vortrag Haus der Technik, 27. Juni 2000, Essen. 7-69. Hannibal, W. and Metzlaw, A. Von der Idee zum Produkt. In: Digitalisierung und Flächenrückführung in der CAD-Prozesskette. Zeitschrift QZ, Qualität und Zuverlässigkeit, 46:7, 2001. 7-70. Fortnagel, M., Doll, G., Kollmann, K., and Weining, H.-K. Aus Acht mach Vier: Die neuen V8-Motoren mit 4,3 und 5l Hubraum. In: Sonderausgabe MTZ Motortechnische Zeitschrift, Mercedes-Benz S-Klasse, 1998. 7-71. Dorenkamp, R., Hadler, J., Simon, B., and Neyer, D. Der VierzylinderPumpe-Düse-Motor von Volkswagen. Sonderausgabe der MTZ Motortechnische Zeitschrift, 1999. 7-72. Aschoff, G., Ebel, B., Eissing, S., and Metzner, F., “Der neue V6-Vierventilmotor von Volkswagen,” MTZ Motortechnische Zeitschrift 60:11, 1999. 7-73. Fuoss, K., Hannibal, W., and Paul, M. Mehrzylinder-Brennkraftmaschine. Patentanmeldung DE 34 44 501, Deutsches Patentamt München, 1993.

7-90. Albrecht, K. H., et al., “Optimierung von Kurbelwellen aus Gusseisen mit Kugelgraphit,” MTZ 47(7/8):277–283, 1986. 7-91. Sonsino, G.M., et al., “Schwingfestigkeit von festgewalzten, induktionsgehärteten sowie kombiniert behandelten Eisen-Graphit-Gusswerkstoffen unter konstanten und zufallsartigen Belastungen,” Gießereiforschung 42(3):110–121, 1990. 7-92. Hagen, W.W., “Neuere Entwicklungen bei geschmiedeten Kraftfahrzeug-Kurbelwellen,” Schmiede-Journal, 14–17, 2001. 7-93. Götz, C., “Wärmebehandlungskriterien bei der Werkstoffauswahl für Kurbelwellen,” MTZ 134–139, 2003. 7-94. Prandstötter, M., Riener, H., and Steinbatz, M., “Simulation of an Engine Speed-Up Run: Integration of MBS-FE-EHD-Fatigue,” ADAMS User Conference 2002, London. 7-95. Menk, W., Kniewallner, L., and Prukner, S., “Neue Perspektiven im Fahrzeugbau—Gegossene Kurbelwellen als Alternative zu geschmiedeten,” MTZ 384–388, 2007. 7-96. Muckelbauer M. and Arndt, J., “Schmiedeteile behaupten sich erfolgreich im Technologiewettbewerb,” Schmiede-Journal, 36–38, 2008. 7-97. Fröschl. J., Achatz, F., Rödling, S., and Decker, M. Innovatives Bauteilprüfungskonzept für Kurbelwellen. MTZ 09.2010, Wiesbaden.

316 | Internal Combustion Engine Handbook

6606_Book.indb 316

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.25 Control Mechanisms for Two-Stroke Engines

7-98. Krivachy, R., Linke, A., and Pinkernell, D. Numerischer Dauerfestigkeitsnachweis für Kurbelwellen. MTZ 06.2010, Wiesbaden, Juni 2010.

7-124. Todsen, U. Interne Schulungsunterlagen Motorentechnik, FH Hannover, Labor f. Kolbenmaschinen, 1996.

7-99. Gümpel, P. and Wägner, M. Harte und verschleißfeste Randschichten auf korrosionsbeständigen Stählen. MTZ 09.2010, Wiesbaden.

7-125. Funabashi, N., et. al., “US-Japan PM Valve Guide History and Technology,” Proceedings of the international Symposium on Valve Train Systems Design and Materials, ASM, 1998.

7-100. ÖKV, Institut für Verbrennungskraftmaschinen, Kraftfahrzeugbau der TU Wien (Veranst.): 22. Internationales Wiener—Motorensymposiuum (Wien 2001). Düsseldorf: VDI, 2001. Tagungsschrift. 7-101. Neußer, Porsche AG, ÖKV, Institut für Verbrennungsmaschinen, Kraftzeugbau der TU Wien (Veranst.): 21. Internationales Wiener Motorensymposium (Wien 2000). 7-102. Wurms, R. Audi AG, Aachener Kolloqioim 2008, Audi Valvelift System. 7-103. Eidenböck, T., Ratzberger, R., Stastny, J., and Stütz, W., “Zylinderkopf in Vierventiltechnik für den BMW Di-Dieselmotor,” MTZ 6:372, 1998. 7-104. Groth, K. Universität Hannover, Institut für Kolbenmaschinen: Grundzüge des Kolbenmaschinenbaus. 2. Aufl. 1983. 7-105. TRW Thompson GmbH & Co. KG: Handbuch. 7. Aufl. 1991. 7-106. Milbach, R. TRW Thompson GmbH & Co. KG: Ventilschäden und ihre Ursache. 5. Aufl. 1989. 7-107. Muhr, T. Zur Konstruktion von Ventilfedern in hochbeanspruchten Verbrennungsmotoren. Diss. RWTH Aachen, 1992

7-126. Rehr, A. Offenlegungsschrift DE3937402 A1, Deutsches Patentamt, 1991. 7-127. Myron, F. Hill. Kinematics of Gerotors. 2. Aufl. The Peter Reilly Company. 7-128. Beck, Th. Beiträge zur Geschichte des Maschinenbaus. Springer Verlag, 1999. 7-129. Lips, W. Strömungsakustik in Theorie und Praxis. 2. Aufl. Expert Verlag. 7-130. VDI Richtlinie 3743 Bl. 1+2. Emissionskennwerte technischer Schallquellen. 7-131. Heckl, M. and Müller, H.A. Taschenbuch der technischen Akustik. 2. Aufl. Springer Verlag. 7-132. Findeisen, D. u. F. Ölhydraulik. 4. Aufl. Springer Verlag. 7-133. Patentschrift FR000000980766A. 7-134. EG-Richtlinie 80/1268/EWG.

7-108. Deutsches Institut für Normung e.V. (Hrsg.). DIN-Taschenbücher, Berlin: Beuth, Zylindrische Schraubendruckfedern aus runden Stäben, DIN 2089 (Teil 1), 7. Aufl. 1984.

7-135. Lamparski, C. Bedarfsgerechte Ölversorgung; Regelölpumpen im Serieneinsatz, Ölkreislauf von Verbrennungsmotoren II, Expert Verlag, 2007.

7-109. Niepage, P. et al., “Meßstellenermittlung und Meßwertkalibrierung zur Spannungmessung an Ventilfedern mittels Dehnungsmessstreifen,” Draht 41(3):333–336, 1990.

7-137. Bensinger, W.-D. Die Steuerung des Gaswechsels in schnelllaufenden Verbrennungsmotoren. Konstruktionsbücher, Bd. 16, Springer Verlag, 1967.

7-110. Yamamoto. Valve Spring Made by Sankos Multi-Arc Wire, Kyoto: Sanko Senzai Kogyo Co. Ltd., 1989. 7-111. VDA-Mitteilungen 2008, Internet: www.vda.de. 7-112. Dolenski, T. Konstruktion eines Hochtemperatur-Stift-ScheibeVerschleißprüfstandes. Diplomarbeit FH Bochum, 1998. 7-113. SAE Valve Seat Information Report, SAE J 1692, Society for Automotive Engineers, Inc., Warrendale, PA, 1993. 7-114. Rodrigues, H., “Sintered Valve Seat Inserts and Valve Guides: Factors Affecting Design, Performance & Machinability,” Proceedings of the International Symposium on Valve Train System Design and Materials, ASM, 1997. 7-115. Dooley, D., Trudeau, T., and Bancroft, D., “Materials and Design Aspects of Modern Valve Seat Inserts,” Proceedings of the International Symposium on Valve Train System Design and Materials, ASM, 1997.

7-138. Holland, J. Die instationäre Elastohydrodynamik. Konstruktion 30, Heft 9, 1978. 7-139. Ruhr, W. Nockenverschleiß—Auslegung und Optimierung von Nockentrieben hinsichtlich des Verschleißverhaltens. FVV-Vorhaben Nr. 285, 1985. 7-140. Holland, J. Nockentrieb Reibungsverhältnisse—Untersuchung zur Verminderung der Reibung am Nocken-Gegenläufer-System unter Verwendung von Gleit- und Rollengegenläufern. FVV-Vorhaben Nr. 341, 1986. 7-141. Brands, Ch. Dynamische Ventilbelastung—Rechnergestützte Simulation der Beanspruchung des Ventiltriebs. FVV-Vorhaben Nr. 614, 1998. 7-142. Dachs, A. Beitrag zur Simulation und Messung von Tassenstößelventiltrieben mit hydraulischem Ventilspielausgleich. Diss. TU Wien, 1993. 7-143. Ruhr, W. Nockentriebe mit Schwinghebel. Diss. TU Clausthal, 1985.

7-116. Motooka, N., et al. Double-Layer Seat Inserts for Passenger Car Diesel Engines, SAE Technical Paper Series 850455, 1985.

7-144. Rahnejat, H. Multi-Body Dynamics. Vehicles, Machines and Mechanisms. SAE International, 1998.

7-117. Valve seat insert information report, SAE J 1692, 30. 08. 1993.

7-145. Beitz, W. and Küttner, K.-H.: Dubbel Taschenbuch des Maschinenbau. Springer Verlag.

7-118. Richmond, J., Barrett, D.J.S., and Whimpenny, C.V. ImechE, C389/057, 121–128, 1992. 7-119. Dalal, K., Krüger, G., Todsen, U., and Nadkarni, A. Dispersion Strengthened Copper Valve Seat Inserts and Guides for Automotive Engines, SAE Technical Paper Series 980327, 1998. 7-120. Rehr, A. Offenlegungsschrift DE 3937402 A1, Deutsches Patentamt, 1991. 7-121. Meinecke, M. Öltransportmechanismen an den Ventilen von 4-Takt-Dieselmotoren, FVV-Abschlussbericht Vorhaben Nr. 556, Institut für Reibungstechnik und Maschinenkinetik, Technische Universität Clausthal, 1994. 7-122. Linke, A. and Ludwig, F. Handbuch TRW Motorenteile, TRW Motorkomponenten GmbH. 7. Aufl., 1991. 7-123. NN. Kupfer-Zink-Legierungen, Messing und Sondermessing, Informationsdruck Deutsches Kupfer-Institut, Nr. I 005.

7-146. Wurms, R., Dengler, S., Budack, R., Mendl, G., Dicke, T., and Eiser, A. Audi valvelift system—ein neues innovatives Ventiltriebsystem von Audi, 15. Aachener Kolloquium Fahrzeug- und Motorentechnik, 2006. 7-147. Schneider, F. and Simmonds, S. MAHLE CamInCam—Neue Freiheit für variable Steuerzeiten am Beispiel eines 8- und 4-Zylindermotors, MTZ/ ATZ Konferenz Ladungswechsel im Verbrennungsmotor, 2008. 7-148. Mohr, U., Hoffmann, H., and Lancefield, T. Modularität im Ventiltrieb, MTZ Konferenz Antrieb von Morgen, 2005. 7-149. Bunsen, E., Grote, A., Willand, J., Hoffmann, H., Fritz, O., and Senjic, S. Verbrauchspotenziale durch Einlassventilhub und Steuerzeitenvariation—ein mechanisch vollvariables Ventiltriebsystem an einem 1.6 l Motor mit Benzindirekteinspritzung, Variable Ventilsteuerung ISBN 978-3-8322-5910-5. 7-150. Schneider, F., Steichele, S., and Ruppel, S. Integration von Zusatzfunktionen in die gebaute Nockenwelle, 3. VDI-Tagung Ventiltrieb und Zylinderkopf Würzburg, 2008.

Internal Combustion Engine Handbook | 317

6606_Book.indb 317

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 7 Engine Components

7-151. Arnold, M., Farrenkopf, M., and Mc Namar, S. Zahnriementriebe mit Motorlebensdauer für zukünftige Motoren. 9. Aachener Kolloquium; Fahrzeug- und Motorentechnik. Aachen: 2000 ika/VKA. 7-152. IWIS-Ketten: Handbuch Kettentechnik. München. 7-153. Fritz, P. Dynamik schnellaufender Kettentriebe. VDI Fortschrittsberichte, Reihe 11: Schwingungstechnik, Nr. 253, VDI-Verlag GmbH: Düsseldorf, 1998. 7-154. Fink, T. and Hirschmann, V., “Kettentriebe für den Einsatz in modernen Verbrennungsmotoren,“ MTZ 62(10):796–806, 2001. 7-155. Hirschmann, V., Bongard, A., and Welke, L., “Neue Zahnkettengeneration für den Einsatz in Steuertrieben von Dieselmotoren,“ MTZ 67(11):878–883, 2006. 7-156. Arnold, M., Farrenkopf, M., and McNamara, St., “Zahnriementriebe mit Motorlebensdauer für zukünftige Motoren,” MTZ Motortechnische Zeitung 62:2, 2001.

7-177. Paffrath, H., Hummel, K.-E., and Alex, M. Technology for Future Intake Air Systems, SAE, März 1999. 7-178. Weber, O., Vaculik, R., Füßer, R., and Pricken, F. Qualitativ hochwertige Akustik von Ansaugsystemen und Kunststoffen—ein Widerspruch?; High Quality Acoustics of Plastic Intake Systems—Vision or Contradiction?; 20. Internationales Wiener Motorensymposium, Mai 1999. 7-179. Pricken, F. Active noise concellation in future air intake systems, SAE, 2000. 7-180. Pricken, F. Sound Design in the Passenger Compartment with Active Noise Control in the Air Intake System, SAE, 2001. 7-181. DIN ISO 362 Akustik, Messung des von beschleunigten Straßenfahrzeugen abgestrahlten Geräusches; Verfahren der Genauigkeitsklasse 2 (ISO 362 AMD 1-1985, ISO 382; 1981). 7-182. Büchler, A. Benchmarking—Polyamides in Automotive, Polymers and e-mobility in the automotive industry, Vol. 9 No. 03/11.

7-157. di Giacomo, T., Schulte, G., Steffens, C., Tiemann, C., Walter, R., and Wedowski, S., “Zahnriemen versus Kette, Studie zum CO2-Sparpotenzial im Steuertrieb,” MTZ-Artikel 70:5, 2009.

7-183. Diez, A. and Baur, M.: Geprägte Stopper—wichtiger Entwicklungsschritt bei Metaloflex-Metalllagen-Zylinderkopfdichtungen. MTZ 9/2004, S. 706–709.

7-158. Affenzeller, J. and Gläser, H.: Lagerung und Schmierung von Verbrennungsmotoren. Springer, 1996 (contains an extensive bibliography).

7-184. Bendl, K., Griesinger, E., and Lieb, F. Kunststoffmodule auch für Nutzfahrzeugmotoren—Gewichtsreduzierung und Kosteneinsparung. MTZ 9/2004, S. 714–718.

7-159. Lang, O.R. and Steinhilper, W. Berechnung und Konstruktion von Gleitlagern mit dynamischer Belastung. Springer, 1978. 7-160. N.N. Gleitlager-Handbuch, Miba Gleitlager AG, 2000. 7-161. Ederer, U.G. and Aufischer, R. Schadenswahrscheinlichkeit und Grenzen der Lebensdauer. TA Esslingen, 1992. 7-162. Arnold, O. and Budde, R. Konstruktive Gestaltung von Lagerungen in Verbrennungsmotoren. HdT Essen, 1999. 7-163. Ederer, U.G. Werkstoffe, Bauformen und Herstellung von Verbrennungsmotoren-Gleitlagern. HdT Essen, 1999.

7-185. Van Basshuysen/Schäfer (Hrsg.). 2006. Dichtsysteme. In: Lexikon Motorentechnik. Vieweg: Wiesbaden. 7-186. ElringKlinger AG. Fachdokumentationen Zylinderkopfdichtungen, Spezialdichtungen, Module- und Elastomer-Dichtsysteme. 7-187. Diez, A., Maier, U., Eifler, G., and Schnepf, M. Integrierte Drucksensorik in der Zylinderkopfdichtung. MTZ 1/2004, S. 22–25. 7-188. Griesinger, E. Ventilhaubenmodule von ElringKlinger—kompaktes Design, vielfältige Funktionen. MTZ 6/2003, S. 504–509.

7-164. DIN-ISO 7146: Schäden an Gleitlagern.

7-189. Walter, G. and Griesinger, E. Kunststoffmodule—Funktion und Ästhetik. In: ATZ/MTZ System Partners 2002, S. 32–37.

7-165. Knoll, G. and Umbach, S. Einfluss des Schmierstoffs auf die Mischreibung in Gleitlagern, MTZ, Wiesbaden, März 2010.

7-190. Diez, A. and Gruhler, T.: Dichtung mit Profil. In: Automobil Industrie Special Mercedes-Benz E-Klasse, Mai 2002, S. 60.

7-166. Damm, K., Skiadas, A., Witt, M., and Schwarze, H. Gleitlagererprobung anhand der Forderungen des Automobilmarkts, MTZ Extra 03.2010, Wiesbaden.

7-191. Zerfaß, H.-R. and Diez, A.: Zylinderkopfdichtungskonzepte für zukünftige Motorgenerationen. MTZ 1/2001, S. 30–35.

7-167. Linke, B. and Buck, R. Aggregatelager mit höherer Lebensdauer durch schwingungstechnische Analyse, MTZ 03.2010, Wiesbaden.

7-192. Jende, S. Robotergerechte Schrauben—Hochfeste Verbindungselemente für flexible Automaten, Techno TIP 12/84. Würzburg: Vogel-Buchverlag.

7-168. Adam, A., Prefot, M., and Wilhelm, M. Kurbelwellenlager für Motoren mit Start-Stopp-Sytem, MTZ 12.2010, Wiesbaden.

7-193. N.N. Industriewerkzeuge—Montagewerkzeuge, company catalog 2000–2001 from Atlas Copco Tools GmbH, Essen.

7-169. Anwender, D., Brodesser, K., and Morgillo, I., “Das variable Ansaugsystem für die V6-Ottomotoren von Audi,” MTZ 65(10):792–796, 2004.

7-194. N.N. Schraub- und Einpresssysteme. Company catalog from Robert Bosch GmbH Automationstechnik, Edition 1.1 (2001), Murrhardt.

7-170. Hummel, K.-E., Huurdeman, B., Diem, J., and Saumweber, C. Ansaugmodul mit indirektem und integriertem Ladeluftkühler, MTZ 11.2010, Wiesbaden.

7-195. Technical Specification ISO/TS 16949: QM systems: Special requirements when applying ISO 9001: 2000 für die Serien- und Ersatzteilproduktion in der Automobilindustrie, VDA-QMC, 2002, Frankfurt.

7-171. Müller, K. and Mayer, W. Einfluss der Ventilgeometrie auf das Einströmverhalten in den Brennraum. 3. Aufl. Wiesbaden: Vieweg, 1999.

7-196. Jende, S. and Mages, W. Roboterschrauben. Wie sollen Roboterschrauben gestaltet sein? Schriftreihe Angewandte Technik (Verlag für Technikliteratur, 1990), pp. 12–18.

7-172. Wild, S. Torque vs. Power—No Conflict with Highly Variable Resonance Runners; Global Powertrain Congress, Detroit 2001. 7-173. Weber, O. and Wild, St.: Leistung plus Drehmoment—optimierte Sauganlage mit voll variablen Resonanzrohren; 22. Internationales Wiener Motorensymposium 2001; VDI Fortschrittsberichte Reihe 12, Nr. 455, Band 2, S. 320–332. 7-174. Alex, M. Akustikoptimierung bei der Filterentwicklung; Haus der Technik Essen, 1996. 7-175. Weber, O. topsys—A New Concept for Intake Systems; SAE 98 Merra. 7-176. Weber, O., Paffrath, H., Beutnagel, H., and Cedzich, W. Thermodynamische und akustische Auslegung von Ansaugsystemen für Fahrzeugmotoren unter Berücksichtigung fertigungstechnischer Belange; 19. Internationales Wiener Motorensymposium, Mai 1998.

7-197. Jende, S. 1986. Automatische Montage hochfester Schrauben— Anwendungsbeispiele aus der Praxis—wt—Zweitschrift für industrielle Fertigung. Springer-Verlag: Berlin; Heidelberg. 7-198. N.N. Informations-Centrum Schrauben—Automatische Schraubmontage, Publisher: Deutscher Schraubenverband e.V., Hagen, 2nd Edition, 1997 (Iserlohn: Mönning-Druck). 7-199. Jende, S. and Knackstedt, R. Warum Dehnschaftschrauben? Definition—Wirkungsweise—Aufgaben—Gestaltung. VDI-Z 128 (1986), No 12. 7-200. Illgner, K.H. and Blume, D. Schraubenvademecum Publisher: Bauer & Schauerte Karcher GmbH, 6th edition. 7-201. Lang, O.R. 1966. Triebwerke schnelllaufender Verbrennungsmotoren. Konstruktionsbücher No. 22. Springer Verlag: Berlin; Heidelberg.

318 | Internal Combustion Engine Handbook

6606_Book.indb 318

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

7.25 Control Mechanisms for Two-Stroke Engines

7-202. Grohe, H. Otto- und Dieselmotoren: Arbeitsweise, Aufbau und Berechnung von Zweitakt- und Viertakt-Verbrennungsmotoren. Kamprath—Reihe concise, Technik, 6th Edition. Würzburg: VogelBuchverlag, 1982. 7-203. VDI: Systematic calculation of highly stressed bolted joints. VDI Guideline 2230, 2002. 7-204. Jende, S., “KABOLT—ein Berechnungsprogramm für hochfeste Schraubenverbindungen, Beispiel: Die Pleuelschraube,” VDI-Z 132(7):66/78, 1990. 7-205. Schraubenberechnungsprogramme: VDI-Software “Schraubenberechnung” 3rd Edition April 2009, Beuth-Verlag GmbH, Berlin (over1000 demo versions issued) SR1-Schraubenberechnung according to VDI 2230; update 2007; Prof. Schwarz, Uni Siegen (over 170 installations). 7-206. Esser, J. Ermüdungsbruch—Einführung in die neuzeitliche Schraubenberechnung, 23rd Edition 1999 (TEXTRON Verbindungstechnik GmbH + Co. OHG, Neuss, 1998. 7-207. Kübler, K.H., Turlach, G., and Jende, S. Schraubenbrevier, 3rd Edition 1990 (KAMAX-Werke Rudolf Kellermann GmbH & Co. KG, Osterode am Harz). 7-208. Scheiding W. Verschrauben von Magnesium braucht mehr als Alltagswissen, Konstruktion und Engineering, 04/01. Landsberg/Lech: Verlag Moderne Industrie. 7-209. Westphal, K. Verschraubung von Magnesiumkomponenten. In: Metall, 56. Jg., Heft 1–2 2002 (Giesel-Verlag, Isernhagen). 7-210. Kübler, K.H. and Mages, W. Handbuch der hochfesten Schrauben. Publisher: KAMAX-Werke, 1st Edition. Essen: Verlag W. Girardet, 1986. 7-211. Jende, S. and Mages, W. Schraubengestaltung für streckgrenzüberschreitende Anzugsverfahren—überelastische Grenzgänger. KEM, Edition 9/1986 Leinfelden-Echterdingen: Konradin Verlag. 7-212. “Grenzen für Grauguss,” Automobil-Produktion, October 2000. 7-213. Hein, M. Deutsches Patentamt, Offenlegungsschrift DE 4324458A1; Az.: P4324458.0, 1.94. 7-214. Weltens, H., Garcia, P., and Neumaier, H. “Neue Leichtbaukonzepte bei Pkw-Abgasanlagen sparen Gewicht und Kosten.” 7-215. Voeltz, V., Kuphal, A., Leiske, S., and Fritz, A. “Der Abgaskrümmer— Vorkatalysator für die neuen 1.0 l- und 1.4 l-Motoren von Volkswagen,” MTZ 60:7/8, 1999.

7-217. Hockel, K., “Der Abgaskrümmer von Personenwagenmotoren als Entwicklungsaufgabe,” MTZ 45:10, 1984. 7-218. Hummel, K.-E., Huurdeman, B., Diem, J., and Saumweber, C., “Ansaugmodul mit indirektem und integriertem Ladeluftkühler,” MTZ 11.2010, Wiesbaden. 7-219. Diez, R., Kornherr, H., Pirntke, F., and Schmidt, J. Effizienzsteigerung durch Zylinderbank- übergreifende Krümmer, MTZ 05.2010, Wiesbaden, Mai 2010. 7-220. Brömmel, A., et al. Elektrifizierung treibt Pumpeninnovationen, MTZ Extra 03.2010 Wiesbaden. 7-221. Keller, P., Wenzel, W., Becker, M., and Roby, J. Hybrid- Kühlmittelpumpe mit elektrischem und mechanischem Antrieb, MTZ 11.2010, Wiesbaden. 7-222. Wickerath, B., Fournier, A., Duran, J.-M., and Brömmel, A. Voll-variable mechanische Kühlmittelpumpe für Nutzfahrzeuge, MTZ 01.2011, Wiesbaden. 7-222. Venedinger, H. J.: Zweitaktspülung insbesondere Umkehrspülung. Stuttgart: Franckh`sche Verlagshandlung, 1947. 7-223. Bönsch, H. W. Der schnelllaufende Zweitaktmotor. 2nd Edition. Stuttgart: Motorbuch Verlag, 1983. 7-224. Küntscher, V. (Hrsg.). Kraftfahrzeugmotoren—Auslegung und Konstruktion. 3rd Edition. Berlin: Verlag Technik, 1995. 7-225. N.N. Hütte; des Ingenieurs Taschenbuch IIA; 28th Edition. Berlin: Verlag Wilhelm Ernst & Sohn, 1954. 7-226. Scheiterlein, A. Der Aufbau der raschlaufenden Verbrennungskraftmaschine. 2nd Edition Vienna: Springer Verlag, 1964. 7-227. Knoll, R., Prenninger, P., and Feichtinger, G. 2-Takt-Prof. List Dieselmotor, der Komfortmotor für zukünftige kleine Pkw-Antriebe; 17th International Vienna Motor Symposium 25–26 April 1996, VDI Progress Reports Series 12 No. 267. Düsseldorf: VDI Verlag, 1996. 7-228. Blair, G.P. Design and Simulation of Two-Stroke Engines. Warrendale PA: SAE International, 1996. 7-229. Meinig, U. Standortbestimmung des Zweitaktmotors als Pkw-Antrieb: Parts 1–4. In: MTZ Jahrgang 62, (2001) Heft 7/8, 9, 10, 11. 7-230. van Basshuysen, R. Zweitaktmotor/Wankelmotor, MTZ Jahrgang 70, Heft 01/2009.

7-216. Eichmüller, C., Hofstetter, G., Willeke, W., and Gauch, P., “Die Abgasanlage des neuen BMW M 3,” MTZ 62:3, 2001.

Internal Combustion Engine Handbook | 319

6606_Book.indb 319

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 320

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8 Engines 8.1 Engine Concepts The engine concept is affected by many factors, which frequently are not freely selectable, such as working cycles (two cycle or four cycle), working processes (diesel or spark ignition—SI), cooling type (water or air), output graduation, cylinder number and arrangement, crank gear configuration, crankcase design, control type, charging, etc. The most important criterion for an engine is its purpose (Figure 8.1), which defines the conditions under which certain requirements must be produced.

Using the laws of similarity mechanics, one can demonstrate that the individual factors of the power Equation (8.1)

P = z⋅

n p 2 ⋅ d ⋅ s ⋅ we ⋅ 4 i

(8.1)

are not independent of one another. For example, not only power, working volume and speed are interlinked but also working cycles, combustion process, cooling type and more. Large absolute outputs can be represented only with large cylinder dimensions (bore and stroke), while work cycle

Figure 8.1 Engine purpose and engine size (source: Zima).

Internal Combustion Engine Handbook | 321

6606_Book.indb 321

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

frequency and high specific work are used for large specific outputs (power/work area and power/motor mass). With increased speed, however, problems will grow. The effects of the masses to be accelerated become more difficult to control as are the thermal stresses, load changes and the injection of fuel. For this reason, speed as a means to increase output has tight limits, at least for series capable and inexpensive passenger car engines. Similar considerations apply to the engine’s working volume. Large working volumes are also limited for passenger car engines due to large cylinder volumes or a large number of cylinders. They increase the engine mass and the required space. But both must be available in a reasonable ratio to the vehicle-related data, such as vehicle mass and available space. Furthermore, large individual cylinder volumes also mean large piston masses to be accelerated which also represents a limitation. Bore diameters between approximately 70 and 110 mm with a nearly square stroke/bore ratio cover the range of today’s standard passenger car engines. The constructive design of engines is basically predefined by the crank gear’s mode of action. Reciprocating piston engines have proven their superiority. Also, this must examined integrally across all important properties. For example, engines with external combustion certainly offer advantages in respect of the individual aspect of thermodynamics but other aspects, such as space, must be assessed as negative factors. The work cycle frequency determines with the number of cycles the engines’ power density; a two-cycle engine produces double the output of a four-cycle engine, at least theoretically, because it has one ignition event for each rotation of the crankshaft. That this is not the case in practice is caused by the lower specific work accomplished by a two-cycle engine. In principle, one can differentiate between single- and multishaft engines; however, multishaft engines are not discussed here. Passenger car engines are manufactured only as single-shaft engines. The development and optimization goals in engine development for passenger car engines are essentially characterized by demands for the following:

•• multihole injection nozzles with air-distributing combustion processes •• high-pressure injection systems with injection pressures of approximately 2000 bar •• predominantly, four valves per cylinder with centrally installed injection nozzle •• air intake systems generating intake twists •• electronic diesel regulation with expanded functionality •• aluminum used as cylinder head material and, increasingly, for the cylinder block •• exhaust gas treatment systems with catalytic converter NOx storage, reduction agent, and particulate filter •• exhaust gas turbocharging with variable turbine geometry and intercooling •• cooled exhaust gas recirculation •• predominantly, four-, six-, or eight-cylinder engines with six-cylinder engines built in V and series design and eight-cylinders in V design •• energy and heat management. •• Spark-ignition engine: •• Engines with intake manifold or direct injection, with direct injection becoming more dominant due to consumption and performance reasons. •• Naturally aspirated engines prevail, with turbo engines with mechanical supercharging, but mostly exhaust gas turbocharging, being used. •• Turbocharging will intensify, in conjunction with a downsizing concept in particular. •• Aluminum used as cylinder head material and magnesium for the cylinder block. •• Variable camshaft shifting systems up to fully variable shifting. •• Controlled three-way catalytic converter with lambda sensors. •• Predominantly four valves per cylinder.

•• improved performance

•• Single-ignition coils and cylinder-selective ignition angle shift.

•• minimized fuel consumption and CO2 emission

•• Cylinder-selective injection.

•• meeting the exhaust gas quality requirements of, for example, EU4, EU5, ULEV

•• Cylinder shutdown in large-volume and multicylinder engines.

•• improved comfort and acoustics

•• Heat management to optimize cooling and engine heating.

•• minimized costs.

•• Controlled intake manifolds for changing pipe lengths in the intake area.

These require the development of assemblies, systems and modules that can meet the aforementioned and somewhat opposed demands. Development steps derived from the objective definition must usually involve a trade-off. Modern engines for individual traffic are characterized by the following overlapping features: •• Diesel engine: •• engines with direct injection

•• Preferably four-, six-, or eight-cylinder engines with six-cylinder engines built in V and series design and eight-cylinders in V design. •• Increased “hybridization” of the powertrain, for example, start-generator, electrically controlled accessories, micro and mild hybrids. •• Energy and heat management.

322 | Internal Combustion Engine Handbook

6606_Book.indb 322

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.1 Engine Concepts

8.1.1 Engine Models

Engines for use in passenger cars are single-shaft engines. Single-shaft engines are crank gears pf reciprocating piston engines featuring one crankshaft. These main types are distinguished as follows: Boxer engine: An engine with the cylinders arranged in one plane with two opposing cylinder banks. They can be also called 180°V engines in which the opposing pistons and conrods work on a common throw or on their own throw. Throw is defined here as the sequence of crankshaft web—crank pin—crankshaft web. The cylinders of adjacent crankshaft throws Figure 8.2 are installed opposite each other. For this reason, they can be arranged more closely to each in crankshaft direction, enabling boxer engines to be built with shorter dimensions than inline engines with the same number of cylinders. They are also not very high. Throw

r Figure 8.3 Air-cooled single-cylinder diesel engine (Hatz).

r = Crankshaft radius Crankshaft web

Crank pin

Figure 8.2 Crankshaft throws of a boxer engine.

Known boxer engines are the air-cooled two-cylinder motorcycle by BMW, the legendary VW Beetle and Porsche engines. Basically, the throw comprises the connecting rod bearing and the crankshaft web connected by the main bearing journal. For inline and V engines, crankshafts with one to 10 throws and partitions at 180°, 120°, 90°, 72°, 60°, 45°, 40° and 36° are common. Usually, crankshafts have bearings after every throw. The conrods may be led either through the piston or laterally through the throw. In latter case, higher demands are made on the accuracy of the stop collars and the widths of the crankshaft webs. Single-cylinder engines: The single-cylinder represents the basic unit of all engine concepts. It is irrelevant as a passenger car power source, however, and is only built for low performance with small cylinder dimensions. As an SI and diesel engine, it is usually air-cooled and is used to power work machines and current generators. As an SI engine, it is also used to power smaller motorcycles and motorbikes. Special measures are required to balance mass effects and to even out the torques (e.g., mass balancing drives). Figure 8.3 shows a section through a single-cylinder engine. Straight engine: The straight (or inline) engine is the standard type. It is created by arranging several cylinders in crankshaft direction. The crankcases of inline engines have a simple design. Straight engines are easy to repair and service. Modern vehicle engines are designed as straight engines with up to six cylinders.

V engines: V engines are created from the combination of two cylinder blocks inclined to each other by a specific angle (V angle), with their crank gears affecting a common crankshaft—two pistons opposing each other in the V on one crankshaft throw. V engines offer benefits with respect to the following: •• High-power density with compact structure and easy access. •• Space requirements in the vehicle are low for V engines (short crank gear length); even six-cylinder engines can be installed transversally in the passenger car. •• The space between and below the engine rows can be conveniently utilized with engine parts (injection pump, exhaust turbocharger, filter, etc.) creating a very compact drive unit. •• Inline engines larger than six cylinders are no longer installed in modern passenger cars (V engines are possible for up to 12 cylinders). •• Fast-running high-performance engines are built as V types from six cylinders on. The disadvantages of V engines are with respect to the following: •• Larger lateral components of the bearing forces in V engines demand a more complex design of the crankshaft main bearing cover in the crankcase. •• The intake manifolds have a more complex design. •• Two “hot” engine sides are given. •• The charging efforts are higher than in inline engines. •• V engines behave less favorably regarding free mass effects compared to inline engines. Despite these disadvantages, the V engine is now a preferred type, in addition to the four-cylinder inline engine.

Internal Combustion Engine Handbook | 323

6606_Book.indb 323

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Another important design characteristic is the conrod-nextconrod realization. From a crank gear point of view, this is the simplest solution (equal conrods and equal bearings). Due to the force engagement being eccentric in crank pin direction (conrod offset), it requires offset divider walls. In “split pin” crankshafts (Figure 8.4), the wider crankshaft webs (for V engines) are offset in circumference direction resulting in better ignition spacing and thus, smoother running. Thrown crank pin

Figure 8.4 Crankshaft with offset crank pin (split pin).

V = 15° Symmetrically shifted cylinders, Volkswagen VR6 d = 81 mm

V = 72° MTU 16 V 595 d = 190 mm

V = 40° MTU 20 V 672 d = 185 mm

V = 90° Pielstick 16 PA4 185 d = 185 mm

An important design parameter of the V engine is the V angle. Selection criteria may be Ignition spacing, mass effects, charging, rotation oscillation behavior, limitation of the engine dimensions in width and height, utilization of the basic powertrain in SI and diesel engines, the number of engine cylinders, and the actually available fabrication equipment. In practice, V angles can have values anywhere from 0° (inline engine) to 180° (boxer motor), (Figure 8.5). Small V angles require longer conrods (smaller conrod ratios λ = r/l) to ensure the necessary free travel of the connecting rods from the cylinders. This yields a higher crankcase with reduced piston-side forces since the angular travel of the connecting rod is shorter. Equal ignition spacing is achieved when the V angle is selected to δ = 720°/number of cylinders (four-stroke engine). For vehicle engines and fast-running diesel engines, the 90° V angle is preferred since it allows the first-order inertial forces to be completely balanced with rotating counterweights; in addition, in eight-cylinder V-90° four-stroke engines, the V angle corresponds to the even angular ignition spacing. If the number of cylinders and V angle do not correspond, an even ignition spacing is still attained by “spreading” the crank pins by the difference between the V angle and angular ignition spacing (split-pin crankshaft, offset crank pin, stroke offset). Accordingly, six-cylinder passenger car and truck engines

V = 45° Deutz MWM 632 d = 250 mm

V = 120° Deutz MWM 816 d = 142 mm

V = 50° Sulzer ZA 40 S d = 400 mm

V = 60° MTU 20 V 1163 d = 230 mm

V = 180° Daimler-Benz OM 807 d = 138 mm V-angles of various engines

Figure 8.5 V-angles of various engines (source: Zima).

324 | Internal Combustion Engine Handbook

6606_Book.indb 324

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.1 Engine Concepts

Cylinders 2,4,6 Cylinders1,3, 5

Cylinders 8,10,12

r Direction of rotation

aS

–q

+q Bank 2

5 3 1

6

11

4

9

10 8

2

Bank angle V- angle Split-pin angle Cylinder distance Bank offset Shifting Conrod length Crankshaft radius Bore Effective stroke

12

B

d

Bank 1

A

RV-angle

Cylinders 7,9,11

aB

aV

l

are being built today with a V angle of 90° (such as Audi, Deutz, DaimlerChrysler), 60° (Ford), and even 54° (Opel) which requires a total crank offset of 30°, 60° or 66°. VR engines: The engine length of inline engines can be shortened by “pulling apart” the cylinders in the plane of the crank circle and “pressing them together” in crankshaft direction. The result is a V engine with a very small V angle (Figure 8.6). For a transversal six-cylinder engine in a passenger car, VW developed a VR6 engine with a V angle of 15°. The free travel of the crank gear was achieved by pulling the cylinder series apart (at constant V angle) so that the cylinder axes did not intersect in crankshaft center but below (offset crankshaft drive). The advantages are as follows: Short engine length by tighter cylinder packing, small engine mass, only one cylinder head and low free mass effects. Disadvantages are caused by the different lengths for intake and exhaust ports, unequal spark plug positions of the two cylinder blocks and from various fire land heights due to the biased piston edge and less favorable conditions for load changes, combustion and pollutant emission.

7

aB aV aS A B q l r d s

mm mm mm mm mm mm mm

72° 15° 12° 65 13 ±12,5 168,5 44,95 84 90,168

Figure 8.7 VW 12-cylinder W engine.

There is a series of other engine designs, such as twin bank engine, opposed-piston engine, radial engine and X engine, that cannot be further discussed in this context.

8.1.2 Differentiating Features of Engine Concepts Regarding the Basic Engine

Different features of engine concepts are discussed as follows:

Figure 8.6 RV arrangement.

W engines: In W engines, the three-cylinder blocks affect a common crankshaft. The individual cylinder blocks may be offset to each other in longitudinal direction resulting in advantages relative to the engine length. Current example is the VW W12 6-liter engine presenting a combination of two V6 engines with a bank angle of 72°, a V angle of 15° and a common crankshaft, Figure 8.7.

•• Crankshaft position: Passenger car engines are usually fitted with crankshafts arranged below. In special cases, such as drives for power generators or specific military application, engines with vertical crankshafts have been built. Modern passenger car engines are fitted only with horizontal crankshafts. •• Cylinder position: In most engines, the cylinders are arranged upright and possibly with a slight incline, Figure 8.8.

Internal Combustion Engine Handbook | 325

6606_Book.indb 325

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

•• Engines with vertical cylinders (occasionally inclined to some degrees). Control and crank gear are easily accessible. The engine oil collects in the lowest area of the oil pan from where it is sucked in and returned into the circuit. Because the height of the engine space in modern passenger cars must be low for a favorable coefficient of air resistance (cw value), inline engines are installed at bias in many cases. Vertically arranged cylinders

Horizontally arranged cylinders

Angularly arranged cylinders

Suspended cylinders

Figure 8.8 Cylinder position: vertical and inclined, horizontal, and suspended.

The incline is usually contingent on the installation situation and space conditions of the vehicle. The pistons affect the horizontal crankshaft. When the crankcase is rotated around the crankshaft, one creates the following: •• Engines with suspended cylinders that have been used only in airplanes. •• Engines with horizontal cylinders (underfloor engines). Underfloor cylinders are used for packaging reasons when low installation heights are required. The engine is rotated by 90° (horizontal) or as V engine with a 180° V angle (boxer engine).

Inline engine

X-engine

V-Engine

Radial series engine

•• Cylinder arrangement: The cylinder arrangement (Figure 8.9) is optimized for low-space requirement, low power-tomass ratio (mengine/P) and dynamic behavior (inertial forces) resulting in many possible combinations. Radial engines are the result when the cylinders concentrically affect a crankshaft throw. Multishaft engines are achieved with cylinder blocks in a polygon with more than one crankshaft. The combination of several inline engines results in V, W, and X engines with a single crankshaft and double-block and H-block engines with two crankshafts. Two or more “stars” combined result in double star and multistar but also multibank inline engines. These considerations in arranging cylinders are as follows: •• The installation conditions and the package limit the engine dimensions. This is the length for a transversal engine in a passenger car, whereas in underfloor engines the height, and in airplane engines, the front face must be considered (due to aerodynamic drag). •• The sensitivity of crankshafts against rotational oscillations which increases with the number of throws. For this reason, inline engines with more than six cylinders are not common. •• Number of cylinders: The basic engine design is the singlecylinder engine. In view of performance, torque, uniformity of torque and speed, passenger car engines are designed only with multiple cylinders. The top number of cylinders is limited by, for example, the application, the available pace, the engine mass, production costs, maintenance efforts and the rotation oscillation behavior of the crank gear. Modern passenger car engines have three to 18 cylinders, motorbikes one to four and commercial vehicle engines four to 12.

W-Engine

Radial engine

Boxer engine

Dual radial engine

Quad radial engine

Figure 8.9 Cylinder arrangement 1 (source: Zima).

326 | Internal Combustion Engine Handbook

6606_Book.indb 326

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.1 Engine Concepts

A further essential characteristic of an engine design is the type series derived from the principally established parameters such as stroke and bore. The basic engine is designed so that, by “adding” further cylinders the number of cylinders, and thus, performance and torque can be adapted to the vehicle requirements, while maintaining, for example, the stroke and bore parameters.

8.1.3 Further Concept Criteria

Considerations so far have mostly related to the mechanical structure of the basic engine. It is, however, possible to present a number of other concept-influencing criteria that would exceed the scope of this contribution. For this reason, only some selected examples will be listed below. Other evaluation criteria taken into consideration during the concept development (for details to the individual elements, see the relevant chapters) may be the following: •• The cooling type, differentiated between fluid cooling and air cooling. The majority of engines used in today’s passenger cars are fluid-cooled, because the cooling effect is more efficient, compared to air-cooling, without having noise effects caused by the blower used for air delivery. •• Mixture formation and combustion processes (heterogeneous mixture formation in the diesel engine and the combustion process derived thereof, internal (inhomogeneous/ homogeneous/stratified) and external mixture formation in the SI engine). •• The type of air intake such as the naturally aspirated engine and the supercharged engine. •• The ignition type such as autoignition (diesel) and ignition initiation by external energy input (SI engine). •• Two-cycle and four-cycle working cycles, although only four-stroke engines are used today as passenger car engines. Within a basic engine concept, further concept-relevant subdivisions may be derived. For example, the manifold construction and function types of a valve timing system— from the fixed assignment of control times across all load and speed ranges on the engine to the fully variable valve actuation in which valve stroke, valve control time and valve opening period are freely selectable. Other examples would be variety of crankcase and cylinder head designs and the oil supply. Reference is made to the corresponding chapters.

8.1.4 Engine Arrangement Concepts in the Vehicle

Engine and vehicle concepts must be harmonized so that the selected motorization in the specific vehicle can realize the specified driving properties. Furthermore, there is a close reciprocal effect between engine and powertrain concept, for example, in the gear drive arrangement. To integrate the engine in an overall vehicle concept, the following issues are at the forefront: •• Arrangement of cylinders: The engine type, such as inline, V, W or boxer, in which the arrangement is determined

by a number of criteria, such as available space, required engine output and the arrangement in the vehicle itself. This also predetermines essential dimensions of the basic engine defining the displacement, the crankshaft position, accessory dimensions, engine bearing, vibration behavior, etc. as influencing factors on the engine concept. •• Engine arrangement in the vehicle: A distinction is made between longitudinal and transversal engine installation. The variant of each can be combined with the position of the engine in the vehicle as front, center or rear engine. These arrangements may then be realized as conventional or underfloor engines. The vehicle specifications determine, to a large extent, the engine arrangement. For example, a demand for a short vehicle length and thus limited engine compartment room had the effect that six-cylinder inline engines have been largely replaced by six-cylinder V engines. Four-cylinder inline engines are frequently transversely installed because this version, in combination with the transmission, results in a relatively short engine length. For this reason, transversally installed engines have prevailed in underfloor engine arrangements. The main advantages and disadvantages of the individual variants are: •• Front engine/transversal mounting: Advantages, such as compact dimensions, short front section and short conduits, have the downside of higher efforts for engine bearing and the available width between the longitudinal beams. This arrangement is particularly well suited for inline engines with three or four cylinders, RV and V6 engines. The use as an underfloor engine is limited to three- and four-cylinder inline engines. •• Front engine/longitudinal mounting: This type can be realized for nearly all engines. Long engines, such as V12 engines, are possible (with two banks of six cylinders each). The larger front section length and the width of the transmission tunnel may be problematic. •• Center engine/transversal mounting: This variant is predominantly suited for three- to five-cylinder inline engines. In addition to the short front section, its excellent axle load distribution must be noted. However, the width of the rear longitudinal beams is an issue, as is the fact that all-wheel transmission is not possible in this variant. Furthermore, center engines are suited only for two-seater vehicles (roadsters). •• Center engine/longitudinal mounting: As a rule, the center engine with longitudinal installation is suitable for all engine designs from three-cylinder inline to V engines to boxer engines. Otherwise, the same advantages and disadvantages as for the transversally installed center engine apply. •• Rear engine/transversal mounting: The advantages of this design are good traction and, with the underfloor variant, a very compact vehicle design. The disadvantages are the necessary width between the rear longitudinal beams and the restrictions in accessibility of cabin and storage space. The underfloor variant is suitable only for

Internal Combustion Engine Handbook | 327

6606_Book.indb 327

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

three- and four-cylinder engines, while the conventional variant is still used for V engines. •• Rear engine/longitudinal mounting: In addition to excellent traction, optimal weight distribution during braking, space utilization (storage and luggage) above the engine are potential benefits. The disadvantages are the engine and transmission length with long projection to the rear, major efforts in conduit installation and an adverse axle load distribution with potential negative affects on driving behavior. Suitable here are engines in inline and V design including boxer engines. Furthermore, the drive concept plays a role in the engine conception for a specific vehicle, differentiated between front, rear and all-wheel drive, and the powertrain concept examining the arrangement of transmission, differential gear, propeller shaft, etc.

8.2 Current Engines 8.2.1 V6 Diesel Engine by Mercedes-Benz

This V6 variant replaces all five- and six-cylinder inline engines in the vehicle series. The essential engine data are presented in Figure 8.10.

intercooling system. Part of the distributor modules of the charge air is an electrically actuated inlet duct shutdown for focused swirl control. In respect to the minimization of noise, measures have been implemented such as a very rigidly designed aluminum crankcase, a balance shaft arranged in V shape, a cylinder head shroud with integrated camshaft bearing, acoustic measures relative to air ducting, chain guide, engine foam covering, etc. Two oxidation catalytic converters are provided for exhaust gas treatment, one of which is installed close to the engine and the other acts as the main catalytic converter. A particulate filter (without additives) may be installed in specific countries.

8.2.2 4-Liter V8 Diesel Engine by Mercedes-Benz

The objectives for the development of the engine were, in addition to an increase in performance and torque, the compliance with Exhaust Gas Regulation Euro 4, the series application of a particulate filter without additives, a decrease in fuel consumption and improved noise properties compared with the previous model. The engine data are presented in Figure 8.11. Number of cylinders/model

–

V8

Bank angle

Degrees

75

Valves/cylinders

–

4

Displacement

cm3

3996

Bore

mm

86

Number of cylinders/arrangement

–

V6

Stroke

mm

86

Bank angle

Degrees

72

Cylinder spacing

mm

97

Valves/cylinders

–

4

Compression

–

17

Displacement

cm3

2987

Engine mass

kg

253

Bore

mm

83

Nominal output at speed

kW at min–1

231/3600

Stroke

mm

92

Nominal torque at speed

Nm at min–1

730/2200

Cylinder spacing

mm

106

Exhaust emission tier

–

Euro 4

Compression

–

18

Conrod length

mm

163

Nominal output at speed

kW/min–1

165/3800 or 173/3600

Nominal torque at speed

Nm/min–1

510/1600–2800 or 540/1600–2400

Engine mass

kg

208

Figure 8.10 OM 642 technical data.

The selected bank angle of 72° represents a compromise between space requirement and crank gear design. The crankcase with a weight of “only” 41 kg features cast-in-place gray-cast sleeves and is manufactured from AlSi6Cu in a gravity-casting process. The drop-forged crankshaft with quadruple bearing is manufactured from 42CrMo4, while the weight-optimized connecting rods are made from the new material 70MnVs4. For the injection system, the third generation of the commonrail technology with Piezo actuator module is used, which permits the realization of up to five injections. The injection nozzle is designed as an eight-hole nozzle. The engine features an electrically actuated EGR valve moving the cooled exhaust gas to the intake port into the

Figure 8.11 Engine data for the 4-liter V8 diesel engine.

Compared to the predecessor model, a significant lowering of the pressure loss could be achieved by dethrottling. The torque progression in the lower torque range could be improved by a prevolume upstream of the inlet into the compressor of the exhaust turbocharger. The guide vanes are positioned with an electric servomotor enabling fast positioning and a high degree of positioning accuracy. The exhaust gas is map-controlled supplied by two EGR valves, with the exhaust gas bypassing the EGR cooler during engine heating. Figure 8.12 shows the structure of the exhaustgas recirculation unit. The increased peak pressure of the engine made it necessary to adapt the cylinder head toward a higher rigidity. This was achieved by in intermediate deck in the water cavity, among others things. For the injection system, a third-generation common rail system with piezo injectors is used, enabling up to five injections per work cycle with seven-hole injection nozzles. Particular importance was given to an acoustic optimization at low speeds and loads. This was achieved, among other

328 | Internal Combustion Engine Handbook

6606_Book.indb 328

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.2 Current Engines

EGR valves Water outlet EGR introduction in the charge air distributor

Change-over damper EGR cooler bypass

Water inlet

Water outlet Exhaust gas inlet left EGR cooler

Figure 8.12 Exhaust-gas recirculation system.

things, by utilizing an extremely rigid crankcase, a larger main bearing diameter, balance shaft, more rigid design of the engine mount, covers and the uncoupling of various components, such as fuel lines and charge air distribution line.

8.2.3 V10 FSI engine by Audi

The engine was designed according to modular principle and proven components of the Audi V family have been utilized. Figure 8.13 shows the main technical data of the engine. Displacement

cm3

5204

Stroke

mm

92.8

Bore

mm

84.5

Cylinder spacing

mm

90

Length/width/height

mm

685/801/713

Valves/cylinders

–

4

Valve diameter/hub inlet

mm

33.85/11

Valve diameter/hub outlet

mm

28/11

Camshaft shifting range

Degrees

42

Compression

–

12.5

Output at speed

kW/min–1

331/7000

Torque at speed

Nm/min–1

549/3000–4000

Firing sequence

–

1-6-5-10-2-7-3-8-4-9

Engine mass

kg

220

Exhaust standard

–

EU IV

Figure 8.13 Audi V10 FSI engine technical data.

The cylinder crankcase was designed as bedplate version with the upper part manufactured as a homogeneous monoblock from AlSi17Cu4Mg and the die-cast bedplate from AlSi12Cu1 with cast-in-place gray-cast cylinder sleeves. The crankshaft from 42CrMoS4 was realized as a split-pin shaft resulting in an even ignition distance of 72 degrees. The free first-order inertial forces are compensated with an opposite-running balance shaft that rotates at the crankshaft speed.

The aluminum-cast piston features a piston head geometry adapted for the combustion process that supports charge movement accordingly. The high-thermal piston load is absorbed by an optimized piston cooling concept. The piston is designed with a cooling channel (salt core); the edge of the recess and the ring groove has been optimized accordingly. The inlet channels feature a separation plate for tumble regeneration. The assembled hollow camshafts are directly mounted in the aluminum and bolted to a longitudinal frame. The intake system is realized as a double-flow type and flow-optimized which made it possible to achieve a total pressure loss of just 40 mbar at an air throughput of 1200 kg/h. A quadruple-shelled controlled intake manifold from die-cast magnesium is used in two lengths. The V10-typical acoustic is filtered with special membrane and foam tuning and led through a “sound pipe” into the vehicle cabin. Fuel is supplied by two demand-controlled single-piston high-pressure pumps generating a working pressure of more than 100 bar. The high-pressure injection valves are manufactured as single-hole twirl valves arranged and realized so that minimum wall wetting can occur. The V10-FSI engine achieves specific values of 63 kW/liter and more than 100 Nm/liter.

8.2.4 1.6-Liter V8 Turbocharged Gasoline Engine by GM

One development objective was, among other things, the achievement of high specific values relative to the performance of 82.5 kW/liter and a torque of above 143 Nm/liter. Figure 8.14 provides the essential engine data. The basic engine with its geometry data is based on the naturally aspirated engine using identical parts. For the turbocharged variant, specifically, the oil/water heat exchanger and the rotational oscillation damper have been adapted and those components that are affected by the higher loads have been optimized accordingly. The inlet manifold, the exhaust manifold

Internal Combustion Engine Handbook | 329

6606_Book.indb 329

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

with an integrated turbocharger, the piston, the underfloor catalytic converter and the two-mass flywheel are some of the components that have been newly developed. Figure 8.15 shows the turbocharger with integrated exhaust manifold. Piston displacement

cm

3

Output [kW]/speed [min–1]

103/4200

Torque [Nm]/speed [min–1]

320/1750–2500

Displacement [dm3]

2.0

Cylinder spacing [mm]

88

1598

Camshaft axis clearance

54.6 1800

Cylinder spacing

mm

86

Injection pressure [bar]

Bore

mm

79

Number of valves/cylinders

4

Stroke

mm

81.5

Compression

16.5

Conrod mass

kg

480

Exhaust gas standard

Euro 5

Piston mass

kg

340

CO2 emission [g/km]

190

Intake valve diameter

mm

31.2

Exhaust valve diameter

mm

27.5

Valve stroke I/O

mm

7.0/7.0

Compression

–

8.8

Maximum output at speed

kW at min–1

132/5500

Maximum torque at speed

Nm at min–1

230/2200–5500

Engine mass

kg

131

Exhaust gas standard

–

Euro V

Figure 8.14 Engine data for the1.6-liter turbocharged gasoline by GM.

Figure 8.15 Turbocharger with integrated exhaust manifold.

An essential tool to achieve the development goals was simulation. In addition to the three-dimensional flow, simulation of the intake manifold and the inexpensive package variant achieved in this manner, complex simulations in connection with the design of the exhaust manifold were applied.

8.2.5 2-Liter 4V-TDI with Common-Rail by Volkswagen

This engine on the basis of the previous 2-liter unit was developed with the goal to reliably meeting the Euro 5 specification and to ensure a future-oriented solution toward Euro 6. This motor, built in series since 2007, is fitted with a common-rail system of the current CRS 3.2 generation. Figure 8.16 provides the essential engine data.

Figure 8. 16 Engine data of the 2-liter V4 TDI.

The engine features four valves per cylinder, which are driven by two camshafts. The valves actuated via roller cam followers are grouped around the centrally located injection valve. Straight-tooth spur pinions connect both camshafts with each other. The valve star is rotated by 90° around the engine longitudinal axis. Compared to its predecessor, an essential modification at the engine is the conversion to the common-rail system. To meet the exhaust gas specification and to improve the acoustics, the system was designed for up to seven injections per work cycle with volumetric metering and pressure regulation at the rail. An eight-hole injection valve with hole diameters of 0.123 mm is used. The improvement in respect to mixture formation results from an adaptation of the intake manifold and a low-temperature EGR. Turbulence and mass flow are set in this engine with constantly adjustable swirl and tumble plates in the intake manifold, depending on the engine operating point. A tangential and a spiral channel are provided for air intake. The tangential channel is used to generate turbulence while the spiral channels service as charge ducts. The low-temperature exhaust-gas recirculation is realized via EGR cooling with a cooling power of up to 8 kW. An electrically operated add-on pump feeds the required cooling fluid into the primary and the EGR cooler. The exhaust turbocharger is fitted with a pneumatically adjustable guide vane adjustment at the turbine side. To reduce the deposit of liquid fuel at the walls, the recess form of the piston in respect to an increased free jet length has been developed further. Local rich zones are reduced in this manner and the generation of a homogeneous mixture is facilitated. A catalytic converter and a particulate filter are used for exhaust gas treatment. The catalytic converter has been designed as a metal substrate to ensure early starting at high conversion rates. The oxidization catalytic converter function in the particulate filter is optimized in respect to thermal resistance; this filter is zonally coated with platinum/palladium. These, but also other measures, resulted in an excellent specific fuel consumption as shown in Figure 8.17. The fuel consumption best point is rated at 196 g/kWh. The output and torque characteristics are presented in Figure 8.18.

330 | Internal Combustion Engine Handbook

6606_Book.indb 330

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.2 Current Engines

Figure 8.17 Consumption map of the 2-liter common-rail crank gear (source: MTZ).

Type

–

Cylinder angle

Degrees

V12-engine 60

Displacement

dm3

5.934

Stroke

mm

91.4

Bore

mm

83

Compression

–

16

Cylinder spacing

mm

90

Main bearing diameter

mm

65

Conrod bearing diameter

mm

60

Conrod length

mm

155

Valve diameter, intake

mm

28.7

Valve diameter, outlet

mm

26.8

Number of valves per cylinder

–

4

Firing sequence

–

1-7-5-11-3-9-6-12-28-4-10

Maximum output at speed

kW at min–1

368/3750

Maximum torque at speed

Nm at min–1

1000/1750–3250

Engine mass

kg

329

Engine length

mm

680

Maximum charge pressure

bar

2.7

Cylinder peak pressure

bar

165

Exhaust gas standard

–

Euro 4

CO2 emission (MVEG test)

g/km

298

Consumption (MVEG test)

l/100 km

11.3

Figure 8.19 Engine data.

Figure 8.18 Output and torque progressions of the 2-liter common-rail crank gear (source: MTZ).

8.2.6 6-Liter V12 TDI Engine by Audi

The most powerful diesel engine in a passenger car according to Audi has been series-produced since 2008 and delivers from approximately 6-liter displacement a torque of 1,000 Nm and 358 kW power at a fuel consumption of 11.3 liter per 100 km in MVEG testing. Some essential engine data are shown in Figure 8.19. The cylinder angle is 60°. Thanks to multistep drive-side arranged chain drive, a short engine length of just 689 mm was achieved. The crankcase is partitioned in the center of the crankshaft. The bearing covers from ductile cast iron (GJS-600) are connected to form a longitudinal frame. Vermicular graphite cast iron (GJV-450) has been selected for the crankcase. All media-carrying components are integrated in the engine block. High bending and torsion change resistant, low stresses on the main bearings from inertial forces, low friction and low excitement of the control and auxiliary units were design criteria for the crankshaft. These requirements were met by, for example, a main bearing journal diameter of 65 mm, a crank pin diameter of 60 mm and a stroke of 91.4 mm. Figure 8.20 shows the chain drive arrangement with four interlaced simplex chains. Both high-pressure injection pumps are driven by two separate drives directly from the crankshaft. An idle gear set drives the two camshaft sprockets.

The cylinder head is designed in three parts. It comprises a cylinder head lower part, a center part and a longitudinal frame with preassembled camshafts. As in other Audi V engines, the valves are actuated by roller cam followers. The intake and exhaust camshafts are designed as assembled hollow shafts. With regard to acoustics, the Al cylinder head shroud is decoupled from the longitudinal frame in the cylinder head via the sealing and bolting concept. Low NOx emissions are possible thanks to a cooled EGR, where the EGR cooler is connected to a separate low-temperature coolant circuit. The exhaust gas is centrally fed by electrically operated EGR valves through integrated channels in the crankcase and in the cylinder heads to the EGR cooler arranged in the V. The EGR cooler can be switched into the no cooling, Average cooling and maximum cooling positions using vacuum-operated flaps. The air gap-isolated exhaust manifold for every bank is made from stainless steel plates. Two intake groups with variable turbine geometries are used for the charging process. A flow straightener at the compressed inlet side and a flow silencer at the compressor discharge port are used for acoustic optimization. Temperature sensors upstream of the turbine prevent a thermal overload, with the permissible temperature limit being 830°C. The combustion process for the engine is similar to the process in a V6 or V8. Using a valve, one of two intake channels of each cylinder can be steplessly closed to generate a high

Internal Combustion Engine Handbook | 331

6606_Book.indb 331

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Figure 8.20 Chain drive (source: MTZ).

turbulence. The compression was lowered to 16, compared to previous models. The cast aluminum piston is fitted with ring carrier cooling duct. Dummy head honing and an optimized set of piston rings contribute to the low oil consumption. The piston recess and the ring section are shown in Figure 8.21.

is a 32-bit processor with 150 MHz clock rate, 136 kB internal RAM, 2 MB internal and 2 MB external flash memory. The exhaust gas system is realized as a double-flow type and is designed for low exhaust gas back pressure and low heat loss. Oxidization catalytic converters near the engine are fitted with platinum coating for fast heating. Downstream installed particulate filters made from SIC substrate have also catalytic effects (platinum/palladium). The particulate filter (DPF) load and their regeneration is monitored and initiated by applying models. There is one model taking the measured pressure upstream of the DPF and also a simulation model taking into account soot ingress and soot burning by oxidization. The regeneration strategy based on the models occurs with up to five map-dependent injections. Figure 8.22 shows a specific consumption map. The optimal value is given as 204 g/kWh. Consumption, output and torque at full load are shown in Figure 8.23.

Figure 8.21 Chain drive (source: MTZ).

The injection pressure is 2000 bar and is achieved via an in-tank pump with a presupply pressure of 1.3 bar and one high-pressure pump for each bank. The piezo inline injection nozzles with eight conical flow-optimized injection holes are connected with the rail via a high-pressure line with a 3 mm inner diameter. Two identical motor control units are used in a master/slave configuration to control the engine. The primary characteristic Figure 8.22 Fuel consumption map (source: MTZ).

332 | Internal Combustion Engine Handbook

6606_Book.indb 332

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.2 Current Engines

Figure 8.23 Output, torque, and consumption characteristics at full load (source attributed to MTZ).

8.2.7 4.8-Liter V8 SI Engine by Porsche (Turbo and Naturally Aspirated Engine)

Two SI engines with direct injection are described, which have mostly identical basic equipment. One engine is offered as a turbocharged version, the other as a naturally aspirated engine. The basic version of both variants is based on a 4.5-liter V8 engine with a crankcase from a hypereutectic Al/Si alloy realized as a closed-deck design. The technology for the gasoline direct injection has been integrated into the cylinder heads (alloy from Al–Si–Mg–Cu). With the advantage oflightweight development, the overall weight was reduced by approximately 30%. Thermal advantages, compared to the predecessor model, were achieved by revising the cooling concept in the area between the webs and the spark plug.

The crankshaft is manufactured from 38MnS6BY and the forged crack conrod from C70S6BY. According to legal requirements, the crankshaft webs have been converted to the lead-free G344 material. Due to the modified combustion process, the pistons had to be redesigned and were given a piston recess to support mixture layering during cold start and subsequent catalytic converter heat-up phase. In this phase, mixture layering is ensured by injection, shortly before TDC into the compression phase. The increased crankshaft rotational oscillations due to rising gas forces have been reduced by implementing a newly designed oscillation damper in the form of a visco damper. To control the charge cycle, a further developed VarioCamplus system has been installed to constantly adjust the above 50° crankshaft angle. Different valve lifting curves, from 10 mm to 3.6 mm in partial load, are achieved at the intake side using adjustable push rods. The proven dry sump lubrication from the predecessor model has been retained, in which a variable oil pump with volumetric flow and pressure-controlled stepping is used. This made it possible to lower the oil pressure to the minimum required value for each operating point. The air gap isolated manifold of the predecessor model has been replaced by a 4-in-2-in-1 manifold. Other development goals that have been met include a lowered pressure loss, uniform flow distribution and weight reduction. The direct supply of fuel into the combustion chamber has been realized with electromagnetically actuated swirl injectors in a volume of range (idle-full load) of above 22. The homogeneous mixture formation is achieved with an intakesynchronous injection or dual injection. The compression ratio was increased by one unit, compared to the predecessor model. Various measures, such as a demand-controlled oil pump, design optimization of the piston ring set, DLC-coated (diamond-like carbon) top piston ring and push rod, resulted in a reduction of the friction mean pressure. Figure 8.24 shows the progression of the friction mean pressure over the speed. The essential engine data are shown in Figure 8.25.

Figure 8.24 Friction mean pressure in both engine variants (source: MTZ).

Internal Combustion Engine Handbook | 333

6606_Book.indb 333

1/19/16 8:37 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Aspirated engine

Charged engine

Number of cylinders

8 (V design)

8 (V design)

Working volume [dm3]

4.806

4.806

Stroke [mm]

83

83

Bore [mm]

96

96

Number of valves per cylinder

4

4

Nominal output [kW] at speed [min–1]

283/6200

368/6000

Maximum torque [Nm] at speed [min–1]

500/3500

700/2250–4500

Compression ratio

12.5

10.5

Fuel [RON]

98

98

CO2 emission to NEFZ (New European Driving Cycle) [g/km]

329–358

358

Exhaust gas standard

Euro 4

Euro 4

Maximum mean pressure [bar]

13

18.2

Maximum injection pressure [bar]

120

120

the intake or compression stroke is executed, providing an option to quickly heat the catalytic converter.

Figure 8.25 Engine data of the 4.8-liter V8 gasoline engine.

A further improvement of engine properties was the result of charge cycle optimization. For this engine, a variable intake manifold with adjustable resonance tube length was developed. Intercoolers with increased block depth increased thermal efficiency and reduced pressure loss. The combustion process is mostly identical for both turbocharged and naturally aspirated engine. The injector is installed laterally below the intake channel Figure 8.26. A recess in the piston surface is intended to support mixture formation during start and during the heating-up phase of the catalytic converter. For supporting mixture formation, this intake channel is designed to enable tumble generation. The “homogeneous” and “stratified” operating strategies are both controlled with the injection time. In homogeneous operation, a single or dual injection is executed in the intake stroke; in stratified operation, the injection is made just before the ignition TDC. After the engine start, a dual injection into

Figure 8.26 Section through the combustion chamber (source: MTZ).

The output and torque characteristics of both engines are shown in Figure 8.27.

8.2.8 Stratified Combustion Process for Fourand Six-Cylinder Intake Gasoline Engines by BMW

Unlike the homogeneous mixture where no gradient of the air/fuel mixture is present locally, the process with stratified charging generates an inhomogeneity relative to the cylinder charge. The naturally aspirated engines, manufactured in series since 2007, are fitted with a jet-controlled combustion process which contributes to consumption reduction in the European driving cycle but also in real driving operations. Second-generation direct injection is used, which is also called high-precision injection.

Figure 8.27 Torque and output (source: MTZ).

334 | Internal Combustion Engine Handbook

6606_Book.indb 334

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.2 Current Engines

The basis of the crank gear is the magnesium crankcase with aluminum sleeves for the six-cylinder engine and an aluminum housing for the four-cylinder engine. An external EGR is used to reduce the NOx emissions in stratified operation. A partial list of the essential engine data for the six- and four-cylinder variants are shown in Figure 8.28.

Type

Inline six-cylinder engine

–

Four-cylinder engine

Stroke

mm

88

90

Bore

mm

85

84

Piston displacement

dm3

2.996

1.995

Compression

–

12

12

Cylinder spacing

mm

91

91

Number of valves per cylinder

–

4

4 125

Effective power

kW

200

At speed

min–1

6700

6700

Effective torque

Nm

320

210

At speed

min

2750–3000

4250

Emission reduction

–

Euro 4

Euro 4

Valve train

–

Roller cam follower, dual VANOS

–1

head is wetted only minimally. These are the prerequisites for an operation with minimum HC emissions. Because this combustion process does not require strong tumble flow, the intake channels are designed as charge ducts. Figure 8.29 shows a section through the combustion chamber.

Figure 8.28 BMW SI engine with stratified charge (extract).

As the injection system, a conventional low-pressure unit with a five bar system pressure and a high-pressure component with a 200 bar injection pressure. The injection valve is designed so that a hollow cone-shaped injection jet is generated. The thermal compensator ensures that a constant needle lift is present at all operating temperatures of the injector. The fast opening and closing of the needle enable immediate successive injection pulses. The injector is arranged central to the combustion chamber, with a slight incline to the intake side. Immediately next to the injector, the spark plug is placed with a slight incline to the other side. With its electrodes, it reaches the recirculation area of the air/gas mixture with an ignitable composition. The piston recess is designed so that a wetting of the piston crown is prevented during stratified charging. The piston

Figure 8.29 Section through the combustion chamber (source: MTZ).

The stratified operation allows quality regulation in partial load operation. This makes it possible to dispense with throttle control in large areas of partial load, effectively reducing charge cycle losses. Particular attention must be paid to the ignition system in a jet-controlled combustion process. Because the spark plug is located at the jet edge near the injection spray, it is subject to high-temperature changes and deposits. For this reason, a surface-gap spark plug with a high self-cleaning capacity is used. The exhaust gas treatment must be effective not only for air-fuel ration at lambda equal 1, but also for lean mixtures. This is achieved with a three-way catalytic converter near the engine and two NOx storage catalytic converters. The operating strategy intends that, after heating of the catalytic converter, a warm-up phase is completed on homogeneous operation. This is followed by the consumption-minimized stratified operation comprising large areas of the European driving cycle. Stratified operation is temporarily interrupted by a regeneration of the NOx storage. Figure 8.30 shows the operating data and consumption map of the 3-liter six-cylinder engine.

Figure 8.30 Operating data and consumption map of the 3-liter six-cylinder engine (source: MTZ).

Internal Combustion Engine Handbook | 335

6606_Book.indb 335

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

In this image, the consumption advantages in the lower partial load, thanks to the stratified process, can be clearly seen. Thus, at the operating point 2000 min–1 and a specific work of 0.2 kJ/dm3, we have a specific consumption of just 295 g/kWh. The output and torque characteristics of the six- and fourcylinder engine are presented in Figure 8.31.

on friction. Further beneficial characteristics are oil circuit optimization with focus on supply volume regulation as required and adjustable piston cooling.

Figure 8.31 Output and torque progression of the six- and four-cylinder engine (source: MTZ).

8.2.9 Four-Cylinder Diesel Engine by Mercedes-Benz

The following are central development objectives of this engine that has been series-manufactured since 2007 have been: •• Fuel consumption reduction at improved driving performance •• Increased torque and performance, compared to the previous model •• Emission tier Euro 5 and potential in regard to Euro 6 •• Engine concept suitable for longitudinal and transversal mounting •• Optimization and standardization of assemblies. Starting with a basic concept, the engine should be expandable with add-on modules, depending on specific requirements. One characteristic of the basic engine is the arrangement of the camshaft drive at the drive end which became necessary for reasons of packaging, Figure 8.32. The camshaft drive is a combination of gear and chain drive resulting in a short motor. A deep connection of the cylinder head bolts and the concomitant dimensional stability of the cylinder sleeve reduces the tangential force of the complete ring set with positive effects

Figure 8.32 Camshaft drive in a four-cylinder diesel engine.

The main technical data of this engine are shown in Figure 8.33. The engine is fitted with two-stage charging, one highpressure and one low-pressure exhaust turbocharger. An adjustable valve is installed in the compressor bypass that opens a parallel air path in the high-performance section. This reduces pressure losses and prevents an overload of the high-pressure charger. An air/air intercooling system with high-cooling power is a necessary prerequisite for the demanded specific engine output. The motor is fitted with a cooled EGR with the recirculated exhaust gas being cooled in an EGR precooler and an EGR main cooler. The exhaust gas may be fed cooled or uncooled through a bypass into the airflow within the engine. The essential characteristics of the injection system are as follows: •• common-rail system with 2,000 bar injection pressure •• up to five injection actions per total injection. This engine meets the Euro 5 exhaust gas standard in effect since September 2009. The defined limit values are met without active NOx exhaust gas treatment.

336 | Internal Combustion Engine Handbook

6606_Book.indb 336

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.2 Current Engines

Type

–

Inline four-cylinder engine

Number of valves per cylinder

–

4

Displacement

cm3

2143

Cylinder spacing

mm

94

Stroke

mm

99

Bore

mm

83

Conrod length

mm

143.55

Maximum peak pressure

bar

200

Charging level

bar

3

Web width

mm

11

Valve angle

Degrees

6

Nominal output at speed

kW at min–1

150 at 4200

Maximum torque at speed

Nm at min–1

500 at −1600–1800

Compression ratio

–

16.2

Maximum effective mean pressure

bar

29.33

Emissions standard

–

EU5

Specific power output

kW/liter

70

Figure 8.33 Engine data.

One objective of the applied thermal management is to prevent coolant motion and thus to achieve heating of the combustion chamber after a cold start as quickly as possible. This process lowers the raw CO and HC emissions. Controlling the coolant temperature within the thermal management to values of up to 70°C also reduces NOx emission. Installed in the C250 CDI, the fuel consumption of the engine is 5.2 liter per 100 km in NEFZ. Figure 8.34 shows the progression of output and torque over the speed.

Figure 8.34 Output and torque characteristics of the four-cylinder diesel engine.

8.2.10 SI Engines with Direct Injection and Dual Charging by VW

The Turbo Spark Ignition (TSI) outputs 125 kW at a working volume of 1.4 liter and represents a combination of direct injection, downsizing and dual charging. The most important engine data are shown in Figure 8.35.

Number of cylinders [–]

4

Cylinder spacing [mm]

82

Displacement [dm3]

1.39

Stroke [mm]

75.6

Bore [mm]

76.5

Compression

10

Output at speed [kW/min–1]

125/6600

Torque [Nm/min–1]

240/1750

Maximum charge pressure (bar absolute)

2.5

Maximum effective mean pressure [bar]

21.6

Fuel consumption [l/100 km]

7.2

Figure 8.35 Engine data.

The high torque at low speeds and the torque progression over the speed are essentially determined by the use of two charging systems: A mechanical supercharging and an exhaust gas turbocharging system. The exhaust-driven turbocharger designed for optimal possible efficiency cannot by itself provide the required charge pressure at low exhaust gas throughputs. In this event, the compressor for mechanical supercharging can be activated by a magnet coupling. It is always active from a specific torque demand and it will be deactivated at approximately 3500 min–1. The compressor designed as a root blower is fitted with an internal transmission. The total transmission relative to the crankshaft is i = 0.2. The compressor operating range is shown in Figure 8.36. The compressor is driven by a five-grooved V-ribbed belt from the water pump. The compressor system is silenced by the reduction of air pulsation and special broadband dampers at intake and outlet, in addition to the acoustic optimization of the compressor’s mechanical components. Compressor and dampers are also encapsulated. The gas paths of air and exhaust are schematically shown in Figure 8.37. Because enrichment for low fuel consumption was omitted to reduce the thermal burden, the turbine side of the charger is subject to temperatures of up to 1050°C, which required a corresponding adaptation of the materials. The turbine housing is made from a high-heat-resistant cast steel and the turbine wheel from a high-temperature-resistant MAR 246 nickel alloy. X45CrSi9.3 is used for the shaft. The ATL water–cooling system is integrated in the enginecooling system. A coolant-trailing pump ensures that the ATL is cooled after the engine has been shut down. The GG cylinder crankcase is realized as deep-skirt housing with 3 mm wall thickness and it weighs 29 kg. The open-deck design is particularly suited for the dual-circuit cooling used (head and block separately) and offers advantages relative to the cylinder pipe deformation. The steel crankshaft contributes to the overall acoustics with the higher E module, compared to a crankshaft from gray-cast iron and the resulting higher rigidity. The piston weight of the cast piston is 238 g. The piston recess with edge for flow control is machined. For the adjustment to ignition pressures toward 120 bar, the wristpin diameter had to be increased from 17 to 19 mm.

Internal Combustion Engine Handbook | 337

6606_Book.indb 337

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Figure 8.36 Compressor operating range (source: MTZ).

Figure 8.37 Gas paths of air and exhaust gas(source: MTZ).

Oil and fuel high-pressure pumps were adapted for the bigger mass throughputs. With the support of the intercooling system, the intake air can be cooled down to 5° above ambient temperature. The engine combustion chamber with gasoline direct injection is designed as a roofed combustion chamber with central spark plug position. In combination with a flat piston recess, a compression of 10 can be achieved. The charging motion valve half-closing the intake ducts in the lower speed range increases the charge motion. From a speed of 2800 min–1, the valves release the entire intake duct profile. To reduce cold start emissions, the catalytic converter is heated by a dial injection (during the intake phase and before the ignition TDC). The inject valve, designed as multihole high-pressure injection valve, is placed at the intake side and features six fuel discharge bores. The injection pressure is 60 bar at idle and up to 150 bar under full load. The measures of the downsizing concept described above resulted in low fuel consumption in wide areas of the map shown in Figure 8.38.

Figure 8.38 Consumption map (source: MTZ).

338 | Internal Combustion Engine Handbook

6606_Book.indb 338

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

The engine output and torque characteristics are presented in Figure 8.39.

Figure 8.40 Beta engine trial vehicle.

Figure 8.39 Output and torque progression (source: MTZ).

Beginning in 2008, the 1.4 liter TSI engine with two exhaustdriven turbochargers and 90 kW will be available. This is classified under the unit described above. A 1.2-liter TSI engine is planned for the end of 2009 which will represent a further step in downsizing direction.

8.3 Motorcycle Engines/ Special Engines Modern motorcycle engines are usually four-stroke engines. The change from two-stroke to four-stroke engines began in the early 1980s. Exceptions are off-road motorcycles such as Motocross and Enduro but also special applications.

8.3.1 Motorcycles for the Road (On-Road)

Unlike racing engines in the on- and off-road sector, engines for road-approved motorcycles must be durable, low vibration, low maintenance, producible at low costs and recyclable. In addition, they must meet increasingly strict legal requirements. For this reason, they cannot be designed according to performance only (Figure 8.40). In the history of motorcycle engineering, many different designs of reciprocating piston engines have been seen, from single-cylinder to eight-cylinder engines. The majority of the products in the market were produced with single-, two- and four-cylinder engines. All of them have their own specific advantages and appeals regarding their performance. Due to the very different fee inertial forces and torques, but also ignition spacing, they also have a very individual vibration and noise behaviors. The importance of the introduction of these forces from the engine into the frame is significantly larger in a motorcycle compared to a car, because, unlike cars, the engines are usually rigidly bolted to the frame as a joint load-bearing element. In addition, some engines must also hold structure components such as rocker bearings.

Comfort Endurance Recycling Weight

Output

Torque Costs

Roadapproved engine

Exhaust gas legislation

Lowmaintenance

Noise legislation

Figure 8.41 Influencing factors on motorcycle engines.

8.3.1.1 Single-Cylinder Engines The single-cylinder engine is the simplest type and is still installed in many light vehicles (Figure 8.42). Typical examples are Enduro and Supermoto vehicles. Its working volume is limited to approximately 800 ccm because, with increasing volume, the moved masses of crank gear and valve drive components become disproportionately heavier. This fact limits the attainable speeds and thus the specific output. Furthermore, the combustion behavior becomes more complex in large cylinder volumes. The flat combustion chambers with unfavorable volume/surface ratio make always higher demands on the ignition phase at low load in the lower speed range. In λ = 1 mixtures and concurrently large residual gas quantity, which is inevitably prevalent due to the valve overlap cross-section required for the high liter performance, clean mixture treatment and accurate positioning of the spark plug must ensure that the mixture ignites reliably and burns at sufficient speed. The exhaust gas values required for Euro 3 can otherwise not be met, because the catalytic converter and drivability would suffer badly. Effects would be a jerking motion during constant driving and in load change situations where a hard restoration of the combustion after a thrust phase could prevent a “clean” curve line in tight radii in particular (mountain hairpin bends).

Internal Combustion Engine Handbook | 339

6606_Book.indb 339

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

1 2 3 4 5 6 7 8 9 10

Oil screen Delivery pump Coarse filter Overpressure valve Rocker arm shaft Transmission Coupling Oil injection nozzles Fine filter Crankshaft

Figure 8.44 Injection principle.

Figure 8.42 Single-cylinder engine with oil circuit. See color section page 1079.

Large single cylinders are not very comfortable due to their very nature. To ensure that engines with more than 600 cm3 for speeds of 8000 min–1 and more are still somewhat comfortable, balance shafts are frequently installed to reduce the first-order inertial forces. However, balance shafts bring weight-related disadvantages and undesirable oscillating masses in particular, degrading the spontaneity of the responses (Figure 8.43–Figure 8.45).

Figure 8.45 KTM 690 Super Moto.

8.3.1.2 Two-Cylinder Engines Swept volumes of 800 ccm and more are usually distributed over multiple cylinders. Two-cylinder engines are very popular. The most commonly selected arrangements of both cylinders are inline, V, and boxer engines. 8.3.1.2.1 Two-cylinder inline engine In a two-cylinder inline engine, two cylinders are arranged next to each other and can be installed longitudinally or transversally in the chassis. The crankshaft may have a 180° or 360° ignition spacing, that is, the pistons move opposingly or concurrently. Both versions have their own inertial force and sound behavior. In the 180° variant, the first-order inertial forces are balanced. However, this results in second-order inertial forces and first-order inertial torque. The 360° variant has first- and second-order inertial forces but no free inertial torques (Figure 8.46). The advantages of the inline engine are its simple structure and the low number of components making it very cost-efficient. The common cylinder head and block and the jointly utilized

Figure 8.43 KTM 690.

340 | Internal Combustion Engine Handbook

6606_Book.indb 340

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

Two-cylinder inline engine

0° 360°

360°-Crankshaft 0°

270°

Zyl 1

Zyl 2

0° 360°

90° 270°

Zyl 1

Zyl 2

90°

90°

270°

180°

Crank diagram

180°

180°

Crank diagram

0° 360°

180°-Crankshaft

0° 360°

0°

90°

270°

270°

180°

Zyl 1 Zyl 2

90° 270°

180°

Crank diagram

Zyl 1

Zyl 2

90°

180°

Figure 8.46 Inertial forces in a two-cylinder inline engine (360°/180° crankshaft).

Crank diagram

valve actuation limit weight and friction points. The same cooling conditions for exhaust and intake manifolds effect a very similar combustion behavior of the cylinders. Modern representatives of this design are currently used in BMW F800 and MZ 1000SF (Figure 8.47 and Figure 8.48).

Figure 8.48 BMW F800 cylinder head.

Figure 8.47 BMW F800 two-cylinder inline engine.

The F800 engine developed by Rotax uses a 360° crankshaft as parallel twin. To nevertheless meet modern comfort requirements, a quite unusual mass balancing concept has been realized in the engine. Between both cylinders, a balance conrod is installed which suspends from a balancing rocker attached at a right angle to the cylinder axis and supported in the engine housing. The kinematics is designed so that the balancing conrod moves in opposition to the two engine piston rods. By guiding over the long lever, approximately straight up and down movements of the small eye of the balancing conrod are obtained, which equalizes the free first-order inertial forces fully and those of the second-order at 70% (Figure 8.49 and Figure 8.50).

Internal Combustion Engine Handbook | 341

6606_Book.indb 341

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Figure 8.49 BMW F800 mass balancing.

Figure 8.50 BMW F800 crankshaft drive.

8.3.1.2.2 V2 engine In a two-cylinder V engine, the two individual cylinders are placed on a common crankcase with offset against each other. It can be designed with various V-angles between 45° and 180°. A typical feature is that the connecting rods are guided on a common crank pin so that the engine width is narrower than the one on the inline engine.

The selection of the cylinder angle depends, among other things, on the stroke/bore ratio, the available space and the application of the specific engine. Decreasing V-angles below 90° reduce the engine length, but the height slightly increases. For this reason, small cylinder angles are preferred when a motorcycle with a short wheelbase is planned. The angle reduction has limits in view of available space. To exclude a contact of the pistons in the gusset area, longer cylinders and connecting rods must be used at small cylinder angles. This drives the moved masses upward which affects the achievable speeds and engine loads. It also increases the height and length of the engine design which interferes with the package of tank and air box. These issues force the designer to compromise. All V engines have first- and second-order inertial forces. The 90° V engines are so popular because these inertial forces can be fully offset by counterweights. Only inertial torques are left which occur due to the connecting rods being placed on a crank pin and thus have a cylinder axis offset. The inertial forces attempt thus to turn the engine around the center between the connecting rods. Because the distance between the connecting rod centers is small, the free inertial torques present in all V engines are also small. The V engine requires separate camshafts for each cylinder head, increasing production costs and weight. Due to additional bearing points, friction losses in the valve train also increase. The cooling situation from the airstream is different for both cylinders. In today’s usually water-cooled engines, this rarely affects cylinder cooling but certainly affects the exhaust cooling. The different hot exhaust gases have also differently large speeds of sound which change the wave movements in the exhaust system, resulting in uneven charge cycle conditions. The uneven ignition spacing also has some adverse effects on the charge cycle but generates a very pleasant “sonic impression.” For this reason and also its comfortable output characteristic, the V engine remains very popular despite its systemic disadvantages. Some typical cylinder angles are closely linked to specific motorcycle manufacturers. Ducati, for example, has always used a 90° angle and calls this engine the L engine. In the early days, the engines were air-cooled and the 90° as a laying L was favorable for the cooling situation. The large form was not that disadvantageous, because, in those days, Ducati favored long wheelbases and large overruns to obtain good driving stability. When transitioning to the water-cooling and four-valve technology of modern Superbike models, the cylinder angle was retained. Typical for the Ducati Twin are also the desmodromic valves allowing very high valve accelerations thanks to the closing cams, thus enabling very high cylinder charges. Due to material improvements in the valve train components however, the advantages, compared to conventional valve trains have become negligible. Even so, the water-cooled Ducati L-Twin with Testastretta cylinder head and liter performance of 140 PS (999) or 150 PS (999R) represents a benchmark in the twocylinder class (Figure 8.51and Figure 8.52).

342 | Internal Combustion Engine Handbook

6606_Book.indb 342

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

Figure 8.51 Ducati 90° V engine.

Figure 8.53 KTM LC8 V75.°

Figure 8.52 Moto Morini 87° V engine.

When designing the LC8 engine, KTM decided on a 75° V engine for its maximum compact size with a short-form length. To improve the mass balancing which is systemically poorer than the 90° engine, the engine is fitted with a balance shaft. Thanks to its very compact outer dimensions, this engine fits not only in the large off-road models (Adventure, Super Enduro) but also in the purely on-road types (Super Duke, Super Moto). KTM will maintain the 75° angle as the successor of the LC8 with larger displacement, which is intended for installation in the RC8 Superbike and will have a significantly higher performance (Figure 8.53). The KTM LC8 is characterized by its very low oscillating mass and low weight, resulting in extreme agility in handling and response behavior. With this engine, KTM is the benchmark for two-cylinder engines (Figure 8.54 and Figure 8.55).

Figure 8.54 KTM 990 Super Duke.

Figure 8.55 KTM RC8.

Internal Combustion Engine Handbook | 343

6606_Book.indb 343

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Aprilia uses a 60° V engine by Rotax for its large four-stroke types, which is a very responsive unit fitted with two balance shafts, one of which is installed before the crankshaft and the other in the cylinder head (Figure 8.56 and Figure 8.57).

Figure 8.58 45° V engine (Harley-Davidson) primary drive (chain).

Figure 8.56 60° V engine (Aprilia Mille).

Figure 8.59 45° V engine (Harley-Davidson) pushrods.

Figure 8.57 Aprilia Mille Factory.

One of the best-known representatives of V engines is Harley–Davidson. They use the 45° angle. Because Harley–Davidson traditionally offers slow-running long-stroke machines with two valves and lateral intake system, the necessarily long connecting rods and the large size are not deleterious. Due to the low required banking freedom, the engines can be installed deeply in the frame. The strong vibrations are filtered out by soft bearings. The primary drive via chain is quite interesting. Two camshafts below actuate the valves with long pushrods allowing only low maximum speeds (Figure 8.58–Figure 8.60).

Figure 8.60 Harley–Davidson softtail deluxe.

Some manufacturers use crank pin offset for engines with cylinder angles below 90° to approximate the mass balancing and sound behavior to the one of a 90° engine. This works quite well in some engines; however, a large crank pin offset bears mechanical risks. For example, high speeds can cause crankshaft breaks when the insert length of the crank pins is insufficiently dimensioned (Figure 8.61).

344 | Internal Combustion Engine Handbook

6606_Book.indb 344

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

Figure 8.62 BMW-R1200 S.

Figure 8.61 V crankshaft with crank pin offset.

8.3.1.2.3 Boxer engine The boxer engine is a model with the cylinders laying opposite each other in one plane. The outside placed pistons act on a crankshaft laying in the center. The main difference to the V engine is that the connecting rods are not acting on a common crank pin but are each on their own. The crank pins are placed with an offset of 180° for the pistons to be symmetrically moved towards each other and away from each other. This form was developed very early because there is no affect by free inertial forces. They are fully offset. the distance of the connecting rods is significantly larger than in the V engine due to the individual crank pins (frequently with a center bearing between the pins). This provides to the same extent higher free inertial torques which cannot be offset in a two-cylinder engine. This is only possible in a six-cylinder boxer engine (Porsche). In practice, a two-cylinder boxer engine does not run more elegantly than a good V engine. The biggest proponent of the boxer concept is BMW AG that has continuously upgraded the air-cooled boxer engine for decades. In regard to speed, and thus performance, the BMW boxer engines have always been limited by their pushrod valve train. This is still the case with the current version with 1.200 cm3 that is used in many models. By reducing the moved masses and concentration on maximum air throughput, a sportive variant can now be offered which achieves higher speeds for the first time. A liter output of 100 PS/l can thus be obtained, representing the limit for air-cooled series engines. The vibration behavior of this sport engine is, however, significantly less comfortable than the more moderate models (Figure 8.62–Figure 8.66).

Figure 8.63 BMW R1200 boxer engine.

Figure 8.64 BMW-R1200 S.

Internal Combustion Engine Handbook | 345

6606_Book.indb 345

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

that are generated by the attempts of the two outer cylinders to turn the engine around the center axis. Because the moved masses are less than in a two-cylinder, the three-cylinder unit can attain somewhat higher speeds. The single stroke volumes, which are larger compared to four-cylinder engines, improve the torque behavior in the lower speed range which, in conjunction with the pleasant sonorous sound, results in a very appealing engine concept.

Figure 8.65 BMW R1200 boxer engine.

8.3.1.3 Multicylinder Engines Due to the smaller single stroke volumes and the related smaller moved component masses, higher speeds can be achieved with an increasing number of cylinders (Figure 8.67). 8.3.1.3.1 Three-cylinder engines For three-cylinder engines, there are fundamentally different crankshaft designs. In the 1970s, Laverda, for example, built a three-cylinder with 180° throw of the crankshaft. The current three-cylinder engines by Triumpf and Benelli (Figure 8.68), work with a 120° ignition spacing. They do not have free inertial forces because they are always balanced with the opposing movements of the pistons. There are only free inertial torques

Type

Series

Boxer

V engine (90°)

Number of cylinders

Crankshaft throw

Ignition spacing

8.3.1.3.2 Four-cylinder engines In the 1960s, Japanese manufacturers in particular, recognized that any further performance increases for racing could only be attained via the speed and thus, a reduction in moved masses. The reduction in the moved masses mandated an increase in the number of cylinders. In addition to some sixand eight-cylinder concepts, the four-cylinder engine proved to be an excellent compromise. In 1968, Honda serialized a four-cylinder inline engine for the first time: the CB750. This engine changed the world of motorcycles forever. The Japanese manufacturers have perfected this concept over the last forty years. Today, they are on a very high level in respect to performance, production and sustainability. Today, the 600 ccm class achieves a liter output of more than 200 PS/l at speeds up to 16,000 min–1. At the same time, running times of 50,000 to 100,000 km are reached. This makes the four-cylinder inline engine the current most successful concept. As in V2 engines, the four-cylinder engine allows various concepts (inline, V, boxer)—all of them are available in the marketplace. The four-cylinder inline engine is the most successful one, being currently used in sport and touring concepts. The advantages of the four-cylinder inline engine are as follows: •• relatively simple design → standardized components •• compact, light-weight engine housing → low weight

First-order inertial force

Second-order inertial force

First-order inertial torque

Second-order inertial torque

1

–

720°

×

×

–

–

2

180°

180°/540°

–

×

×

–

2

360°

360°

×

×

–

–

3

120°

240°/240°

–

–

×

×

3

180°

360°/180°

–

×

–

–

4

180°

180°/180°

–

×

–

–

4

90°

90°/270°

–

–

×

×

6

60°

120°/120°

–

–

–

–

2

–

360°

–

–

–

×

4

180°

180°/180°

–

–

–

×

6

120°

120°/120°

–

–

–

–

2

–

270°/450°

× (compensable with counterweight)

–

Very small

Very small

4

90°

270°/90°

× (compensable with counterweight)

–

Very small

Very small

6

120°

150°/90°

–

–

×

Figure 8.66 Inertial forces and torques of various types.

346 | Internal Combustion Engine Handbook

6606_Book.indb 346

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

16000 15000

Four-cylinder

14000

Nominal speed [1/min]

Two-cylinder

12000

9000

Triumpf Speed Triple

Yamaha FZR1000

Triumpf Daytona 675

KTM Super Duke

11000 10000

Suzuki GSXR750

Yamaha R1

Three-cylinder

13000

Yamaha R6

Singlecylinder

BMW R 1200S

MV Agusta F4

KTM LC4

Suzuki Hayabusa Kawasaki ZZR1400

8000 7000

BMW F800 Ducati GT1000 Moto Guzzi Norge

6000

40

50

60

70

Four/five-valver

Two-valver

Yamaha MT03 Yamaha MT01

5000

Benelli TNT1130

80

90 100 110 Specific output [kW/L]

120

130

140

150

160

Figure 8.67 Specific output.

•• short and rigid crankshaft → high speeds •• identical conditions for charge cycle range → high performance •• smooth running → great comfort. Except in GP1 engines, a 180° crankshaft with firing sequence 1-3-4-2 or 1-2-4-3 is used currently. This ensures even ignition spacing of 180° and the first-order inertial forces and firstorder and second-order inertial torques can be offset. Only second-order inertial forces are left which can also be offset with a balance shaft (Figure 8.69). Another option is the noneven crankshaft with 90° throw. With this design, the first- and second-order inertial forces are fully offset but first- and second-order inertial torques will occur. This type is no longer used in modern series engines despite its advantages with respect to the inertial forces. The reasons for this are disadvantages in charge cycles and the uneven ignition spacing causing major rotation irregularities and thus major rotational oscillation excitations (Figure 8.70).

Figure 8.68 Benelli 1130 TNT.

Four-cylinder inline engine 180°-Crankshaft

270°

90°

180°

0° 360°

0° 360°

0°

270°

90°

90° 270°

180°

Crank diagram 1. Order

180°

Crank diagram 2. Order

Figure 8.69 Crank diagram 180° four-cylinder shaft.

Internal Combustion Engine Handbook | 347

6606_Book.indb 347

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Four-cylinder inline engine 90°-Crankshaft

270°

0° 360°

0° 360°

0°

90°

270°

180°

180°

Crank diagram 1. Order

Modern four-cylinder inline engines can be built very compactly, by smart arrangement of the accessories and the gear shaft enabling a favorable mass concentration and short wheelbases. Modern materials and manufacturing processes allow very light engine housings. Because of the compact engine, the intake paths can be designed very straight line and can be given a large air box. Small single-stroke volumes with correspondingly small and light components and a very rigid short crankshaft enable extremely high speeds. Yamaha achieves speeds of above 16,000 min–1 in the current R6 super sport which actually is a series model. With a liter performance of 211 PS/l, the R6 achieves values that were some years ago only possible for racing bikes with a short service life. If only the performance aspects are taken into consideration the fourcylinder inline engine remains clearly unbeaten. However, motorcycling is emotional to a large extent, something these perfect driving machines are lacking. This is the reason why we still have motorcycle manufacturers who produce different concepts, such as V2 engines.

Figure 8.71 Yamaha R1 section.

90°

90° 270°

180°

Crank diagram 2. Order

Figure 8.70 Crank diagram 90° four-cylinder shaft.

Off the racetrack, there are still sufficient riders for whom emotion and sound are more important than the engine output. More than four cylinders have very limited use today. The smoothness can be still increased as a matter of fact, by using a six-cylinder inline engine with complete balancing. This makes their application quite useful in comfort vehicles such as the Honda Goldwing. However, the engine output can no longer be raised by speed increases because the crankshaft becomes longer and thus more sensitive to rotational oscillations. This and the increase in width and weight prevent an application in super sport motorcycles (Figure 8.71 and Figure 8.72).

Figure 8.72 Yamaha R1.

348 | Internal Combustion Engine Handbook

6606_Book.indb 348

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

8.3.1.4 Development of Power To increase the power, all individual, controllable parameters must be examined. Power is defined as follows: Pe = i × pme × V h × n. To increase power, we can increase either the displacement, the mean pressure, or the speed. A displacement increase is not always useful in a motorcycle engine due to the disadvantage of increasing moved masses and package dimensions. If the displacement remains the same, we can only increase the mean pressure and the speed (Figure 8.73).

Figure 8.73 Full-load engine test stand.

To increase the mean pressure, the volumetric efficiency must be increased. Volumetric efficiency is the ratio of the fresh gas mass remaining in the cylinder after a charge cycle relative to the theoretical fresh gas mass. λ l = mFZ/ρ 0 × V h. This can be achieved with charging effects. One option is charging systems such as mechanical or turbochargers. The appeal of this solution is a very high power density. This means that a small and light engine with high torque and high power can be used. Mechanical chargers may be designed as displacement chargers, such as root blowers or radial compressors or chargers, with internal compression such as a screw compressor. Large space requirements and a relatively high weight argue against the use of mechanical chargers in motorcycles, as do the very high drive power at high speeds and the large speed range of the motorcycle engine. The intercooling system required for good efficiencies would have room only in rare cases. For this reason, several manufacturers in the 1980s produced engines with turbocharging. The problems arise from an nonstatic behavior that is nearly uncontrollable. Unlike the diesel engine which always runs at full filling, the throughput air masses vary in a ratio of 1:40 to 1:50 in the SI engine. For this reason, the charger requires quite some time to build pressure when transitioning from partial load to full load. In conjunction with low engine oscillating mass, this results

in an unsatisfactory response behavior and disharmonious driving in motorcycle turbo engines. At this point in time, almost exclusively naturally aspirated engines are used in motorcycles, which use antivibration effects of the ram pipe and resonance tube supercharging. The piston triggers during load change phase a vacuum wave in the resonance tube that travels through the intake manifold in the direction of the air box or air filter. This wave is reflected as a pressure wave at the open funnel and travels toward the engine. If the pressure wave arrives at exactly the moment when the inlet valve is closed, fresh gas may still be pushed into the cylinder even if the piston after BDC is already moving upward. This can achieve a supercharging action of up to 15%. These processes are, of course, very dependent on the intake manifold geometry (pipe length to be traversed) and the kinetic energy of the gas flows. They, on the other hand, are significantly affected by the intake manifold diameter and the excitement by the piston which, of course, depends on the speed. Because the timing of the camshafts is usually fixed, the charging effects are very much speed-dependent, enabling the generation of positive effects for the volumetric efficiency in various speed ranges. The options for increasing efficiency and thus mean pressure, however, are very limited in naturally aspirated motorcycle engines. The charge changing organs and resonance tube lengths in a motorcycle are mostly designed for nominal speed anyway. Modern engines have been optimized with intensive charge cycle simulation calculations and work with little throttling at the intake and pressure side. The mean pressure can be increased to a larger extent only by variabilities in the charge cycle system. Systems, such as intake manifold length adjustment, cam phasers, valve lift switching and valve systems, are proven in series production and efficient in the automotive area. Complexity, space requirement, speed continuity and costs argue against these measures in motorcycles. For this reason, any major increase in performance can be almost exclusively reached by raising the nominal speed. A nominal speed raise requires numerous measures because the inertial forces in the engine increase quadratically with the speed. Hence, the main objective in power development must be always the minimization of the moved masses in the engine. In the coming years, more efforts will need to be made to enable an increase in performance despite the increasing demands in respect to noise and exhaust gas emissions of the future (Figure 8.74–Figure 8.76). The engine manufacturers will need to use variabilities in the valve train and intake manifold area and controllable valve systems in the exhaust train, because most measures for exhaust and noise improvements are counterproductive to higher performance.

Internal Combustion Engine Handbook | 349

6606_Book.indb 349

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

. Spec. output [PS/L]

Effects of European legislation Noise measuring

Noise limit New limits for accelerated drive past

New measuring procedures for accelerated drive past (target acceleration)

(80 dB → 77 dB) ca. 2009

2007

2008

2009

ca. 2010

2010

Exhaust gas limit Euro-4 limits (reduction to approx. 50% of the EU3 values) ca. 2011 Application of all measures

2011

Conventional development

2012

Figure 8.74 Power development.

Figure 8.75 Exhaust gas control valve, Yamaha R1.

8.3.1.5 Stroke–Bore Ratio Long-stroke designed engines are more effective in respect to the combustion process (stroke–bore ratio > 1). Favorable combustion chamber shape, good surface/volume ratio with lower heat losses and shorter flame paths result in excellent combustion efficiency, high knock resistance and low pollutant emission. For these reasons, some passenger car engines are designed with long strokes. With increased specific outputs and the necessary high speeds, the ratio shifts toward short-stroke models. Several reasons speak against a long-stroke design in motorcycle engines with high-specific output at high nominal speed: •• little possible valve faces → limited liter performance •• large piston speeds → long-term durability •• large inertial forces → mechanical strength •• large size → high center of gravity. The advantages of a short-stroke engine outweigh the disadvantages in combustion when specific outputs and correspondingly high speeds are increasing. For this reason,

Figure 8.76 Intake manifold control, Yamaha R1.

modern motorcycle engines are designed with short strokes. The exceptions are engines based on traditional engine concepts (Harley–Davidson, Buell or retro types such as the Yamaha MT01). These engines are frequently still designed as twovalve engines and thus limited in speed and liter performance (Figure 8.77).

350 | Internal Combustion Engine Handbook

6606_Book.indb 350

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

1,3

Harley-Davidson Screaming Harley-Davidson Triumpf Sportster Thunderbird BMW F800 MZ 1000S Benelli TNT1130

1,2 1,1

Stroke-bore ratio

1,0

Kawasaki ZZR1400

0,9

Triumpf Daytona 675 Yamaha R1 Suzuki GSXR750

Yamaha FZ1

Yamaha R6

0,8 0,7

Yamaha MT01 Yamaha MT03 Moto Ducati Guzzi 0,5 Norge GT1000 BMW R 1200S 0,6

0,4 0,3

Suzuki Hayabusa KTM Super Duke

40

50

60

70

80

Ducati 999R

90 100 110 Specific output [kW/L]

8.3.1.6 Valve Train The high-specific output of motorcycle engines requires a design of the charge changing organs for very high nominal speed. To optimize the filling and minimize throttling losses, very large intake valves, valve strokes and throttle valves are required. Air ducting and intake channels must have large cross sections and should be designed to be as straight as possible. The speed level increasing with raising specific power requires a reduction of moved masses in the engine. With increasing cylinder volumes, the charge changing control devices are usually the first to reach their geometrical limits. To keep the forces and accelerations at camshafts and valve actuation within sustainable limits and to precisely maintain the control times, the moved valve train masses must be minimized and the stiffness of the valve gear must be maximized (Figure 8.78). Depending on the application, several valve control types are suited for the valve train in motorcycle engines. OHC drives are often used in simple and light vehicles with single cylinders. The advantages of this concept are costs, space, low maintenance and a relatively low-speed level. Only modern DOHC drives make sense for modern high-performance engines with nominal speeds significantly above 10,000 min–1. The engine speed strength and precision outweigh costs and space situation (Figure 8.79). Increasing valve diameters can be partially compensated by a reduction of the valve shaft diameter and the valve length. Steel valves may be replaced by those made from titanium to obtain a specific weight of 4.5 g/cm3 instead of 7.85 g/cm3.

OHC DOHC

120

130

140

150

160

Figure 8.77 Specific output, in relation to the stroke–bore ratio.

This can be easily achieved at the intake side, due to the lower thermal stresses, but is a significantly more expensive solution.

Figure 8.79 Valve train, Suzuki GSX-R 1000.

The service life is also limited, when compared with the steel versions. To reach speeds of more than 10,000 min–1, push rods or cam followers are the preferred valve actuators as they combine rigidity with low weight. Both work with a sliding contact, for which reason that the cam followers are frequently manufactured with ultra-hard DLC coatings. These hard carbon coatings (diamond-like carbon) are increasingly used in high-speed engines due to

Moved mass

Stiffness

Engine speed strength

Costs

Package

Valve play adjustment

HVA capability

Rocker arm

–

–

–

+

+

+

–

Rocker arm

○

○

○

+

+

+

–

Cam follower

○

○

○

○

+

+

+

Pushrod

+

+

+

–

–

–

+

Cam follower

+

+

+

–

○

○

+

Valve actuation OHV

MV Agusta F4

Figure 8.78 Comparison of valve train parameters.

Internal Combustion Engine Handbook | 351

6606_Book.indb 351

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

engage in the gate of the camshaft controller. By turning the controller, the sliding wheels can be shifted in both directions. This is only possible because all gears are straight-toothed.

their low-friction and wear resistance properties. With the correct design, the coating of one friction partner can lower the friction losses of such a sliding pair to approximately 1/10. When both friction partners are coated, friction can be reduced by another 50%, an effect utilized in Formula 1 and Grand Prix racing. Wristpins, piston rings and camshafts are well suited for this type of coating. The pneumatic system as valve return spring widely used in the Formula 1 may have the largest speed potential but cannot be realized in series vehicles due to cost factors and other issues. The low-friction roller levers frequently utilized in automotive engineering cannot be used for reasons of weight and space. Even valve actuators with hydraulic valve lifters cannot be used in high-speed engines due to their reduced rigidity, engine speed strength and the larger moved masses.

8.3.1.7.3 Material and manufacturing Due to the demands on weight and size, mostly alloyed casehardened steels with high purity are used. Depths of harness between 0.7 and 1.1 mm and HRC values in the 59–63 range allow the required surface pressings. All gears and the drive shaft on which usually the first gear is placed are forged or cold extruded. Pockets and meshes are either raw or machined (with undercut), depending on the manufacturing process and precision. The meshing is milled, struck, ground or power-honed, depending on the sound and operational performance required, Tip reliefs or tip edge fractures help to compensate tooth curving under load. Precision lubrication of the travel gear groups is also standard (Figure 8.80).

8.3.1.7 Transmission Currently, most motorcycle engines are using constant-mesh transmissions. This type requires the least space and is thus predestined for use in motorcycles. The low-width facilitates the position on the motorcycle, supports the banking freedom and, of course, reduces the engine weight.

8.3.1.8 Clutch As a rule, disk clutches running in an oil bath are used in motorcycles. These multiplate clutches allow low hand forces by using relatively low exterior diameters.

8.3.1.7.1 Function All gears are constantly in contact and can be switched without the use of synchronizer rings. Due to identical speed of the gears, switching is usually performed without prior coupling and it is even possible to change gears while under load.

8.3.1.8.1 Function Friction disks (engage in the basket) and steel disks (engage in the follower) are pressed together by spiral springs. Amplifier systems on the basis of climb surfaces will reduce the hand force required. These systems are acting on both sides, that is, the moment to be transferred is increased under load and reduced in slide operation. This reduction causes the so-called antihopping effects in a motorcycle, that is, the rear wheel does not “stamp.”

8.3.1.7.2 Switching operation For this, the sliding wheels on which the claws (used as coupling) are placed must be shifted axially. The claws engage in the pockets of the floating wheels, creating the positive connection. The wheels are shifted by shifter forks that use pins to 11

3

14 22 12

6

23

13

5

2 10 Bearing, flush set 18

9 15

8

6.G

5

2.G

17

7

16

4

1.G

19

26

24

5.G

5

21 27

4.G

3.G

1

20

15

27 7

25

Pos. 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27

Bill of material

Drawing number 60033_Drive shaft 60033_Output shaft 0405222613 0405242813 0405263113 0417280150 0417300151 0471300150 0618091312 60033001000 60033002000 60033003000 60033005000 60033006000 60033010000 60033011000 60033012000 60033013000 60033014000 60033015000 60033016000 60033041000 60033042000 60033043000 60033044000 60033045000 60033046000

Designation Transmission output shaft M600 Transmission output shaft M600 Needle roller and cage assembly K22x26x13 TN with groove Needle roller and cage assembly K24x 28x13 F-81900-50 Needle roller and cage assembly K26x31x13 F-95085 Circlip Seeger SW28 Circlip Seeger SW30 Retaining ring DIN 471 - 30 x 1,5 Drawn cup needle roller bearing with open end HK 0912 B Needle cage, slotted Fixed 2nd gear Sliding 3rd/4th gear Idler 5th gear Idler 6th gear Output shaft Idler 1st gear Idler 2nd gear Idler 3rd gear Idler 4th gear Sliding 5th gear Sliding 6th gear Thrust washer 22.2/30.2/1.5 Thrust washer 28.3/35.75/1.5 Thrust washer 20.2/34/1 Thrust washer 24.5/35.75/1 Thrust washer 26.2/36/1.5 Thrust washer 30.2/39/1.5

Pcs. 1 1 1 1 4 1 2 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 2 1

Figure 8.80 Transmission and clutch.

352 | Internal Combustion Engine Handbook

6606_Book.indb 352

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

8

7 6 1 5 2 4

3

8 7 6 5 4 3 2 1

2638 P2-7504 5750 2633 P1-6803 P1-6804 2711/E P1-6805

543.32.012.000 548.32.080.000 598.32.602.03 581.32.081.000 546.32.003.000 548.32.002.000 575.32.013.000 548.32.001.072

Disco Condotto (Sp. 1.2), Driven Plate (Th. 1.2) Molla, Spring Vite M8x20, Screw M8x20 Scodellino, Cup Piatto Springdisco, Pressure Plate Tamburo Condotto, Driven Hub Rondella, Washer Assieme Campana, Housing Clutch Assembly

8 6 1 6 1 1 1 1

Figure 8.81 Clutch.

In addition, within the first disk, diaphragm springs are installed that are intended to support a soft engagement. The transmission of power into the transmission is created by a follower installed on the drive shaft. The coatings required for the transfer of moment is usually of an organic material (paper, cork, etc.) and is glued to aluminum carriers. The spacing of the trapezoid fields allow oil removal and will reduce the adhesion of the friction coating plates to the steel plates. The oil is mostly used for heat transport. Followers, pressure cap and basket are manufactured from aluminum and partially surface-treated (nickel-plating, hardcoated, PTFE (polytetrafluoroethylene) on climb surfaces. The basket is mounted on the clutch wheel and can pivot, thus allowing dampening generated either by radially arranged spiral springs or by rubber elements (Figure 8.81).

Figure 8.82 Racing operation.

8.3.2 Off-Road Motorcycles 8.3.2.1 Motocross Motocross is defined as driving on unpaved ground, usually enclosed circular courses with jumps that may be quite extended. During championship events in the USA, Motocross events are held in the open air while Supercross events happen in halls (Figure 8.82 and Figure 8.83). The premier difference is the length and layout of the course and the track conditions. The outdoor tracks are between 1.5 and 4 km and feature fewer spectacular jumps than the demanding short tracks in halls. Another factor is the ambient conditions outdoors under which the motorcycle and its engine have to survive. They can vary from very dusty and hard to muddy, slippery and wet.

Figure 8.83 Racing operation.

Internal Combustion Engine Handbook | 353

6606_Book.indb 353

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

The motorcycles developed for this purpose are subject to various rule sets but not homologated for road traffic. For the engine design, the following marginal conditions are taken into consideration: The engine should be operable in nearly any position for at least a short time. Above average clutch wear and resulting heavy contamination with oil must not have adverse effects. A rugged engine design is the prerequisite to counter rough handling in respect to impacts and sucked-in dust and sand. With these requirements, two-cycle engines with four-speed gears have been developed. 8.3.2.1.1 The origin of off-road sport The two-cycle single-cylinder engine was the unbeatable and traditional drive for the classic displacement variants of the racing regulations. Other engine concepts with the same displacement have not much of a chance against the very high-specific power yield of a two-stroke engine designed for racing. Only strong pressure and lobbying by the major Japanese manufacturers who want to start exclusively with four-cycle engines in off-road competitions for the reasons of product and marketing strategies, eventually changed the racing regulations for a four-cycle concept. This means that at the beginning of the new millennium, racing regulations were modified so that vehicles with four-cycle engines with close to or exactly double the displacement can start as vehicles with two-cycle engines in the individual classes. Today, 125 ccm two-cycle engines start against 250 ccm four-cycle engines and 250 ccm two-cycle against 450 ccm four-cycle engines in Enduro and Motocross competitions. The predominance of the two-cycle engine in off-road sport has been eliminated; however, the two-cycle engine is again fashionable after the original four-cycle boom, because it is still competitive under these uneven conditions (and cannot be bettered at the same displacement). In the E2 and E3 Enduro classes in particular, the two-cycle engines are very competitive. The cause is the very specific advantages of the two-cycle concept that became apparent, depending on the competition type, the road conditions and the driver’s requirements. The low weight and the lower purchasing and maintenance costs are also essential arguments for the two-cycle engine. In hobby and club racing and training events in particular, but also in youth sport or private teams with poorer financial backing, the cost factor plays an important role. 8.3.2.1.2 Demands on an off-road competition engine The least weight possible, a high-power yield with simultaneously good power delivery that can be translated into traction is essential characteristics to achieve the required driving dynamic. The optimal driving dynamic, which can only function in a perfect symbiosis between frame, engine and driver, is ultimately responsible for fast lap times. The vehicle must be adaptable to these demands as much as possible in order to adjust to the actual conditions (competition, course and driver).

This includes the following features as well: •• adjustable engine characteristic •• low-space requirement and integrability of the engine •• position of the rotating motor masses (handling vs. driving stability) and optimal harmonization of the mass moments of inertia to the engine characteristics (traction, drivability vs. response behavior and aggressiveness). 8.3.2.1.3 Types of two-cycle motocross and enduro engines Fifty-ccm automatic types (Figure 8.84 and Figure 8.85) are used in junior Motocross events and training races. These are air- or water-cooled engines fitted with a single-stage automatic centrifugal clutch for power transmission. The power yield can be up to 12 PS at 11,500 min–1.

Figure 8.84 50 SX motorcycle.

Figure 8.85 Mini in action.

8.3.2.1.3.1 65 ccm MC In 65-ccm MC capacity class, water-cooled two-cycle engines are used. Power is transmitted through a multidisk oil bath clutch and a mesh-controlled six-speed gear. At this point in time, approximately 16 PS are realized at 11,500 min–1; a higher power yield would impair the drivability.

354 | Internal Combustion Engine Handbook

6606_Book.indb 354

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

8.3.2.1.3.2 85/105 ccm MC (Figure 8.86) Here we have water-cooled two-cycle engines with exhaust timing using speed-dependent control valve positioning.

Figure 8.86 85 SX motorcycle.

The control valve is actuated by a centrifugal force regulator. Power is transmitted through a multidisk oil bath clutch and a mesh-controlled six-speed gear. Approximately, 26 PS are currently realized at 11,500 min–1. Output can be well transmitted during driving because the exhaust timing ensures a flatter power rise, that is, a higher torque in the lower and medium speed ranges. 8.3.2.1.3.3 125/200/250/300 ccm MC/E (Figure 8.87) Here we have a water-cooled two-cycle engine with exhaust time using speed-dependent control valve positioning for the outlet duct. In addition, the two auxiliary outlets are controlled by the rotary slide valve. The control valve and both auxiliary outlets are moved by a centrifugal force regulator and lever system or mechanical unit. Power is transmitted through a multidisk oil bath clutch and a mesh-controlled six- or five-speed gear. At this point int time, approximately 40 PS at 11,500 min–1 in a 125 ccm series model and 51 PS at 8500 min–1 in a 250 ccm series model can be realized. Output can be well transmitted during driving because the exhaust timing ensures a flatter power rise, that is, a higher torque in the lower and medium speed ranges. 8.3.2.1.4 Series solutions in two-cycle motocross and enduro competition engines Motocross and Enduro engines are designed to be as light and compact as possible; however, manufacturing costs must be controlled in series production. 8.3.2.1.4.1 Engine housings, water pump cover, clutch intermediate cover, clutch basket, clutch follower and clutch pressure plate:

Figure 8.87 Section through a125SX engine.

Die casting lends itself as the manufacturing process because very thin wall thicknesses (up to 2 mm), high accuracy combined with a high degree of prefabrication (which means that many mechanical machining steps can be omitted), very inexpensive blank and machining costs, and high levels of process reliability and quality are possible. The disadvantage is the high price of the die casting tools (approximately 300.000 € per tool which amortizes only from a certain number of units and the inflexibility in the design (a lost core process is not possible). 8.3.2.1.4.2 Engine covers For the reasons mentioned, ignition cap, clutch cover and controller covers are also cast in a die cast process but with MgAl9Zn1 magnesium alloy as the material. Magnesium die cast components have approximately 35% less weight than those of aluminum. The covers do not require mechanical finishing; cast sealing ring grooves accept molded sealing rings from NBR 70 (nitrile butadiene rubber) or Viton which, however, require prior powder coating because magnesium is highly corrosive. This fact also prevents the use of magnesium in components with coolant ducts because the coolant would cause the magnesium to corrode quickly.

•• Al226 GD-AlSi9Cu3 die casting alloy

Internal Combustion Engine Handbook | 355

6606_Book.indb 355

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

8.3.2.1.4.3 Cylinder, cylinder head (Figure 8.88) •• Gravity die casting of 12CuMgNi alloy with artificial aging Low-pressure gravity die casting lends itself for some components because lost cores are required (water space, overflow channels, outlet channel) on the one hand and the blanks are more exact and less expensive than sand-cast components on the other hand. The reason for this is the steel die with its cavities representing the outer form of the components. In sand casting, every outer form must be individually shaped in sand and is subsequently lost. In addition, the sand cores are more precisely placed in the steel die and the nonmold dependent geometries are also more precise than those in a sand-casting process. The running surface of the cylinder is coated with Nikasil, a galvanically applied nickel matrix with incorporated, very hard silicon carbides, honed and finally given a so-called cross-hatch finish that improves the oil absorption capacity of the cylinder wall.

8.3.2.1.4.6 Transmission Cold extruded gearwheels from 17CrNiMo6, high resistance, carburized. Blanks have a strain-hardened texture and high degree of prefabrication, the switching meshes and pockets are ready pressed. After machining, the components are electrochemically deburred to decrease the notch effect of the machining striae. Modern two-cycle Motocross and Enduro engines are fitted with sophisticated control mechanisms and channel arrangements. State-of-the-art is, for example, an exhaust gas gate with different actuations (Figure 8.89).

Figure 8.88 Engine section 125/200. See color section page 1080.

8.3.2.1.4.4 Pistons •• G-AlSi18CuMgNi die casting, low-pressure die casting, without lost cores The cast die is made completely from steel. The silicon content of the alloy is decisive for the wear behavior but also for achieving a thin piston wall. As a rule: The lower the Si content, the higher the ductile and fatigue strength and the better suited for light, thin-walled pistons which, however, will reach the wear limit at the piston skirt faster. Forged pistons are usually manufactured with maximal 14% Si because a higher portion will cause the material to crack during forging. 8.3.2.1.4.5 Crankshaft, connecting rod “Built” crankshaft with forged crankshaft webs made from 42CrMoS4, a tempered steel that can be easily machined with high strength and good ductile strength, non-hardened. Crank pin from 16MnCr5, carburized for high wear resistance with ductile core. Connecting rod from 15CrNi6, copperplated and thus partially carburized at the wear surfaces of the conrod eyes.

Figure 8.89 KTM exhaust timing unit.

From these essential characteristics, single-cylinder fourcycle engines have been developed after pressure was put on the regulations. They are and have been realized with, for the most part, rolling bearings. 8.3.2.1.5 Types of four-cycle motocross and enduro engines The realized displacements of 250, 450, and larger than 475 ccm result from the previously described FIM regulations for Enduro and Motocross competitions. Most manufacturers use the same basic engine type for each displacement for

356 | Internal Combustion Engine Handbook

6606_Book.indb 356

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

both Motocross and Enduro and add some features that are specifically designed for each application. A displacement >475 ccm is always a derivation of the 450 ccm engine with usually an enlarged bore (Figure 8.90 and Figure 8.91).

large and plain bearing in the small eye, and a piston forged from aluminum.

Figure 8.90 KTM 450 SX-F.

Figure 8.92 Crankshaft drive, valve train KTM 450SX-F.

Figure 8.91 KTM 450 SX-F right-side view.

8.3.2.1.5.1 Engine housing, cylinder, engine cover Vertically partitioned engine housings made from aluminum, with crank and gearing space and fitted coated closeddeck cylinder. Engine cover made from magnesium (if not water carrying). 8.3.2.1.5.2 Crankshaft drive, mass balancing (Figure 8.92) The crankshaft drive comprises a built crankshaft, mounted in anti-friction bearings with ball and/or rolling bearings, a single-part connecting rod with needle bearing in the

The piston diameters in Motocross engines are approximately 76 mm in the 250 ccm class and approximately 97 mm in the 450 ccm class. With the controlled speeds of 13,500 and 12,000 rpm, respectively, piston speeds of approximately 25 m/sec are the result. For friction, weight and space optimization, the pistons are realized as single-ring pistons. Short connecting rods with a push-rod ratio of approximately 0.30 ensure compact construction. The oscillating mass on the crankshaft webs is an essential component in the design of the engine characteristic. Reductions result in an improved handling of the overall vehicle and improved response behavior, increases will cause better traction and less chance of the engine stalling. Mounting on rolling bearings is selected for safety reasons, because due to frequent crashing in off-road operations, the engines are frequently forced to temporarily run without pressure oil supply. To reduce engine vibration, 450-ccm engines are designed with a balance shaft. 40–50% of the oscillating masses are usually balanced on the crankshaft but only 20–30% on the balance shaft due to space restrictions. In some of the 250-ccm engines, a balance shaft is omitted. This is possible due to a consequent reduction of the oscillating masses (currently approximately 240 g) and optimized mass balancing.

Internal Combustion Engine Handbook | 357

6606_Book.indb 357

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

8.3.2.1.5.3 Cylinder head, valve train, timing drive Due to high speeds and high valve accelerations and the resulting need for valve train rigidity, DOHC systems with cup valve or cam follower drive are mostly used in Motocross engines. Because the users of Endure engines do not demand very high speeds, some of them are still designed as SOHC with rocker arms. The advantages of the SOHC version comprise a more compact, lighter construction with lower maintenance and service requirements. All engines are realized as four- or five-valve models. To minimize mass, the valves are mostly made from titanium and MMC (metal matrix composite) is partially used for the spring collars. The timing drive is usually designed with toothed chain and hydraulic or mechanical chain tensioner. 8.3.2.1.5.4 Transmission and clutch The transmissions are designed as constant mesh to a large extent. Enduro and Motocross use six-speed transmissions in the 250-ccm class. Enduro transmissions have a larger angle to achieve a higher top speed for liaison sections during racing and for driving at walking speed in extreme conditions. Engines of 450 ccm are designed as four- to five-speeds for Motocross and five- to six-speeds for Enduro machines. The clutch is always designed as a multidisk oil bath clutch. Due to the extreme stresses in off-road operation (running with dragging clutch, impact after jumps at high differential speeds), clutch and transmission are designed to be very rugged.

8.3.2.1.5.6 Lubrication system To design the engines to be as compact and low-weight as possible, the manufacturers try to use as little oil as possible (1.2 to 1.5 liters). For this advantage, one accepts very short oil change intervals (5 to 15 operating hours). Series engines are realized with one or two oil circuits for engine block/cylinder head and transmission/clutch. The advantage of two circuits is a cleaner engine oil without clutch abrasion for engine block and cylinder head, the disadvantage is higher costs. The pressure circuit comprises an Eaton pump, pressure control valve and oil filter. Oil lines are usually integrated and the oil filter is always integrated in the housing and/ or the covers. To remove the oil from the engine block, Eaton suction pumps are used or diaphragm valves using the pump effect of the piston, or a common engine and transmission space with suction and separate oil tank. 8.3.2.1.5.7 Cooling system Due to the high-power density (approximately 130 PS/l at 450 ccm and 160 PS/l and 250 ccm), all engines are water-cooled. Enduro engines with their wide range of applications are fitted with a thermostat and retrofittable fan, while Motocross engines have neither thermostat nor fan. In most types, the water pump is placed on the balance shaft before the engine block; it may be also placed on the camshaft or a timing drive intermediate gear. 8.3.2.1.5.8 Starter unit For easier starting, all Enduro machines are now equipped with an E starter. Because multiple starting in difficult conditions will often fully drain the battery, a kickstarter is installed as a backup system. For weight reasons, an E starter is mostly omitted in Motocross (weight advantage of approximately 2 kg) (Figure 8.93 and Figure 8.94).

126

5,7

93,2

85

100

143,5

8.3.2.1.5.5 Mixture preparation Although all manufacturers work on electronic fuel injection system (EFI) for off-road bikes, the majority still use mostly flat side carburetors. Their advantages are high reliability in extreme applications, simple adaptability during engine tuning and changed ambient conditions (temperature and air pressure), at lower overall vehicle weight and lower manufacturing costs. EFI systems, however, offer significantly more options to design the engine characteristic. EFI became a

necessity in the Enduro sector due to the new exhaust gas laws in force since December 2007.

Figure 8.93 KTM 250 SX-F engine section.

358 | Internal Combustion Engine Handbook

6606_Book.indb 358

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Ø76

8.3 Motorcycle Engines/Special Engines

19,5

98

8.3.2.2 Enduro and Rally Unlike Motocross motorcycles, the Enduro motorcycle is an all-terrain vehicle that is approved for road use. To achieve road approval, performance reduction is often required to meet the sound limits (Figure 8.95 and Figure 8.96).

27,4

Figure 8.94 KTM 250 SX-F engine section.

At this point in time, these are realized with various exhaust silencer systems, influencing the ignition timing and slider stops. A considerable higher output can be assumed off public roads with a competition exhaust gas system and ignition curve and without slider stop. As additions to the aforementioned competition and sport Enduros, a wide range of road, travel, rally and hard Enduros were developed over time for a variety of applications and events and more or less off-road capability (Figure 8.97 and Figure 8.98).

Figure 8.95 Erzberg. Figure 8.97 KTM 525 EXC.

Figure 8.96 Paris–Dakar. Figure 8.98 KTM Rally.

Internal Combustion Engine Handbook | 359

6606_Book.indb 359

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Because these Enduro models have originated from different frame and engine families, two engine concepts are known. Firstly, there is the derivation from the Motocross engine and the one derived from a roadworthy engine (Figure 8.99).

The demands made by the legislation were thus not unreasonable and resulted in these changes: •• 1999 EURO-1, introduction of uniform regulations for emission limits •• 2002 EURO-2 for mopeds •• 2003 EURO-2 for motorcycles •• 2006 EURO-3 for motorcycles. In the last 12 years, major progress has been made in respect to exhaust emissions. The limits for the pollutants CO, HC and NOx were dramatically slashed across Europe. Overall, the limits for HC and CO were lowered by 95% and those for NOx by 50% since 1994 (Figure 8.100 and Figure 8.101). [g/km] 25 20

Figure 8.99 KTM 990 Adventure.

15

In both cases, the only possible options are changes to the cooling and oil circuits and to the transmission tuning.

10

8.3.2.3 Trial Similar to Motocross bikes, trial bikes do not usually have road approval and are developed purely as sport machines (trials are homologinated like Enduros). The displacement classes (50, 80 125, 250, and 320 ccm) similar to Enduro have developed from two-cycle engines and have been supplemented by four-cycle engines only since 2005. In this sport, the vehicle weight (70 to 80 kg) in addition to good throttle response and good low-speed behavior plays a decisive role. The technical modifications of the engines that have mostly been developed from Motocross types, focus on increased oscillation mass, different transmission tuning and, possibly, a more compact arrangement of the entire engine package.

0

8.3.3 Legislation 8.3.3.1 Exhaust Gas Emissions The first exhaust gas standards for motorcycles were introduced in 1994. Motorcycle riders and manufacturers responded less than positively to the new legislation. They were of the opinion that “those few bikes” would not have any significant effects on total emissions. However, the following factors must be taken into consideration as they were published by the Federal Environment Ministry some years ago: •• Motorcycles have a share of approximately 2.5% of the total kilometers traveled per year but 15% of the total HC emissions. •• The HC emission by all motorcycles in summer is approximately the same as the emissions by the total fleet of passenger cars with “G” catalytic converters and 20 times the mileage. •• The start and evaporative emissions represent more than 40% of the total two-wheel emissions.

CO

HC NOx

–95 % (CO) –95 % (HC) –50 % (NOx)

5 1994 ECE-R 40

1999 EURO-1

2003 EURO-2

2006 EURO-3

Figure 8.100 Exhaust limit reduction since 1994. [g/km] 25 –63 % (CO) –50 % (HC)

20 15

HC

NOx

–58 % (CO) –67 % (HC)

10

–64 % (CO) –70 % (HC) –50 % (NOx)

5 0

CO

1994 ECE-R 40

1999 EURO-1

2003 EURO-2

2006 EURO-3

Figure 8.101 Reduction phases.

Not just the limits themselves but also the exhaust test cycles for the determination of the exhaust values became more strict. Exhaust values are determined at a dynamometer at which the specified load/speed cycles can be accurately driven. In this process, the exhaust is removed at the discharge port of the muffler and fed into a CVS system (constant volume sampler). The exhaust gases are diluted in a dilution tunnel and a specified part is stored in three bags. In addition to the test cycle, the measuring system also analyses the bag content and determines the integral exhaust values by comparing them with the limits, but also the temporally dissolved pollutant values in order for an analysis as to where the motorcycle works well or poorly in respect to the emission values. Bag 1 contains the exhaust gas from the start and warm-up phase, bag 2 the gas from city cycles in

360 | Internal Combustion Engine Handbook

6606_Book.indb 360

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

warm conditions and bag 3 the gas from an overland tour at speeds of up to 120 km/h. The results enable the engine developer to work focused on improvements (Figure 8.102 and Figure 8.103).

Figure 8.103 CVS system.

Thus, the engines had 390 seconds to warm up and to heat the catalytic converter to its operating temperature (>250°C). Since 1 January 2006, EURO-3 is valid across Europe. Compared with EURO-2, the emission limits have been reduced by approximately 50%. 8.3.3.1.1 New limits •• CO < 2.0 g/km •• HC < 0.3 g/km •• NOx < 0.15 g/km.

Figure 8.102 Exhaust gas dynamometer.

Until 2006, the EURO-2 emission cycle was valid. The EURO2-cycle comprises six inner-city triple blocks with the exhaust gas removal being started only after the first inner-city block.

The measuring cycle has also been tightened. The measuring cycle for motorcycles with more than 150 cm3 displacement is not aligned with the new European driving cycle (NEDC) for passenger cars and is called the “NEDC Motorcycle.” It consists of two parts: Part 1 comprises six inner-city cycles (ECE R40 cold) as required previously. Part 2 simulates an extra-urban cycle (EUDC—extra urban drive cycle) intended to capture somewhat higher speed and load phases. A significant difference to the EURO-2 cycle is the start phase because the exhaust is measured from the beginning and thus records the entire cold-start phase as well (Figure 8.104).

130 Sampling

120 110 Part One

100

Part Two

Speed: [km/h]

90 UDC Cycle - 6 Phases IECE15 of 195 s = 1170 s km 1.013 · 6 Phases = 6.078 km Max. Speed: 50 km/h Average Speed: 19 km/h

80 70 60 50 40 30

EUDC Cycle of 400 s km 6.955 Max. Speed: 120 km/h Average Speed: 62,5 km/h

20 10 0

1

101

201

301

401

501

601

701

801 Time [s]

901

1001

1101

1201

1301

1401

1501

Figure 8.104 Euro-3 cycle for motorcycles above 150 ccm displacement.

Internal Combustion Engine Handbook | 361

6606_Book.indb 361

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

UDC Cycle - 6 Phases IECE15 of 195 s = 1170 s km 1.013 · 6 Phases = 6.078 km Max. Speed: 50 km/h Average Speed: 19 km/h Sampling

60

50

Speed: [km/h]

40

30

20

10

0 1

101

201

301

401

501

601

701

801

901

1001

1101

Time [s]

Figure 8.105 Euro-3-cycle for ≤150 ccm motorcycles.

For motorcycles with maximum 150 cm3 displacement, the Euro-3 cycle differs in not containing the overland tour. Only six ECE R40 urban cycles are measured (Figure 8.105). Because the emission measurement cycle describes the running behavior in an insufficient way, the European and international committees are working on the establishment of a world motorcycle testing cycle (WMTC). It comprises fewer stationary and more dynamic phases making the compliance with the limits even more difficult. For the WMTC, there is another vehicle classification according to displacement and maximum speed: •• For Class 1, part 1 is driven cold and hot and weighted at 50% each. •• Class 2 drives part 1 cold and part 2 hot, with part 1 weighted with 30% and part 2 with 70%. •• Class 3 drives part 1 cold, part 2 hot and part 3 hot, with the parts weighted with 25%, 50%, and 25%.

140

The actual cycle which is not yet fully adopted is shown in Figure 8.106. The WMTC classification is shown in Figure 8.107. Vmax [km/h] 3.2 3.1

< 130

2.2

United Nations

< 115 < 100 < 60 < 50

2.1 1.1

1.3 1.2 Displacement

Figure 8.107 Classification: WMTC cycle. See color section page 1080.

Phase 3

130 120 110

Phase 2

100

Speed: [km/h]

90

Phase 1

80 70 60 50 40 30 20 10 0

1

101

201

301

401

501

601

701

801

901

1001 1101 1201 1301 1401 1501 1601 1701

Time [s]

Figure 8.106 WMTC cycle.

362 | Internal Combustion Engine Handbook

6606_Book.indb 362

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

The Euro-3 limits can no longer be met without exhaust gas treatment using a catalytic converter. The engineers’ main tasks for reaching the exhaust gas limits are as follows: •• minimizing raw gas pollutant emissions •• exhaust gas treatment •• ECU application strategy. 8.3.3.1.2 Minimizing raw gas pollutant emissions Minimizing the raw gas pollutant emissions is realized with the following measures implemented in the engine: •• internal residual gas recirculation •• charge movement •• mixture formation

Some options and variabilities for improving exhaust emissions that are conceivable for motorcycles are described as follows: 8.3.3.1.3 Cam phaser, camshaft shifter One variability is the camshaft adjuster that is well established in cars and which can be designed as a two-stepped or continuous adjuster (Figure 8.108). It enables the rotation of the inlet and/or outlet camshaft against the crankshaft resulting in changed valve actuation times. The valve stroke progression remains unchanged, however. Using an inlet cam phase, the inlet closing can be adjusted at full load when speed is increased, positively influencing the air expenditure and the volumetric efficiency.

•• combustion chamber design. 8.3.3.1.2.1 Internal residual gas return To minimize raw gas pollutant emissions, a compromise is attempted between full-load torque and output on the one side, and combustion in the partial load range at the other side. High-power yield is possible only at high speeds in motorcycles. This requires a design of charge changing organs (valve diameter, duct diameter and geometry, valve control times, intake manifold length and diameter) for low flow losses at nominal speed. This design has inherent disadvantages in the entire partial load range and in low load range near idle speeds in particular. The large valve overlap cross-sections generate a high internal exhaust gas recirculation rate with the undischarged residual gas of a cycle flowing through the vacuum at the intake side into the intake manifold when the inlet opens and is then again taken in during the next cycle. The large duct diameters return low intake speeds of the fresh mixture at low-charge density in partial load and low excitations due to low speeds. This causes a nearly absent charge movement resulting in a poor mixture of residual gas and fresh mixture. The result is an inhomogeneous mixture that is difficult to ignite due to the high portion of residual gas. In addition, the flame transport speed is very low which causes the already very slowly occurring ignition to move relatively slowly into the combustion chamber. Under these circumstance, combustion may occur far beyond 90° after TDC with the result that the outlet will reopen during this delayed combustion. This causes a loss in work and efficiency, excessive raw gas pollutant emissions and, due to the high exhaust temperatures, a badly loaded catalytic converter. In lowest load, the residual gas portion can raise so that ignition or burning is no longer possible. This effect can cause the engine to occasionally stop or even stall, in particular in engines with little oscillating mass. Due to the desire for continually increasing performance and the concomitant tighter emission laws, motorcycle engines will be fitted with variabilities. This is the only way to combine good flow properties and also good partial load behavior. Type and number of the variabilities will be limited, compared to car, due to available space, weight, engine speed strength, and costs.

Figure 8.108 Stroke selector.

On the other hand, adjusting the intake opening time can adjust the valve overlap cross-section in partial load and in the critical low-load ranges, close idling in particular. With these measures, good burning stability, good emission behavior and even improved consumption can be obtained in the near-idle range. 8.3.3.1.4 Fully variable valve trains (Figure 8.109) If a variation of the valve stroke progression should be obtained in addition to the temporal shift of the valve opening time, only a system with variable valve stroke can be used. The fully variable systems, such as Valvetronic by BMW, VVH (German acronym for variable valve stroke), by META and similar systems are very complex and insufficiently speedstrong due to the large moved masses (rolling contact and similar) for an application in a motorcycle engine. 8.3.3.1.5 Valve stroke selector The less complex systems of a valve stroke selector such as VTEC by Honda or VarioCam+ by Porsche are deemed to be more suitable under the aforementioned considerations.

Internal Combustion Engine Handbook | 363

6606_Book.indb 363

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Valve stroke, valve opening time, IO and IC can be adjusted as two-step switches which are fully sufficient for motorcycle application.

the intake flow and also generates a focused swirling flow in the cylinder space, effecting a significantly better charge movement and thus improving the mixture treatment and accelerating combustion.

Figure 8.109 Stroke selector.

Reliable and clean combustion required a high flame speed. The flame speed is comprised from the pure burn rate of the flame and the flame transport speed. The burn rate is, among other things, very dependent on the air ratio λ and the residual gas content. A lean mixture with λ = 1.25 burns much slower than a rich mixture with λ = 0.85. Large quantities of residual gas also decelerates the burning process. In λ =1-operation, an engineer has few options to influence this behavior. The flame transport speed, on the other hand, is the speed at which the flame is carried forward by the moving cylinder charge. This charge movement in the combustion chamber can be controlled in a targeted fashion by implementing various measures. 8.3.3.1.6 Tumble For example, injection of the fresh mixture over the top half of the intake valve plate may generate a tumble flow in the cylinder space, that is, a turbulent flow diagonally to the cylinder axis (Figure 8.110). In a pronounced form, the tumble can significantly reduce the burning time and improve consumption and exhaust gas values. The motorcycle channel geometries and the low-charge density, however, will allow only relatively low-energy tumble flows in the near-idle range where the charge movement is required. The options for a burn behavior improvement are severely limited. 8.3.3.1.7 Channel disconnection (Figure 8.111) If separate throttle valves are intended for the intake valves, the inflow can be influenced in respect to speed and orientation. One intake channel can be disconnected in the low-load range. The total mixture mass flows through an intake valve which, due to the smaller cross-section, significantly increases

Figure 8.110 Tumble channel.

Figure 8.111 Channel disconnection.

8.3.3.1.8 Valve disconnection Instead of closing a throttle valve to shut down a channel, the intake valve may also be shut down. This is a very efficient method but requires switches in the very fast-moving components of the valve train. 8.3.3.1.9 Bypass systems Another option is small bypass bores in the intake area of the cylinder head, which are aligned with the intake valve. At low load, the throttle valve is closed and the mixture is

364 | Internal Combustion Engine Handbook

6606_Book.indb 364

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

forced through these bores. Thanks to the high flow speeds, stable swirl and tumble flows can be generated. Due to the small cross-sections, a more sensitive control is possible, compared to the throttle valve. The application is, however, restricted to small map area.

the limits, the engine system has to be upgraded and the strategies must be revised.

8.3.3.1.10 Combustion chamber design The combustion chamber is formed from the cylinder head and the piston. The objectives in its design are as follows: •• compact form, favorable surface/volume ratio → good η i (inner efficiency) •• central spark plug position → uniformly short flame paths •• few obstacles for the mixture intake → good filling •• few obstacles for the flame distribution → good combustion •• gap minimization → Avoidance of the flame becoming extinguished. Earlier two-valve cylinder heads frequently features semispherical or lens-shaped molded combustion chambers. These forms are very compact and have little heat loss due to their surface/volume ratio thus enabling good inner efficiency. The disadvantage of this design is the low compression ratio that can be achieved with a flat piston. To reach a high compression ratio (ε > 11) as is required for efficiency increases, one has to use piston heads with superstructures. However, this results not only in an increased piston weight due to the necessary valve seat pockets but also in fissured combustion chamber surfaces which means that their enlarged surfaces will adversely affect the efficiency. Modern motorcycle engines are usually designed as multivalve engines. The majority of the engines feature four-valve cylinder heads. The number and size of the valves determines to a large extent the combustion chamber shape. The result is a rather flat, roof-shaped combustion chamber that stays quite compact with small valve angles and has an acceptable surface/volume ratio. With flat pistons and little valve seat pockets, compression ratios of approximately ε = 12 can be reached. Using larger valve seat pockets, the compression ratio can be increased to approximately 13.5 in thermodynamically-satisfactory combustion chambers. 8.3.3.1.11 Exhaust gas treatment For EURO-2, it was often considered good enough to use a passive secondary air system to meet the limits. By guiding air into the exhaust system, the CO and HC emissions generated in idle and thrust phases were after burned in the exhaust system. Uncontrolled small catalytic converters with 100–200 cpsi (cells per square inch) were also frequently used (Figure 8.112). During cold start, the fuel conversion of the secondary air system supported catalytic converter heating. However, such a simple system cannot meet the EURO-3 limits values. To ensure that the exhaust gas values are below

Figure 8.112 Catalytic converters.

8.3.3.1.11.1 Catalytic converter For a sufficiently large converting surface, a catalytic converter cell density of 200–300 cpsi (cells per square inch) is necessary. With an increased number of cells, the surface disproportionally increases, as does the surface, as does the exhaust gas backpressure, negatively affecting the maximum moment and, most of all, the maximum performance. This can be compensated in part by a larger catalytic converter cross-section. For these reasons, the catalytic converter volume must be approximately 0.5–0.8 l/l of the working volume, in order to not only achieve a sufficient convection surface but also to minimize the exhaust gas backpressure. The position of the converter is important for quickly reaching the actuation temperature (Light off) (Figure 8.113). Thus, the catalytic converters will move away from the muffler toward the engine in motorcycle engines, as they already do in passenger car engines. The cell numbers of the converters will increase to 400 cpsi. 8.3.3.1.11.2 Mixture preparation A warmed-up catalytic converter can convert, when exactly maintaining a stoichiometric mixture (λ = 1), up to 98% of the pollutants into CO2 and water. The difficulty is the exact maintenance of λ = 1, which makes a mixture preparation control and thus an electronically-controlled injection system, mandatory. This is the only option to quickly respond to lambda fluctuations. The lead variable used is the signal of the lambda probe that detects the oxygen content in the exhaust gas. The motor controller or ECU (electronic control unit) evaluates the lambda probe’s signal and use the injection quantity to generate a stoichiometric mixture. In stationary machines, this works very well when the control algorithms are attuned to the charge changing system. However, two motorcycle operating states differ strongly from the ideal stationary operation: •• start (cold, warm, repeat) •• acceleration and deceleration.

Internal Combustion Engine Handbook | 365

6606_Book.indb 365

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Acceleration and deceleration In normal operation, the exact maintenance of the fuel-air mixture is extraordinarily difficult due to the engine behavior that is much more dynamic than that of a car. The low oscillating mass of motorcycle engines leads to very large speed gradients, which makes it difficult to determine the actual speed at the time of injection. Furthermore, the lambda probe always only displays the state of the previous cycle and the ECU always lags behind. If one would use only the map values and the lambda regulation, the mixture would become extremely lean during sudden acceleration and over-rich during transition into thrust phase. Besides driving issues, this would also affect the exhaust gas conversion. To minimize these inadequacies, the ECU has functions for enrichment in acceleration, taking each running state and the driver’s intervention (change gradient of the gas handle position) into account. The Euro-3 harmonization has some inherent target conflicts: •• The increasing number of cells increases the active catalytic converter surface but also the exhaust gas back pressure that impedes improving the full-charge performance. •• The near-engine position of the catalytic converter necessary for its fast actuation can lead to major overheating problems during high speed at full charge or in thrust phases with mixture enrichment. •• Lambda1 operation and thrust shut-off lead to a deterioration of the drivability in low-charge ranges and charge changing phases.

Figure 8.113 Near-engine cat position (Kawasaki ZX 10R, Yamaha R1).

 tart S If the engine is started in a cold state, several adverse conditions are present: •• The lambda probe is still cold and not ready for operation. •• The ECU can pre-regulate only with the injection values stored in the engine temperature data. A regulation is not yet possible. •• All engine components are cold. •• The fuel vaporization in the intake section is severely limited, causing strong wall coating formation. For this reason, the mixture must be heavily enriched in the cold-start phase to ensure that the engine starts. It is important that the enrichment is quickly reduced to λ = 1 with rising engine temperature and ready lambda probe. •• The catalytic converter is still cold. •• As long as the temperature of the catalytic converter is below 250°C, exhaust gas conversion is not or only slightly possible.

8.3.3.1.12 ECU application strategy and objectives 1.  Catalytic converter warm-up to operating temperature as quickly as possible The objective is to reach high exhaust gas temperatures in the cold start and warm run phase, in order to quickly heat all heat-absorbing components, specifically, the catalytic converter. For this purpose, igniting may be initiated very late, which brings a late and lagging combustion with it which partially progresses into the exhaust phase. In addition, for the first cold start phase, a very rich mixture is provided, supporting combustion into the exhaust system. Furthermore, the rich mixture can be after burned by blowing air into the exhaust system resulting in the development of high exhaust gas temperatures in the catalytic converter. A time and temperature function will then adjust the mixture toward a lambda = 1. As soon as the catalytic converter has reached 250°C, a high conversion rate can be obtained in λ = 1 operation. 2. Maintaining λ = 1 as long and as accurately as possible The exhaust gas treatment in a new catalytic converter at operational temperature can convert more than 97% of the exhaust pollutants into CO2 and water at a combustion air ratio of λ = 1. Such reduction cannot be obtained with combustionimproving measures. Hence, it must be attempted to maintain lambda = 1 even in dynamic operation. A good control algorithm makes this possible at constant runs. Problems arise during cold start, idle and in acceleration phases where enrichment is mandatory to realizing good driving behavior. To reach this compromise, the acceleration enrichment must be adjusted very carefully.

366 | Internal Combustion Engine Handbook

6606_Book.indb 366

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

3.  Overrun fuel cut-off Combustion should not occur in thrust phases with closed throttle valve. For this reason, it is intended to shut off fuel during thrust. The transition from drive operation to thrust operation and vice versa is a very sensitive area, however. In the lowest charge range, ignition and combustion are sluggish. In a high-performance design with large overlaps, cycles with very poor combustion or even stops can occur causing the engine to jerk. In tight turns with closed throttle valve, the restart of fuel injection generates a hammer, bringing disturbance into the running behavior. This trend is exacerbated with high-performance throttles with large diameters. In lowest charge phases, a change of the throttle valve angle of 1° can mean a multiplication of the power applied. In idle or near-idle range, only the addition of a second throttle valve or a bypass system will help to improve the driving behavior. In future, E-gas systems will become common, as they are in cars. With an E-gas system, a mechanical connection between gas handle and throttle valve is eliminated. At the gas handle, a potentiometer will accept the driver’s request and transfer to the ECU. Here, the information is processed, compared with the current driving situation and adjustment information is sent to an electronically controlled throttle valve. Such a system has a number of advantages: The “response behavior” of the gas handle can be stored by map and modified. This means that the sensitivity can be adjusted depending on charge, speed and velocity. In near-idle urban operation, where even small changes to the throttle valve position have major effects, a less sensitive response behavior can be set, significantly facilitating the application for good driving behavior. At higher speeds and charges, responding may be set to fast and thus support sporty driving. Even excessively fast opening of the throttle valves leading to insufficiently quick adjustment of the injection volume with the engine choking can be avoided by adjusting the throttle valve opening speed to the acceleration enrichment. An E-gas system can also be

integrated into safety-relevant systems such as anti-slip and wheelie controls. In the Moto GP and Superbike top racing series, such systems are already used with great success. 8.3.3.1.13 Evaporative emissions In future, in all states of the United States and in Europe, not only the exhaust gas emissions will be limited, as is already the case in California. The motorcycle system will then be considered holistically. This means that even fuel emissions emitted through the tank surface, the tank cover, the hoses and all fuel-carrying components must be minimized. California has already limits for evaporative emissions. They are determined in the SHED test. The SHED chamber is HC-proof chamber (SHED = Sealed Housing for Evaporative Detection) into which the motorcycle is sealed for the test (Figure 8.114). Using an established testing procedure, various vehicle states are generated and the arising vaporizations are measured. The limit is 2 g/test. Figure 8.115 schematically describes the process during an SHED test. 8.3.3.2 Noise Emissions Since 1986, the limits for motorcycle noise emissions have been successively lowered. This trend will continue although the exact date for the introduction of new limits and the precise definition of the then-tightened measuring conditions are not known at this point in time (Figure 8.116). The current measuring procedure for determining noise emissions is defined as follows (Figure 8.117): •• 20 m long and 15 m wide measuring path (ISO asphalt) •• two microphones in a distance of 7.5 m to the left and right of the measuring path •• arrival in the measuring path at constant 50 km/h •• full charge acceleration from 50 km/h in second and third gear •• noise limit: 80 dBA.

Figure 8.114 SHED test stand.

Internal Combustion Engine Handbook | 367

6606_Book.indb 367

1/19/16 8:38 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Start Tank filling to 50 % of the test fuel nominal volume: Indolene or Phase 2 CARB gasoline

Drain and refill fuel max. 60 min Precondition at chassis dynamometer

Phase 1 + 2 of FTP 75 (cold start + stabilized phase)

max. 5 min Cold stop phase

12 to 36 hours Tank filling to 50 % of the test fuel nominal volume: Indolene or Phase 2 CARB gasoline

Drain and refill fuel Warm-up phase in SHED HC emissions caused by temperature fluctuations

Duration: 1 hour; separate heating pads for fuel and vapor, 1 thermocouple each

0–60 min Cold start exhaust gas test

Phase 1 + 2 of FTP 75 (cold start + stabilized phase)

10 min Hot start exhaust gas test

Phase 1 of FTP (hot start)

max. 7 min Hot evaporation phase in SHED chamber

1 hour

End

Figure 8.115 SHED test process.

Noise limits for motorbikes Motor bikes > 175 cm 3 86 dBA

86 dBA

Future limits

83 dBA 80 dBA 80 dBA

1984

1988 1990

1993

76–77 dBA

2006 2009–2011 ?

Figure 8.116 Noise limits development since 1984.

No limits are defined for stationary noise of road-approved vehicles. The measurement is made in a distance of 0.5 m at

a 45° angle to the exhaust outlet and the values are recorded (Figure 8.118). The noise measuring procedures for accelerated passing which presumably becomes obligatory sometime between 2009 and 2011 is closely based on the aforementioned procedure. The decisive difference is the acceleration phase; previously optional, the acceleration phase in the measuring path is now specified by the motorcycle’s power-to-weight ratio. The necessary curve for determining the target acceleration is still under discussion by the EU expert panels. With this measure, the European legislative body tries to exclude manipulations of noise test detection by the manufacturers. In the past, it could not be dismissed that a control device detects the noise measuring procedure and, for example, closed the second electronically controlled throttle valve so far when passing through test section, that the acceleration and the related noise generation remained low.

Photocell Radar

Mic. left

7m

10 m

10 m

7m

0m Mic. right

Figure 8.117 Noise measuring system.

368 | Internal Combustion Engine Handbook

6606_Book.indb 368

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

•• engine vibrations (engine order or multiples) are transmitted to other components radiating the noise. 45+ 5°

5

≥3

0,

≥3

≥3

≥3

Figure 8.118 Boundary conditions for stationary noise measurement.

The intake sound was also strongly dampened by the mostly closed valve. The noise level was higher in real driving operation. In addition to the change in the measuring procedure, the limits for noise emission will be lowered from 80 dB to presumably 76–77 dB. The new limits and the tougher procedures will both lead to very major development efforts by all motorcycle manufacturers. In future, muzzle noises (intake and exhaust sounds) must be heavily dampened. It will be very difficult to achieve this without losses in performance and torque. The silencer volumes will have to be designed with much larger dimensions. The air box, too, will become larger to attenuate the amplitudes of the pressure pulsation. The muzzles of the straight large-volume intake paths required for a high power yield must then be attached either covered, or fitted with resonator elements or variabilities such as valve systems. In most cases, these measures will not be sufficient to comply with the limit values. In addition, at very low muzzle sounds, the engine sounds become more predominant. Clanging, rattling, grinding and clattering of the engine dampen the feeling of quality. That is why the noise emissions of engine and transmission, road noise and noise radiation of the excited components must be significantly reduced, compared to what is acceptable today. Mechanical engine noises are as follows: •• piston canting •• timing drive (gears, chains, tensioner) •• valve train (valve contact, push rod, cam follower) •• primary drive •• transmission chatter (tooth face play) •• oscillating engine surfaces •• vibrations •• inertial forces •• inertial torques •• combustion noise •• chains knocking on the lever •• run-in sounds (pinion, sprocket) •• sound conduction tracts

In future, the acoustics specialists or NVH engineering (NVH = noise, vibration, harshness) must be involved in the design phase at an early stage to minimize, in close collaboration with the structural strength specialists, the vibration amplitudes in the individual components. Potential sound conduction tracts can be avoided, interrupted or dampened at an early stage. To determine the individual portions of the noise sources to the overall noise level, traditional noise level measurements are completed at the specific component while all other components are isolated. Modern measuring procedures such as array measurement, laser vibrometers or holography may also be used. With correct handling, it is possible to very quickly determine conspicuous noise sources. Depending on the frequency range examined, either of these measuring procedures may be successful.

8.3.4 Racing Engines 8.3.4.1 125 and 250 2T for GP In both classes—125 ccm and 250 ccm categories—prototypes and replicas manufactured in small series are used. In the 125 class, the maximum permissible vehicle weight plus driver including his clothing is 136 kg. In the 250 class, the vehicle weight alone is limited to at least 98 kg. Leaded fuel is prohibited. The engine’s technical design is very specifically focused on optimum power yield over the entire race distance, low weight, good integrability into the frame, tunability and adaptability to each course, driver and meteorological condition. Water-cooled single or two-cylinder two-cycle engines are used, with or without electric or pneumatic outlet control. The number of cylinders is limited by the regulations and is restricted to one cylinder in the 125 class (until the late 1980s, two cylinders were allowed) and to maximal two cylinders in the 250 class. Engines (125 class) reach an output of more than 50 PS and 250 class one of above 100 PS at approximately 13,000 min–1 both and over speed up to 14,200 min–1. Depending on the race course, top speeds of up to 240 km/h (125) and 280 km/h (250) can be reached (Figure 8.119). In the 250 class, V but also inline arrangements (only KTM at this time) are common, with the cylinders in a two-stroke engine each requiring a dedicated separate crankcase; “real” V engines with a single crank pin for both connecting rods are thus not possible. The engine characteristic can be optimized with an ignition offset of both power units that differs from optimal mass balancing. An ignition timing offset from the first to the second cylinder of 90° crankshaft angle are common in modern 250 racing motorcycles. Power is transmitted via multidisk dry disk clutches (better efficiency than oil bath clutch) over jaw-type switched six-gear cassette gearbox (changing individual gear steps or translation ratios possible without engine disassembly) onto the drive

Internal Combustion Engine Handbook | 369

6606_Book.indb 369

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Figure 8.119 GP 250 racing engine (Aoyama).

chain and the rear wheel. The water pump is driven electrically (fewer friction losses) or mechanically via the crankshaft. The crankcase is also frequently water-cooled in a partial range. Today, all engines have balance shafts to significantly reduce engine vibrations. This is necessary because the engines are rigidly bolted within the aluminum frame to act as joint load-bearing elements. Frame cracks and vibrations at the handlebars and footrests interfering with the driver would be present without balance shafts. The mixture treatment is realized using one or two (250 class) slide carburetors and additional, electronically-controlled fuel addition with solenoid valve (power jet) or injection nozzle up- or downstream of the carburetor. Thanks to this additional tuning option, the engine can be actively and performance-optimized injected via the carburetor because it is possible to add additional fuel to the lean-tuned engine in critical charge conditions. This helps to prevent piston seizures or jamming which would be unavoidable with a lean performance-optimized carburetor adjustment. These issues arise under high speeds and low carburetor slide opening in particular, which is, when little vaporization cooling at the piston head and poor mixture oil supply at the cylinder walls are present, or with knocking engine. In addition, this system is a good tuning instrument to improve the driving behavior of the engine when opening the gas gate in the marginal area at the curve crest or when accelerating out of the curve. The intake is realized as an air box in which the carburetor(s) is or are installed. Depending on the speed, this air box builds a ram pressure that improves the filling and charges the engine. The ignition timing is variably adjusted over a map, depending on the speed, gate position and, in part, other influencing factors. An electronic, freely programmable ignition system is used. A data recording function records the most important measuring data during driving, which are essential for tuning and improvement of the vehicle. They include the following: •• velocity •• rotational Speed •• gear detection

•• gate position •• exhaust and water temperature •• air box pressure and temperature •• engine block temperature •• front and rear wheel speed (spin) •• various frame and chassis data. Figure 8.120 provides a short sample description of the KTM V4-GP1 engine. Model, V-angle

–/°

V4/75

Firing sequence

–

1-4-2-3

Crankshaft throw

° Crankshaft angle

360

Cylinder spacing

mm

94

Bore/stroke

mm

84/44.6

Displacement

cm3

989

Conrod length

mm

96.5

Valves/cylinders

–

4

I/O valve face ratio

–

n/a

Compression ratio

–

14:1

Mixture preparation

2 injection nozzles per cylinder

Valve train

Cam follower operation, pneumatic valve spring

Timing drive

Gearing drive

Cooling system

Cross-flow water cooling

Lubrication

Integrated dry sump lubrication

Transmission

6-gear cassette-type transmission

Mass balancing

95% balancing of first-order forces with counterweights at crankshaft webs and opposite-running rotating balancing wheels, 100% balancing of first order torques at the balancing wheels

Clutch

Dry clutch

Engine management

McLaren Electronics

Engine weight

kg

58

Figure 8.120 KTM V4-GP1 engine data.

370 | Internal Combustion Engine Handbook

6606_Book.indb 370

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

8.3.4.2 GP1 The MotoGP class has the following boundary conditions: •• solely prototypes with a displacement of maximal 990 ccm •• motorcycle overall weight: two- and three-cylinder engines—135 kg four- and five-cylinder engines—145 kg

engines with six and more cylinders—155 kg

•• naturally aspirated engines •• maximum tank size 24 liters, 2005—22 liters. 8.3.4.3 Construction The engine housing halves manufactured in a sand-casting process (Figure 8.121) from G AlSi7MgCu 0.5 are horizontally split. In this split plane, the bearing system of the drive shaft for the gear drive is installed. The housing top is realized as a closed-deck design and unites the desired light-weight construction with the required increased structural rigidity. The cylinder running surfaces are nikasil-coated to meet the high tribologic stresses. For each cylinder bank, six tie rods ensure the connection between cylinder head and housing top. Due to the high-peak combustion pressures and the inertial forces caused by the high-speed concept, a bed plate construction ensures, in combination with double bolting, that the design can withstand the increased stresses. Primary objective is a rigid design of the main bearing race and the direct guidance of the force lines from the cylinder head over the tie rods of the crankcase top to a stiff connection to the bed plate construction.

design characteristics, oil suction pumps, oil pressure pump and mechanical oil/air centrifugal separators have been installed in the housing bottom. A further motorcycle-specific characteristic is the cast-on gear housing with integrated bearing seats, with the design as a fast-change gear for racing being given. The core component of the crank gear (Figure 8.122) is the triple bearing of the crankshaft, which has been machined from solid, with four counterweights and gas-nitrided main and conrod bearings. Due to the gas-dynamic advantages in the overall tuning of intake and outlet system, a crankshaft throw of 360° is provided. The excellent high bending and torsion resistance and the end-to-end light-weight construction at the crank gear guarantee a high natural frequency and the basis for the high-speed concept.

Figure 8.122 Crankshaft triple bearing.

Figure 8.121 Section through the KTM V4-GP1 engine block.

In this manner, the peaks in surface pressings during full-charge operation, the tensions and the specific bearing deformations can be reduced. The construction-oriented implementation of the required integrative solutions has been realized in the housing with the water duct in the V of the cylinder banks. Similar to these

Balancing the inertial forces and torques is realized with imbalances at the crankshaft webs and two additional balancing wheels at the lateral surfaces of the crankcase. With this arrangement, first-order inertial forces can be offset to 95% and first-order inertial torques to 100%. When examining the two first orders, the overall stress on the engine suspension points is comparably lower than in the four-cylinder inline engine (Figure 8.123). The cylinder head (Figure 8.124) is manufactured from G AlSi7MgCu 0.5 in a sand-casting process. It is designed as a four-valve concept with central spark plug position. The two camshafts have triple bearings and are additionally supported by a needle bearing. The drive of the mechanical fuel pressure pump is integrated via the exhaust camshaft. The valve seat insert and valve stem guide are manufactured from copper beryllium. The demand for a compact, combustion-optimized combustion chamber, high compression and maximal possible utilization of valve cross-sections resulted in a slightly radial arrangement of the valves. The pneumatic spring is integrated

Internal Combustion Engine Handbook | 371

6606_Book.indb 371

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

6000

Overall force over the crankshaft angle (V4/75 Var_30)

6000 5000

Overall force [N]

Overall force [N]

5000 4000 3000 2000

4000 3000 2000 1000

1000 0

Overall force over the crankshaft angle (Inline 4)

0

FMAG front

90

180 270 Crankshaft angle [degrees]

FMAG rear bottom

360

FMAG rear top

0

0

FMAG front

90

180 270 Crankshaft angle [degrees]

FMAG rear bottom

360

FMAG rear top

Figure 8.123 Engine suspension stresses. See color section page 1080.

via cast-on cylinders and pressure lines for a leakage-caused supply with nitrogen and provision of the required system pressure. Proper bearing of the cam followers is assured by bolted-in bearing blocks. The functions of the head gasket are assumed by gas-filled steel rings towards the combustion chamber and additional O-rings for a tight passage of the cooling water and oil flows. The upper bearing of both camshafts per cylinder bank is realized directly with the cylinder head shroud milled from aluminum. This constructive solution allows for an ultra-stiff camshaft bearing and a compact design.

stresses. The designed selection of low gear ratios, together with a stiff gear design and in conjunction with the housing design, enable an operation of up to 18,000 rpm with a respectable safety distance to the system’s natural frequencies. The excitation arises from the engine first and third order of the camshaft torques.

Figure 8.125 Timing drive designed as gear drive.

Figure 8.124 Cylinder head.

The high demand on the accuracy of the kinematic valve stroke to be mapped and the demand for meeting the control times in dynamic operation, result in the mechanical design of the timing drive as a gear drive (Figure 8.125). The conceptual determination of the valve train and the resulting options in the design of the valve accelerations lead to high dynamic

The first rotary natural frequency of the system is at approximately 1,150 Hz, which corresponds in a third-order engine excitation to an engine speed of 23,000 rpm. As an unrestrictedly suitable control concept for charge changing, the valve train was realized as a cam follower drive with pneumatic valve spring (illustration), corresponding to the high-speed demands on a racing engine. At the forefront are, on the one hand, the free design of the kinematic valve stroke progression to optimize the volumetric efficiency and, on the other hand, the mechanical ruggedness of the system. The realization of the desired valve accelerations and dynamic behavior of the valve train result in a consequent

372 | Internal Combustion Engine Handbook

6606_Book.indb 372

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

reduction of mass in the individual components. The valves and pneumatic pistons are manufactured from titanium and the valve seats and valve guide from copper beryllium. The oscillating masses and essential data of the valve train are shown in Figure 8.126.

Valve

Unit

Inlet

Exhaust

g

19.25

17

Valve lock

g

 0.45

 0.45

Adjusting shims

g

 0.85

 0.85

Pneumatic piston

g

 6

 5.35

Oscillating cam follower portion

g

 8.46

 8.46

Total oscillating mass

g

35.01

32.11

Valve stroke

mm

14.5

12.1

Maximum valve acceleration

mm/rad2

77

62

Opening time at1 mm stroke

° Crankshaft angle

n/a

n/a

Figure 8.126 Valve train specific parameters.

The use of pneumatic valve springs crucially reduces the moved overall mass with simultaneous optimization of the dynamic vibration behavior. Further advantages are realized for the spring characteristics by varying the pneumatic pressure. The layout of the pneumatic system is shown in Figure 8.127. An integrated high-pressure receiver (250 bar) in the V of the cylinder banks serves as a pressure accumulator that uses a mechanical two-stage valve and a damping volume to supply the system pressure (13 bar) to the pneumatic cylinders. The nitrogen acts here similar to a pneumatic valve spring. The design-related leakage losses are offset by the low-pressure system via throttle cross-section toward the cylinder volume.

Figure 8.127 Pneumatic system of the valve actuation.

To optimize the tribologic behavior of the valve train, camshafts and cam followers are DLC-coated. Fresh oil lubrication is ensured through jet nozzles that supply the contact camshaft to the cam follower. The embodied integrated dry sump lubrication features a pressure and an intake circuit. This modularly designed pump string comprises a pressure pump and two suction pumps from wear-resistant aluminum and is installed in parallel to the crankshaft axis in the lower crankcase. The pressure pump sucks oil via an intake snorkel from the dry sump and transports it through the oil pressure control valve, the oil filter and the oil/water heat exchanger to the consumers. The oil pressure control valve and the oil filter are realized as “cartridges” and can be easily replaced. For the distribution, a system of separate oil lines is provided with consumer-dependent throttle cross-sections for optimizing the oil flow volumes. The main and conrod bearings are separately supplied with oil to ensure reliable and especially even build-up of the oil film. The main bearings are supplied from an oil passage within the V, while the conrod bearings and the conrod eye are centrally supplied via a sliding ring seal at the crankshaft. In the cylinder head, the cam follower axles and the contact cam to follower are lubricated through jet nozzles, as are the camshaft bearing points. The friction in the transmission is also reduced by focused jet nozzles towards the tooth face contact. One suction pump each evacuates the closed crankcase of cylinder 2/4 and cylinder 1/3 and the separate suction of the cylinder heads 1/2 and 3/4, with the residual oil volume flowing pressure-free from the transmission space into the sump. This sucked off air/oil mixture is combined at the pressure side of the suction pumps. The air is reliably separated from the oil in two separators one behind the other. The first stage integrated into the pump string and driven by the same (Figure 8.128) is designed as mechanical centrifugal separator. The separated air escapes axially and is again separated from tiny

Figure 8.128 Pump string oil system.

Internal Combustion Engine Handbook | 373

6606_Book.indb 373

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

oil droplets in a small cyclone separator. The mostly “cleaned” oil flow escaping radially from the centrifuge is also fed into a cyclone for further fine air separation. The degassed oil is now sent through a system of slug plates into the dry sump. The separated airflows back into the air box. Throttle cross-sections are arranged at the cylinder heads to provide ventilation to set the required vacuum in the suction circuit. The basic concept of the cylinder head coolant circuit is designed as transversal flow cooling. The water transport from the intake into the cylinder jacket and the drain from the cylinder heads is integrated in the cylinder bank V according to the central demands of the specification. The throughput cross-sections and the subsequent water transport and water jacket design allow intensive cooling of the critical areas in the cylinder head. The spark plug area and the cooling between the outlet ducts in particular must be mentioned. The circuit in the racing motorcycle closes over the radiator and the water pump which is designed as an axial/radial wheel with three-dimensional vane geometry. The design value of the water pump has been specified with 180 l/min and 2 bar delivery pressure at rated speed. The turbo-engine was optimized for efficiency for speeds in the driving range. The particular demands relative to power boosting and yield in racing resulted in the arrangement of two injection nozzles per cylinder (Figure 8.129). The cylinder-selective, delivery-specific adjustment of the mixture composition enables greatest possible freedom in the overall tuning and thus a high potential for meeting the drivability criteria which, apart from an “harmonious” full-load area, is essentially defined by dynamic charge changes in full-load operation.

The electromagnetically controlled injection valves by Magneti Marelly in the intake duct are mostly adjusted for their function in idle and partial load situations. They enable a faster adjustment of the mixture composition in chargechange processes and in dynamic operation. The “top feed” injection nozzles above the intake funnels homogenize the fuel/air mixture in full load and ensure efficiency-increasing combustions and rising power yield. A graphic representation of the fuel system in Figure 8.130 shows that the filtered fuel is presupplied with defined initial pressure by an electric supply pump and an intermediate tank to the mechanical high-pressure pump. This very compact gear pump is driven by the exhaust camshaft of the rear cylinder bank. The pressure control valve regulates the injection pressure to approximately 12 bar in the current stage of development. The engine management has been developed in collaboration with McLaren Electronics and is specially adapted for racing-oriented demands. In addition to the standard tasks of a map-controlled injection and ignition control, the ECU also assumes data recording, traction control and regulating strategies for the electronic pneumatic control system. The crankshaft throw of 360° crankshaft angle and 75° V angle favor the system of two separately routed two-in-one exhaust systems (Figure 8.131). The advantage of this exhaust routing is, on the one hand, the achievement of a compact package at the motorcycle, and, on the other hand, fluid-mechanical and gas-dynamic criteria. The symmetry of the firing sequence per cylinder bank significantly simplifies an adjustment of the efficiency-optimized overall system.

Figure 8.129 Throttle valve body with injection bars.

374 | Internal Combustion Engine Handbook

6606_Book.indb 374

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

Fuel circuit

Throttle Catch tank (0,5 bar presupply pressure) Presupply pump

“Top feed” injection nozzles

p Niede rdrucksensor

Filter Tank

Mech. delivery pump

Fine filter

Intake line Low-pressure sensor High-pressure line Bypass

p

Pressure sensor

T

Temperature sensor

High-pressure line Intake manifold injection nozzles

Figure 8.130 Concept representation of the fuel system.

Figure 8.132 Cassette-type transmission with bearing shield.

Figure 8.131 Exhaust system in a racing motorcycle.

Because the currently applicable sound pressure level limits of 130 dBA do not require silencing, very low-pressure losses are attainable. The clutch is designed as dry disk clutch. The “dry” working excludes an additional rise in temperature and oil contamination by abraded material. The transmission unit dimensioned as a replaceable cassettetype transmission features a sequential six-gear mechanism. Unlike traditional motorcycle transmissions, the shifting elements are now located at the output shaft. The shift forks engage in the driven sift sleeves and ensure the positive connection of the individual gear steps. All gearwheels are placed on needle bearings, enabling a reduction of moved masses for the shifting action and thus obtaining reduced shifting times (Figure 8.132).

8.3.5 Special Applications 8.3.5.1 Snowmobile One of the oldest special designs for leisure applications are the engine-driven sleds or snowmobiles invented in the 1920s in the United States, often called Ski-Doos which became a generic term. In 1922, Joseph-Armand Bombardier built the first “ski dog” snowmobile which became “ski-doo” due to a typographical mistake. They were motorized mostly with two-stroke engines because their power-to-weight is unsurpassed until the present time. Similar to motorcycles, these machines also have several segments and applications. The most important segment is the “mountain segment.” As the name indicates, this sled is used in the mountains and in deep snow in particular, and the weight lying on the front runners is a decisive factor. Engine outputs of 120 PS (88 kW) and more offered in this segment.

Internal Combustion Engine Handbook | 375

6606_Book.indb 375

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

The so-called Laker is a sled for long straight paths at the highest speeds possible, which makes it perfect for driving on frozen lakes with their smooth surfaces. Speeds above 200 km/h at powers of 150 PS (110 kW) and more are possible. The Utility segment is used in various areas such as forestry, lift and restaurant supply in winter sport centers, transport of accident victims from remote areas and much more. Engines used in this segment can be from 40 PS (30 kW) air-cooled two-strokes to 100 PS (75 kW) water-cooled four-stroke types. 8.3.5.2 Jet Skis or PWC (Personal Water Craft) Jet skis are relatively small watercrafts made from fiber-glass reinforced plastic without board walls for the transport of one (standing or seated) person or two to four persons (seated) on inland and coastal waters. The watercraft is driven by a combustion engine. The craft propulsion and control is realized with a water jet drive, the so-called jet pump; the engine does not feature a transmission for manual shifting. Jet skis are heavily motorized (some of them up to 164 kW), are very agile and can reach high speeds (up to 140 km/h). In the early days, they were driven by a derivation of the sled engine, as was the Sea-Doo by BRP/Rotax. The toughened exhaust emission regulation has led to the development of a four-stroke engine in this application as well. The R-1503 boat engine originates in the Rotax 4-TEC series. Engines of this series are used in various leisure segments such as watercraft, motorized sleds, all-terrain vehicles and motorcycles. They have a multitude of identical parts and identical technologies. The 4-TEC engines are light-weight, powerful drive aggregates with short-stroke design, four-valve technology, fluid cooling, gasoline injection and innovative technical details that are customized for each market segment. The R-1503 engine is a three-cylinder engine with 1,500 cm3 displacement. It meets all exhaust emission and noise regulations valid for boats worldwide since 2006. 8.3.5.2.1 Engine concept The engines of riding watercrafts are longitudinally installed in the craft interior. The crankshaft directly drives the jet pump of the craft drive over an output shaft (Figure 8.133).

Due to the dynamic driving of the craft, extreme listing of up to 45° to the side but also the stern and bow, which means that the demands on the oil circuit significantly differ for those in road vehicles. The oil system of the R-1503 engine was designed with this in mind. Even the possibility of a craft rollover must be taken into consideration. The engine remains oil-tight and can be subsequently restarted. Furthermore, water (up to 70 liters) will penetrate the craft interior through the ventilation ducts due to the dynamic running behavior, which must be taken into account when designing the discharge system and the piston-cylinder path pairing. The entire engine is designed for marine application (plug, engine control unit, bolts, cast aluminum materials). The engine package follows the given ergonomic contour of the craft. For this reason, the cylinders are slanted by 19° to the exhaust side. The engine power build-up follows the characteristic of the jet pump for craft drive. The maximum speed is 7300/min and maximum torque is supplied only at 7000/min to match the jet drive. Other demands on the engine resulted from Coast Guard regulations. For example, the surface temperature, including that of the exhaust system, must be maximal 90°C resulting in a “wet” exhaust system. Figure 8.134 summarizes the most important geometrical values of the engine. Figure 8.135 shows the longitudinal and transversal cross-section of the engine. The design features below characterize the engine: •• cylinder head, crankcase top and exhaust manifold from aluminum •• SOHC camshaft with drive over single-roller chain, four valves per cylinder •• actuation with aluminum rocker roller with hydraulic valve clearance compensation •• die-cast crankcase top with cast-in rough cast liners •• lost-foam cast crankcase bottom with integrated dry sump oil cycle •• forged steel conrod with fracture-split large conrod eye •• weight-optimized piston with low fire land height, fitted with three piston rings •• balance shaft to compensate the first-order moment excitations •• plastic intake pipe with integrated flame lock •• oil separator module from aluminum with integrated bilge pump, oil filler plug and blow-by shut-off valve •• freshwater-cooled exhaust manifold.

Figure 8.133 The outer contour of the engine is determined by the craft.

8.3.5.2.1.1 Cylinder crankcase The cylinder crankcase is manufactured from aluminum in two parts. The separating plane between top and bottom part lays on the crankshaft center. The top part in open-deck design is manufactured from AlSi9Cu3 in a die cast process. To guarantee process-reliable castability, it was important to integrate as few functions as possible. Contrarily, a lostfoam process was selected for the bottom part, enabling the integration of the oil duct channels, the dry sump oil tank and the partition of the crank chambers.

376 | Internal Combustion Engine Handbook

6606_Book.indb 376

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

Displacement Bore Stroke Stroke-bore ratio Cylinder center distance Conrod length Compression ratio Valve train Valve angle/inlet/outlet Valve diameter inlet/outlet Maximum intake valve stroke Inlet opens Inlet closes Max. exhaust valve stroke Outlet opens Outlet closes Valve overlap

[cm3] [mm] [mm] [-] [mm] [mm] [-] [°] [mm] [mm] [°KW vor OT] [°KW nach UT] [mm] [°KW vor OT] [°KW nach UT] [°KW]

1,5 l-Engine

1439,8 100 63,4 0,634 110 120 10,5 4V SOHC, Rocker roller 17/18 38/31 10 10 45 9,4 50 5 15

Figure 8.134 R-1503 engine data.

Figure 8.135 R-1503 longitudinal and transversal section.

Internal Combustion Engine Handbook | 377

6606_Book.indb 377

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

The lost-foam process also enables optimized wall crosssections for weight reduction, minimizing machining and assembly tasks and a precise and constant quality of the cast components. To obtain the required strength, AlSi10Mg(Cu) was used as material (Figure 8.136).

Figure 8.136 Top part of the cylinder crankcase.

8.3.5.2.1.2 Cylinder head The four-valve cylinder head is manufactured from an aluminum-silicon alloy in gravity die casting. To homogenize the material structure and to increase rigidity, the cast component is subjected to subsequent heat treatment. The single-art cylinder head can be very cost-effective when being manufactured thanks to the camshaft design for longitudinal insertion (Figure 8.137). The designers ensured a minimum of sealing joints and leak potential for the entire engine design. The chain well is typical for this as it avoids three-ways T-joints between cylinder, cylinder head and timing drive cover (Figure 8.138).

Figure 8.138 Timing drive end of the cylinder head.

8.3.5.2.1.3 Valve train The valve train concept shown enables an identical-part concept for single- and multi-cylinder engines and a maintenance-free valve gear for the entire engine service life at low height (Figure 8.139).

Figure 8.137 Short building cylinder head.

Figure 8.139 Compact valve train. See color section page 1080.

378 | Internal Combustion Engine Handbook

6606_Book.indb 378

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

The camshaft is made from C53F in a forging process and longitudinally inserted into the cylinder head. Two rocker rollers actuate the valves for each cylinder for the exhaust valve and a forked rocker roller is used for both intake valves For reasons of weight and costs, the rocker arms are made from AlSi11CU2(Fe) in a die casting process. Hydraulic compensators arranged at the valve-side end of the rocker arms are used to offset the valve play. The cams are inductively hardened, their design being realized by a proprietary calculation system. The ITI-SIM simulation program has been used for calculating the dynamic behavior of the valve springs. The valve stroke progression was optimized for the contract force between camshaft and the roller of the rocker arm and the contact speed. The shaft diameter of the weight-optimized intake and exhaust valves is 6 mm. For each valve, two valve springs are used to avoid valve chatter up to a total speed of 8500/min. The design has been experimentally verified at the component test stand by using a laser vibrometer to measure the valve movement. Nimoni valves are used to deal with the high thermal stresses on the exhaust valves. The rocker arm bearings are on a common axle which is, pre-assembled with the rocker arms, placed from above on the cylinder head and bolted four times. This enables a simple and process-reliable preassembly of the cylinder head. The hydraulic compensators and the camshaft bearings are supplied with oil over the rocker arm axis. 8.3.5.2.1.4 Cooling system The engine features a closed coolant circuit with thermostat control and an open cooling circuit for the exhaust system. The coolant flow for the closed coolant circuit is provided by a water pump installed at the front side of the timing drive cover and driven by the balance shaft running at engine speed.

Coolant tank

The water pump housing made from plastic bears the 87°C thermostat and all house connections. The water pump transport the coolant through a duct enclosed in the timing drive covers into the cylinder crankcase. The primary feed is realized at the engine intake side; an additional channel at the exhaust side enables an auxiliary fluid feed. The primary flow enters the cylinder head at the exhaust side of the cylinder crankcase through calibrated bores in the metal head gasket. To ensure efficient cooling in the cylinder head at a highspecific engine power, the cross-flow principle is applied. The coolant is collected in a cast-in channel above the intake channels and sent to the thermostatic valve. Besides temperature, pressure and volumetric flow measurements, optical analysis with a transparent rapid prototyping head was used in the experimental development of the cooling water flow design. With the thermostatic valve closed, the coolant is fed directly to the intake side of the water pump. When the switching temperature is reached, an increasing volume of the coolant flow is returned to the water pump over the patented water/ water heat exchanger installed in the craft’s hull (Figure 8.140). The arrangement of the engine in the sealed hull and the high-specific engine power require an oil/water heat exchanger which is installed at the bottom part of the crankcase and supplied with coolant from a bypass. In order that the surface temperature permitted for watercraft of maximum 90°C is not exceeded, a water jacket is required for the exhaust system up to the first silencer. Fresh water flows through this cooling water jacket (open coolant circuit), forced by the backpressure built in the craft’s jet pump. A portion of the freshwater is additionally injected into the first silencer to cool the exhaust gases. 8.3.5.2.1.5 Oil circuit and crankcase ventilation The operating conditions for watercraft contain listing positions of 45° around all axes at full engine power and rollovers.

Water temperature sensor activates monitoring beeper and limp home mode when temperature exceeds 100 °C (212 °F)

Blend hose from cylinder head to coolant tank Waterpump housing including thermostat (operates at 87 °C/188 °F) and waterpump impeller Ride plate (operates as radiator)

Water flows to ride plate

Water flows to oil cooler

Oil cooler Water return from oil cooler Water return from ride plate

Figure 8.140 Closed coolant circuit. See color section page 1081.

Internal Combustion Engine Handbook | 379

6606_Book.indb 379

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Neither event must damage the engine and cause oil to escape. For these conditions, a special dry sump oil circuit has been developed, with the dry sump oil tank integrated in the bottom part of the crankcase. The crank chambers of the individual cylinders are fully sealed and connected with the oil tank by a cast-in return duct (Figure 8.141).

and to the camshaft bearings. Blow-by gases are fed into the oil tank together with the return oil from the crank chambers. From there, they enter a specially developed and patented oil separator module (TOPS—Tip Over Protection System). This module comprises a cyclone separator, the oil filler plug and an electromagnetic two-way valve for the blow-by channel. In a currentless state, at engine standstill or rollover, this valve closes all lines from the oil tank and the blow-by channel to the airbox. This prevents oil from escaping into the intake system and from the tank into the cylinder head and timing drive space. In the event of a defective valve, a spring-loaded valve disk ensure the functioning of the oil circuit. Downstream of the oil separator, the blow-by gas is led into the intake tract and a bilge pump sucks the separated oil via a throttle from the cyclone separator to return it into the tank. The entire system was adjusted at the gravitational balancing test stand and a special rollover test stand. 8.3.5.2.1.6 Intake system The following criteria were taken into account in the development of the intake system: •• minimum pressure loss to achieve the required power

Figure 8.141 Sealed crank chamber for oil return.

The oil return from the crank chambers into the oil tank is forced by the pump effect of the pistons during upward stroke. The returning oil from the cylinder head is fed over the chain case into a sealed space in the timing drive cover and from there returned to the oil tank by means of a bilge pump. The intake point of the oil pressure pump is located centrally in the oil tank. The oil pressure pump itself is installed in the timing drive cover and driven by the balance shaft. The oil filter is also placed in the timing drive cover and can be easily reached from the craft seat opening. The jet nozzles for piston cooling and the crankshaft and balance shaft bearings are supplied by the main oil passage. A throttled riser leads to the hydraulically dampened chain tensioner, the rocker axle, the hydraulic compensators in the rocker arms

•• avoidance of water intake from the craft interior, even in the event of multiple “rollovers,” of flames exiting in the craft interior according to US Coast Guard regulations. •• integration of flame arrestor in the intake collector (plenum) to avoid •• compliance with statutory noise regulations and optimization of the subjectively sensed quality of the intake sound in respect to the “sporty” criterion. The air box is located in the bow area of the craft below the steering bar. From there, air is taken under the intake cover serving as splash water protection. The patented two-chamber system of the air box was optimized for acoustics and water separation (Figure 8.142). For this purpose, the constant intake of water and a single intake of a larger water volume (gushing) was simulated in two standardized test procedures. Thanks to

System with integrated water separator

Figure 8.142 Intake system with integrated water separator. See color section page 1081.

380 | Internal Combustion Engine Handbook

6606_Book.indb 380

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.3 Motorcycle Engines/Special Engines

separating plates, the reliable separation of 2 liters of gushing waters was achieved at fully opened throttle valve. Separated water is led through drain valves into the craft interior. From the air box, the intake air travels through a connecting hose into the throttle valve located at the side of the intake collector which is realized as a double-shell friction-welded unit. Fin-shaped wire-knit grids and a perforated support plate are used to avoid a flame distribution in the intake system (flame arrestor) and promote the atomization of the water sucked in. 8.3.5.2.1.7 Timing drive space The timing drive space contains these components and functions: •• camshaft drive via bushed roller chains •• balance shaft drive via straight-toothed gears •• drive of the pressure oil and water pump from the balance shaft •• starter crown ring with starter gear •• AC generator •• trigger wheel for the engine control with inductive sensor •• crankshaft drive with serration toothing and movable output drive bearing and sealing unit •• pressure oil pump and oil filter •• extraction point of the de-pressurized oil from the cylinder head and the timing drive space. The demand for an integration of all components represents a large challenge for designers but allows for an extremely short engine form. The movable output drive bearing for the output shaft permitting sways of up to ±5° in particular must be mentioned. The low-maintenance requirements of the serration toothing on the output drive are ensured by the provision of lubricating oil. An integrated worm conveyor on the output shaft bushing prevents oil build-up in the sealing units and minimizes so the risk of escaping oil. 8.3.5.2.1.8 Electronic engine control To meet future exhaust gas regulations, reduce fuel consumption, meet specific performance requirements and optimal engine response over the entire speed and charge area, a compact engine controller was developed in collaboration with Continental Automotive (formerly Siemens VDO Automotive). The engine control system’s characteristics are as follows:

All cylinder-specific functions are processed crankshaftsynchronously. All input variables are recorded by highly integrated electronic modules. The external components are actuated by integrated high-power output transformers. The functionality of these engine controls meets and, in individual areas, exceeds the Automotive Standard. The monitoring of the tip over protection system (TOPS) represents a specific marine requirement: The engine control unit processes the information provided by a craft position sensor and, in the event of a rollover, closes the rollover valve to prevent oil from escaping. The engine control system also monitors the function of the electric components and ensures reliable running by activating an emergency mode in the event of a sensor signal failure. The engine control unit and the entire engine wiring harness are premounted on the engine and the engine can thus be supplied in a ready and fully tested state for installation in the craft. A single plug provides the connection to the craft’s electrical system. 8.3.5.2.1.9 Engine acoustics The acoustic requirements relative to subjective sound impression and legal provisions for watercraft (pass-by test at maximum speed or 70 km/h) have been taken into consideration from the early stages of the engine development. The selection of a three-cylinder inline engine supports a sporty sound character by formation of 1.5th engine orders. By arranging the generator and the camshaft and balance shaft drives at the output side, these components act as oscillating masses and reduce the introduction of rotational oscillation into the craft’s drive. Particular importance was given to the structural rigidity of the components. For example, by integrating the oil tank into the crankcase bottom part, a very stiff engine was created. The rigidity of large-surface add-on components such as the plastic intake collector was also optimized by calculation or experimentally with a laser scanning vibrometer (Figure 8.143).

•• 16-bit μ -controller •• flash memory •• CAN and K-line communication interface •• active anti-knock control •• sequential multi-point fuel injection •• cylinder-sequential calculation of the pre-ignition •• active idle control •• charge calculation from combining intake vacuum and throttle valve position •• integrated engine immobilizer •• start monitoring (avoidance of a starting process at running engine) •• electronic speed monitoring.

Figure 8.143 Surface velocity of the plastic intake collector. See color section page 1081.

Internal Combustion Engine Handbook | 381

6606_Book.indb 381

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

Torque R-1503 Torque 2-stroke drive Output R-1503 Output 2-stroke drive

0

1000

2000

3000 4000 5000 6000 Rotational Speed [1/min]

7000

Thanks to these measures, the engine achieves sound levels of 106 dB(A) at 6000/min and 110 dB(A) at 7300/min. This means the sound level is approximately 10 dB(A) lower than two-cycle marine engines in the market and is in the range of good-quality four-cylinder car engines. Figure 8.144 shows a comparison with the most powerful two-stroke Sea-Doo drive regarding the torque and output progression. With the R-1503, the Sea-Doo GTX 4-Tec achieves a top speed of 90 km/h and acceleration of 0–61. (200 ft) in 4.9 seconds.

8.4 Rotary Piston Engine/ Wankel Engine 8.4.1 History

The history of the rotary piston engine is inseparably associated with Felix Wankel, which is the reason for the alternative Wankel engine designation. Wankel was born on 13 August 1902 in Lahr, Baden, and was fascinated by machines for his entire life without ever having had any technical training. Wankel was not a scientist thinking in abstract patterns but a tinkerer with a very limited relationship with mathematics: “The formulas are irritating.” Nevertheless, Wankel became the father of the rotary piston engine. In 1954, the first four-cycle engine with rotating pistons intended for installation in a vehicle was created. The Wankel design was a charge blower for a two-stroke engine with 50 ccm displacement which debuted in 1956 with a world record: The NSU engine drove the “flying lawn chair”, a cigar shape on two wheels, to a top speed of 196 km/h. In 1957, the first rotary piston combustion engine ran in the laboratory and was celebrated as a revolutionary development in professional circles. The DKM 54 test engine developed by Wankel in collaboration with NSU ran evenly and for some minutes in February 1957. Following some design changes towards the end of 1957, the 250 ccm engine outputs 29 PS at

8000

9000

10000

Figure 8.144 R-1503 torque and output progression, compared to two-stroke drive.

17,000 rpm and, for a short time, even 22,000 rpm are measured. Four of these engines were built, and one of them can be seen in the Deutsche Museum in Munich, Germany. Together with the entrepreneur Ernst Hutzenlump, Wankel founded the Patentverwaltungsgesellschaft Wankel GmbH, a patent management company. This means that Wankel is one of the very few inventors who can actually live carefree thanks to license revenues. In 1958, Curtis Wright, an American aircraft engine manufacturer, joined Wankel and built rotary piston aircraft engines under license. The first passenger cars with rotary piston engines appeared in 1960 on German roads as NSU “test princes.” In 1963, the first Wankel series car, an NSU Spider celebrated its premiere at the International Car Exhibition in Frankfurt, Germany. With a chamber volume of 500 ccm, its rotary piston engine generated 37 kW or 50 PS. One year later, the engine is mass-produced. In 1967, the Mazda Cosmo is the first Wankel car with double-disk engine. In 1968, NSU builds the Ro 80 with double-disk engine, 1-liter chamber volume and 81 kW/110 PS. The front-wheeldriven sedan with a top speed of 180 km/h runs extremely smoothly but is very susceptible to failure. In the early 1970s, potential licensees are waiting in line at Wankel. Wankel concludes agreements with Daimler Benz and VW, Rolls Royce and Porsche, General Motors and Ford, Nissan, Mazda and Yamaha, Toyota, American Motors, Krupp and all major motorcycle manufacturers. The licensing revenues are considerable. Trouble arises in 1974. Although the problems with “chatter marks” at the housing interior surfaces and those with the apex seals are resolved, the expectation that the rotary piston engine can be produced at lower costs than the reciprocating piston engine is not met. Rising fuel costs during the first energy crisis and tightened exhaust emission regulations in the US halt the further development of the Wankel engine. General Motors and Daimler-Benz abandon advanced Wankel projects. Peugeot stops the bi-rotor production of their subsidiary Citroën in 1975 which had been started only in 1974. Two

382 | Internal Combustion Engine Handbook

6606_Book.indb 382

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.4 Rotary Piston Engine/Wankel Engine

years later, Audi terminates the production of the Ro 80 that they had assumed from NSU. Of all original licensees, only Mazda installs a now mature rotary piston engine in the RX-7 sport coupe. As the only motorcycle manufacturer, the British company Norton continues to use the Wankel principle in police motorcycles for the domestic market. But Wankel does not just build for the automotive and motorcycle sector. In 1976, rotary piston engine with 220 kW/299 PS drives the “Zisch” motorboat across Lake Constance at more than 100 km/h. In 1978, Wankel succeeds with sealing the innovative DKM 78 two-stroke rotating piston engine which, compared to traditional four-stroke rotary piston engine (KKM), has a significantly higher output at lower consumption in a smaller construction volume. On October 9, 1988, Dr. h.c. Felix Wankel dies after a long illness in Heidelberg, Germany. Mazda Motor Corporation assures that they will continue to build engines without valves and connecting rods according to the Wankel principle. This promise has been kept by the Japanese manufacturers. Since 1961, Mazda has built more than two million rotary piston engines—most of them for the RX-7. The modern Renesis engine drives the Mazda RX-8.

8.4.2 General Functionality of a Rotary Piston Engine

The rotary piston engine fundamentally differs in its functionality from all conventional combustion engines. In traditional reciprocating piston engines, a translational movement is converted into a rotational movement at the crankshaft. The combustion chamber is located at one end and the crankshaft at the other. The up and down movements and the rotations of the crankshaft generate strong oscillations which must be balanced with a mass flywheel. Another disadvantage is the many movable components of a reciprocating piston engine, which are subject to heavy stresses and thus high wear. The rotary piston engine does not have these disadvantages. The piston also called a rotor is triangle-shaped in the Wankel engine, with three convex sides of equal length. At the three vertices and the flanks, that is, all contact surfaces, the rotor is sealed against the housing so that gas cannot flow from one working chamber to the next. Sealing elements are inserted in the three corner edges and the side surface of the piston. The sealing of the Wankel engines represented a difficult problem for the longest time. A series of measures has now eliminated all leakages. Sealing bolts designed as short cylindrical components under laid with small washers converge at the ends of the apex seals with the lateral sealing strips. The rotary piston engine is an inner-axis system with an axis-parallel position of the rotary axes of two rotating bodies of revolution. The piston rotates in the stator, an oval housing that is slightly reduced in the center. During piston rotation, the three corner are constantly touching the housing wall, which means the piston center point describes a closed circle during the rotation.

The epitrochoid that is the basis the rotary piston engine can be generated in different ways. For example, it is generated by rolling one circle on another circle with doubled radius For this purpose, a point selected within the rolled circle is constantly marked. The radius of the basic circle corresponds to the distance from the center point of the rotating piston to one of its corners (generating radius = R). The distance of the selected point (curve-generating point) from the center point of the rolled circle corresponds to the eccentricity. If the rolling circle rolls within the basic circle, a hypotrochoid is generated. If the point lays on the circumference of the rolling circle, epi- or hypocycloids are generated. The rolling circle may also hang above the basic circle, similar to an internally toothed annulus above an externally toothed smaller gear, and can thus be compared to the internal-axis principle of a rotary piston engine. The trochoid actually generated in the engine does not correspond to the mathematically calculated curve, however. It is shifted by a small value to the outside for the sealing strips to be able to follow the trochoid contour with less wear. The dimension for the equidistant corresponds to the radius of the rounded strip apex. The rotor moves eccentrically in the housing so that the three corners of the rotor always follow the housing wall at every rotation. In the rotor itself, an annulus with internal toothing is installed that rolls down on a toothed gear attached to the engine housing side. This toothing is necessary for the rotor to be constantly supported by its internal toothing on the fixed toothed gear during the rotation and simultaneous exercising a rotational movement on the eccentric shaft. Between the three flanks of the rotor and the inner housing surface, three working chambers are thus created with constantly changing volumes during one rotor rotation. This function eliminates a crankshaft and valves; the only moving parts are the rotary piston and the eccentric shaft. These characteristics result in the low weight and small dimension of the Wankel engine. The rotor is the power generating and the eccentric shaft is the power-releasing component in a rotary piston engine. The eccentric shaft is comparable to the crankshaft of the SI engine. Piston annulus and fixed pinion have a tooth ratio of 3:2, that is, the piston rotates at two-third of the angular velocity of the eccentric shaft. At a two-disk Wankel engine, the eccentric shafts offset by 180 degrees cause better running smoothness than the single-piston version. A three-disk rotary piston engine has a similar running smoothness as an eightcylinder reciprocating piston engine. By arranging several engine cells in line, large outputs in small engine dimensions can be realized with little construction effort. While a standard four-stroke needs two up and down movements for a single work cycle, the rotary piston engine completes all four cycles in a single rotation of the rotary piston. Almost no unbalance forces occur because the center of gravity of the piston moves at little distance around the rotary axis and the piston is thus dynamically balanced.

Internal Combustion Engine Handbook | 383

6606_Book.indb 383

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

8.4.3 Four-Stroke Principle

The operation of a rotary piston engine corresponds to the four-stroke SI engine principle. Hollow spaces are generated because the three rotor corners always touch the stator. During the piston revolution, its three edges, together with the housing wall form three chambers (A, B, C) with variable volumes, in each of which a complete four-cycle process is executed during one piston revolution, that is, intake, compression, ignition and exhaust. The intake and exhaust opening are shaped as slots and are opened and closed by the piston itself during the rotation. Due to the superimposed circular and rotational piston movements, the sickle-shaped chambers change their volumes. Thus, three of four work cycles are always executed simultaneously in the three chambers, and after each complete piston rotation, the engine has completed three times the entire four-cycle SI engine process.

thus be opened significantly earlier than in previous designs (Figure 8.147). In return, the exhaust openings with nearly double the size are released later and with lower flow resistance resulting in a longer exhaust cycle and a significantly improved thermal balance.

1. Cycle (intake) As soon as a rotor corner releases the intake slot by sweeping past, the gasoline–air mixture flows into the next chamber, and the chamber volume increases because of the rotor movement. 2. Cycle (compression) With the continued rotor revolution, the volume of the chamber containing the mixture decreases and compresses the gasoline-air mixture. 3. Cycle (ignition) The compressed mixture is ignited. During the combustion, the fuel–air mixture expands and rotates the piston which then drives the eccentric shaft. 4. Cycle (exhaustion) The first sealing strip of the rotor sweeps past the exhaust slot to release it. This working cycle occurs simultaneously in all chambers.

8.4.4 The Rotary Piston Engine of the Passenger Car Renesis

Despite its well-known advantages, only the Japanese manufacturer Mazda retained the rotary piston engine principle in the automotive sector. This modern crank gear is used in the RX-8 sports car and is called the “Renesis,” (Figure 8.145(a) and (b)). The name is composed from the abbreviation for RE (rotary engine) and “genesis” and is intended to demonstrate that Mazda has redesigned and revolutionized the familiar design forms of the rotary piston engine. The Renesis is a further developed version of the MSP-RE (Multi-Side-Port Rotary Engine) which Mazda presented for the first time in the concept sports car RX-01 during the Tokyo Motor Show in 1995. The Renesis fundamentally differs in essential design characteristics from traditional rotary piston engines. The exhaust channels usually attached on the trochoid housing in conventional rotary piston engines, are located in the stator’s side wall, (Figure 8.146). This arrangement avoids the undesirable overlap of the openings of intake and exhaust channels and thus considerably increases efficiency. Furthermore, the intake openings are larger by 30% and can

Figure 8.145(a) Mazda Renesis cutaway model Unlike conventional rotary piston engines, the Renesis features lateral intakes and exhausts. Thus, three ignitions occur with every full revolution of the piston. This means that the torque progression of a rotary piston engine is significantly more uniform than a single-cylinder SI engine in which only one ignition for every two crankshaft revolutions occurs. Renesis engine in the RX-8 Variant

STD Power

Type

Rotary piston, 2 rotors

High Power

Displacement

654 cm3 per rotor

Mixture preparation

Electromagnetic pump

Compression ratio

10:1

Ignition system

Fully-electronic

Maximum output

141 kW at 7000 min–1

170 kW at 8200 min–1

Maximum torque

220 kW at 5000 min–1

211 kW at 5500 min–1

Fuel

Lead-free 95 RON

Emissions standard

Euro IV

Cooling system

Water-cooled

Figure 8.145(b) Technical data of Renesis engines in Mazda RX-8.

Figure 8.146 Exhaust channel arrangement.

384 | Internal Combustion Engine Handbook

6606_Book.indb 384

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.4 Rotary Piston Engine/Wankel Engine

30 % increase

Increased Intake Port Area

Figure 8.147 Intake openings.

8.4.4.1 Lateral Exhaust Mazda uses exhausts that are laterally placed in the housing. Neither intake nor exhaust ducts are led through the jacket, that is, from the periphery, in the Renesis engine. The advantage of the peripheral intake, the high output, is compensated by its disadvantages, the large overlap and the deceleration jerking. The advantages of the lateral exhaust, that is, no overlap of intake and exhaust, no deceleration jerking, improved mixture preparation and easier oil sealing of the rotors are the opposing arguments. The disadvantage of the lower filling level can be compensated by an exact design of the intake paths and slot parameters. Hanns–Dieter Paschke, an engineer with NSU, remarked on the advantages of a lateral exhaust technology in the late 1950s. The series implementation of a lateral exhaust technology has been made possible only by the use of ceramic port liners. Port liners are ceramic inserts which are positioned in the mold in the foundry and which are cast around by liquid aluminum. The base material is aluminum titanate, the stoichiometric mixed phase of aluminum oxide and titanium oxide. Important properties of the material are its low thermal conductivity, very low heat expansion coefficient and the related very high thermal resistance and open porosity. This porosity is of essential importance for application in engine manufacturing. It results from a special effect during the cooling down period when critical internal stresses are generated, causing the formation of microscopically small cracks. During warming up, some of these cracks in the material heal themselves. Port liners are used to line the lateral exhaust ducts in the Renesis engine. Because ceramic is a poor heat conductor, only a fraction of the exhaust heat is emitted into the lateral components. On the contrary, most of it is transported into the externally placed exhaust duct. This idea sounds simple but until now, it had not been possible to process heat-stressed ceramic components in iron casting so that they would be fully seated in position over the course of an engine’s lifespan. Mazda has overcome this problem. The stator running surface has been made from a chromium molybdenum alloy. Molybdenum is one of the few materials not generating chatter marks. In the early years in particular, these grooved wear signs on the running surface caused by the

friction vibrations of the apex seals used to be a symptomatic weakness of the rotary piston engine. The jacket itself is cast from aluminum. The wall behind the running surface is designed thick-walled to the cooling water carrying recesses. Along with massive ribbing, this results in excellent rigidity in longitudinal direction. The running surface is interrupted only by four small holes: two for the spark plugs and two for the oil supply to the sealing strips. The oil supply has two functions: It is essential for lubrication and, at the same time, ensures tightness. The oil for the lubrication of the corner edge seals of the piston is directly applied to the inner walls of the combustion chamber. Thanks to the decision to use short oil paths and selection of suitable nozzles, the Renesis consumes only approximately half as much oil as conventional rotary piston engines. The top spark plug ignites through a firing channel, (Figure 8.148). The lower plug is placed in the compression compensation point with nearly equal pressure between the individual chambers and does not require a firing channel.

Figure 8.148 Section through the engine block.

8.4.4.2 Variable Intake Control and Electronic Throttle Valve Rotary piston engines are usually fitted with one exhaust duct each on the outside of the trochoid housing. The Renesis, however, is fitted with two lateral exhausts per rotor, each featuring a diameter double the size of a conventional exhaust. This configuration not only improves the exhaust gas flow but also allows the delayed opening of the exhaust duct. The intake openings are released earlier than before while the exhaust openings are released later: The result is an extended ignition cycle and a higher thermal efficiency—both improving the consumption. The combustion chamber recess has been deepened compared to its predecessor, making the combustion chamber much more compact. In addition, the lateral exhaust in the Renesis

Internal Combustion Engine Handbook | 385

6606_Book.indb 385

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

prevents the escape of unburnt hydrocarbons from the combustion chamber into the exhaust ports. Rather, the residual gases are transported into the next combustion cycle and burned, significantly reducing emissions (Figure 8.149). The innovative fuel/oil-sealing system contains separating valves and has been specifically developed for the lateral exhaust configuration. The almost hermetic sealing greatly improves output, consumption and emission levels.

Current Peripheral Exhaust

Side Exhaust Rotor

Rotor Exhaust Port

Exhaust Port

Trochoid Unburned Gas

Trochoid

Unburned Gas

Figure 8.149 Lateral exhaust advantage.

The Renesis works with a variable 6PI intake system (six port induction) with three intake channels for each of the two rotors (Figure 8.150). An electric motor actuates the rotary-disk valves at the intake duct of each rotor which utilize the dynamics of the incoming air for charging and thus rise the filling level. Furthermore, the Renesis is fitted with an electronic throttle valve translating the commands given by the engine control unit. This enables a highly accurate and direct valve control. Finally, the newly developed intake manifold from plastic is lighter and designed for optimal flow properties to reduce air resistance and intake losses to a minimum.

The Renesis features innovative injection nozzles for ultra-fine atomization of the fuel. Small high-performance spark plugs ensure a better ignition of the mixture. This combination of ultra-fine atomization and powerful ignition results in nearly perfect combustion and, thus, directly to higher efficiency at low emissions. The double-wall exhaust manifold keeps the exhaust gas temperatures high and shortens the cold-run phase of the two-stage catalytic converter. The new flat wet sump lubricating system features and oil pan with just 40 mm depth. This is only half the dimension of a conventional rotary piston engine. One of the major advantages of the rotary piston engine is that the eccentric shaft is placed higher than crankshaft of a reciprocating piston engine, that is, above the sump, thus preventing friction losses. Furthermore, the pump losses are lower than with dry sump lubrication. The system also uses a specially shaped baffle chamber to control the oil flow and ensures that the oil does not accumulate at one side during extreme lateral acceleration. Mazda abandoned trials with dry sump construction after costs, weight and reliability analyses for the selected solution. Hence, the entire powertrain—and thus the vehicle’s center of gravity—can be placed lower reducing the moment of inertia in curves by up to 15%. The lateral exhaust technology works acoustically as well. Unlike rotary piston engines with peripheral exhaust channels, the Renesis generates clear and transparent sounds in high frequencies and a sonorous sound in lower frequencies. This means that the Renesis not only features a highly uniform power development but it also sounds as is to be expected from a sports car engine. A turbocharger may also be implemented. To a limited extent, the rotary piston concept is suited for a diesel engine. Although the highest possible compression ratio of approximately 1:12 is insufficient for an autoigniter, diesel operation is possible with charger and auxiliary ignition.

8.4.5 Hydrogen Rotary Piston Engine

Shutter Valve Figure 8.150 6PI geometry (six port induction).

The topic hydrogen as drive energy for a rotary piston engine is in the practice testing phase. NSU completed its first positive trials in the early 1970s. In 1991, the first unit was presented with the HRX engine series. Until today, the bivalent rotary piston engine Renesis Hydrogen works with hydrogen direct injection in the Mazda RX-8 Hydrogen RE (Figure 8.151). For the operation with hydrogen in particular, the rotary piston engine offers additional benefits. Unlike the conventional reciprocating piston engine, the rotary piston engine has isolated chambers for the intake and the combustion tract. This gives it an inherent structural advantage when hydrogen is used. The separate arrangement of spark plugs and injection nozzles also offers advantages: Because hydrogen gas has a very low density. the hydrogen rotary piston engine uses two injection nozzles for each rotor which directly inject the hydrogen during the intake cycle, in order to reach the hydrogen volume optimal for combustion and to obtain optimal combustion. In a standard reciprocating piston engine, this would not be possible for reasons of space, because intake and exhaust valves, injection nozzles and spark plugs have to share the

386 | Internal Combustion Engine Handbook

6606_Book.indb 386

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

8.4 Rotary Piston Engine/Wankel Engine

Hydrogen

Electronically controlled hydrogen direct injection

Peak edge

Air

Exhaust gases Spark plugs

Corner edge seal

Rotors

Electronically controlled hydrogen direct injection High-pressure hydrogen tank

Gasoline tank

Hybrid engine

Gasoline tank

MagLean burn Nox catalytic converterer

Three-ways catalytic converter

144 Volt battery

Turbocharger with e-motor support Inverter

Figure 8.151 Mazda RX-8 Hydrogen RE.

available space. In addition, unlike with the reciprocating piston engine, spontaneous combustion of hydrogen gas at still hot parts cannot happen: The compression and the combustion process happen physically separated in different chambers, and, due to the direct injection, the hydrogen is introduced into the relatively “cool” intake chamber without risk. The eccentric shaft of the rotary piston engine rotates by 270° for each work cycle, while the crankshaft of a conventional reciprocating piston engine rotates only by 180°. This promotes a thorough mixing of hydrogen and air at a simultaneously high mixture flow intensity. In a gasoline operation, the engine uses the conventional system with lateral injection nozzles.

8.2. Hadler J. 2007. TDI-Motor mit Common-Rail-Einspritzung von Volkswagen, MTZ 11/2007.

Bibliography

 oll, G., Fausten, H., Noell, R., Schommers, J., Sprengel, C., Werner, P. Der D neue V6-Dieselmotor von Mercedes-Benz, MTZ 09/2005 Jahrgang 66.

8-9. Hahne, B., Neuendorf, S., Paehr, G., Vollmers, E. Neue Dieselmotoren für Volkswagen-Nutzfahrzeug-Anwendungen, MTZ 01.2010, Wiesbaden, Januar 2010.

8.1. Königstedt, J., Der Neue, UA. 2006. V10-FSI-Motor von Audi, 27. Internationales Wiener Motorensymposium, Wien.

8-10. Eiser, A., Böhme, J., Ganz, M., Marques, M. 2010. Der neue 2,5-l-TFSIFünfzylinder für den Audi TT RS, MTZ 05.2010, Wiesbaden.

8.3. Bach, M., u.a. Der 6,0-Liter V12-DTI-Motor von Audi, Teil 1, MTZ 10/2008. 8.4. Bauder, R., u.a.: 6,0-Liter V12-DTI-Motor von Audi, Teil 2, MTZ 11/2008. 8.5. Baretzk, U., u.a. 2007. Der V12-TDI für die 24h von LeMans—Sieg einer Idee, Wiener Motorensymposium. 8-6. Schwarz, C., u.a. Die neuen Vier- und Sechszylinder-Ottomotoren von BMW mit Schichtbrennverfahren, MTZ 05/2007. 8.7. Schommers, J., u.a. Der neue Vierzylinder-Dieselmotor für Pkw von Mercedes-Benz, MTZ 12/2008. 8.8. Schommers, J., u.a. Der neue 4-Zylinder Pkw-Dieselmotor von Mercedes-Benz für weltweiten Einsatz, 17. Aachener Kolloquium, 20008, Aachen.

Internal Combustion Engine Handbook | 387

6606_Book.indb 387

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 8 Engines

8-11. Herrmann, D. 2010. Der Abtrieb des Audi A1, ATZ extra, Wiesbaden. 8-12. Brinkmann, C., Baur, G., Geywitz, G., Fronemann, J., Heubach, W., Königstedt, J., Schwarnberger, A. 2010. Die neue Generation der V8 FSI-Motoren von Audi, 5. Emission Control 2010, Dresden.

8-17. Bauder, R., Fröhlich, A., Rossi, D. 2010. Neue Generation 3,0-l-Motors von Audi, MTZ 10.2010, Wiesbaden. 8-18. Kahrstedt, J., Zülch, S., Streng, C., Riegger, R. 2010. Neue Generation des 3,0-l-TDI-Motors von Audi, MTZ 11.2010, Wiesbaden.

8-13. Lechner, B., Kiesgen, G., Kriese, J., Schopp, J. 2010. Der neue MiniMotor mit Twin-Power-Turbo, MTZ 07–08.2010, Wiesbaden.

8-19. Apfelbeck, L. 1997. Wege zum Hochleistungs-Viertaktmotor. Motorbuch Verlag, Stuttgart.

8-14. Waltner, A., Lückert, P., Doll, G., Kemmler, R. 2010. Der neue 3,5-l- Ottomotor mit Direkteinspritzung von Mercedes-Benz, MTZ 09.2010, Wiesbaden.

8-20. Hütten, H. 1998. Motorradtechnik. Motorbuch Verlag, Stuttgart.

8-15. Bick, W., Köhne, M., Pape, U., Schiffgens, H.J. 2010. Die neuen Tier-4-lMotoren von Deutz, MTZ 10.2010. Wiesbaden.

8-21. Nepromuk, B., und Janneck, U. 2006. Das Schrauberhandbuch. Delius Klasing, Bielefeld. 8-22. Stoffregen, J. 2010. Motorradtechnik. Wiesbaden: Vieweg+Teubner.

8-16. Doll, G., Waltner, A., Lückert, P., Kemmler, R. 2010. Der neue 4,6-lOttomotor von Mercedes-Benz, MTZ 10.2010, Wiesbaden.

388 | Internal Combustion Engine Handbook

6606_Book.indb 388

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9 Tribology 9.1 Friction

•• coolant pump •• alternator •• injection pump

9.1.1 Characterizing Features

The utilizable power at the output shaft of the internal combustion engine (effective power Pe) is lower than the internal power to the piston (indicated power Pi). The difference is referred to as the friction Pr

Pr = Pi − Pe

(9.1)

The frictional power includes the losses of the individual engine components, such as the crank gear (crankshaft, connecting rods, piston with piston rings), the valve train including the timing gear, and the requisite auxiliary drives. The internal power also takes into account the losses due to the charge exchange. Thereby the operating status and, consequently, the drive power of the auxiliary units are defined differently in the various standards [9-1]. The frictional power reduces the available existing engine power at the output shaft and, thereby, also influences the fuel consumption of the engine. Analogous to the effective and indicated mean pressure, the mean friction pressure pmr is utilized to compare various engines with different displacement volumes.



⎛ P − P ⎞ ⎛ Pr ⎞ pmr = pmi − pme = ⎜ i e ⎟ = ⎜ . ⎝ i ⋅ n⋅VH ⎠ ⎝ i ⋅ n ⋅VH ⎟⎠

(9.2)

The friction of a complete engine includes the frictional power and/or drive unit powers of the individual components: •• Crank gear consisting of •• crankshaft main bearing with radial shaft seal rings •• connecting rod bearings and piston group (piston, piston rings, and wristpin) •• possibly existing mass balancing systems. •• Valve train and timing gear •• Auxiliary equipment such as •• oil pump with possibly existing oil pump drive

•• radiator fan •• vacuum pump •• climate compressor •• power steering pump •• air compressor.

9.1.2 Friction States

Depending on the lubrication prevailing at the various friction points in the engine, different friction states occur. The most important are as follows: •• Solid friction (Coulomb’s friction)—friction between solids without a fluid intermediate layer •• Boundary friction—friction between solids with an applied solid lubricant layer without a fluid intermediate layer •• Mixed friction—fluid friction and solid friction or boundary friction occur simultaneously; the lubricant layer does not completely separate the two friction layers from one another, and a certain contact occurs •• Fluid friction—(hydrodynamic friction) a liquid (or gaseous) substance between the two friction layers completely separates the two from each another. In the internal combustion engine, the movement of the friction surfaces against one another creates the hydrodynamic supporting effect of the intermediate substance. The occurrence of the different friction states is explained below using an example. In a hydrodynamic plain bearing, the different friction states occur as the engine passes through the engine revolution band. The Stribeck curve in Figure 9.1 illustrates the relationship between the coefficient of friction μ and the shaft speed n or the sliding velocity v at constant temperature (or constant viscosity η ).

Internal Combustion Engine Handbook | 389

6606_Book.indb 389

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

y = Constant

Coefficient of friction µ

Friction when idling (Start-up friction)

Limiting friction

Mixed friction Fluid or viscous friction Notching point

p = Constant upper lower operating limit operating limit Sliding speed n Fluid or viscous friction Dry friction

Figure 9.1 Stribeck curve [9-2].

The overall friction is made up of the two portions of solid friction (or boundary friction) and fluid friction. The static friction exists at a standstill. Solid friction and/or boundary friction initially exist with lower revolutions, and then, the mixed friction band occurs in which the friction decreases with increasing engine revolutions with a corresponding increasing buildup of a hydrodynamic supporting film. The breakaway point in this model represents the point at which the hydrodynamic supporting film can completely separate the surface roughness of the two friction partners. The engine speed at which this state achieved is also referred to as the transitional speed at which the minimum friction occurs. At engine speeds above the transitional speed, fluid friction occurs, and the friction increases again because of the increasing shear rates. Increasing loads on the friction pair, or decreasing viscosity of the fluid, shifts the transitional speed upwards and extends the range of mixed friction. Operating states on the left-hand branch of the Stribeck curve are unstable, as a brief variation such as an increase in engine revolutions or reduction in the load leads to a significant rise in the coefficient of friction and, hence, to an automatic amplification of the fault. For this reason, the operating point of a friction pair in continuous operation must be sufficiently far away from the breakaway point on the right-hand branch of the Stribeck curve.

9.1.3 Methods of Measuring Friction

The exact calculation of the friction losses involves a great deal of work. There are various ways of determining the friction, although the majority of these exhibit significant inaccuracies. The following methods are commonly utilized for calculating the friction [9-3], [9-5]:

•• The rundown method: Here, the engine is switched off after stabilization at an operating point, and the change in the rpm is measured as a function of time. The friction moment or mean friction pressure is then calculated using the mass moment of inertia of the moving masses. •• The shutoff method: On multiple-cylinder engines, the fuel supply to one of the cylinders is shut off, and this cylinder is then dragged along by the other working cylinders. The friction loss can be determined from the change in effective engine power before, and after, the fuel shut off. •• The Willans lines: The fuel consumption of an engine is plotted on the Y-axis against the mean effective pressure pme for various engine speeds. The intersections with the negative pme axis are then determined by linear extrapolation of the values down to fuel consumption zero; these can be roughly regarded as the mean friction pressures at the respective engine speeds. •• The motoring method: The engine is motored on a test rig by an external motor. The motoring power required to drive the engine is regarded as the friction loss. With this method, either the engine can be motored at operating temperature and measured immediately after shutting off the fuel supply, or it can be conditioned via external thermostat installations. •• The strip method: Strip measurement is a special form of motoring that is utilized to measure the friction losses of the various engine components, such as the friction of the engine, the valve train, and the auxiliary drives. The designation derives from the method where the engine is dismantled (stripped) step by step on a motoring test rig. The friction losses of the individual components are determined from the difference between the measured values with, and without, these components. The total friction of the engine is obtained by addition of the values for the individual components. •• The indication method: This method can be used to determine the friction of an engine in motoring mode. The integration of the measured cylinder pressure over a working cycle supplies the indicated work Wi which, when referred to the displacement volume, gives the indicated mean pressure pmi. If the mean effective pressure pme, calculated from the torque measured at the drive shaft is subtracted from this, then we obtain the mean friction pressure pmr. •• Special measuring method: Apart from the friction measuring methods described previously, there are a large number of other processes for evaluation, for example the friction of individual components during operation. Torque measuring flanges can be utilized to execute this for shaft-driven components [9-2], [9-5]. There are various facilities for measuring piston frictional force for piston groups [9-9]. A crucial aspect for the precision and reproducibility of the individual methods and, therefore, for the comparability of various measurements is strict compliance with the boundary conditions. It is essential for all measurement processes that the lubricating oil and coolant temperatures of the engine

390 | Internal Combustion Engine Handbook

6606_Book.indb 390

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

of the piston can therefore influence the calculation of the mean indicated pressure and, thus, also of the mean friction pressure. Fundamental studies have shown that an error of just 0.1°kW in the TDC position of the crankshaft can affect the calculated mean friction pressure by more than 10%, depending on the engine load. A direct comparison of the different measurement methods is not possible, since the different boundary conditions influence the measurement results. This is illustrated as an example of a diesel engine with direct injection in Figure 9.2. The fluid temperatures have been kept the same for the complete series of tests: 90°C oil temperature in the main gallery and 90°C coolant outlet temperature. A good correlation for the whole rpm range is obtained between the results of the strip measurements and the motoring measurement (the charge cycle losses were determined by indexing and deducted). The different friction values discovered with the motored engine are attributable to the following influences: •• The lubricant film temperatures in the engine are higher, despite the same temperature in the main oil gallery. •• The combustion results in higher temperatures at the piston group and cylinder barrel. •• The lateral piston forces change due to the gas pressure. •• The load conditions of the injection pump change.

rpm variation

PMR friction mean effective pressure [bar]

are set to less than ± than 1 K. This is generally only possible with precise external thermostatic equipment. Of the possibilities described for determining pmr, the first three are already subject to significant inaccuracies from the principle of the process and are therefore only suitable for the identification of trends. With the motoring method, the problem is that the inertia moment of a complete engine includes not only the mechanical engine friction and the drive power of the auxiliary drives, but also the charge cycle losses and that, without additional indexing, no distinction can be made between the friction and the charge cycle losses. However, since the charge cycle losses react very sensitively to changes in ambient conditions on the test rig, or to minor differences in the intake and the exhaust systems, the comparability of different engines is rather restricted with this method. With the strip method, the boundary conditions can be set very accurately using external systems, so that a good reproducibility and comparability of the results can be achieved. Characteristic for the strip method is the fact that the engine is always driven via the output shaft. This has the advantage over other measuring methods, in that the boundary conditions for the components under consideration are as close as possible to the conditions in the complete engine itself, and a good transferability of the results is guaranteed. At the same time, this results in the limitation to the application of the strip method for determining the friction losses of any particular parts of the rotating engine: a functional (in the sense of the motoring operation) configuration of the engine must be possible with, and without, simultaneous movement of the parts under consideration. As a consequence, this means that the friction values measured for a component always also include the friction attributable to the drive; these are also eliminated when the components are removed. For example, the determination of the friction in the valve train also includes the friction generated in the timing belt or timing chain. This is also expedient in that the power losses can be allocated to the component in question and that the load and any dynamics affect the level of the power losses. The indexing method demands a higher measuring complexity in order to obtain reliable results. One of the greatest influences is that the individual cylinders can exhibit significant differences in their mean pressure with multiple cylinder engines. For this reason, a simultaneous pressure measurement on all the cylinders is necessary. This causes considerable measurement complexity in practice. Furthermore, the complexity is increased by the fact that even minor errors in the top dead center (TDC) positioning, and deviations in the pressure measurement from the calibration curve of the pressure sensors, can cause a significant difference in the pmi-value, and errors in the torque measurement distort the pme value. Therefore, great demands have to be made on the accuracy of the indexing and the torque measurement, as the result of the subtraction (of the mean friction pressure) is more than one power smaller than the initial parameters, so that the percentage errors are multiplied by a factor of ten. Even minor deviations in the determination of the (TDC)

3,0

Oil temperature/coolant temperature Complete engine: Crank gear, Valve train, Oil pump, Water pump, Generator, Injection pump, Power steering pump, Vacuum pump

2,5 2,0 1,5 1,0

Fired, full load Fired, zero load Towed Stripped

0,5 0,0

0

1000

2000

3000

4000

5000

Engine speed [min –1]

Figure 9.2 Comparison of different measurement processes on a passenger car diesel engine with direct injection.

9.1.4 Influence of the Operating State and the Boundary Conditions

The operating state of the engine and the boundary conditions, under which the engine is operated, have a significant influence on the frictional behavior. The most important parameters are described in the following sections. 9.1.4.1 Run-In Status of the Internal Combustion Engine In the first hours of operation, an adaptation of the friction partners takes place at the individual sliding points and therefore with a smoothing of the surface unevenness. This

Internal Combustion Engine Handbook | 391

6606_Book.indb 391

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

Variation of the oil viscosity grades at 2000 min –1

pmr [bar]

0,4 0,3

Valve train

0,2

Oil viscosity classes SAE 15W40 SAE 5W40 SAE 0W30

0,1 0,0 35/35°C 90/90°C 120°C/110°C Oil temperature/coolant temperature

pmr [bar]

2,0 1,5

Crank gear

0,5 35/35°C

4 3 2 4000 2500

Oil tempera

ture/coolant

80

100 120 temperatu re [°C]

m

1000 60

rp

40

e

20

gin

0

[1

/m

0

in]

1

Figure 9.4 Influence of the fluid temperatures on the friction [9-4].

9.1.4.4 Engine Operating Point The engine operating point influences the friction both via the parameter “engine speed” and via the load. The influence of the engine speed is attributable to the increase in the sliding speeds at the friction points of the individual engine components. Increasing engine load has the following effects: •• higher gas pressures and, thus, higher lateral piston forces, contact pressures of the piston rings, bearing loads, and forces for opening the exhaust valves •• locally higher component temperatures and, hence, a possible increase in deformation •• locally higher lubricant temperatures and, hence, a change in the friction status at the corresponding lubrication points •• possibly modified drive power of the injection pump.

1,0 0,0

5

En

9.1.4.2 Oil Viscosity The change in the shear forces means that the viscosity of the lubricant has a significant effect on the conditions at the lubrication point. Operating the internal combustion engine with lubricating oils of different viscosity results in change in the frictional status when the other boundary conditions remain unchanged. A lower viscosity of the lubricating oil means a lower load-bearing ability of the lubrication gap and, therefore, a reduction in the lubricant film thickness. This is also associated with an increase in contact between solid objects in the mixed friction zone. Depending on the boundary conditions, the friction then drops if the hydrodynamic friction portion predominates or increases if the solid contact rises sharply. The behavior of different oils with different viscosities is illustrated in Figure 9.3 for a car gasoline engine at 2000 min–1 rpm. With the boundary conditions prevailing here, a reduction in friction with decreasing oil viscosity was observed in the crank gear. In the valve train, this reduction in friction is only observed at low temperatures. At higher temperatures, on the other hand, the friction increases because of the mixed friction conditions in the valve train caused by the lower oil viscosities. This change also has additional effects on the lubrication system and the oil pump drive power, as oil pressures and oil volumetric flows in the lubrication system are influenced by the various components and by the friction of the oil pump.

9.1.4.3 Temperature Influence The operating temperature of the internal combustion engine, that is, the temperatures of the components and the oil and coolant, influences the friction. The reasons for this are, first, the change in viscosity of the lubricant and, second, the change in the clearances in the various friction pairs. The effects of the changes in the fluid temperatures in the temperature range between 0°C and 120°C are shown in Figure 9.4. Even at fluid temperatures of approximately 20°C, the friction losses are already doubled compared with an engine at an operating temperature (90°C). This is one of the reasons for the increase in fuel consumption after a cold start, and for short journeys, when the engine is not at operating temperature.

Friction mean effective pressure [bar]

process involves a certain amount of wear and increases the friction loss of the engine. Thus, the running-in process takes place at different speeds for the different friction pairs and is completed in modern car engines after approximately 20–30 operating hours but, in certain individual cases, only after more than 100 operating hours, so that the engine reaches a constant friction level. This remains more or less constant until the engine components reach their service life limits, once again leading to an increase in the friction.

90/90°C

120°C/110°C

Oil temperature/coolant temperature

Figure 9.3 Influence of oil viscosity on friction.

The effect of the influences of engine load and engine speed on the friction behavior of a car with a gasoline engine is shown in Figure 9.5. The measurements collected on the motored engine with loads between 0 bar (zero load) and full load are also compared with the results from drag measurements (pme corresponds to the drag torque). The measurements in

392 | Internal Combustion Engine Handbook

6606_Book.indb 392

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

motoring mode at zero load show a good correlation with the measured values of the drag measurements at 0 bar.

Complete engine

2,5 2,0 1,5 1,0 0,5 0,0 –4

–2

0

2 4 6 PMR Load [bar]

8

10

Figure 9.5 The dependence of the friction on the engine load and speed increases in the low-sliding speeds due to the larger piston side forces. at high engine revolutions sinks the friction with increasing load. The reasons for this are the higher oil temperatures at the cylinder barrel at high engine powers, despite the same main oil temperature, and the partial compensation of the inertial forces in the crank gear by gas forces.

The main influencing parameter is the engine revolutions. The engine friction increases at higher engine revolutions. At moderate engine revolutions, the engine load has only a very minor influence on the friction, that is, the effects shown have only a minor influence, or compensate each other in this rpm range. At an rpm of 1000rpm, the friction increases with increasing load.

9.1.5 Influence of Friction on the Fuel Consumption

The mechanical efficiency η m of an internal combustion engine is defined as the ratio of mean effective pressure pme and indexed mean pressure pmi.



⎛ p ⎞ ⎛ p − pmr ⎞ hm = ⎜ me ⎟ = ⎜ mi ⎝ pmi ⎠ ⎝ pmi ⎟⎠

(9.3)

From this relationship, it is clear that at low engine loads, that is, low mean effective and indicated pressure, the mechanical efficiency drops. The spreads of mean friction pressures of modern gasoline and diesel car engines are shown in Figure 9.6. At 2000rpm, with values 0.53–1.1 bar for gasoline engines and 1.02–1.4 bar for diesel engines, including injection pump, the friction losses at full load are as high as 10% of the indexed

PMR friction mean effective pressure [bar]

PMR friction mean effective pressure [bar]

1000 1/min Oil temperature/coolant temperature = 90°C 2000 1/min 3000 1/min Fired, load variation 4000 1/min Fired, zero load 5000 1/min Towed 6000 1/min 3,0

power. In part-load operation, the mechanical efficiency drops so that the influence of friction on the fuel consumption continues to rise. A reduction in friction, therefore, offers a significant fuel saving’s potential and presents a worthwhile development objective. The span in each case between the engine with the highest, and the engine with the lowest friction, not only means an increase in fuel consumption, but also a reduction in the maximum power.

3,0

Fired, full load

Crank gear, Valve train, Loaded oil pump, Loaded water pump, Unloaded generator

Oil:15W40 Oil temperature/ coolant temperature: 90 °C

Gasoline Diesel (direct injection)

2,5 2,0 1,5 1,0 0,5 0,0 0

2000

4000 6000 Engine speed [ –1]

8000

Figure 9.6 Scatter band of friction during fired engine operation (car engines) [9-5].

The development of friction over time is examined below, taking as an example the four-cylinder spark-ignition (SI) engine. Figure 9.7 illustrates the development of the mean friction pressure pmr on the basis of studies in drag mode at 2000rpm. The first thing that is obvious is that the spread of the values has a very large bandwidth, although a downward trend is noticeable that is marked clearly by the regression line. In particular, the friction behavior of the gasoline engine has been significantly improved in recent years. In purely statistical terms, the friction of a 2 l, 4-cylinder gasoline engine has been reduced by approximately 20% during a period of around the last 10 years. The extrapolation of the regression lines, however, results in an unrealistic reduction in the friction for the future. The reduction in fuel consumption as a function of the mean friction pressure with the engine at operating temperature and an engine speed of 2000rpm is illustrated in Figure 9.8. The hypothetical case of a frictionless engine enables a saving in fuel consumption of approximately 21% for the gasoline engine and approximately 26 % for the diesel engine. With the conventional measures used today (component optimization, roller valve train, modified pumps, heat management), approximately 30% of this potential may be utilized.

Internal Combustion Engine Handbook | 393

6606_Book.indb 393

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Reduction in fuel consumption

Friction mean effective pressure [bar] at 2000 1 / min

Chapter 9 Tribology

1,2

1,0

0,8 FEV 2009

0,6

0,4 1998

2000

2002

2004

Year of manufacture

30% Diesel Engine 20% Spark-Ignition Engine 10%

0% 0%

20%

40%

60%

80%

100%

Friction reduction

Figure 9.8 Influence of the friction reduction on fuel consumption (friction considered at n = 2,000rpm) [9-7].

9.1.6 Friction Behavior of Previously Introduced Combustion Engines 9.1.6.1 Breakdown of Friction When considering the friction losses of an engine, it is not only the total value but also the breakdown of the friction between the various components which is of crucial importance. A common method employed for this is the strip method, which is described in detail in the following sections. Before the actual strip measurements, the complete engine with intake and exhaust sections is motored (“complete engine”). The drive torque measured here includes not only the mechanical engine friction, but also the charge cycle losses. During this measurement, the oil pressures at the oil pump outlet, in the engine gallery as well as when possible in the cylinder head together with the oil volumetric flows through the engine, are recorded for every working point. The cooling system is subjected to external pressure for a constant pressure at the inlet to the coolant pump. The recording of these boundary conditions subsequently allows for the boundary conditions

2006

2008

Figure 9.7 Development of friction in 4-cylinder gasoline engines (1.6 l–2.2 l displacement).

on the complete engine to be set exactly in the individual strip steps. Following the recording of the boundary conditions on the complete engine, the measuring program for determining the friction of the individual components is then executed. The stripping steps, which have to be executed, are described as follows: 1. The cylinder head is removed to determine the engine friction. To maintain the strain conditions of the engine block in the area of the bolts, the cylinder head is replaced by a plate with rounded cylinder openings. The gas chamber is therefore open in this series of measurements; the piston is not subjected to gas forces. All the auxiliary drives will also be removed. The oil pressure in the main gallery is set for the engine operation—on the basis of the measurements performed for the complete engine or according to data from other sources—using an external hydraulic oil supply. 2. Arrangement of the piston and connecting rod for determining the crankshaft-bearing friction. The influence of the rotating masses is compensated by attaching “master weights” on the connecting rod bearing pins. The setting for the oil pressure in the engine gallery is also hereby executed—as in 1)by utilizing the external hydraulic oil supply. 3. Measurement of the friction losses from the crankshaft (including “master weights”) with valve train. The setting for the oil pressure in the engine gallery is executed here again—as in a)—by utilizing the external hydraulic oil supply. 4. Measurement of the friction losses from the crankshaft (including “master weights”) with the oil pump. The setting for the oil pressure in the engine gallery is also hereby executed—as in a)—by utilizing the external hydraulic oil supply. The engine’s own oil pump returns the oil

394 | Internal Combustion Engine Handbook

6606_Book.indb 394

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

directly back to the oil pan, in a separate hose circuit via a variable throttle, that regulates the oil pump pressure. The oil pump pressure is also set for the working point according to the pressures previously measured.

The result of a strip measurement for a modern car gasoline engine is illustrated in Figure 9.9. The percentage breakdown of the friction proportions is illustrated in Figure 9.10. The definition of the reference parameter total friction includes the auxiliaries necessary for the engine operation—oil pump and coolant pump under load as well as alternator not under load, and components not purely for comfort, such as power steering pump or air-conditioning compressor. Component-specific spreads can subsequently then be elaborated from the measurement results for the individual components. By comparing the measurement results for individual components with the corresponding spreads and, therefore, with the latest technical standards, it is also possible to identify potentials for a reduction in the friction loss and to selectively exploit these potentials by optimization processes. Figure 9.11 illustrates the mean friction pressure of the stripped engine components over the displacement volume for an engine speed of 2000rpm and an oil/coolant temperature of 90/90°C. The spreads in these figures illustrate that the displacement volumes above 1.5 liters have no practical effect on the level of mean friction pressure for a stripped complete engine. This can be attributed to the fact that the power demand for different power units is dependent on the size of the vehicle and is not further reduced but, also therefore that the upper displacement volume limit for the smaller, car engine families

5. Measurement of the friction losses of the crankshaft (including “master weights”) with coolant pump, alternator, power steering pump, and air-conditioning compressor, including tensioner and guide pulley(s). The setting for the oil pressure in the engine gallery is executed—as in 1—by utilizing the external hydraulic oil supply.

PMR friction mean effective pressure [%]

PMR friction mean effective pressure [bar]

The friction losses of the piston/connecting rod bearings, valve train, oil pump, and auxiliary units are determined from the differences between the results for the individual series of measurements. Furthermore, the sum of the determined values for the individual components results in a friction value for the whole engine, which is then referred to as “stripped complete engine.” This describes the purely mechanical friction losses of the engine, without the gas cycle losses. A further detailing of the measuring program, for example, the determination of the friction of individual or all piston rings, or the breakdown of the valve train friction between camshaft bearing friction and valve actuation, is possible by including further stripping steps. On the other hand, a measurement of all the components is not absolutely essential when only individual units are to be considered.

2,0

Water pump and generator Oil pump Valve train Piston and connecting rods

1,5 1,0 0,5

Crankshaft 0,0

2000

0

4000

6000

Engine rpm [min-1]

Oil: SAE 15W40 Oil temperature/ coolant temperature: 90 °C

Figure 9.9 Friction breakdown in a modern, car gasoline engine [9-7].

Power steering pump Climate compressor Unloaded Fuel pump Water pump and generator Oil pump Valve train

120 100 80

Piston and connecting rods

60 40 20

Crankshaft 0

0

2000

4000

Engine rpm [min –1]

6000

Oil:SAE 15W50 Oil temperature/ coolant temperature: 90 °C

Figure 9.10 Percentage breakdown of the friction modern, car gasoline engine [9-7].

Internal Combustion Engine Handbook | 395

6606_Book.indb 395

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

PMR friction mean effective pressure [bar]

Scattering range Serial petrol engines

Date: 1/2003 - 1/2009 Engines: 53 Oil: 0W30 - 15W40 Temperature: 90 °C

Complete Engine (stripped)

Crank gear, Valve train, Loaded oil pump, Loaded water pump, Unloaded generator

Figure 9.11 Friction for a gasoline engine depending on the cubic capacity.

Piston displacement [cm3]

lies at approximately 1.5 l. Because of the identical parts in the engine families, the engines are designed for this largest variant so that the smaller engines in the family have certain friction disadvantages. 9.1.6.2 Crank Gear The crank gear of an internal combustion engine consists of the crankshaft, including the radial shaft seal rings, and the friction of the piston group and connecting rods. By utilizing the strip method, the crank gear can be further split into the friction of the crankshaft, and the friction of piston group and connecting rods. 9.1.6.2.1 Crankshaft The crankshaft friction is determined using master weights and includes the radial shaft seal rings. If one plots the mean friction pressure of the crankshaft against the engine speed and extrapolates the values up to a theoretical engine speed of 0rpm, then the Y branch received as a result can be roughly interpreted as the friction portion of the radial shaft seal rings that is relatively independent (of the engine speed).

The resultant value correlates with the measurement values from the separation of the radial shaft seal rings by stripping. The friction moment of an individual main bearing referred to by its diameter illustrated in Figure 9.12, for an engine speed of 2000rpm, can be calculated from the friction values for the crankshaft. The representation is for the measured values for a large number of engines, as well as the regression lines for different engine concepts. The spread of the measured values around the respective regression lines illustrates that further parameters influence the friction, in addition to the mainbearing diameter. These include et alia the bearing geometry, the bearing play, deformations or alignment deviations of the bearing race, as well as differences in the friction of the radial shaft seal rings. 9.1.6.2.2 Connecting rod bearing and piston group The friction of the piston group, including connecting rod bearings, can be determined by the subtraction of the friction values for the crankshaft from the friction values for the engine. A further breakdown is difficult to achieve with the strip method, as the connecting rods and piston group cannot

Friction torque/main bearing [Nm]

Main bearing diameter [mm] 30

1,00

40 45

50

55

60

65

70 2000 1/min Oil:SAE 15W50 Oil temperature/ coolant temperature: 90 °C

0,75

Regression line R Otto engines Regression line diesel engines Regression line Otto V-engines

0,50

0,25

0,00 0

100

200

300

Main bearing diameter 3 [cm3]

400

Figure 9.12 Friction per crankshaft main bearing over main-bearing diameter3 [9-5].

396 | Internal Combustion Engine Handbook

6606_Book.indb 396

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

Dynamic friction force-measuring system - PIFFO Piston friction force

Dynamic Friction force [N]

Piston ring package: Version B: Basic version Version C: optimized pre-tensioning and ring height

Fired, full load 2000 1/min Oil temperature/ coolant temperature: 90 °C

Piston ring variants: 1, 2, 3, 4 Ancillary conditions: Dragged, full load Temperature: 90 °C

V2

V1

V3

0,2

V4 V3 0,1 1000 1/min 2000 1/min 0,0

0

1

2

3

• p0 [N/mm2] Total piston ring surface pressure

Figure 9.14 Piston ring friction as a function of the preload [9-9].

9.1.6.2.3 Mass balancing Mass balancing is the term used to refer to the measures employed for partial or complete balancing of the inertial forces, and torques on crank drives. An additional mass balancing is employed in many cases in car engines to improve the comfort ratio. The friction losses of the mass balancing gearing are influenced by the following: •• the order of the inertial forces or torques to be balanced and, hence, the number and speed of the balancing shafts

Version C

100

0,3

Friction mean effective pressure [bar]

be operated independently from each another. The friction of the connecting rod bearings can be determined using a practically friction-free aerostatic piston guide [9-8]; however, the effort involved is very high. The breakdown of piston and piston rings, or the separation of single piston rings, is possible but it must be remembered that the removal of piston rings significantly changes the lubrication conditions of the piston and the other rings. As illustrated previously, the friction of the piston group has a very large proportion of the total friction in an internal combustion engine. Great importance has therefore to be attached to its optimization, in order to attain the goal of a low-friction engine. A wide range of measuring systems have been developed in order to measure the friction behavior in the piston group [9-9], or for being able to record the friction influencing parameters such as the cylinder deformation in motoring mode [9-10]. The direct measuring of the piston friction force in motoring mode provides, as illustrated in Figure 9.13, the curve of the friction force over the crankshaft angle, which thereby enables detailed conclusions to be drawn up for the frictional processes between the piston and the cylinder wall as well as, in the event of force peaks, an indication for possible wear.

•• number, design, and diameter of the bearing locations •• losses in the drive of the mass balancing elements.

0

–200

Version B

–100 0

180

360

540

720

Crankshaft angle [degrees]

Figure 9.13 Friction force curve of the piston group in motoring mode [9-7].

The influence of various parameters, such as piston photomicrograph, piston clearance, and piston ring pretension, can be examined in dragged and motoring mode. A variation in the piston ring surface pressure (piston ring tangential stress referred to the bearing piston ring surface) is illustrated in Figure 9.14. One can clearly observe the significant influence of the sum value on the measured friction. A comparison of a two-ring piston with the conventional three-ring piston of similar piston geometry and mass, and the same sum value of the surface pressure, that is, higher surface pressure of the individual rings for the two-ring piston, showed no significant differences in the mean friction pressure.

The balancing of the free second-order inertial forces in four-cylinder engines requires two balance shafts, which rotate at twice the crankshaft speed and, hence, exhibit unfavorable boundary conditions with respect to the friction behavior. Mass balancing gearings already built for four-cylinder engines exhibit friction values of 0.05–0.16 bar at 2000 min–1, which can correspond to as much as 18% of the total friction in the engine. 9.1.6.3 Valve Actuation (Valve and Control Engines) The friction of the valve train can be determined with the strip method from the difference between the measurement for the crankshaft with valve train and timing gear and the measurement for the crankshaft. A further separation, for example, of the friction in the valve actuators or the camshafts is possible but, in the analysis, it has to be remembered that the timing gear dynamics are influenced and, hence, that the friction behavior changes. Various valve train concepts are employed in modern car engines. Figure 9.15 Illustrates the example of a multivalve engine, which shows that these concepts also have a considerable effect on the friction behavior of the valve trains. In valve trains with sliding tappets, the hydraulic valve clearance

Internal Combustion Engine Handbook | 397

6606_Book.indb 397

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

PMR friction mean effective pressure [bar]

Sliding actuation with hydraulic valve clearance compensation Sliding actuation without hydraulic valve clearance compensation Rolling contact Multi-valve petrol engines

Sliding actuation / roller actuation with / without HVA DOHC - Cup / tilt / cam follower

Date: 1/2003 - 1/2009 Sliding contact: 11 Rolling contact: 38

Drive through chain / synchronous belt

Oil: 0W30 - 15W40 Temperature: 90 °C

Valve train Valve actuation, Camshaft bearing, Shaft sealing ring, Drive unit including tensioning and guide rollers

Engine rpm [min-1]

Figure 9.15 Comparison of different valve-drive concepts.

compensation increases the friction, by the way of the additional friction caused by the pressure of the hydraulic element in the area of the cam base circle and the larger moving masses. Valve trains with roller tappets generally exhibit very favorable frictional behavior. The unfavorable system dynamics of the timing gear associated with roller tappets, however, frequently necessitate higher pretensions in the timing gear. This can then result in increased friction, especially with chain drives [9-11]. The breakdown of the friction within the valve train is necessary for the implementation of effective optimization measures. Figure 9.16 illustrates this distribution for various valve train concepts. Sliding tappets exhibit the largest portion in the contact area of cam to tappet. This is because of the high contact forces and high relative velocities between cam and tappet. A reduction in friction can be achieved through a lowering of the contact forces by

Friction mean effective pressure [bar]

0,4

0,3

Valve lead

Light construction

reducing the valve spring forces. With unchanged maximum engine speed, however, a reduction in the moving masses in the valve train is indispensable. The other possibility is to reduce the relative speeds by using a roller between the cam and the tappet. 9.1.6.4 Auxiliary Units In addition to the crank gear and the valve actuation, a modern internal combustion engine also has a large number of auxiliary units. These are necessary for a trouble-free operation of the internal combustion engine and also for fulfilling the additional functions, for example, safety and exhaust gas cleaning or complying with the rapidly growing comfort demands of the vehicle owners. Examples of the functions of the auxiliary units are as follows: •• assurance of the engine’s proper mechanical function in all operating states of the automobile: lubrication oil pump,

Camshaft bearing Valve guide Valve lead EHD-Contact

Roller bearing

0,2

0,1

0,0 2000

4000 Engine rpm [min–1]

6000

Figure 9.16 Division of friction in the valve train [9-11].

398 | Internal Combustion Engine Handbook

6606_Book.indb 398

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

coolant pump, fuel supply system, radiator fan, mechanical turbocharger •• assurance of proper supply of electrical energy to the engine, and the automobile, in all operating states by utilizing an alternator •• creation of an additional exhaust gas cleaning facility: secondary air pump, catalytic converter preheating •• provision of auxiliary energies to cover enhanced passenger comfort and safety requirements: power steering pump, airconditioning compressor, vacuum pump, starter, antilock brake system, traction control system, level control system. Depending on the operating state, the drive of these auxiliaries consumes a large proportion of the mechanical energy provided by the internal combustion engine in the modern series application. The drive power required for these auxiliaries thus represents a mechanical loss and can be assigned to the friction loss. Various definitions make allowance for these auxiliaries in different ways. Our purpose here is not to consider the definitions, but rather to look at the fundamental relationships with respect to the friction of the auxiliary units. As this plays a significant role in the fuel consumption of the vehicle, this aspect is becoming more and more important as a considerable increase in the energy demand is to be expected for the future due to additional, or more powerful, consumers. This section gives an overview of the auxiliary units of a modern internal combustion engine. In view of the large number of such auxiliary units, we can only look here at the auxiliaries necessary for operation of the engine and the auxiliaries with the highest drive powers. We also take a brief look at the large number of components driven electrically and not directly by the internal combustion engine. The power supply to these components must not be neglected when considering the alternator drive power. In modern engines, the auxiliaries are driven almost exclusively with a constant gear ratio to the crankshaft; this means that the speed of the individual auxiliaries is proportional to the crankshaft speed. The spread of the auxiliary speeds (ratio

of maximum to minimum auxiliary unit rpm) is defined by the spread of the speeds of the internal combustion engine due to the fixed transmission ratios. An adequate power output of the individual auxiliaries even close to the engine idle speed determines the transmission ratio. On the other hand, the power to be output from the crankshaft via the belt drive or chain drive increases with the engine revolutions, even if the power provided by the auxiliaries on the secondary side is not required. However, the individual power demands of the auxiliaries are not directly dependent on the engine revolutions. Therefore, the direct drive represents a compromise between benefits and costs. In the evaluations below, a distinction is made between the following definitions of power: •• Auxiliary power: engine power is required to drive the auxiliary units. •• Exhaust output: the power output exhibited by the auxiliary units (e.g., electrical energy or flow energy) •• Power demand: power output of the auxiliaries needed to cover the power demand of the engine or the automobile. Figure 9.17 illustrates the mean friction pressure of the auxiliary units, which is necessary for the engine operation: oil and coolant pumps deliver according to the engine-operating point, the alternator is driven but does not supply any electric power. The sum of the most favorable individual values for different engines illustrates that there is still adequate optimization potential. 9.1.6.4.1 Oil pump Modern four-stroke engines are lubricated by a forced feed lubrication system. The following major components are essentially supplied with lubricating oil by the oil pump: •• crankshaft main and connecting rod bearings •• piston spray jets •• valve train and drive (camshaft, tappets, gear wheels, etc.) •• turbocharger

PMR friction mean effective pressure [bar]

Scattering range Serial petrol engines Date: 1/2003 - 1/2009 Engines: 53 Oil: 0W30 - 15W40 Temperature: 90 °C

Ancillary drive unit

Loaded oil pump, Loaded water pump, Unloaded generator (each incl. drive)

Engine rpm [min-1]

Figure 9.17 Friction of the power take-offs.

Internal Combustion Engine Handbook | 399

6606_Book.indb 399

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

•• additional lubrication points depending on the engine design. The task of the engine oil circuit is thereby to •• ensure a supporting oil film on all sliding surfaces under all operating conditions, to effectively prevent mixed friction and the related wear •• prevent localized overheating of components, and the resulting damage, by ensuring an adequate heat removal

shifter) as well as high revolutions for sufficient oil supply to the dynamically, highly loaded cam rod bearings [9-15], [9-16]. Figure 9.19 shows oil flow rates and oil pressures in a lubrication system. It is necessary that the operation of the engine at all operating points of the minimquired oil pressure to ensure a correct operating level of lubricant supply, for example, without tappet rattling and cavitation is reached or exceeded in the connecting rod bearing.

•• prevent corrosion. The oil pumps generally used in the motor vehicle engines are crescent-type or trochaic pumps driven directly by the crankshaft, or externally geared pumps and trochaic pumps driven via a step-down gear and auxiliary drive. The drive power of the pumps differs significantly, depending on the drive system and pump type. The various optimization steps described in references [9-12], [9-13], [9-14] enable the pumps to be individually improved and adapted to the engine requirements. A feature common to all pump types is, as illustrated by the spread of the oil pump mean friction pressures in Figure 9.18, the increase in the drive power at high engine speed. In the majority of oil pump operating ranges, the widely utilized, but energetically unfavorable bypass control, results in lower efficiencies. Adequate lubrication, in other words a certain minimum oil pressure, must exist under all engine operating conditions because without this minimum pressure, engine damage can result within a very short time. The oil pump is, therefore, designed for the least favorable case, that is, high oil temperature and an engine with a long service life and, therefore, large clearances. Other designs are low revolutions to ensure the oil supply of the hydraulic valve clearance compensation elements (hot idle), and oil pressure-controlled actuators (e.g., camshaft

Conrod bearing

6

Oil pump 4

Main bearing Cylinder head

2 0 60

Oil flow rate [L/min]

•• prevent or remove deposits

Oil pressure [bar]

•• pick up particles (soot and/or wear particles) and keep them in suspension

(1) Pressure valves (cylinder head) (2) Pressure relief valve (pump) (3) Hydraulic valve clearance compensation (4) Exhaust turbochargers (5) Vacuum Pump (6) Camshaft bearing (7) Piston cooling nozzles (8) Conrod bearing (9) Main bearing

50 40 30 20

(1) (2) (3) (4) (5) (6) (7) (8) (9)

10 0

0

1000

2000

3000

4000

5000

Engine rpm [1/min]

Figure 9.19 Oil pressure and oil flow in the lubrication circuit.

The oil displacement of the engine increases less sharply with increasing engine speed than the delivery of the oil pump that increases more or less in proportion to the increase in engine speed. Therefore, a part of the flow rate at medium and high speed will be mostly recycled through a bypass valve to the suction side of the pump.

Directly driven PMR friction mean effective pressure [bar]

Indirectly driven Serial gasoline engines

Date: 1/2003 - 1/2009 Directly driven: 23 Directly driven: 30 Oil: 0W30 - 15W40 Temperature: 90 °C

Oil pump

Loaded oil pump, Regulating valve (active) incl. pump drive,

Engine rpm [min-1]

Figure 9.18 Friction of various oil pumps (loaded).

400 | Internal Combustion Engine Handbook

6606_Book.indb 400

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

Apart from the demand-oriented adaptation of the lubrication system, and the detail optimization of the oil pump to the requirements of the engine, variable-displacement pumps have a great potential for reducing the drive power of the oil pump. Possibilities for adapting the delivery of oil pumps to the necessary demand are concepts with a variable delivery chamber volume that are, however, generally highly complex and expensive, as well as controlling revolutions by separation of the pump speed from the engine speed. 9.1.6.4.2 Coolant pump The coolant pumps for internal combustion engines are predominantly centrifugal pumps, which are designed to provide an adequate coolant throughput to dissipate the heat, both at low engine speeds and high engine load (e.g., driving uphill with a trailer) and at rated output. With speed control of the pump dependent on the temperature of the components, or of the coolant, for example, via an electric drive, the temperature level of the walls surrounding the combustion chamber and, hence, the efficiency of the engine at part load could be increased, the warm-up time of the engine shortened, and the drive power of the coolant pump reduced even at high engine speeds. The engine speed-proportional drive results in high deliveries at high engine speeds that, with an unfavorably designed coolant circuit, leads to high-pressure losses and, hence, to high drive powers [9-17]. This offers an optimization potential in the design of the circuit with a correspondingly modified coolant pump [9-18]. 9.1.6.4.3 Alternator High-performance and low-maintenance AC claw-pole alternators with a rated voltage of 14 V are today almost exclusively utilized for providing electrical energy in private vehicles. The efficiency of the alternators is currently limited to a maximum of 60% to 70% and are achieved at low engine speed and with high loads on the alternator. Frequently, however, alternators are operated at high engine speeds and with low load and, therefore, with low efficiencies of between 20% and 40%.

The electrical load requirements installed in the automobile have risen drastically in the last 40 years from around 0.2 kW to 2.5 kW and are expected to rise in the next 20 years to roughly 4 kW. The forecasts based on the Prometheus projects go even further. Here, around 8 kW of electric power will have to be provided up to the year 2010, when the power limits of the normal 14-V-alternators of approximately 3 to 5.5 kW will be exceeded [9-19]. Figure 9.20 shows the friction of various generators without electrical power output. Startergenerator systems with higher electric power outputs, and 42 V output voltages, are expected for the vehicles of the future. Considerably, higher electric powers are a precondition for the transition to electric motors for various components (e.g., oil pump, coolant pump, electromechanical valve train). The power consumption of the consumers required to maintain the engine function is more or less independent of the vehicle operation. The electric power demand for all the other consumers, on the other hand, in particular for comfort functions, depends to a very great extent on the operating conditions (Summer, Winter, day, or night). Overall, there is a spread in the total power demand from roughly 300 W up to 1200 W for a middle-class vehicle, depending on the operating conditions and frequency of operation. Due to physical relationships can be at constant generator weight, the output power when the engine is idling and the maximum power does not independently determine [9-20]. This unfavorable scenario is exacerbated by the increasing power requirements at idle speed, and the effort to further reduce the idling speed of fuel economy reasons. As a consequence of this, thought must be given to the alternator concept and to the drive management of the alternator. A characteristic parameter in the alternator design is the two-third engine speed at which the alternator can output two-third of its maximum power. The transmission ratio between alternator and engine is normally selected so that the alternator runs at two-third of the engine speed at engine idle speed and thereby ensures the power supply to the engine and the automobile. Alternator optimization goals are a high efficiency in all operating ranges, a low starting speed, and a high-power

PMR friction mean effective pressure [bar]

Scattering range Serial petrol engines Date: 1/2003 - 1/2009 Engines: 53 Oil: 0W30 - 15W40 Temperature: 90 °C

Generator, unencumbered Generator without charge controller, Drive including generator cooling

Engine rpm [min-1]

Figure 9.20 Friction of car generators (unloaded).

Internal Combustion Engine Handbook | 401

6606_Book.indb 401

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

output. The current should, therefore, rise sharply above the starting speed (1000 to 1500 1/min) so that a high power can be output to the running consumers even in the lower engine speed range. The largest share of the generator losses at full-load operation make particularly the iron and copper losses in the stator and the friction and fan losses, while the diode and excitation losses are relatively small [9-21]. Since the power output above a generator speed of 5000 1/min only slightly increases, an operation is recommended in generator speeds between 2000 1/min and 5000 1/min. 9.1.6.4.4 Fuel injection pump The fuel injection pump serves to inject the fuel through an injection nozzle directly into the combustion chamber toward the end of the compression stroke. Depending on the design of the injection volume and the engine operating point, the injection pressure lies between 50 and 200 bar for SI engines with direct injection and over 2000 bar for diesel engines. Figure 9.21 shows the friction of a distributor injection pump of a diesel engine with direct injection. Between zero load and the maximum position of the fuel quantity positioner, the friction values are quadrupled. The friction value occurring at full-load make up a major portion of the total friction in a diesel engine and are one of the main reasons for the increase in the engine friction between zero load and full load in diesel engines.

PMR friction mean effective pressure [bar]

Oil temperature/coolant temperature: 90 °C 0,4

0,3 Full load adjustment Idle adjustment

0,2

0,1

0,0

0

2000

4000

6000

Engine rpm [min –1]

Figure 9.21 Friction injection pump (comparison full load no load).

9.1.6.4.5 Air condition compressor The development of air conditioning in automobiles began in the United States in the 1960s. Whereas only 20% of the automobiles in the North American market were equipped with air conditioning in 1965, this figure had already risen to 80% by 1980. The desire of the Japanese automobile manufacturers to conquer the North American market led to the Japanese discovering air conditioning for themselves so that, by as early as 1985, the percentage of automobiles in Japan with air conditioning was higher than that in the United States. A similar albeit delayed development has been observed since the end of the 1980s in Europe, too,

but without the market penetration of the United States being achieved until now. The cooling capacity demand for air conditioning in automobiles is dependent on the solar irradiation and the ambient temperatures. The average duration of the air conditioners in Europe is about 23% (about US 42%) and the average required cooling capacity 1 to 2 kW (US 4 to 5 kW) [9-22]. Of all the auxiliary units of the air-conditioning compressor has the greatest power, which can be up to 11 kW at high speeds depending on the compressor design and operating condition. The average drive power is depending on the duty cycle between 180 W and 2000 W. Air compressors are usually driven by a belt from the engine speed proportional to give a function of the cooling capacity of the engine speed results, but while the demand is almost independent of speed. The air-conditioning compressors are designed for maximum required cooling capacity that has to be available even at low engine speeds (when driving in town with a high percentage of engine idle time). At higher engine speeds, the compressors are, consequently, over dimensioned and have to be controlled. In many cases, this is achieved with an electromagnetic clutch via which the compressor can be switched on and off. Current developments are now moving away from the pressure-controlled compressor and increasingly toward volumetric flow-controlled compressors that reduce the excess power by varying their displacement. Energetic but no significant improvement to be expected since the regulated compressor stays on longer and even with low cooling capacity the mechanical losses, especially at high speeds, significantly are [9-23]. For use in compact economy cars, especially in Japan, also more compact and lightweight compressors have been developed (for example vane and scroll compressors) in recent years. 9.1.6.4.6 Radiator fan The radiator fan has to ensure an adequate flow of air through the heat removal (radiator) to dissipate the heat at high load and low vehicle speeds. Earlier the fan was driven directly by the engine at a speed proportional to the engine speed. In modern designs, temperature-controlled drive systems with a clutch (electric or hydrostatic drives) are utilized. These reduce the required drive power by comparison with the rigid drive by between 25% and 50%. Electrically driven radiator fans are switched on when needed, depending on the coolant temperature. A switching hysteresis of around 10° prevents continuous switching on and off. An electric fan is in operation when driving in town traffic for between 30% and 40% of the time. At higher vehicle speeds (main roads, highways), the airflow through the radiator is normally sufficient for heat removal, even without the fan. At low engine speeds, viscous fans require less drive power than electric fans. This is because of the higher drive efficiency of the viscous fan at low engine speeds by comparison with the electric fan where the alternator efficiency also has to be taken into consideration. During the warm-up phase or part

402 | Internal Combustion Engine Handbook

6606_Book.indb 402

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.1 Friction

load of the engine and at higher engine and vehicle speeds, the electric fan has the advantage over the viscous fan that it can be switched off when there is an adequate flow of air through the radiator. However, both drive systems still have a significant potential for reducing the transmission losses. 9.1.6.4.7 Power steering pump Power-assisted steering systems, which a few years ago were still reserved for the luxury-class automobiles, are available today even for compact models. The trend toward broader tires and, hence, to increased steering effort, the more direct function of the power steering system, and the resulting improved handling of the vehicle has led to a significant increase in the market share of automobiles with power steering in recent years. The steering assistance is provided by the oil pressure supplied by the power-steering pump and controlled at the steering gear according to the power assistance currently required. For cost reasons, a vane pump with bypass control is predominantly utilized as power-steering pumps in seriesproduction vehicles. The pressure requirement in the hydraulic system is dictated by the vehicle speed and the steering angle of the wheels. In present-day systems, maximum pressures of up to 130 bar occur in some cases at standstill with maximum steering angle. With increasing vehicle speeds, however, the required steering assistance drops sharply. The minimum pressure of the power steering system required to overcome the flow losses of the steering system when traveling straight ahead is vehicle and steering specific and lies in the order of 2 to 5 bar. The displacement of a power steering pump has to be sufficiently high at low engine speeds, and high steering speeds, to ensure the steering assistance. For the design conditions, this means engine idle speed with the automobile at standstill and high steering speed on a dry road. These conditions occur during vehicle operation, particularly when parking or maneuvering. At higher engine speeds, a multiple of the useful oil flow is discharged as oil loss via the flow controller. The driver power of the pump increases proportionally to the engine speed. The maximum possible drive power does not normally occur in practice, because high pressures in the steering system and high engine speeds do not occur simultaneously. The required drive power of a power steering system is heavily dependent on the pump speeds and system pressures dictated by the vehicle operation. Typical driving power of conventional steering systems when driving straight ahead on average 250 and 1200 W. Through the use of controlled power steering pumps, such as intake manifold regulated radial piston pumps, the driving power can be significantly reduced. Great potential is offered here by electric power steering systems requiring average drive powers of only 100 to 200 W that have been used in recent years in small- and medium-sized series production vehicles. 9.1.6.4.8 Vacuum pump On engines with throttle-free load control, a vacuum pump is employed to generate a vacuum (e.g., for the brake booster).

The friction of normal vacuum pumps lies between 0.01 bar at low engine speeds and 0.04 bar at high engine speeds.

9.1.7 Method of Calculating the Friction on the Example of the Piston Group 9.1.7.1 Modern Computer Simulations The optimization of engine components can be ideally performed computationally. Against this background, mixed models of computation of linear finite element method (FEM) and multibody simulation have prevailed in most cases for the dynamic simulation. The strength of this model lies in that the number of degrees of freedom, for describing the structure stiffness, can be drastically reduced without reducing noticeably the calculation accuracy. Explicit FE analyses permit direct evaluation of the component stresses. However, these have the major disadvantage that the calculation times are orders of magnitude higher. For detailed analyses of bearing tribology, the use of an elastomer-hydrodynamic (EHD) calculation method is recommended. These are able to conclusively solve the interaction between the hydrodynamic bearing capacity of the lubricant film and the local resilience of the structures. Modern commercial multipurpose applications also represent the core of numerical friction analysis in the engine calculation. MKS system of units (MKS) and FEM are therefore indispensable in today’s development process. The piston with rings and cylinder tube contact represents, for example, a highly complex tribological system, which accounts for the bulk of the engine friction. The increasing performance requirements for this module make it ever more imperative to resolve in more detail the performance limit with modern simulation techniques. Against this background, detailed simulation model are developed now [9-24]. These models are constructed primarily on commercial software, which will be extended to specific problems via user-programmed subroutines (Figure 9.22). User subroutine

Structure

Temperature

Lubrication

Topology

Materials Micro geometry

Tube deformation

Figure 9.22 Detail model to simulate the piston group [9-24].

Internal Combustion Engine Handbook | 403

6606_Book.indb 403

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

From a physical point of view, known approaches are used: The Reynolds differential equations in the field of hydrodynamics. The approaches according to Greenwood–Tripp for the description of mixed friction and wear, or the Labyrinth Model from Eweis for the gas dynamics in blow-by. Other models for the notation of the lubrication film proportion in the complete system allow statements regarding oil supply and oil consumption. The macrogeometry is three dimensionally displayed by the FE model. The microgeometry (such as the crown of the piston rings) and the topology (such as the honing structure) are described by the subroutines. The guidance of the piston rings (including shock) is only possible via contact formulations. Hence, multipart oil control rings, including all contact surfaces, can also be modeled physically. The main evaluation parameters result from the calculations: Friction, wear, gas leakage, and lubricant emission and all movement dimensions, for example, the secondary piston movement or the piston ring dynamic. Based on detailed dynamic measurements, by means of specially developed special measurement techniques, there can then finally be an assumption of the different test supports for a high-quality basic setup of all modules. Thus, a powerful tool for targeted and efficient numerical optimization of the individual components is available—a common method today.

9.2 Lubrication 9.2.1 Tribological Principles

Engine technology is based on machine elements of different kinds that, linked by form and function, act on and influence one another by the following: •• kinematics: generation, transmission, and inhibition of movement •• kinetics: power is transferred by contact interfaces •• transmission and conversion of mechanical energy •• transport processes: transportation of liquid and gaseous media. An important role in these processes is played by tribology*—according to DIN 50323, “tribology is classified as the science and technology of interacting surfaces influencing each other in relative motion. It encompasses the entire field of friction and wear, including lubrication, and includes the respective interracial interactions between both solids, and between solids and liquids or gases.” This lubrication hereby enables, improves, and ensures the lubrication function, efficiency, and service life of the components, functional groups of the engine and of the entire drive system. In the field of their interactions, tribological systems can be reduced to a basic structure (system elements) (DIN 50320):

basic surface, mating surface, intermediate substance (particles, fluids, gases), and ambient medium (Figure 9.23).

Stress collective

Structure of the tribological system Ambient medium

Counter body

Precursor

Basic element

Surface changes (Wear manifestations)

Wastage (Wear-measure)

Wear characteristics

Figure 9.23 Schematic: tribological system [9-39].

Tribologic stresses result from the movement process, effective forces (normal force), speeds, temperatures, and the duration of the load. 9.2.1.1 Friction Friction is a complex phenomenon that is not easy to understand without further consideration. It is ambiguous because it prevents movement as well as actually making movement possible. There is no firm hold without friction—but also no movement away from the hold! “Friction is an interaction between material areas of bodies in mutual contact. It opposes their relative movement. In the case of external friction, the areas of the substance, which are in contact, belong to different bodies but, in the case of internal friction, they belong to one and the same body.” (DIN 50 323 Part 3). Friction depends on both the state of movement of the friction partners, adhesive friction (static friction, striation) and motional friction (dynamic friction), and on the type of relative movement of the friction partners: •• Sliding friction: sliding, translation in the contact surface, relative movement of the sliding partners •• Rolling friction: rolling, rotation about an instantaneous axis in the contact surface •• Sliding friction: combined sliding and rolling with microscopic or macroscopic proportions of sliding. Friction is also dependent on the unit status of the substance areas involved: •• solid-state friction •• fluid or viscous friction

* tribos (Greek) reiben und -logie (Greek) Suffix feminine nouns with the meaning of teaching, patron, science

•• gas friction •• mixed friction.

404 | Internal Combustion Engine Handbook

6606_Book.indb 404

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.2 Lubrication

Friction in the engine is, however, undesirable because the mechanical energy already “generated” with poor efficiency is converted again into thermodynamically “lower valency” heat. By reducing the viscosity and load-bearing strength of the lubricant, this heat then impairs the function of components. In extreme cases, damage can occur because of the bearings running warm or hot. Solid-state friction is based on the following mechanisms: •• Adhesion and shearing: formation and destruction of adhesive connections in the contact surfaces •• Plastic deformation: deformation due to relative tangential movement •• Scoring: •• Sliding partners of different hardness, the rough peaks of the hard partner penetrates into the softer partner or/and •• A hard particle between the sliding partners penetrates into one or both •• Deformation: elastic hysteresis and damping •• Energy dissipation: frictional energy (mechanical energy) is transformed into heat and is lost. Static friction exists when a body is pressed onto its counterpart under the effect of a resulting force and adheres at rest. Static friction is the basis for the transmission of power between all rigid parts of the engine permanently joined by bolts, clamps, or compression fittings, such as crankcase and cylinder head, crankshaft and drive flange, or mounting bore and bearing. The critical coefficient for static friction μ R for such connections depends on material pairing, the surface characteristics and the tribological conditions (lubrication); it is therefore not a material property but rather a system property [9-28]. For sliding friction (friction of movement), the fluid friction is particularly important in engine technology; this presupposes lubrication. The relevant friction conditions for machine parts are represented in the Stribeck curve named after Richard Stribeck (1861 bis 1950) as follows:

Friction losses are recognized with the mechanical efficiency. As the quotient of the effective power Pe and the indicated power Pi the mechanical efficiency includes all the mechanical losses from the piston to the crankshaft flange. Furthermore, it also takes into account hydraulic losses (splash losses) as well as the drive powers of the ancillary machines necessary for the operation of the engine. The mechanical efficiency of engines lies—as a rated output—in the range from 75 to 90% and drops sharply at part load. 9.2.1.2 Wear “Wear is a progressive material loss from the surface area of a solid body, which is caused by mechanical effects, that is, the contact and relative movement against a solid, liquid, or gaseous counterpart” (DIN 50320). Wear impedes functions and shortens service lives but, as part of the gradual use, it is unavoidable in the operation of any machine. Wear occurs when two friction bodies (basic and mating surfaces) are moved relative to one another under the effect of force—continuously, in oscillation, or intermittently. Hereby, the structural properties, strengths, hardness form, and surface geometry all have an influence on the wear. The wear process comprises several components that occur individually, or in differing combinations, with one another: Shearing, elastic and plastic deformations, as well as boundary surface processes. As a result, particles are released from the basic and mating surfaces and, in turn, increase the wear (Figure 9.24). For engine operation, it is the wear rate that is important, that is, the speed at which the wear develops the following: •• Degressive: Running-in processes during which roughness, unavoidable in production, is smoothed out and the bearing surfaces of the partners are increased •• Linear: Normal operation during which the wear increases steadily, but only slightly •• Progressive: Self-propagating, the rate of wear accelerates so that functional faults quickly occur and lead to damage. Wear in engines is predominantly caused by the following:

•• Solid-state friction with direct metallic contact between the sliding partners.

•• Sliding wear with dry contact and with boundary and mixed friction (incomplete separation of basic and mating surfaces).

•• Boundary friction when the sliding partners are covered with traces of the lubricant.

•• Vibrational wear (typical): fretting (fretting corrosion, fretting corrosion).

•• Mixed friction as the coexistence of solid-state and liquid friction when the lubricating film is partially interrupted between the sliding surfaces.

•• Fluid friction (complete separation of basic and counter body).

•• Elastohydrodynamic lubrication: If high pressures exist between the sliding partners, the pressure in the oil film increases the viscosity of the oil. This is why—despite essentially unfavorable conditions—a sufficient minimum lubricant film thickness is obtained (e.g., Contraform contacts: gear pairs, cam/cam follower, etc.). •• Hydrodynamic lubrication: Fluid friction with complete separation of the sliding partners from one another by a lubricant film.

•• Cavitation: the formation of cavities because of localized low pressure in a fluid with subsequent implosion of the vapor bubbles. This causes damage to the adjoining surfaces; hydrodynamic properties deteriorate: exposure of solid bodies to liquids containing particles (e.g., lubricants or fuels with foreign particles or gas flows with particles (exhaust gas with combustion residues)); this causes erosion on the material surfaces. •• Wear due to impingement. •• Wear due to corrosion. Wear expresses itself in the engine as a reduction in cross section, changes in surfaces, functional deterioration because

Internal Combustion Engine Handbook | 405

6606_Book.indb 405

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

Wear mechanisms Abrasive wear

Abresive wear

Microploughing

Surface distress

Microcutting Tribochemical reaction

Figure 9.24 Wear mechanisms.

of increased clearances, reduction in overlaps, and impairment of the geometry and kinematics. The results could be increased friction, seizing as well as overload, and vibration fractures. Wear in the engine is mainly caused by the following: •• overloading •• insufficient lubrication as a result of •• lack of lubrication or/and •• unsuitable, unapproved or old oil •• unfavorable operating conditions •• malfunction or failure of engine components. Wear occurs predominantly on the following functional groups: •• Crank gear: pistons, piston rings, cylinders, bearings, and shafts •• Gearing drive: gear wheels •• Control system: cams and cam followers, valves, valve seats and valve guides, belt drives.

9.2.2 Lubrication System 9.2.2.1 Lubrication Lubrication is the coating or wetting of sliding partners with a lubricant; this can hereby refer to “liquids, gases, vapors, that is, fluids, plastic substances, and solids in powder form.” The lubricants have a function for •• power transmission •• reduction in friction and wear •• precision sealing: parts sliding on and inside one another can, in principle, be sealed purely by means of a lubricant film •• damping of impact and vibration •• reduction of noise •• cooling, that is, dissipation of friction heat

•• cleaning, that is, discharge of all kinds of particles •• corrosion protection. The lubricant is considered to be a machine element; in the bearings, it transmits the component forces by utilizing lubricating films with layer thicknesses of just a few 1000s of a millimeter. This ability can be traced back to the viscosity, that is, the lubricant’s ability to resist a change in shape and/or property. The separate fluid particles rub together; tangential stresses (shear stresses) are created at their contact surfaces, whose magnitude is dependent on the speed rate perpendicular to the flow direction dv/dz (Shear rate) and a material characteristic of the fluid, the kinematic viscosity η (viscousness), (Newton’s shear stress theory). The kinematic viscosity, in turn, is dependent on the lubricant, its temperature and pressure as well as from the shear rate (Figure 9.25). The shear stresses execute friction work in the sliding direction (dissipation work); this kinetic energy, which is transformed into heat, is then “lost.” Fluid friction has a negative effect in machine operation: It uses up mechanical energy and heats up the lubricant; this reduces the load-bearing capacity of the lubricating film. This frictional heat must be dissipated and this, therefore, necessitates additional design, construction and operational coats and efforts. In the worst case i.e. mixed friction, this will lead to wear on the sliding partner—up to and including seizure. However, a fluid could not transmit forces without inner friction. 9.2.2.2 Components and Function A lubrication system can be considered to be a system of lubricant conveying services, such as pipes or hoses, pumps, filters, heat transmitters and control and regulating elements in an arrangement relative to each other. These are particularly the: oil reservoir (oil pan), oil pump(s), oil heat exchanger, oil filter, control and regulating valves, filling mountings and monitoring the oil volumes (oil level) and oil volumetric flow (oil pressure).

406 | Internal Combustion Engine Handbook

6606_Book.indb 406

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.2 Lubrication

z v + dv t

t

v

t = h · (dv/dz)

z

x

dv

dz

9.2.2.3 Engine Lubricating Oil Circuit The intake screen of the oil pump is located at the lowest point of the oil pan, so that the oil supply is ensured even when the vehicle is at an angle. A positive-displacement pump—driven via gear wheel, chain, and synchronous belt, or mounted directly on the crankshaft—forces the engine oil through the filter and, depending on the design of the lubricating oil system, through a heat exchanger into the main oil line. A pressure relief valve, which is located on the pressure side, allows oil to bypass when the set pressure is exceeded. The control bores are designed in such a way as to level out pressure peaks and suppress pressure fluctuations. The discharged oil either runs off freely or is returned to the intake side of the pump so that it does not become enriched with air. From the pump, the oil passes into the filter. As a protection against overloading due to excessive oil pressures for example, during cold starting, the pump has a bypass valve; a nonreturn valve prevents the oil from running back when the engine is at standstill (Figure 9.26).

x Figure 9.25 Shear and shear rate [9-40].

One must differentiate between •• Fresh oil lubrication—or total loss lubrication: The oil is pumped from an oil reservoir to the individual consumers in this case. It has to be ensured that clean, cool oil is supplied to the consumers at all times. The oil consumption can be kept to a minimum with careful metering. The fresh oil lubrication method is used in two-stroke gasoline engines with fuel injection. •• Mixture lubrication: This method of lubrication is used today predominantly for small two-stroke engines. The lubricating oil is added to the gasoline in a particular ratio (1:50 and/or 1:100). The oil enters the cylinder, together with the fuel, on the intake stroke and into the engine block with the overflow. The discharged oil lubricates the bearing and the cylinder wall. Lubricating oil also enters the exhaust system with the scavenging air: This increases the oil consumption and reduces the exhaust gas quality. •• Forced-feed lubrication: Four-stroke engines and two-stroke diesel engines are generally lubricated by this method. A pump delivers the oil from a tank via a system of pipes to the consumers, and from there it flows back pressure-free to the tank. •• Dry sump lubrication: Dry sump lubrication is used for conserving space (installation space) or for special operating conditions (off-road vehicles, sports cars). A suction pump draws the oil into a separate tank and, from there, it is returned by a pressure pump to the oil system after cooling and filtration. The suction stage and pressure stage of the pump are often designed together.

The bypass valve secures the oil supply

The return valves prevent the draining of the filter.

Figure 9.26 Bypass valve and nonreturn valve (check valve) for an oil filter (Volkswagen).

The primary function of the oil filter is to protect the sliding partners from foreign particles in the oil. For this purpose, the filter must be installed upstream of the consumers so that the full oil flow passes through the filter (full-flow circuit). To relieve the full-flow filter and reduce its soiling, part of the oil is branched off from the main flow and is passed through a bypass filter—an oil centrifuge or a fine filter (Figure 9.27). Bypass filters are not an alternative to oil changes, but they can neither replace consumed additives nor filter fuel, water, and acids out of the lubricant [9-29]. If the engine oil is subjected to high-thermal loads, then it must be cooled separately, either with a water–oil heat exchanger or an air–oil heat exchanger. The oil heat exchanger is normally installed downstream of the filter to minimize the pressure loss in the filter with the still warm and, therefore, low-viscosity oil. For optimum protection of the engine, however, the filter should be located downstream of the heat exchanger—therefore immediately in front of the oil consumers.

Internal Combustion Engine Handbook | 407

6606_Book.indb 407

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

Main power circuit Surroundings valve

Main bypass circuit Surroundings valve

Full flow filter

Oil cooler

Full flow filter

Oil cooler

Throttle Pump M

Pressure control valve Smudges

Pump M

Oil pan

Smudges

Oil pan

The oil then passes via the main oil channel from the filter and/or heat exchanger into the oil consumers. The crank gear is supplied with oil from the main oil channel through bores in the crankcase intermediate walls and in the main-bearing shells. The oil passes through bores in the crankshaft to the connecting rod bearings and from there—depending on the design—through a bore on the connecting rod to the piston pin bearing (Figure 9.28).

Lubrication of control system

Riser duct to the cylinder head Main oil channel Piston cooling Lubrication of the engine Oil pump

Bypass filter

Oil filter

Strainer

Figure 9.28 Lubricating oil circuit (schematic) of a gasoline engine car (Volkswagen).

Centrifugal force has to be overcome in order to supply the oil to the main-bearing pins. The supply from the bore in the main-bearing pin to that of the crank pin or to the pinion pin bearing is enhanced by

Figure 9.27 Main flow and main/ bypass flow filtration [9-41].

the centrifugal force, or by the oscillating movement of the connecting rod. In principle, one main bearing should supply only one crank pin with oil. In high-performance engines, the oil circuit is divided into two channels; one supplies the camshaft control with oil under high pressure and the other supplies the camshaft bearing and push rod with oil under low pressure [9-30]. The oil supply for engine parts, such as belt tensioning bearings, and for engine accessories, such as exhaust-driven turbocharger, injection pumps, etc., comes directly via oil channels. Components not connected to the oil supply system, such as rocker arm contact surfaces or the flanks of gear wheels, are lubricated indirectly by the spray oil in the crankcase. Under critical conditions, separate spray nozzles ensure an adequate supply of oil. The valve guides are also lubricated by sprayed oil, whereby the oil supply to the guides is limited or metered by valve shaft seals. Today’s trend is toward more or less integrated oil lines, and short oil paths with low-pressure losses (hydraulic losses) (Figure 9.29). Engines with high-specific output no longer have a substitute for piston cooling. Lubricating oil is diverted from the main flow and then injected through injection nozzles against the underside of the piston, or into piston cooling channels for the piston cooling. Pressure-controlled valves prevent heat being unnecessarily drawn from the piston when the engine, and hence, the oil is still cold. Spraying the underside of the piston through bores in large conrod eyes is rather disadvantageous, as the cooling oil has to be additionally transported through the crankshaft. As delivery begins only when the engine is started, there is a danger that the oil consumers will not receive oil, or too little oil, during the first few revolutions of the engine. For this reason, nonreturn and/or check valves are fitted in supply lines and oil galleries in cylinder heads, from which the collected oil can flow quickly to the consumers (Figure 9.30).

408 | Internal Combustion Engine Handbook

6606_Book.indb 408

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.2 Lubrication

Cylinder cutout Cyl. 7 to 9

Cyl. 10 to 12

3/2-way valve

3/2-way valve Secondary circuit pump

Hydraulic tappets

Hydrostößel

Steuerung

Control system Camshaft adjustment

Camshaft adjustment 3/2-way valve

3/2-way valve Oil/water- heat exchanger

Crank gear Piston cooling

Filter

Thermostat and safety valve Register pump

Oil-air heat exchangers

Bulkhead Rotation Oil

Baffle Oil return duct

Air

Figure 9.31 Leading the return oil out of the cylinder heads of the Audi V6 Biturbo.

The oil in the oil pan is kept away from the crank gear by oil baffle plates (oil windage tray), so that the crankshaft cannot become immersed in oil because of the sloshing of the oil caused by the vehicle movement (Figure 9.32).

Oil pan Figure 9.29 Oil circuit of a V12 Gasoline engine with cylinder shut-off (Mercedes-Benz).

The electrically driven lubricating oil pilot pumps, normally used on larger gasoline and large diesel motors, cannot be used in motor vehicle engines because of the design complexity, the additional weight, and the cost. Low oil levels and frequent oil circulation result in increased foaming of the oil. The upper limit for the gas content is considered to be 8%. Centrifugal separators and/or low-level oil return lines are utilized to counteract foaming. As a result, the gas content can be reduced to below 4% (Figure 9.31).

Oil gallery

Figure 9.32 Oil baffle plate (oil windage trap) of a four-cylinder car engine (Opel-Ecotec).

9.2.2.4 Oil Pumps Recirculating positive-displacement pumps—gear and ring gear pumps—of various designs are used for vehicle engines:

Figure 9.30 Arrangement of the oil gallery in the cylinder head of a gasoline car engine (Ford).

Internal Combustion Engine Handbook | 409

6606_Book.indb 409

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

Filling piece Eccentricity

Eccentricity

Filling zone Gear pump

Displacement zone

Displacement zone Internal Gear Pump (Sickle pump)

External gear pumps, internal gear pumps (crescent pumps), and ring gear pumps (rotor pumps). These pumps are compact, have good efficiency, display a good intake behavior and are suitable for a wide range of viscosities of the fluids to be pumped. The change in volume necessary for pressure boosting with positive-displacement pumps is affected by the meshing of the gear wheels. The supply volume is calculated from the tooth geometry and the pump speed (Figure 9.33). The evaluation criteria for oil pumps is the delivery characteristics, efficiency, sensitivity to cavitation, noise development, installation size, weight, and manufacturing costs. Important factors are a low intake head and a rapid pressure buildup in the oil circuit. The transport losses have to be covered, and the centrifugal force in the main-bearing pins, and the flow resistances of the oil consumers (bearings), have to be overcome. The pressure losses from the pump to the cylinder head are in the order of approximately 1.5 to 2 bar. The flow velocity of the lubricating oil in the lines should not exceed 3 to 4 m/s. Oil pumps are mounted on the crankshaft, on the engine block or in the oil sump. Mounting on the crankshaft permits a simpler design and is cheaper (roughly 50% less expensive than installation in the sump), but it also forces larger impellers and higher pump speeds to be used than is really necessary. The power consumption is therefore—irrespective of the pump design type—considerably higher. Furthermore, The wobbling of the crankshaft must be compensated for—either by mounting the inner rotor in the pump house for ring gear pumps or by centering the inner rotor on the crankshaft as an alternative method [9-31]. If the pump is located in the oil sump, then the intake head is lower and the pump draws in oil better on starting up. In addition, one can select lower pump speeds (e.g., lowering to 1:1.5), which reduces the drive power required. One disadvantage hereby is the complexity of the drive with chain, synchronous belt, gear, or worm drive. The delivery characteristic of recirculating positive displacement pumps is dependent on the pump speed. With increasing pump pressure, the volumetric efficiency drops due to the leakage losses. The oil demand from the engine, however, is more or less independent of the engine speed, so that with increasing engine revolutions the difference between delivery and demand becomes even larger. The requirements

Filling zone Gear pump (Gerotor pump)

Figure 9.33 Design types of engine oil pumps (schematic).

for the individual oil consumers are different: The bearings require a specific oil volumetric and hydraulic actuators a specific pressure. For the camshaft adjustment mechanism, for example, higher deliveries are required; a dedicated secondary pump must be provided for the cylinder shutoff. The design of the pump for a minimum oil volumetric flow at (hot) idle speed—that is, low engine revolutions and low viscosity of the oil—means that with increasing engine revolutions, oil has to be bypassed above a certain counter pressure so that roughly 50% of the hydraulic energy is transformed into heat. One has to make a distinction between control valves that are controlled directly by the system pressure and valves that are controlled indirectly, that is, from the system pressure and a given control pressure, controlled regulating valve (Figure 9.34).

Figure 9.34 Directly-controlled regulating valve (Mercedes-Benz).

For additional consumers, such as an exhaust-driven turbocharger, more oil must be delivered. In addition, a reduction in the engine idle speed to lower the engine losses results in a significantly increased delivery at high engine revolutions. The disparity between the oil to be delivered at low engine revolutions, and the oil volume actually required

410 | Internal Combustion Engine Handbook

6606_Book.indb 410

1/19/16 8:39 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.2 Lubrication

at high engine revolutions, becomes even greater. For this reason, one tries to adapt the pump characteristic curve by regulating the pumps more efficiently to the oil requirement of the engine by utilizing sequential pumps, by varying the eccentricity on the pumps with internal gearing, by utilizing intake control for ring gear pumps, by axial displacement of the secondary gear on pumps with external gearing, or by isolating the pump drive from the engine revolutions by utilizing electric drives for the pump. Such solutions demand, however, a careful comparison of the design complexity and the additional weight and costs against the power savings that can be achieved. For four- to six-cylinder engines, the oil demand is 40 to 100 I/min, and 8-cylinder engines require around 100 to 120 l/min. One calculates the crankshaft main bearings for car engines with 3 l/min per bearing, the connecting rod bearing 4 to 5 1/min per bearing, the piston cooling requires 1.5 to 3 1/min per nozzle, the cylinder head about 12 1/min. However, 50% to 60% of the oil volumes are spilled off. Engines with an aluminum engine block require slightly more oil as the clearances increase with the temperature because of the greater thermal expansion. The delivery pressure is thereby approximately 5 bar. The drive powers of oil pumps for four- to six-cylinder engines lie in the range from 0.5 to 2 kW, for larger engines up to 5 kW. 9.2.2.5 Oil Monitoring The oil supply must be monitored as it is a vital component for the engine. As a rule, the pump counter pressure is utilized as a monitoring parameter. This is problematical in that it is not just the oil volumetric flow which is the physically relevant parameter, but that of the dependent parameter for the pump counter pressure, that serves as the monitoring parameter. On the one hand, this increases as a square of the flow velocity (in line with the volumetric flow) and, on the other hand, it is also dependent on the flow resistance. With increasing temperature, the resilience (viscosity) of the lubricant decreases—so that more oil has to be delivered to maintain the specified control pressure. If the line becomes clogged, then the flow resistance increases so that, despite a lower oil volumetric flow, the pressure does not decrease. If, on the other hand, the coefficient of resistance drops because of an increase in bearing play, although more oil flows through the bearings, the pressure drops and there is an incorrect

signal for “Low oil.” For this reason, the oil pressure should be monitored at the end of the line, for example, downstream of the last crankshaft bearing or in the cylinder head. Because the engine operator cannot keep an eye on the oil pressure gage the whole time, they often notices a drop in the oil pressure only when it is too late, namely from the generally disastrous consequences. For this reason, the drop in oil pressure should also be signaled acoustically. Further monitoring parameters are the oil temperature and oil level. Sensors are used for this purpose; it must also be possible to check the oil level manually using an oil dipstick with marks for the maximum and minimum oil levels. 9.2.2.6 Oil Burden The burden on the engine oil has increased continuously over the course of time: Because of smaller oil filling volumes, increasing powers as a result of higher engine revs and turbocharging, because of more compact engines (downsizing, particularly with the V-type engine), through more complex designs, longer inspection and oil change intervals and because of widely and frequently changing engine loads and speeds. Furthermore, aerodynamically optimized body forms allow the temperature in the engine compartment to increase. The oil burden can be expressed in figures with various coefficients (Figure 9.35), for example, oil-filling volume/swept displacement or oil filling volume/power. More precise information is given by the oil burden coefficient:

Oil burden =

⎛ Engine [kW] ⎞ ⎝⎜ ×Oil Change [km]⎠⎟

⎛ (Oil volume + Refill volume ⎞ ⎜⎝ per oil change interval) [l]⋅1.000⎟⎠

(9.4)

Two such coefficients are compared in [9-32]: Oil burden coefficient

kW ⋅ km/l

Ford Taurus 1949

11.5

Audi Quattro 1987

277.2

9.2.2.7 Oil Consumption The oil volume in the oil tank (oil pan) decreases during the course of the operating life because of oil losses and oil consumption. Oil losses occur when oil escapes between the rigid and moving separating parts of the engine. These can be: Connection from the crankcase to the oil pan and cylinder head, the connection from the cylinder head to the cylinder

Year

1937

1940

1951

1960

1970

1980

1990

2000

Type

Super 6

Kapitän

Kapitän

Kapitän

Commodore

Commodore

Omega

Omega

Piston displacement

dm3

2.5

2.5

2.5

2.5

2.5

2.5

2.6

2.6

Output

kW

40.4

40.4

42.6

66.2

88.2

110

110

110

Rotational speed

rpm

3600

3600

3700

4100

5500

5800

5600

5600

Oil filling

l

5

4

4

4

4.5

5.75

5.5

5.5

Oil burden

kW/l

8.1

10.1

10.65

16.55

19.6

19.1

20

20

Oil filling/swept displacement

l/dm3

2

1.6

1.6

1.6

1.8

2.3

2.1

2.1

Oil change Interval

km

2000

2000

3000

7500

10,000

15,000

Figure 9.35 Coefficient for 2.5-l Opel engines.

Internal Combustion Engine Handbook | 411

6606_Book.indb 411

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 9 Tribology

head shroud, the connections between oil filter and oil cooler, as well as leaking oil drain plugs and crankshaft seals. The actual oil consumption results from internal leaks because of burning and/or evaporation of oil. Such leaks are caused by worn piston rings or piston ring grooves, mirroring in the upper area of the cylinder tracks, excessive clearance between valve stem, and valve guide or leaks in the turbocharger. The oil consumption can only be estimated roughly because it depends on a large number of parameters, which change during the course of the engine life. 0.1 to 0.25 (0.5) l per every 1,000 km can “normally” be assumed for car engines. A constant oil level does not always indicate that there is no oil consumption, as the oil consumption can—especially with diesel engines—be “balanced out” by ingress of fuel.

9-4. Koch, F, Haubner, F, Schwaderlapp, M. 2000. Thermomanagement beim DI Ottomotor—Wege zur Verkürzung des Warmlaufs. 22. Internationales Wiener Motorensymposium, Wien.

9.2.2.8 Oil Change The oil, as a medium for lubrication, is subject to a huge number of changes during the engine operation. These necessitate the periodic replacement of the oil fill (oil change). The oil change periods have been significantly increased during the last decade. Criteria for the oil change are the content of liquid and solid foreign matter, the exhaustion of the additive effectiveness, and any impermissible changes in the viscosity. The filters also have to be changed at the same time as the oil is changed. The oil change intervals are specified by the engine manufacturers, depending on the engine type (gasoline engine, diesel engine), engine model, service life in km or miles, operating time in months, and the respective operating conditions; they vary widely for car engines from (5000 km), 15,000 to 20,000 km (30,000 km). These intervals must be strictly observed! The old oil must be disposed of in the prescribed manner. More recently, the development has been towards flexible, load-dependent oil change intervals from 20,000 to 40,000 km—which corresponds to 1 to 2 years. The decisive factor for the oil change interval is the condition of the oil. The oil quality deteriorates during engine operation through oxidation, formation of organic nitrates, reduction in the effectivity of the additives and additionally, for diesel engines, soot ingress. Determining factors hereby are the engine size, that is, the load on the engine, the operating conditions (cold start, hot running), and the oil quality. A sensor is used to monitor the operating temperature of the engine, the oil-filling level, and the oil quality, whereby the dielectric constant is regarded as a criteria for the condition of the engine oil [9-33].

9-10. Speckens, FW, Hermsen, F, Buck, J. 1998. Konstruktive Wege zum reibungsarmen Ventiltrieb. In: MTZ 59, 3.

Bibliography

9-1. Pischinger, S. 2007. Vorlesungsumdruck Verbrennungskraftmaschinen, 26. Aufl. Selbstverlag. 9-2. Affenzeller, J, Gläser, H. 1996. Lagerung und Schmierung von Verbrennungsmotoren: Die Verbrennungskraftmaschine. Neue Folge Band 8, Springer Verlag. 9-3. Pischinger, R, Kraßnig, G, Taucar, G, Sams, T. 1989. Thermodynamik der Verbrennungskraftmaschine: Die Verbrennungskraftmaschine. Neue Folge Band 5, Springer Verlag.

9-5. Koch, F, Hermsen, FG, Marckwardt, H, Haubner, FG. 1999. Friction Losses of Combustion Engines—Measurements, Analysis and Optimization Internal Combustion Engines Experiments and Modeling. Capri, Italien. 9-6. Schwaderlapp, M, Koch, F, Bollig, C, Hermsen, FG, Arndt, M. 2000. Leichtbau und Reibungsreduzierung—Konstruktive Potenziale zur Erfüllung von Verbrauchzielen. 21. Internationales Wiener Motorensymposium, Wien. 9-7. Haas, A. 1987. Aufteilung der Triebwerksverluste am schnellaufenden Verbrennungsmotor mittels eines neuen Messverfahrens, RWTH Aachen, Diss. 9-8. Koch, F, Geiger, U, Hermsen, FG. 1996. PIFFO—Piston Friction Force Measurement During Engine Operation. SAE-Paper 960306. 9-9. Koch, F, Fahl, E, Haas, A. 1995. A New Technique for Measuring the Bore Disortion During Engine Operation, 21st Int. CIMACCongress, Interlaken.

9-11. Haas A, Esch. T, Fahl, E, Kreuter, P, Pischinger, F. 1991. Optimized Design of the Lubrication System of Modern Combustion Engines. SAE Paper 912407. 9-12. Haas, A, Fahl, E., Esch, T. 1992. Ölpumpen für eine verlustarme Motorschmierung. Tagung “Nebenaggregate im Fahrzeug.” Essen. 9-13. Fahl, E, Haas, A., Kreuter, P. 1992. Konstruktion und Optimierung von Ölpumpen für Verbrennungsmotoren. Aachener Fluidtechnisches Kolloquium. 9-14. Maaßen, F. 1997. Pleuellagerbetrieb bei verschäumten Schmieröl. RWTH Aachen, Diss. 9-15. Esch, T. 1992. Luft im Schmieröl—Auswirkungen auf die Schmierstoffeigenschaften und das Betriebsverhalten von Verbrennungsmotoren. Lehrstuhl für Angewandte Thermodynamik, RWTH Aachen. 9-16. Haas, A, Stecklina, R, Fahl, E. 1993. Fuel Economy Improvement by Low Friction Engine Design. Second International Seminar “Worldwide Engine Emission Standards and How to Meet Them,” London. 9-17. Haubner, F, Klopstein, S, Koch, F. 2000. Cabin Heating—A Challenge for the TDI Cooling System, SIA-Congress, Lyon. 9-18. Bolenz, K. 1991. Entwicklung und Beeinflussung des Energieverbrauchs von Nebenaggregaten. 3rd Aachener Kolloquium Fahrzeug- und Motorentechnik. 9-19. Gorille, I. 1989. Leistungsbedarf und Antrieb von Nebenaggregaten. 2nd Aachener Kolloquium Fahrzeug- und Motorentechnik. 9-20. Henneberger, G. 1990. Elektrische Motorausrüstung. Wiesbaden, Braunschweig: Vieweg Verlag. 9-21. Schlotthauer, M. 1985. Alternativantriebe für Nebenaggregate von Personenkraftwagen. In: Antriebstechnik 24, Nr. 8. 9-22. Fahl, E, Haas, A, Esch, T. 1990. Tagung “Dynamisch belastete Gleitlager im Verbrennungsmotor.” Esslingen. 9-23. Maaßen, F. et al. 2004. Simulation und Messung am Kurbeltrieb. 13. Aachener Kolloquium Fahrzeug- und Motorentechnik. 9-24. Maaßen, F. 2007. Hybride Analyseverfahren für die moderne Mechanikentwicklung. MTZ Motortechnische Zeitschrift, 68, 6. 9-25. Deuss, T, Ehnis, H, Freier, R, Künzel, R. 2010. Reibleistungen am befeuerten Dieselmotor—Potenziale der Kolbengruppe, MTZ 05.2010, Wiesbaden, Mai. 9-26. Lückert, R, Bargende, M, Pischinger, S, Grebe, UD, Junker, HK, Esch, HJ, Göschel, B. 2010. Reibungsoptimierung—Wo hat sie noch Sinn—Forum der Meinungen, MTZ 06.2010, Wiesbaden. 9-27. Schmid, J. 2010. Reibungsoptimierung von Zylinderlaufflächen aus Sicht der Fertigungstechnik, MTZ 06.2010, Wiesbaden.

412 | Internal Combustion Engine Handbook

6606_Book.indb 412

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

9.2 Lubrication

9-28. Czichos, H, Habig, KH. 2003. Tribologie Handbuch 2. Aufl. bearb. von Erich Sntner und Mathias Woydt. Wiesbaden: Vieweg.

9-37. Reinhardt, GP. u. a. 1992. Schmierung von Verbrennungskraftmaschinen. Ehningen: expert-Verlag.

9-29. Greuter, E, Zima, S. Motorschäden, 2. Edition. Würzburg: Vogel Buchverlag.

9-38. Treutlein, W. 1997. Schmiersysteme in: Handbuch Dieselmotoren (Hrsg. Mollenhauer, K.) Berlin: Springer.

9-30. Porsche 911. Sonderausgabe ATZ/MTZ.

9-39. Norm DIN 50320 Verschleiß (Begriffe).

9-31. Eisemann, S, Härle, C, Schreiber, B. 1994. Vergleich verschiedener Schmierölpumpensysteme bei Verbrennungsmotoren. MTZ 55, 10.

9-40. Zima, S. 1999. Kurbeltriebe, 2. Aufl. Wiesbaden: Vieweg.

9-32. Eberan-Ebenhorst, CGA. 2000. von: Motorenschmierstoffe als Partner der Motorenentwicklung., in: Schmierung von Verbrennungkraftmaschinen. Lehrgang TA Eßlingen. 9-33. Warnecke, W, Müller, D, Kollmann, K, Land, K, Gürtler, T. 1998. Belastungsgerechte Ölwartung mit ASSYST. MTZ 59 7/8. 9-34. Affenzeller, J, Gläser, H. 1996. Lagerung und Schmierung von Verbrennungskraftmaschinen. Die Verbrennungskraftmaschine—Neue Folge. Band 8. Wien: Springer. 9-35. Fuller, D. D. 1960. Theorie und Praxis der Schmierung. Stuttgart: Berliner Union.

9-41. Motorenfilter, Die Bibliothek der Technik 31. Landsberg/Lech: Verlag Moderne Industrie, 1989. 9-42. Maaßen, F. 1997. Pleuellagerbetrieb bei verschäumten Schmieröl. RWTH Aachen, Diss. 9-43. Kahlenborn, M. et al. 2010. Die Wälzlagerung im Verbrennungsmotor als Maßnahme zur Reduzierung des Kraftstoffverbrauchs. 22nd International AVL Conference “Engine & Environment.” 9-44. Schöffmann, W. et al. 2010. Hochleistung und Reibungsreduktion— Herausforderung oder Widerspruch? Zukünftige Diesel- und Ottomotoren auf Basis einheitlicher Familienarchitektur. 22nd International AVL Conference “Engine & Environment.”

9-36. Gläser, H. Schmiersystem in: Kraftfahrzeugmotoren, 3. Aufl. (Hrsg. Küntscher, V.), Berlin: Verlag Technik.

Internal Combustion Engine Handbook | 413

6606_Book.indb 413

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 414

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10 Charge Cycle The term charge cycle is, hereby, understood as the exchange of the cylinder charge. Besides the control elements being located in the cylinder head, the charge cycle is substantially influenced by the connected intake manifold and/or exhaust system. These, thereby, determine the quality of the fresh gas supply and the removed exhaust gas. The quality of this process is decisive for internal combustion engines, because it substantially affects the maximum output and the maximum torque, as well as fuel consumption, exhaust quality, and running behavior. Several factors influence the charge cycle, such as the valve timing, valve lifting curves, design of the intake and exhaust systems, flow loss, wall temperatures in the ports and the combustion chamber, environmental temperature, and pressure. The quality of the charge cycle can be described by the indices air expenditure λ a and the volumetric efficiency λ l:

ll =

ll =

VGZ ⋅ rGZ VGZ ⋅Tth ⋅ pZ = Vh ⋅ rth Vh ⋅TZ ⋅ pth

(10.3)

where VG and VGZ, respectively, describe the supplied mixture volume and the mixture volume remaining in the cylinder after the charge cycle. The effective output and, therefore, the torque of an engine at a constant speed depend on the mean effective pressure. The following formula for the mean effective pressure: pme = heZ ⋅ ll ⋅ HGZ



(10.4)

produces the effective output by also considering the scavenging loss, the pressure loss, and the heat absorption during induction as follows:

mG mK + mL = mth Vh ⋅ rth

(10.1)



mGZ mKZ + mLZ = mth Vh ⋅ rth

(10.2)

The effective efficiency η eZ, and the lowest calorific value η eZ, refer to the composition of the cylinder charge after IC. The following holds true for the torque:

la =

cylinder, and by the pressure loss. Assuming ideal gas, the following holds true for the volumetric efficiency

where mG is the quantity of residual mixture (fuel mKZ and air mLZ) fed to the cylinder and mGZ is the quantity of mixture remaining in the cylinder after the charge cycle. They are in proportion to the mixture quantity mth that could theoretically fill the cylinder. The air expenditure, therefore, provides more information on the intake system and the intake process, while the volumetric efficiency characterizes the fresh charge quantity [that is, after inlet closes (IC)] actually remaining in the cylinder, therefore the efficiency of the charge cycle. These two charge quantities differ by the amount of scavenging loss flowing into the exhaust from the intake during the valve overlap phase. When valves are actuated with a low valve overlap, the approximation λ a ≈ λ l holds true, otherwise λ a > λ l holds true. In the charge cycle, an important role is played both by the heat absorbed by the fresh charge in the intake system and



Pe = i ⋅ heZ ⋅ ll ⋅ HGZ ⋅VH ⋅ n

M=

1 i ⋅ heZ ⋅ ll ⋅ HGZ ⋅VH 2⋅p

(10.5)

(10.6)

The individual factors are mutually influential. The volumetric efficiency is particularly influenced by the engine revolutions. On the one hand, the throttling loss in the lines rises with the revolutions, while, on the other hand, the gasdynamic processes play a substantial role. The efficiency in the closed combustion chamber η eZ increases with the volumetric efficiency, because friction loss at constant speeds is constant, similar to throttling loss. For this reason, η eZ also depends on the revolutions. In general, the maximum for the term η eZλ ln needs to be obtained for maximum output, and the maximum for the term η eZλ l needs to be obtained for maximum torque. This means that the two optimum values are separate from each other in two different, narrow

Internal Combustion Engine Handbook | 415

6606_Book.indb 415

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

The charge cycle in the four-stroke process is the expulsion and intake tact. These occur sequentially as a result of the displacement effect caused by the piston. The inlet and outlet of the cylinder must be periodically opened and closed by actuators. The actuators must particularly satisfy the following requirements: •• a large opening cross section •• short-time requirements for the opening and closing procedure •• a flow-promoting design •• high sealing efficiency during the compression, combustion, and expansion phases •• high resistant strength. Figure 10.1 shows two designs for actuators for four-stroke engines. Lift valves ensure a simple and secure seal, whereby the cylinder pressure reinforces the sealing efficiency. The fast acceleration and deceleration that occur during the stroke, therefore, result in a great deal stress from the inertial forces on the valve gear. In addition, the gripping forces can also be lost at high engine revolutions. Rotary-disk valves have short opening and closing times and no inertial force. However, there are problems because of the higher temperatures and thermal expansion, sealing, and operational safety (jamming and seizing). The customary design method for controlling the timing today is with lifting valves [Figure 10.1 (left)].

Figure 10.1 Lift valve and rotary-disk valve timing [10-1].

10.1.1 Valve Gear Designs

Mushroom-head valves are almost exclusively used for fourstroke engines, and sometimes for two-stroke engines, for controlling the charge cycles. The required actuating mechanism, including the valves themselves, is termed the valve gear. The complete valve arrangement is the drive through the camshaft (CS), which runs at half-crankshaft speed in fourstroke engines.

•• number of valves per cylinder (Figure 10.2) •• the position of the CS. Two Valves

Three Valves

Four Valves

Five Valves

Six Valves

Seven Valves

Number of intake valves

1

2

2

3

3

4

Number of exhaust valves

1

1

2

2

3

3

Figure 10.2 Valve arrangements.

Doubling the number of intake and exhaust valves to two is now a sufficiently tried and tested measure for improving the volumetric efficiency, and for reducing the charge cycle work by using larger cross-sectional flows. The advantages compared with more complex valve gear are increased specific performance and the reductions in the specific fuel consumption, with an enhanced influencing on the combustion. When this technical approach is pursued, the question must be raised of whether the conventional four valves per cylinder represent an absolute or relative optimum. In this regard, Aoi [10-9] investigated four- to seven-valve arrangements. The following terms must be defined in this context: •• valve area: circular area of the valve openings per cylinder •• valve opening area: lateral surface when the valves are open. Assuming the same cylinder diameter, the five-valve arrangement shows the largest valve opening area, whereby at this juncture only refers to the intake valves that have the predominant influence on the sought effect (Figure 10.3). Given the same pressure ratio, this arrangement has the highest flow rate and the best volumetric efficiency. With the same valve opening areas, the cylinder diameter could be somewhat smaller for five valves than for four valves. The more compact combustion chamber of the five valves, therefore, has advantages for output. Intake valve

Exhaust valve

14 1

12

10 2

9 8

10

8

0,32

0,35

0,35

0,37

4 V.

5 V.

6 V.

7 V.

Input valve area 1 x 1/100 [mm²]

10.1 Gas Exchange Devices in Four-Stroke Engines

The various valve trains can be divided according to

Input valve opening area 1 x 1/100 (at max. piston displacement) [mm²]

engine speed ranges, which is why, with the conventional engines [with neither a variable valve actuation (VVA) nor a controlled intake manifold], a compromise always has to be made between torque and output.

sve/dve

Number of valves

Figure 10.3 Influence of the number of valves on the intake valve area and the intake valve opening area [10-9].

416 | Internal Combustion Engine Handbook

6606_Book.indb 416

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

Nevertheless, four-valve gasoline engines have become widely accepted in passenger cars. This is primarily because the improvement attained with five valves, rather than four valves, is not worth the effort for most applications. This commences with the valve guide in the cylinder head, and then continues with the mechanical valve train components. The lack of space in the cylinder head, because of new developments such as dual ignition or direct fuel injection, also represents a problem that is difficult to solve. 10.1.1.1 Number of Valves per Cylinder Figure 10.4 shows a four-valve engine with a radial valve arrangement and a five-valve engine with a roofed combustion chamber. 10.1.1.2 Camshaft Position 10.1.1.2.1 Bottom camshaft The CS lies below the dividing line between the head and the block (Figure 10.5). Standing valves [Figure 10.5(a)] that can be actuated directly by tappets, however, produce an inferior combustion chamber (knocking and hydrocarbon emissions); this design is antiquated.

Overhead valves [Figure 10.5(b) and (c)] require a tappet, pushrod, and valve rocker arm to be actuated. The valves can be arranged in parallel [Figure 10.5(b)] or V-shaped [Figure 10.5(c)]. 10.1.1.2.2 Top camshaft CSs above the head/block dividing line are generally used in modern, fast-running gasoline and diesel engines. The valves can be actuated through a valve lever and/or rocker arm, cam follower, or tappet (Figure 10.6). This is advantageous in that dispensing with the pushrod and tappet, or the valve lever or cam follower, reduces the unevenly moved mass and the elasticity of the valve gear. In today’s conventionally used valve gears, the transmission elements (rocker arm, cam follower, valve lever, tappet, and so on) are pressed under spring force (valve spring) against each other or against the cam when the valve is open. This grip can be lost at high engine revolutions. This does not hold true for desmodromic valves where lifting from the control cam is avoided by means of a second cam (Figure 10.7); this makes valve springs unnecessary. Valve clearance is also required here. This solution has not been implemented because of the effort and expense involved (manufacturing and servicing).

Figure 10.4 The four- and five-valve engine [10-1].

Rocker arm

Push rod

Tappet A

Camshaft

B

C

Figure 10.5 Valve train with a bottom CS. (a) Standing valves. (b) and (c) Overhead valves.

Internal Combustion Engine Handbook | 417

6606_Book.indb 417

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Valve levers/ Cam follower

Rocker arm

Figure 10.7 Desmodromic valve [10-5].

10.1.2 Components of the Valve Train 10.1.2.1 Camshaft The CS transmits the torque introduced from the CSs drive through the individual cams to the tappets. In addition to the valve train cams, the CS can have additional cams for actuating injection pumps (single pumps and pump-nozzle elements) or engine braking systems. Based on their manufacturing features, CSs can be divided into •• cast •• forged •• assembled CSs. CSs must be subjected to a heat treatment after they are formed to give them the required strength and tribological

Tappet

Figure 10.6 Valve train with an upper CS [10-6].

properties. In the case of clear chill casting, the CS is hardened in one working step by rapid cooling (quenching) the casting mold. In the case of centrifugal casting, the metal flows into a rotating permanent mold and hardens under the effects of centrifugal force. The CS is usually cast hollow to save weight. In the case of assembled CSs, the cams are manufactured separately from the shaft body and permanently joined later. Manufacturing them separately allows the materials to be adapted to function manufacturing processes and stress. Colddrawn structural steels (for example, St52K) or alloyed steel (for example, 100Cr6) can be used. Case-hardened steels (for example, 16MnCr5) are used for the cams. The accepted forms of joining in series production are friction-lock connections by shrinkage, or by the hydraulic expansion of the tube from internal pressure, and a shape-type process. With shape-type connections, projections are created by roughing the tube at the attachment locations. The cam is given an internal splined profile and is pressed on with controlled force (KRUPP-PRESTA procedure). The additional advantage of the assembled CS is the potential small cam spacing (multiple valves) and the weight reductions of up to 40%. However, the transmittable torque is limited by the technical joining methods. Multipart CSs are frequently used for large engines. Individual CS segments are screwed together to create CSs for engines with different numbers of cylinders. The bearing positions for the CS plain bearings, used on all CSs, are ground directly on the tube in the case of assembled CSs. The cam profile preparation is also created by grinding. With a rolling contact, a negative radius of curvature (concave cam) of the cam profile is necessary to attain the desired valve train kinematics. With set minimum grinding disk diameters, the negative radius of curvature can limit the valve train kinematics. Extremely small curvature radii can be created by belt sanding the cam profile. The alternating loads from injection pumps and the valve train generate flexural and torsional vibrations in the elastic CS. Torsional vibrations, in particular, generate angular deviations and, hence, deviations in the control and injection time between the first and the last cams. To minimize vibrations, the CS should be very rigid with comparatively low

418 | Internal Combustion Engine Handbook

6606_Book.indb 418

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

inertia (hollow shaft). Torsional vibration resonances can be calculated from the natural torsional frequency of the CS that arises within its speed range. Particular attention must be given to resonances that arise in low-seated engines with long CSs. In certain unfavorable cases (long cylinder benches), torsional vibration dampers must be placed on the free end of the CS.

Valve spring collar Valve keeper units

Valve stem seal Valve spring

10.1.2.2 Camshaft Drive In addition to rare special designs (vertical shaft drive and preengaged drive), there are three conventional options for driving the CS with the crankshaft:

Valve shaft guide

•• gear wheels •• chain with a gear

Valve

•• synchronous belt. Gear wheels are mainly used for a bottom-mounted CS; the design becomes very complicated when gear wheels are used for an overhead CS. Chains and gear wheels, as well as synchronous belts, are exclusively used for overhead CSs in today’s designs (Figure 10.8). A tensioning device is necessary for both types of drive units. Synchronous belts made of plastic with long fibers are less noisy and cheaper than chain drives. While chains have to be lubricated, the synchronous belt needs to run in an oil-free area.

Camshafts Slip rail Chain tensioner

Auxiliary drive Crankshaft Chain drive

Camshaft Timing belt tensioner

Valve seat insert Figure 10.9 Valve and valve components [10-8].

The valve shaft seals, with spring-loaded elastomer sleeves, must, on the one hand, provide sufficient shaft lubrication and, on the other hand, also prevent the penetration of excessive lubricating oil. Light-metal cylinder heads are provided with pressed-in valve guides and valve seat rings (made of special bronzes or alloyed cast iron) that are also frequently used in gray cast iron cylinder heads. Valves are components that are subjected to high thermal and mechanical loads, as well as corrosion. The mechanical stresses arise from the valve head bending under pressure because of ignition and forceful contact when closing (impact). These stresses can be countered by providing the head with the appropriate strength and shape. The valves, with their large surface, absorb heat from the combustion chamber. The top of the exhaust valve is also heated during opening by the exiting hot exhaust. In the valve, the heat primarily radiates to the valve seat, and a small part flows over the skirt of the valve guide. Intake valves can reach the temperatures of 300–500°C, and exhaust valves reach 600–800°C. A typical temperature distribution is shown in Figure 10.10. If the seal on the valve

Auxiliary drive Crankshaft

Timing-belt drive

700

685

635

660

690

685

620

675

10.1.2.2.1 Valve Figure 10.9 shows a valve with its installed components. The seat surfaces and the shaft ends of valves, made of heat- and wear-resistant alloys (such as Cr–Si or Cr–Mn steel), are either only hardened or reinforced with hard metals. The valve is chromed on the skirt.

660

675

700

Both the drive units must be encapsulated for protection, as well as to avoid lubrication loss.

620

650

650

Figure 10.8 CS drives [10-1].

565

590

Figure 10.10 Temperature distribution in the exhaust valve [10-1].

Internal Combustion Engine Handbook | 419

6606_Book.indb 419

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

seat is not perfect during the combustion phase, then local overheating and melting occur that cause the valves to fail. To improve the heat conduction through the skirt, it is designed hollow and filled with sodium when it has to meet particularly high demands [Figure 10.11 (left)]. The movement of the sodium, that is liquid at temperatures above 97.5°C, enhances the transfer of heat. This can, therefore, lower the valve temperature by up to 100°C. The seat can be reinforced by welding on stellite to reduce the wear [Figure 10.11 (right)].

only weakened slightly, because the round counter-bore has a slight notch effect. 10.1.2.2.2 Valve spring Cylindrical or conical steel springs and pneumatic springs can be used as valve springs (Figure 10.13 and Figure 10.14). They primarily differ in the way they transmit force along the spring path. Whereas the cylindrical steel spring usually has a linear characteristic curve, the conical steel spring has a progressive characteristic, and the pneumatic spring has a strongly progressive characteristic curve (Figure 10.15). The progressive characteristic curve can achieve a favorable behavior with high revolutions. Because of the expense and the required supply of compressed air, pneumatic springs have only been used to date in motor sport.

Filled with sodium Seat facing

Figure 10.11 Exhaust valve cooled with sodium [10-1].

The material for the valves must be very heat resistant and scale resistant. Special steels can be used, as well as titanium. Valve seat inserts are installed in the cylinder heads to counteract wear. A seat ring must always be provided for light metal cylinder heads (alloyed centrifugal casting and austenitic cast iron in special cases with heat expansion coefficients approximately as high as light metal). In the case of engines subject to high stress, and for exhaust valves in gray cast iron-cylinder heads, seat rings made of alloyed centrifugal castings are used. The valve seat inserts are either pressed in or shrunk on. To prevent local temperature differences in the valve head, as well as uneven wear, the valve should slowly rotate during operation. This movement can be supported by valve rotating mechanisms between the valve spring and the cylinder head (rotovalves, rotocaps, and rotocoils) that convert the pulsing spring force into small rotary movements. The rotary movements are transferred to the valve through the valve spring and the spring collar. The spring collar is affixed to the valve stem with clamping cones (Figure 10.12). The skirt is, thereby,

Figure 10.12 Affixing the spring collar with clamping cones [10-5].

Cylindrical spring

Conical spring

Figure 10.13 Cylindrical and conical steel springs.

10.1.2.2.3 Valve rockers and valve levers 10.1.2.2.3.1 Rocker arm A rocker arm with push rods is used for bottom-mounted CSs and with valves arranged in a V shape for overhead CSs. Because of the strong contact pressure exerted on the pivot, the bearing must be especially rigid. For the rocker arm ratio i = l2/l1 (Figure 10.25), the values of 1–1.3 are recommended as a compromise for less surface pressure on the tappet, less moved mass, and high rigidity. The force of the rocker arm is transmitted to the valve along an axial path, as much as possible, to keep lateral forces from acting on the valve stem and, thus, preventing increased wear of the valve guide. At one-half of the valve stroke, the center of rotation of the rocker arm should be perpendicular to the valve axis at the height of the shaft end to attain the least possible displacement of the rocker arm and valve in relation to each other (favorable sliding conditions). The force-transmitting spherical or cylindrical surface should be applied to the rocker arm but not the valve. The rocker arm end is hardened for the reasons of wear. Figure 10.16 shows valve rocker designs. Rocker arms are usually cast or forged. Economical and light, but less rigid, are rocker arms pressed from sheet metal. It is advantageous to

420 | Internal Combustion Engine Handbook

6606_Book.indb 420

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

Valve closed

Valve open

Pressure approx. 95 bar Temperature approx. 300 °C

Discharge

Aeration approx. 15 bar with check valve

Overpressure valve

Figure 10.14 Pneumatic spring.

Force

Linear spring characteristic curve (cylindrical spring) Progressive spring characteristic curve (conical spring) Strong progressive spring characteristic curve (air spring)

A

E

Spring deflection

Figure 10.15 Spring characteristic curves.

Figure 10.17 Valve gear with rocker arms and hydraulic valve clearance compensation [10-10].

set the valve play at the resting rocker bearing. With a forged rocker arm, the setting screw is normally on the rocker end, which increases the moved mass of the valve gear. Figure 10.17 shows a valve gear with hydraulic valve clearance compensation integrated in the rocker arm. The compensation element is supplied with lubricating oil through the rocker arm shaft and the holes in the rocker arm. Valve play adjustment

Forged / Molded

Figure 10.16 Rocker arm [10-5].

Pressed from sheet metal

10.1.2.2.3.2 Valve levers (cam follower) The valve lever is exposed to a much lower degree of force than the rocker arm. The influence from changes at the bearing point is less; it is possible to install an automatic valve play adjustment system in the lever bearing in valve levers, without substantially changing the overall elasticity of the valve gear. The design of two valve levers is shown in Figure 10.18. A possibility for reducing friction loss, especially at low speeds, is provided using the so-called roller cam follower. A roller finger follower on a needle bearing is, hereby, used at the contact point between the cam follower and the CS. This can reduce the moment of friction of the valve gear by as much as approximately 30% in comparison to a sliding rocker arm arrangement (Figure 10.18). Figure 10.9 to Figure 10.15 show a spread range that illustrates the advantages of the roller cam follower in regard to reduced friction. The reduction of the valve gear friction, however, also reduces the damping of the oscillating torque introduced from the cam force and, hence, increases the load

Internal Combustion Engine Handbook | 421

6606_Book.indb 421

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

10.1.2.5 Tappet The tappets in push rod engines [Figure 10.5(b)] must guide the push rods and absorb the transverse force that arises from the sliding of the cams. In overhead CSs with a pushrod drive (Figure 10.6), the pushrod has to keep the lateral force away from the valve guide. Normal tappet designs for push rod engines are shown in Figure 10.19. Flat-based tappets and bucket tappets can be dismantled from both upwards and downwards. Roller tappets are used for maximum loads (diesel engines are subject to greater loads).

Hydraulic valve clearance compensation (not moving)

Roller actuator with needle bearing

Rolling contact

Pan tappet

Flat tappet

Mushroom tappet

Roller tappet

Figure 10.19 Tappet for valve gear [10-5].

Sliding contact

Figure 10.18 Valve gear with valve level (cam follower).

on the CS drive. Under certain circumstances, the subsequently required stronger chain or belt tensioning devices (tensioning pulley, slip rail, and damping elements) can compensate for the friction advantages gained in the valve gear.

Figure 10.20 shows a push rod, which is almost exclusively used for overhead CSs with a tappet drive. The tappet diameter is determined by the maximum tappet speed. The surface pressure between the CS and the tappet is determined by the cam width. Because the cam and the tappet must glide on each other under high surface pressure, the materials for the two elements must be harmonized. The combination of hardened steel/white-hardened gray cast iron is quite suitable. Frequently, one allows the tappet to be rotated on its own axis to prevent uneven wear. For this reason, it is offset against the center of the cam by 1–3 mm. In addition to rigid tappets,

Adjusting disk (valve play adjustment via selecting the disk thickness) Lifting tongue Valve body

Figure 10.20 Pushrod without hydraulic compensation.

422 | Internal Combustion Engine Handbook

6606_Book.indb 422

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

Oil reservoir (4) External housing Leak gap (5) Internal housing Valve cap (3) Return spring Oil inlet borehole

Oil carryover Piston Valve ball (1) Valve spring (2) Oil supply groove

Figure 10.21 Pushrod with hydraulic compensation.

there are also tappets with automatic play adjustment (refer to Figure 10.21). The play is, hereby, kept constant with the amount of oil retained in the high-pressure chamber. If the valve play is too great, then oil flows through the ball valve (1)–(3) from the reservoir (4); if it is too low, then the excess oil exits through the leak gap (5). In addition to easier servicing by dispensing with the play setting, this system also prevents noise. The disadvantages are, however, the large mass, low rigidity, and problems starting the engine after long periods of rest because of insufficient oil supply. Today, with the exception of engines for motor racing, bike racing, and high-performance diesel engines, tappet engines almost exclusively use tappets with automatic play adjustment. Engines with valve levers, valve rockers, or cam followers have the hydraulic valve clearance compensation adjusted with additional inserted elements.

Pre-tappets

x

Velocity



dx dx da NW = ⋅ = x ′ ⋅ w NW dt da NW dt

where ω NW is the angular speed of the CS.

(10.7)

aNW

x = x ′ vNW

aNW

Acceleration

The valves must open and close quickly for a good charge cycle. However, the inertial force of the valve train needs to be considered in the design. Figure 10.22 shows a typical path for the cam stroke, the cam speed ẍ, and the cam acceleration ẍ over the angular displacement of the cam angle. These quantities correspond to the respective quantities of the valve movement. The cam stroke, or the cam contour, is composed of the cam lobe and the cam body. In the area of the cam lobe, the stroke speed x is slow so that the common changes in the valve play do not generate strong impact pulses. The main cam body determines the opening cross section for the charge cycle. The completion is formed by a deceleration corresponding to the cam lobe. The stroke characteristic is a function of the CS angle α NW. The stroke speed ẋ is, therefore, taken as

x = x (aNW)

x

10.1.3 Kinematics and Dynamics of the Valve Train

x! =

Main tappets

x

x = x ′′ v2NW

aNW

nNW = Constant

Figure 10.22 Kinematics of the cam [10-1].

At a constant CS angular speed, the following results for the stroke acceleration x!! =

2 d2 x d 2 x da NW 2 = ⋅ = x ′′ ⋅ w NW 2 2 dt da NW dt 2

(10.8)

In Eq. (10.7) and Eq. (10.8), x′ and x″ are speed-independent functions that are determined only by the geometry of the cams. The cam shape is also influential for the characteristic of the valve movement. Figure 10.23 shows the relationship between the stroke characteristic and the cam shape in connection with a flat-based tappet. In Figure 10.23, the rotation of the cam has been replaced by the swing of the tappet in the

Internal Combustion Engine Handbook | 423

6606_Book.indb 423

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

opposite direction with a standing cam. The cam shape is the envelope curved end of the tappet-sliding surface. One can replace the cam drive with thrust cranks for a kinematic investigation, whereby their articulation corresponds to the respective contact point B of the curvature mean point M of the cam contour. x′ (rotated vector) and x″ depend on the crank length (rM) and the position of the currently valid thrust crank.

l1

l2

JK mF mV

Fred

mSt

x mred

x

R+

x

x A - Valve train

B

mred

FN

mStö

aNW

FN

B - Replacement system

x ′′

Figure 10.24 Rigid valve train. x′ rM

If one reduces all the quantities on the cam side, then the equation for the cam force is

M

R

FN = Fred + mred ⋅ x!!



(10.10)

Equation (10.10) corresponds to the replacement system in Figure 10.25. The following condition must be fulfilled for the gripping force: Figure 10.23 Kinematics for the tappet stroke.

One can clearly see that the distance of the cam contact points B from the middle of the tappet is proportional to the speed. The tappet diameter must, therefore, be adapted to the maximum stroke speed. It is important that friction is always present between the cam and/or tappet or cam follower. Furthermore, there must also be friction between the valve and tappet and/or rocker arm or cam follower for the valve to follow the cam stroke. The valve stroke may be recalculated corresponding to the rocker arm or the valve lever ratio i = l2/l1. The force between the cam and the tappet must be evaluated to test the grip. The inertial force and the spring force must also be considered. With a valve gear corresponding to Figure 10.24, the following results for the force on the cam FN: 2 ⎛l ⎞ l ⎡ J m FN = FF ⋅ 2 + ⎢ mStö + mSt + 2K + mV × ⎜ 2 ⎟ + F l1 ⎢⎣ l1 2 ⎝ l1 ⎠

Fred + mred ⋅ x!! > 0



⎛l ⎞ ⋅⎜ 2 ⎟ ⎝ l1 ⎠

2

⎤ ⎥ ⋅ x!! (10.9) ⎥⎦

(10.11)

or otherwise expressed as x!! <

Fred mred

(10.12)

The characteristics of the tappet acceleration determine whether or not the valve lifts. Figure 10.25 shows the acceleration of the cam stroke ẍ over the cam angle displacement of the cam for two CS rotational speeds nNW1 and nNW2. If the characteristic of ẍ intersects the curve −Fred/mred, then the gripping force has been interrupted. This can only occur in the deceleration period of the main cam. There is always a revolution speed above which lifting can occur. Pre-tappets

Main tappets

nNW2 > nNW1

where FF = valve spring force

Acceleration

mF = mass of the valve spring (only one half is used, because it rests on one side of the cylinder head)

x nNW1

JK = moment of inertia for the rocker arm

aNW

mStö = tappet mass mSt = push rod mass mred = reduced mass

–

Fred mred

Displacement

mV = mass of the valve Fred = reduced spring force.

Figure 10.25 Lifting conditions for a valve train.

424 | Internal Combustion Engine Handbook

6606_Book.indb 424

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

The valve train should, therefore, be designed so that the valve is not lifted at maximum CS speed (= ½ crankshaft speed). This is attained with a small moved mass (mred) and high spring force.

3

p 2

Aö

10.1.4 Design of Gas Exchange Devices in FourStroke Engines 10.1.4.1 Charge Cycle Energy Loss The loss of charge cycle energy causes a reduction of the indicated work and, hence, the indicated output; this causes the specific consumption to rise. It occurs either as expansion energy loss at the beginning of the charge cycle [between exhaust opens (EOs) and BDC (before dead center)] or as increased pumping work during the charge cycle. Pumping work represents the work required by the piston to draw a fresh charge during the intake cycle into the combustion chamber and to expel the exhaust from the combustion chamber during the exhaust cycle. Correspondingly, the pumping work can be divided into intake work and exhaust work. In calculations, the pumping work is represented with the aid of the average effective pump pressure. During intake, pressure losses occur at several locations, which can lead to increased intake work: (1) flow loss when the medium enters and leaves the intake system; (2) pressure drop in the lines because of bends and rough surfaces; (3) pressure loss in the air filter, at the airflow sensor, at the throttle valve; and (4) loss at the valves. Assuming quasi-stationary flow, the overall pressure loss in the intake system toward atmospheric pressure can be described by the sum of the individual losses of the various components (Heywood):

4

Es

Eö

1′

pu

1 As

OT

UT

v

Figure 10.26 Intake energy loss without throttling under full load [10-1].

Partial load n = 2000 1/min pme = 1 bar

p bar

Aö

Eö 1 Es 0,45 0

As OT

UT

Figure 10.27 Intake energy loss with throttling under partial load [10-1].

2



⎛A ⎞ ∆p = ∑ ∆pi = ∑ xi ⋅ r ⋅ vi2 = r ⋅ vK2 ⋅ ∑ xi ⋅ ⎜ K ⎟ ⎝ Ai ⎠ i i i

(10.13)

where ξ i is the loss coefficient, vi is the local flow speed, and Ai is the smallest flow cross section of the respective component. This makes it clear that, to achieve less pump work in the charge cycle, greater flow cross sections are desirable, and that the pressure loss depends on the average piston speed vK and/ or the rpm, that is, it increases with engine revolutions. The cross-sectional flow can be increased by increasing the size of the geometric opening cross section (valve stroke, valve seat ring diameter, and number of valves). Increased intake work primarily arises from throttle control (TC) while operating under a partial load. In gasoline engines, the amount of charge in the partial-load range required for the desired load is attained by adjusting the throttle valve, that is, by changing the flow cross section. The piston must aspirate corresponding to the pressure loss at this point against a lower pressure than the atmosphere (the absolute pressure of the intake manifold pipe drops). In the idling range, the increased intake work can be up to 30% of the work accomplished by the engine (Figure 10.26 and Figure 10.27).

The charge cycle energy losses are, respectively, represented by the hatched area (without throttling) and/or by the hatched and dotted areas (with throttling). From the time point EO to BDC, there is a resulting loss of expansion work. The expulsion of the exhaust gas produces loss from exhaust work. Upon the induction of the fresh charge with a vacuum, intake work is expended. Throttling loss occurs during throttling, and must be compensated in addition to the expansion and flow losses from the actuators. In a best case scenario, the losses (hatched area) to the left of the compression line in Figure 10.27 are avoided by controlled intake without throttling, such as by infinitely variable valve trains (VVTs). In this case, the supplied amount of charge is controlled by adjusting the valve timing (IC is very important in this instance) or—depending on the variability of the system—by the variable stroke of the intake valve. Increased pump work results not only from the intake of fresh air at a vacuum but also during the expulsion of exhaust gas. Although the combustion gases are at a higher pressure than atmospheric pressure, they cannot leave the cylinder at the right time through the outlet and the exhaust system without work being done by the piston (that is, before the end

Internal Combustion Engine Handbook | 425

6606_Book.indb 425

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

of the expulsion cycle). The exhaust counter-pressure has a decisive influence on this process (Figure 10.28).

L2

D2

5

p [bar]

D1

L1

6

Motor

2

V

4 3 L1 D1 L2 D2 V

1 0

0

1

2

3

4

5

6

7

8

9

Figure 10.28 Exhaust work.

The time point EO is of great importance from the point of view of the charge cycle. The time point always represents a compromise. When the EO is later, more expansion work is gained; therefore, consumption is lowered. At higher revolution speeds, however, greater exhaust work is required for the exhaust gas to leave the cylinder within the shorter period. This leads to increases in consumption. With an early EO, less exhaust work is necessary, because the cylinders can be purged more easily and quickly. However, expansion work is lost, and the thermal load on the exhaust valve increases (Figure 10.29). 6 5

1

2

LR Resonance pipe length (2) DR Resonance pipe diameter (2) VR Resonance volumes, Resonance tank (1) VA Compensation volumes, Collector tank (3) DK Throttle Valve

Figure 10.30 Schematic diagram of the ram tube charging and resonance charging.

10.1.4.2.1 Ram tube charging The ram tube effect is based on the vacuum wave triggered by the descending piston, which travels in the intake manifold pipe opposite the direction of flow to the common plenum chamber, and is reflected there at the open tube end. The overpressure wave that arises in this manner increases the cylinder charge by increasing the pressure gradient through the intake valve. This effect is particularly useful briefly before the intake valves are closed while the piston is ascending. The pressure wave prevents the expulsion of the fresh charge from the combustion chamber into the intake manifold pipe and/or generates a charging effect. Corresponding to the acoustic design, the pressure wave requires the following time with the speed to leave and return in the ram tube:

4

t=

2 ⋅ LIntake a

(10.14)

The inlet time [from inlet opens (IOs) to IC] should average one-third of the time required for an engine revolution at a given speed:

3 2 1 0

Vibrating pipe length Vibrating pipe diameter Aeration pipe length Aeration pipe diameter Air distributor (Index = Type)

10 11 12

V/Vc 6000 min–1, 9,4 bar, Full Load, AÖ 54v.UT, AS 6 n. OT

1

3

DK

2

p [bar]

Resonance charging

Ram pipe charging

0

1

2

3

4

5

6

7

8

9

10 11 12

V/Vc 6000 min–1, 9,4 bar Full Load, AÖ 54 v. UT, AS 6 n. OT

Figure 10.29 Expansion energy loss.

10.1.4.2 Intake Systems Both in an intake system and in the exhaust system, gas-dynamic processes occur that are based on the periodic excitation of the piston and natural frequency of the system. These can be used to improve the charge cycle process. These gas-dynamic effects in the intake system can be fundamentally differentiated into ram tube and resonance effects. A schematic illustration of both the intake manifold systems is shown in Figure 10.30.

t≈

1 3⋅n

(10.15)

This, therefore, allows the optimum length of the induction pipe to be determined at a given speed n:

LIntake ≈

a 6⋅n

(10.16)

Hence, the intake manifold pipe length is the determining quantity for the ram tube effect. Corresponding to the acoustic design, there is a preferred speed for each intake manifold pipe length, for which there is maximum air expenditure. This has been fundamentally demonstrated in engine tests, in which only the manifold intake pipe length was varied [10-11]. Figure 10.31 shows the influence of the intake manifold pipe length on the maximum mean pressure. A shorter intake manifold pipe shifts the torque peak in the direction of higher speeds and vice versa.

426 | Internal Combustion Engine Handbook

6606_Book.indb 426

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

eff. Mean pressure pms [bar]

L1 = L1 = L1 = L1 =

220 mm 420 mm 750 mm 1150 mm

In real engine operation mode, however, the influence of the intake manifold pipe length is more complex and partially overlaps with the influence of other intake-side parameters. For example, in addition to the pressure characteristic before the closing intake valve, the charge cycle is strongly influenced by the formation of a free vibration in the intake manifold pipe in the period between IC and IO in correlation with the intake vibration that forms in the period between IO and IC. A fixed intake manifold pipe length is, therefore, only advantageous within a specific revolution range. At higher revolution speeds, a short intake manifold pipe length is desirable, and at lower revolution speeds, a long pipe is desirable. Engines are, therefore, designed with a multistage manifold pipes, that is, the intake manifold pipe length is adapted to the engine speed (Figure 10.32). When the throttle valve is open, the intake wave coming from the cylinder is reflected at this point (high revolutions from 4000 rpm). At speeds up to 4000 rpm, the throttle valve is closed (long intake manifold pipe). Figure 10.33 shows a further developed three-stage controlled intake manifold

D1 = 40 mm, VA = 1,2 dm3, L2 = 420 mm, D2 = 70 mm

12 11 10 9 8 7

1000

2000

3000

4000

rpm n [min–1]

5000

6000

Figure 10.31 Influence of the intake manifold pipe length L1 on the maximum mean effective pressure over the speed.

300

Engine torque [Nm]

270 240 210 180 150

Flaps closed, long intake manifold Flaps opened, short intake manifold

120 0

0

2000 4000 3000 Engine rpm [min-1]

8000

Figure 10.32 Intake system with two-stage controlled intake manifold pipe; principle (Audi V-6).

Flap two

Flap one Under 3360/min: both flaps closed

3360 to 5200 /min: Flap one opened

Over 5200/min: Flap two opened

Figure 10.33 Intake system with three-stage controlled intake manifold pipe.

Internal Combustion Engine Handbook | 427

6606_Book.indb 427

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

pipe. Recently, stage-less variable intake manifold pipes have also been used. While the time of the waves depends on the intake manifold pipe length, the amplitude of the wave is influenced by the intake manifold pipe cross section. The flow speed in the intake manifold pipe rises with the rpm, so that the amplitude also rises correspondingly (Figure 10.34). Sufficiently high amplitudes to yield a corresponding recharging effect at low speeds can be created with a small intake manifold pipe cross section. At high speeds, however, the cylinder charge falls with a small flow cross section. A good cylinder charge at high speeds, therefore, requires a large intake manifold pipe cross section.

Flap opened

Air Expenditure

1,0

Flap closed

0,9

0,8

0

1000

2000

3000

4000

5000

6000

rpm [min–1]

Figure 10.36 Influence of the port closing on the air expenditure [10-1].

Air Expenditure

D1 < D2 < D3

Rotational speed

Figure 10.34 Air expenditure as a function of the pipe diameter [10-7].

When there are several intake valves, such as those used in four-valve engines, the intake manifold pipe cross section can be adapted as a function of the load and the rpm by closing a port (Figure 10.35). At low rpm and a low load, only the primary port is effective. As the rpm and the load increase, the secondary port is actuated. Primary duct Secondary duct

10.1.4.2.2 Resonance system With a resonance charge, the charge effect is generated by an oscillating container-pipe system. The periodic intake cycles of the individual cylinders, therefore, cause oscillating pressure through short intake manifold pipes in the container, which then increase the pressure gradient between the inlet port and the combustion chamber at the beginning and end of the intake phase. This oscillating pressure, which substantially increases the air expenditure, has a definite maximum when the excitation by the cylinder corresponds to the natural angular frequency of the container-pipe system. An optimum condition for exciting oscillations is when the individual intake phase is offset by a 240° crankshaft angle (CA), that is, by three cylinders per resonance container. When the intake valve is open, the system vibrates similar to a Helmholtz resonator. Vibrations, hereby, arise when the air column in the inlet port moves against the rigid air in the cylinder, and the entire system functions like a springmass system. The air in the cylinder can be regarded as the spring, and the air column can be regarded as the mass. The natural frequency of a Helmholtz resonator can be determined as follows: f=

Injection valve

Cut-off flap

Figure 10.35 Intake system with port closing [10-1].

At lower rpm, the cylinder charge is better when the shutoff valve is closed (Figure 10.36). In addition, introducing a specific charge motion (turbulence) can be generated with the inflow to improve the mixture. This increases the efficiency during the partial-load operation, especially when the engine is operated with a lean mixture (lean-burn engine).

a 2⋅p

AIntake LIntake ⋅VBE

(10.17)

where AIntake is the cross-sectional area of the induction pipe and VBE is the container volume. In transferring the Helmholtz equation (10.17) to the internal combustion engine, Engelman [10-12] used the compression chamber for the volume VBE plus the half stroke volume of a cylinder, and created the following simple relationship for the resonance speed nres in a system consisting of a cylinder with an intake manifold pipe: nres =

15 ⋅ a AIntake LIntake ⋅ (Vc + 0.5Vh ) p

(10.18)

This, therefore, allows the natural frequencies of the Helmholtz resonator effect to be precisely described for a cylinder with an intake manifold pipe. If there are several

428 | Internal Combustion Engine Handbook

6606_Book.indb 428

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

Vibration pipe

Main collector

Resonance pipe

Solenoid valve Vacuum pressure can DKS

calculations. The air mass controlling is implemented with the central throttle valve. The gain in torque from such a system is shown in Figure 10.38. 110

Gain via resonance system

100 90

Torque [%]

cylinders, the overlapping of the waves influences the results, and describing the phenomenon is confronted with many larger obstacles. This vibrational behavior is also noticeable with closed intake valves. The manifold volume, hereby, acts as the resonance volume. With this design (volume), the natural frequency of the system can be varied, so that it increases the air expenditure at certain speeds—when an overpressure wave arrives in the intake manifold port—briefly before the IC of the intake valve. The resonance charge is of particular importance in combination with turbocharging to compensate for the poor low torque at low revolution speeds. In addition, there is the possibility to combine ram tube charge and resonance charge for six- and twelve-cylinder engines. Therefore, at low speeds, the resonance vibration in the container is exploited, while short intake manifold pipes at higher speeds contribute to the increase in the air expenditure as a ram tube system. Figure 10.37 schematically illustrates a combined ram tube and resonance charge with a six-cylinder engine.

80 Switching flap opened

70 60

Switching flap closed 0

1000

2000

3000

4000

rpm [min–1]

5000

6000

7000

Figure 10.38 Torque characteristic of an inline six-cylinder gasoline engine with a resonance system [10-7].

10.1.4.3 Exhaust Systems The exhaust system fulfills three tasks: (1) it influences the power characteristic of the engine; (2) it reduces exhaust noise; and (3) it reduces the pollutants in the exhaust with an installed catalytic converter. These tasks cannot be fully separated from one another. The noise damping always influences the power characteristic, generally in an undesirable manner; conversely, maximum performance exhaust systems are often too loud. The sound pressure at the exhaust valve lies between approximately 60 and 150 dB(A). This needs to be reduced to the legally prescribed value (Figure 10.39).

HFM Engine exhaust LLS

Resonance control flap Resonance collector

Turbulence collector Turbulence borehole Ø: 6 mm

Figure 10.37 Illustration of the intake manifold system of an inline six-cylinder engine [10-7].

The adaptation is implemented by opening or closing the resonance control valve. In the torque position, the resonance control valve is closed, so that two three-cylinder intake manifold systems with long pipes are active. In the output position, the resonance control valve is open, and the intake manifold module works for all the six cylinders as a ram tube system, which is then fed from the entire upper manifold range with short ram tubes. The cross section and lengths can be tailored and optimized for these effects by using 1-D

Gas vibration

Pu 1 bar Pü As

Aö

Pü = Excess pressure Pu = Vacuum pressure

Vacuum pressure

Figure 10.39 Decrease of gas-column vibrations in a muffler [10-2].

Similar to the processes at the fresh gas side of a reciprocating piston engine, the transient flow behavior is also found in the exhaust system. When the exhaust valve is opened because of the overpressure in the cylinder, and later by the upwards-moving piston, an overpressure wave is induced

Internal Combustion Engine Handbook | 429

6606_Book.indb 429

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

that continues toward the tailpipe. Pressure and speed waves are reflected at the open pipe ends and are returned as an aspiration wave. This supports the charge cycle by lowering the exhaust counter-pressure when the pipe lengths in the exhaust system are dimensioned correctly. Contrastingly, a returning overpressure wave can hinder the exit of fresh gas that is already in the cylinder. This mechanism is primarily exploited during the operation of two-stroke engines [10-3]. 10.1.4.3.1 Designs There are two basic muffler designs: (1) the resonator-type muffler and (2) the absorption-type muffler. Frequently, combinations of the two types are used (Figure 10.40), which reduce noise within the relevant range of 50–8000 Hz. Depending on the engine design (displacement, output, supercharging, number of valves and cylinders, and so on), a certain minimum volume is required for the reflection and/or absorption range (or several mufflers can be used: front, central, and rear mufflers). Reflection

f0 =



c0 /2 ⋅ p A/(l + 0.7 ⋅ d)⋅V

(10.19)

A problematic side effect of reflection-type mufflers is the excitation of vibration, which the wall structure of the muffler experiences from the pulsing exhaust flow. The resulting structure-borne sound can increase the noise emitted from the muffler. This can be counter acted by selecting a sufficient wall thickness for the intermediate plates in the muffler, by using a sufficiently rigid construction of the overall muffler structure, and by using an outer double-layer jacket with or without an absorbent intermediate layer. 10.1.4.3.2 Overall system Figure 10.41 shows the basic construction of an exhaust system for a four-cylinder engine. When a single catalytic converter is used, it is necessary to connect the exhaust pipes of the individual cylinders. The exhaust from all the cylinders runs through the collection manifold that contains a central exhaust gas oxygen sensor to measure the integral air to fuel ratio.

Interference l 1, d 1 l 2, d 2 l 3, d 3 Catalytic converter

Resonator

Throttle

Absorption

Figure 10.40 Combined muffler system [10-2].

With absorption-type mufflers, the flow of gas is guided through the muffler, whereby the gas-guiding pipe is perforated. The area between the jacket and the perforated pipe is filled with an absorbent material. The pulsing flow of gas can expand through the perforation into the area filled with the absorbent material. A majority of the vibration energy is, thereby, attenuated by friction and converted into heat. The flow of gas that exits the muffler is, therefore, largely pulsefree. The absorption-type muffler is especially distinguished by a good damping suppression in the frequency range of >500 Hz and its low exhaust counter-pressure. In reflection-type mufflers (also termed interference mufflers), the damping is suppressed by being diverted, by the changes in the cross section, and by the partitions inside the muffler. The corresponding chambers and alterations in the cross sections must be precisely harmonized with each other. Interference will then occur when the sound waves extinguish each along two paths of different lengths (by being 180° out of phase). This principle is particularly effective in the range of <500 Hz. Pressure peaks of extremely loud vibrations build in resonators [Figure 10.40 (left)] that have a particularly low flow loss. The frequency at which a resonator is effective depends mainly on the dimensions (length l, diameter d, and cross-sectional area A) of the pipe extending into the resonator volume V. The resonance frequency f0 can be calculated according to the following equation:

Exhaust Y-Branch Manifold pipe

Front pipe

Middle silencer

End silencer

Collector pipe

1st. Intermediate pipe

2nd. Intermediate pipe End pipe

Figure 10.41 Schematic design of an exhaust system [10-4].

The combined reflection/absorption-type muffler and/or the combined reflection/branch resonator muffler are preferred to minimize exit noise. Based on the transient flow process, the exhaust system can, given a suitable design, clearly improve the charge cycle similar to the intake manifold system [10-4]. The exhaust system largely affects the charge cycle properties with three influential factors that mutually affect one another: •• gas-dynamic effects •• exhaust work •• residual share of gas in the exhaust. The exhaust work is determined by the flow properties of the exhaust system. The flow properties and the gas-dynamic effects in the exhaust system largely determine the residual exhaust gas of the cylinder charge when operating under a full load, which, in turn, strongly influences the combustion properties. The ignition conditions that change with the residual exhaust gas, the inner efficiency, and, hence, the torque behavior are significantly influenced by the adapted ignition timing points.

430 | Internal Combustion Engine Handbook

6606_Book.indb 430

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

10.1.4.3.3 Design criteria In addition to the requirements for noise suppression and exhaust treatment, there are certain design criteria for the exhaust system related to the charge cycle.

the exhaust manifold to the Y-pipe. The remaining portion of the original pressure pulse that passes along the exhaust path l3 is reflected at the transition to the catalyst housing as a vacuum.

10.1.4.4 Even Distribution The exhaust pipes that can be assigned to the individual cylinders at the exhaust manifold must have pipes with equal lengths and cross sections. In view of the options within the vehicle interior, the elbows of the exhaust manifold and the pipe connections should be designed similarly. These requirements also apply to the Y-pipe.

Catalytic converter housing

l3

10.1.4.5 Exhaust Counter-Pressure Level To achieve a low exhaust counter-pressure, superior flow properties should be sought for the cylinder head exits and the exhaust system. The exhaust counter-pressure cannot be reduced to zero because of the flow resistance of the catalytic converter and the basic function of noise suppression, because noise suppression always involves an irreversible conversion of energy that is manifested by the exhaust counterpressure behavior.

Collector pipe

Y-Branch pipe l2 Cylinder. 2 and 3 Exhaust Manifold l1 Cylinder. 4

10.1.4.6 Gas-Dynamic Effects The exhaust system should support the charge cycle in the defined rpm ranges for pipe length, cross section, and pipe branching.

Early-break pressure impulse with mainly positive amplitude Late-break pressure impulse with mainly negative amplitude

2,0

n = 3000 min–1

A2 1,5

Pressure [bar]

10.1.4.6.1 Gas-dynamic processes The exhaust gas that is under high pressure in the combustion chamber causes a pressure wave when the exhaust valve is opened, which then makes the exhaust pulsate at a high amplitude. According to acoustic theory, the pressure amplitude advances at the speed of sound through the exhaust line and is reflected at the open pipe end as a negative pressure amplitude. If it is at the exhaust valve at the approved time, then the negative pressure amplitude can support the charge cycle by extracting residual gas from the combustion chamber. Real implemented exhaust systems have different reflection locations in the exhaust line, from the cylinder head to the entrance in the catalyst housing, because of the individualized pipe branching. Figure 10.42 schematically illustrates the pressure wave for cylinder 1 in an exhaust system from Figure 10.41. After passing along the exhaust path l1, the positive pressure curve meets the first reflection location, where the pressure pulse is divided according to the design of the pipe branches and the pipe cross sections of the exhaust manifold and the Y-pipe. At a correspondingly sharp branching angle, a small amount of the pressure pulse with a primarily positive amplitude passes through the exhaust line l1 of cylinder 4, and is reflected from the closed exhaust valve as a mainly positive pressure pulse. Another part of the pressure pulse is reflected from the pipe branch as a vacuum pulse, and returns against the main direction of flow back to cylinder 1. Most of the original pressure passes along the exhaust path l2 of the Y-pipe up to the pipe branch at the manifold, where a division of the positive pressure pulse occurs similar to the transition from

Cylinder head Cylinder. 1

A1 1,0

A5

aC 0,5 AÖ

lC = C · ∅t

A3

UT

∅t

= aC(n · 360°)

C = Sound velocity

A4 OT EÖ AB

AÖ UT Cyl 4

Figure 10.42 Left: schematic illustration of the reflection locations. Right: pressure curve in the exhaust manifold (100 mm after the exhaust valve) [10-4].

The rise in the positive pressure triggered by the opening exhaust valve starts at A1. The rise in pressure to the maximum A2 depends mainly on the function of the lifting valve. The further course of the pressure curve from A2 to A4, for the maximum of the reflected vacuum characteristic, depends on the design of the exhaust system. The characteristic length lC, that remains constant for the respective exhaust system,

Internal Combustion Engine Handbook | 431

6606_Book.indb 431

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

independent of the working point, can be calculated from the CA aC that extends from A2 to A4 by considering the rpm and speed of sound. The pressure curve from A4 to A5 is characterized by the overlapping wave movements in the exhaust system. The basic characteristic is similar to any respective exhaust systems, and is almost independent of the working point. At A5, the pressure of cylinder 4 begins to rise at the measuring sensor after passing through l1 of cylinder 4 and l1 of cylinder 1 up to the measuring sensor according to the EO of cylinder 4 [10-4]. The engine characteristics are strongly influenced by the location and characteristic curve of the pressure from A3 to A5 and/or the characteristic length lC. In principal, a minimum pressure during the valve overlap is always advantageous. The characteristic length lC essentially depends on the exhaust pipe lengths l1 and l2, the ratio of the Y-pipe diameter d2 = constant

to the exhaust intake manifold diameter d2/d1, and the design of the transition from the exhaust manifold to the Y-pipe. As the sum of l1 + l2 increases and the diameter ratio d2/ d1 decreases, the characteristic length lC increases, because the main reflection site is farther from the inlet port. This is also shown by the following experimentally determined pressure characteristics of three different exhaust system variants (Figure 10.43). 10.1.4.7 Valve Control Timing Valve control timing is always a compromise, because the engine operates within wide rpm ranges and load ranges. Because of the factors described in Chapter 10.6.1, one cannot simultaneously optimize the charge cycle for maximum torque and maximum-rated horsepower without additional features, such as the CS adjustment system, the control cam system,

lges = l1 + l2 + l3

l1 = constant

Variante A:

l2 = 0,4 l1

d2/d1 = 1,25

lges = 2,3 l1

Variante B:

l2 = 0,8 l1

d2/d1 = 1,25

lges = 2,3 l1

Variante C:

l2 = 0,4 l1

d2/d1 = 1,0

lges = 2,3 l1

Pressure[bar]

1,5

1,0

aCA

0,5

n= 6000 min–1

aCB aCC AÖ

UT

EÖ

OT

AS

UT

2,5 Variante A Variante B Variante C

Pressure [bar]

2,0

1,5

1,0

aCA aCB aCC

0,5

AÖ

UT

n= 2000 min–1 EÖ

OT

AS

UT

Figure 10.43 Pressure curve in the exhaust manifold (pressure sensor 100 mm after the exhaust valve) [10-4].

432 | Internal Combustion Engine Handbook

6606_Book.indb 432

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

or a multistage manifold. The offsetting of the valve control timing is related to these factors, whereby the terms “early” and/or “late” indicate a relative position to the basic control times that are indicated as the degree of CA relative to the closer dead center. 10.1.4.7.1 Exhaust opens The exhaust usually opens in gasoline engines at 50–30° CA from BDC shortly before the end of the expansion cycle. This control time represents a compromise between a gain in the expansion work and the greater exhaust work. If the EO is pushed toward “late” (that is, the EO occurs closer to the BDC), then the working gas can expand longer and exerts work on the piston, the thermal efficiency increases, and the consumption falls. A longer expansion leads to lower hydrocarbon emissions and to lower exhaust temperature. At greater rpm and loads, the exhaust work substantially increases at the start of the expulsion cycle. This, in turn, increases the consumption. A late EO is primarily relevant for partial loads, and its influence on full loads is slight (Figure 10.44). 6

p [bar]

5 4 3 2 1 0

0

1

2

3

4

5

6

7

8

9

10 11 12

V /Vc

2000 min–1, 10.37 bar, Full Load, AÖ 56° cw UT, AS 10° n. OT 2000 min–1, 10.63 bar, Full Load, AÖ 42° cw UT, AS 10° n. OT

Figure 10.44 Expansion work gain by shifting the EO toward the direction “late.”

When the EO shifts in the direction of “early,” the opposite, respectively, occurs: (1) expansion work is lost; (2) the thermal efficiency drops; and (3) the fuel consumption increases. The hydrocarbon emissions and the exhaust temperature rise. However, less exhaust work is required, because the cylinder pressure is always at a higher level and the exhaust leaves the cylinder more quickly. An important factor is that the consumption will be increased at a partial load. Another fact is that the thermal load on the exhaust valve rises with an early EO and, therefore, increases the material requirements for wear resistance. The pressure loss during expulsion is also dependent on the characteristic of the valve lifting curve of the exhaust valves. When the valve stroke rises strongly during opening, it is easiest way for the exhaust to exit the cylinder. For this reason, the required compromises with two exhaust valves are less critical than with only one exhaust valve. When there are two exhaust valves, there is an increase in the effective opening

area available for expulsion at a faster rate. The exhaust can, therefore, leave—as it has a higher available pressure—the cylinder more quickly at the beginning of the expulsion cycle. This, therefore, reduces the exhaust work for the piston. 10.1.4.7.2 Exhaust closes A common approach to exhaust close (EC) is 8–20° CA after TDC (top dead center), which indicates the end of the valve overlap phase. In addition to IO, EC is the control time that can be used to control the length of the overlap. At low rpm and load points, the EC controls the amount of exhaust drawn back by the exhaust system, and at higher load levels and speeds, it controls the residual gas that can be expelled. Under a full load, the cylinder can be thoroughly purged by a late EC, which, thereby, increases the volumetric efficiency. This is ised for engines with a higher rated horsepower, such as sports engines. An increasingly greater portion of the fresh charge flows through the cylinder without participating in combustion (scavenging loss from short-circuit flow), which, thereby, increases consumption and the hydrocarbon emissions. Under a partial load, an increasingly greater portion of the exhaust is drawn back (internal exhaust-gas recirculation) by the suction of the piston, wherein substantial advantages for consumption and emissions can be achieved. The last part of the exhaust is always relatively rich in noncombusted hydrocarbons, as the combustion is incomplete in the cylinder charge zones close to the wall. This component will be expelled relatively late. If this component in the exhaust is recombusted, then the consumption is reduced, and there are fewer hydrocarbon emissions. Because of the diluted charge, the combustion temperature is lower, which, thereby, reduces nitrogen emissions. Another consideration is that the fresh mixture becomes homogenized because of the hot residual gases and, thereby, results in a better mixture. There is less additional intake work with a later EC. This occurs for two reasons: (1) because the drawn back exhaust component expands in the cylinder and supports expansion and (2) when there is more residual exhaust gas in the cylinder charge, less throttling for load control is required to compensate for this quantity while retaining the load. This leads to additional reductions in the consumption. The restriction on internal exhaust-gas recirculation is determined by the residual gas compatibility during combustion. With an early EC, the combustion gas cannot leave the cylinder at the right time (exhaust lock-up), so that the residual exhaust gas in the cylinder rises. This causes a drop in the volumetric efficiency and the rated horsepower. The scavenging loss is lower, which slightly lowers the consumption. In this case as well, the last component of the exhaust is recombusted, which can have advantages for consumption and emissions under a partial load [the nitrogen oxide (NOx) emissions are reduced because of the low combustion temperature]. The exhaust remaining in the cylinder continues to flow—partially guided by the piston—very strongly into the intake manifold pipe, which, therefore, improves the mixture preparation. Because there is a continually smaller area for expelling the exhaust after a certain piston position, the exhaust work will

Internal Combustion Engine Handbook | 433

6606_Book.indb 433

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Source: Mercedes-Benz

10.1.4.7.3 Inlet opens The control time IO is commonly set between 20 and 5° CA before TDC for gasoline engines. It is also important as the beginning of the valve overlap phase, like the EC, for regulating the residual amount of gas in the fresh charge under partial loads and for scavenging the residual gas under full loads. In this role, it has a substantial influence on idling quality. The duration of the valve overlap phase is shortened with a late IO. Under a partial load, this produces a charge that is less diluted with exhaust, which, therefore, increases the speed of combustion. Under such conditions, the rpm can be lowered during idling, which, therefore, reduces the consumption. Given the lower residual exhaust gas and the fast combustion, the combustion temperature increases and the emissions of NOx increases. The hydrocarbon emissions can be lowered, whereby the following conditions must be considered. Because the intake valve opens later, the flow in the cylinder is faster at a specific piston position, which increases the flow within the cylinder. This, in turn, leads to an improvement in the mixture preparation and combustion is more thorough, which shortens the combustion or ignition delay, as well as the length of combustion. When the IO is late, the intake work increases, because a vacuum is generated in the cylinder in the first phase of intake. This leads to an increase in the consumption. Under a full load, the mean effective pressure is less, because the air expenditure is lower. With an early IO, the valve overlap phase is lengthened, and a particularly large amount of exhaust returns into the intake manifold pipe under a partial load. This has a negative influence on combustion, because the mixture becomes inhomogeneous and burns more slowly. However, this effect can also be put to positive use for intake manifold injection with throttle-free load control (VVA). Because of the lack of intake manifold pipe vacuum with throttle-free load control, there is frequently insufficient mixture preparation, which causes the combustion to last longer and be incomplete. Fuel deposits can also form close to the valve. These deposits can be vaporized by the returning hot exhaust and be sucked back inside, which, therefore, heats the intake manifold pipe wall and improves the mixture. Investigations have shown (Göbel and MTZ) that this method can positively influence mixture preparation despite the reaction-inhibiting higher residual exhaust gas, which, in the final analysis, therefore, enhances the reactivity of the mixture.

10.1.4.7.4 Inlet closes The valve timing element that is primarily responsible for the torque and the power characteristic is the IC. It usually lies between 40 and 60° CA after BDC, and it influences the charging of an engine much more than the other control times. The characteristic quantities, such as torque and output, are primarily determined by the IC. Offsetting the IC in the late direction to a time optimized for the maximum torque yields greater air expenditure and volumetric efficiency at higher speeds. A higher rated horsepower is correspondingly attained with a late IC. As illustrated in Chapter 10.6.2.1, the gas dynamic effects at higher speeds play the most important role (especially the recharging effects), whereby, when the IC is offset, the most important task is to exploit these effects by capturing the overpressure wave in the cylinder. At lower speeds and under a full load, a long opening time has a negative influence on the torque. Because the intake valve is closed later, a greater amount of charge is pushed back into the intake manifold pipe by the piston. This is countered by a lower pulse because of the lower gas speed, which, in turn, lowers the volumetric efficiency. The influence of the control time IC on the air expenditure with full load is shown in Figure 10.45. By offsetting the inlet CS by a 20° CA toward late, the air expenditure is clearly reduced at a low rpm. At the nominal rpm, the air expenditure is contrastingly increased by approximately 8%. This applies for an engine as an eight-cylinder gasoline engine with four valves per cylinder.

1,0

Air Expenditure

be increased. At the end of the expulsion cycle, the residual gas can also be compressed by an early EC, which slightly increases the consumption. An early EC is limited by the increased exhaust work, a fresh charge diluted with exhaust, and an inhomogeneous mixture from a strong inflow of exhaust into the intake manifold pipe. When the dynamic effects in the exhaust system can be optimized, the efficiency of the expulsion can be improved if a vacuum wave reduces the static pressure in the exhaust port shortly before EC, and, thereby, sucks the exhaust out of the cylinder.

0,8 Aligning the intake camshaft after „early“

0,6

Aligning the intake camshaft after „late“

0

1000

2000

3000 4000 5000 Engine speed Adjustment angle : 20° KW

6000

7000

Figure 10.45 Air expenditure with variable IC.

Under a partial load, a late IC lowers the intake work, because the charge is aspirated with less throttling. This leads to lower consumption. The thermal efficiency of the process is lower, because the effective compression becomes increasingly low. The combustion temperature is reduced by the low peak pressure, which, in turn, reduces the NOx emissions. With VVA and a late IC, the engine can be operated without throttling. The goal is either to achieve a higher rated horsepower or to reduce the consumption under a partial load. Under a partial load, the excess charge is returned by the piston into the intake manifold pipe during the compression cycle.

434 | Internal Combustion Engine Handbook

6606_Book.indb 434

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.1 Gas Exchange Devices in Four-Stroke Engines

Because of the throttle-free load control, less intake work is required. Similar to previously described, this, therefore, reduces the consumption, the thermal efficiency, the consumption temperature, and the nitrogen-oxide emissions. The limit for a late IC is the drop in the thermal efficiency and the worse mixture preparation in the intake manifold port because of the lack of a vacuum (lower gas speed). If there is an early IC and conventional valve actuation, then the intake phase becomes shorter, which, therefore, reduces air expenditure. Under a full load and at higher speeds, this reduces the volumetric efficiency and yields a low-rated horsepower. However, because less charge is returned into the intake manifold pipe at low speeds, the volumetric efficiency and the torque increase. Under a partial load, the required load can be attained with less throttling because of the shorter intake phase, which reduces the amount of intake work. This has a positive effect on the consumption. With VVA and an early IC, the load no longer has to be controlled by throttling, but it can be regulated by the selected IC valve timing. The goal can be either to increase the torque under a full load or to reduce the consumption under a partial load. As soon as the amount of charge is in the cylinder that is required for the load, the intake valve will be closed. The piston is still moving toward BDC in this phase, and a vacuum is generated in the cylinder. Because the load control is executed as throttle-free, the amount of intake work is much lower than when throttling is used to control the load, and this, thereby, reduces consumption. The difference in pressure between the intake manifold and exhaust systems is low, and only a slight amount of its exhaust is sucked back by the outlet. Assuming that the IO timing is at a conventional position, and that the overlap phase is not long, an early IC produces stable combustion under low loads and low rpm. The limitation for a low IC is the mixture formation. Because the intake ends earlier than the BDC, there is a frequently negligible charging movement in the cylinder during ignition, which can make combustion longer and incomplete after a long combustion delay. This can produce greater hydrocarbon emissions and increase consumption, despite the low amount of work involved in the charge cycle. Furthermore, there is the danger of fuel condensation in the cylinder from the charge cooling because of the generated vacuum. As mentioned in Chapter 10.1.4.7.3, the mixture is insufficiently prepared in the inlet port because of the absence of a vacuum, which makes the aspirated mixture inhomogeneous. Fuel deposits could form close to the valve. 10.1.4.8 Flow Cross Sections To achieve high volumetric efficiency and low work losses during a charge cycle, the control valves must have a large geometric opening cross section. The characteristic curves of the opening cross sections of the intake and exhaust valves correspond to the valve lifting curves (Figure 10.46). The valve stroke and opening cross section for the intake valve are greater than for the exhaust valve. The opening cross section is made even greater, because the intake valve is larger than the exhaust valve (intake valve diameter > exhaust valve diameter).

h

Exhaust

Inlet

UT

UT

OT

Exhaust Aö

Inlet As

Eö UT

OT

Es UT

°KW

Figure 10.46 Valve lifting curves and valve cross sections [10-1].

The flow cross section at the valve strongly influences the charge cycle. The flow cross section is smaller than the geometric cross section because of the influences of the hydrodynamic processes (Figure 10.47).

Geometrical cross-section A

h a

Cs Poppet valve

d Flow cross-section A

Figure 10.47 Flow cross section and valve stroke [10-1].

Both the geometric opening cross section and the flow cross section are ring areas that surround the valve axis corresponding to the valve seat angle α . The valve stroke is the perpendicular distance from the valve head to the valve seat. Assuming an isentropic flow at the valve seat, the theoretical speed ciS results in the flow cross section AS. Because of frictional influences, the actual speed cs is less than ciS. The following holds true for the mass flow at the valve:

! = V! ⋅ r = AS ⋅ cS ⋅ r = y ⋅ A ⋅j ⋅ ciS ⋅ r m

(10.20)

where

ρ = density in the flow cross section ψ = jet contraction (constriction number) φ = friction coefficient.

Internal Combustion Engine Handbook | 435

6606_Book.indb 435

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

The following holds true for the isentropic flow cross section AiS: AiS = y ⋅j ⋅

r ⋅A riS



! = AiS ⋅ ciS ⋅ riS m

T1, p1 = thermal condition before the measurement arrangement, for example, in a collection tank p2

= pressure in the cylinder

ṁ

= mass flow, for example, measured with an orifice.

The measurement can be done using both suction or pressure (compressed air). The isentropic flow cross section AiS can be calculated from the recorded measured values. The following, therefore, holds true: k −1 ⎡ ⎤ ⎛ p2 ⎞ k ⎥ 2 ⋅k ⎢ ciS = ⋅ RL ⋅T1 ⋅ 1 − ⎜ ⎟ ⎢ ⎝ p1 ⎠ ⎥ k −1 ⎢⎣ ⎥⎦

AiS AV

(10.25)

where AV is the valve area corresponding to the internal valveseat diameter, and α V does not provide any information on the quality of the charge cycle. A measure of the valve flow in a given engine and, hence, for the charge cycle is the flow factor

(10.22)

The evaluation of an isentropic flow cross section AiS of a valve as a function of the valve stroke is determined in a stationary flow test. A flow is guided through the cylinder head, or a corresponding model, and quantities are measured for different valve strokes (Figure 10.48):



aV =

(10.21)

where ρ iS is the density of a given isentropic flow in the flow cross section. One, therefore, obtains the following equation for the mass flow:

The flow factor of the valve α V is used to evaluate the quality of the actuators:

aK =



AiS AK

(10.26)

where AK is the piston surface area. Figure 10.49 shows the flow factor as a spread for modern engines as a function of the intensity of a tumble flow. The values for the VW-FSI [1.4 L with direct injection (DI)] are shown as solid dots and open squares for open and close charge movement (LBK) valves in the inlet port. α K is very suitable for comparing different engines with the same average piston speed. Reference values for the inletside flow speed of the engine given a maximum valve stroke hV,max are •• gasoline engine •• two-valve:

α K = 0.09–0.13

•• four-valve:

α K = 0.13–0.17

•• diesel engine (10.23)

•• two-valve:

α K = 0.075–0.09

•• four-valve:

α K = 0.09–0.13.

and 0,25

1

⎛ p ⎞k riS = r1 ⋅ ⎜ 2 ⎟ ⎝ p1 ⎠

where κ = 1.4 for air. Discharge duct

Intake duct Inlet board

p1 T1 ∅p

Valve poppet alignment Duct

LBK closed LBK open

0,20 Conventional systems

0,15

Tumble system

0,10

p1T1 p1T1

Valve p2

Cylinders For air-throughput measuring Induction-based Pressure mode operation

Flow rate aK



(10.24)

0,05

0

2

4

6

8

Tumble intensity Measured air-throughput

Figure 10.49 Flow as a function of the tumble intensity.

Pressure mode

Figure 10.48 Measuring arrangement for determining the flow [10-1].

10.2 Calculating Charge Cycle

In terms of approximation, AiS is independent of the set pressure ratio p2/p1 in the stationary flow test. In addition, AiS can be transferred to real engines even though the flow is transient, because a quasi-stationary calculation is permissible, given the short throttling sites in the direction of flow.

The simulation of the engine combustion process, especially combined with a 1-D simulation of the gas dynamics in the intake and exhaust systems, is today a generally accepted tool for predicting the output data of engines in the design phase or during construction. It is also used for analyzing the charge cycle and the thermodynamic process of engines running on

436 | Internal Combustion Engine Handbook

6606_Book.indb 436

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.2 Calculating Charge Cycle

a test bench. Especially for the last application, it can, when used correctly, offer information that could not otherwise be experimentally determined except at great expense. Because of the complexity of the charge cycle process, a huge amount of effort goes into its theoretical analysis. Depending on the respective question, a certain amount of simplification is required. For this reason, various calculations for special applications have been developed for analysis and simulation. A distinction is drawn among purely thermodynamic 0-D models, 1-D models that couple 0-D analysis with gas dynamics in the intake and exhaust systems, and 3-D spatial models (CFD (computational fluid dynamics)). Whereas a 1-D analysis makes it possible today to describe the entire engine from the air filter to the exhaust system, and, thereby, offers a temporal description and spatial (1-D) description along the pipes of the processes, the 3-D CFD calculation is limited to spatial (3-D) and temporal analysis of the processes in the subsystems of the engine because of limited computer capacity.

To be able to determine the mass flow in the inlet and outlet, information must be obtained concerning the states at the inlet and outlet of the cylinder. Physically, strong 3-D flows occur through the valves that manifest the jet disintegration and zones of turbulence. From a simplified perspective, the 0-D and 1-D models assume that the flow through these throttle sites is quasi-stationary. In this case, quasi-stationary means that the status vector at the inlet and outlet areas of the throttle site (Figure 10.51) does not change within a unit of time of the calculation, and that the change over time of the vector results from the sequence of different stationary states. Because the throttle site does not extend infinitely, this analysis is more applicable for smaller throttle sites in the direction of flow in comparison to the connected pipes (Figure 10.51).

10.2.1 The Filling and Emptying Method

The easiest way to describe the charge cycle in a real engine is the filling and emptying method. Because the spatial gradients of the status variables are not covered by this method, the filling and emptying method belongs to the 0-D methods of calculation. Despite this simplification, it is still sufficient in most cases for comparisons and an initial evaluation of the charge cycle. In the filling and emptying method, the intake line and the exhaust line in the cylinder are viewed as containers whose contents are characterized by pressure, temperature, and material composition (Figure 10.50).

pFg sFg

TFg

EV

AV

pAg

TAg

sAg

sZ

pZ TZ

mZ VZ

Figure 10.50 Model of the filling and emptying method.

The filling and emptying method is based on the first main law of thermodynamics: d ( mz ⋅ u) dV dQw dme dma = −pz ⋅ −∑ +∑ ⋅ he − ∑ ⋅ ha (10.27) da da da da da

(P,T,u)E

(P,T,u)A

Figure 10.51 Status variables at a throttle site.

Given these assumptions for this model, and the basic equations of the 1-D model, stationary flow can be used to calculate the status vectors (p, T, u)E and (p, T, u)A at the edges of the pipes. Using the continuity and the energy equation for 1-D stationary flow, we obtain the St. Venant’s theoretical flow equation that holds true when an isentropic, loss-free change in state in a flow cross section arises after the inlet and outlet surfaces of the quasi-stationary throttle sites. Because, however, the change in state is not isentropic and the pulse attenuates, this approach must be corrected. A stationary measurement is required that quantifies the thermodynamic effect of the flow phenomena that causes the pulse to attenuate. This pulse attenuation manifests itself thermodynamically by an irreversible increase in the entropy of the fluid. The mass flow passing through the throttle site in an irreversible flow is smaller than the mass flow that would result with a loss-free flow. This loss is measured with the aid of the flow coefficient α that is defined as the ratio of the actual mass flow to the theoretical (isentropic) mass flow. The mass flows at the inlet and outlet are, therefore, calculated as follows:



! = Aeff ⋅ p01 ⋅ m

2 ⋅y R ⋅T01

(10.28)

where

Aeff = a ⋅

dvi2 ⋅ p 4

(10.29)

Internal Combustion Engine Handbook | 437

6606_Book.indb 437

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

and the flow function ψ in the subsonic range is



110000 108000

Intake manifold pressure [Pa]

y=

2 c+1 ⎡ ⎤ c ⎢ ⎛ p2 ⎞ c ⎛ p2 ⎞ c ⎥ ⋅ ⎜ − ⎟ ⎜ ⎟ ⎢ c − 1 ⎝ p01 ⎠ ⎝ p01 ⎠ ⎥ ⎢⎣ ⎥⎦

(10.30)

106000 104000 102000

and in the transonic range is

100000

1

y = y max

⎛ 2 ⎞ c−1 c =⎜ ⋅ c+1 ⎝ c + 1 ⎟⎠

(10.31)

The flow coefficient, α , thereby changes itself with the valve stroke, and is experimentally determined using stationary flow experiments. 10.2.1.1 Principle for Calculation The goal of the calculation is to determine the characteristics of the pressure, temperature, mass, and composition of the cylinder charge, and the characteristic curve of the mass change influenced by valves as a function of the characteristic of the CA during the charge cycle phase. These quantities cannot be measured, or can only be measured with great effort. Only the pressure can be indicated by means of a quartz sensor. The characteristic curves of these quantities are, therefore, calculated from a starting point using numeric integration. The initial values of the pressure, temperature, mass, and composition are determined at outlet opens by measuring or estimating, and their differential changes are calculated from this starting point using basic thermodynamic equations. On this basis, a suitable integration is applied step by step, until all the values are known up to the time of ICs. 10.2.1.1.1 One-dimensional gas dynamics The filling and emptying method is a quasi-stationary singlezone model. In this instance, quasi-stationary means that transient processes are viewed as stationary for short intervals; that is, the individual quantities (pressure and temperature) are dependent only on time but not location. Dynamic influences such as pressure pulses that, for example, arise in the ram tube charge and the resonance tube charge cannot, of course, be included. The amplitudes and phase angles of the oscillations can support the charge cycle at a certain rpm and hinder them at other speeds. The characteristic of the volumetric efficiency is essentially determined by means of the rpm and torque characteristic of the engine. These oscillations are excited by pressure waves that arise when the valves are opened and closed, and when piston motion occurs. Figure 10.52 illustrates the pressure curve determined with the aid of induced low pressure in the intake manifold of a slow one-cylinder four-stroke engine at 3000 rpm. At the beginning of the intake process, the downward movement of the piston generates a vacuum wave at the intake valve. This vacuum wave advances to the air filter that acts as an open pipe end. It is reflected as an overpressure wave, returns to the intake valve, and reaches it at IC (Figure 10.52).

98000 96000 94000 92000 90000

0

60

120 180 240 300 360 420 480 540 600 660 720

AÖ

EÖ

AS

Crankshaft angle [°KW]

ES

Figure 10.52 Pressure characteristic in an intake manifold at 3000 rpm.

In the 1-D simulation of the engine intake flow, the overall engine system is divided into individual abstracts, that is, simplified elements such as the cylinder (C1), air filter (P11), orifices (SB1, R1, and SB2), and pipes (1–4) (Figure 10.53 and Figure 10.54). R1

SB1 1

Pl1 2

C1 3

SB2 4

Figure 10.53 Schematic representation of an entire engine system.

This is done by assuming that the flow in the overall system can be described by a 1-D transient pipe flow in the pipe elements, and by a 1-D quasi-stationary throttled flow in the components that connect the pipe elements. The 1-D transient analysis within a pipe element assumes that the status quantities, such as pressure p, density ρ , and speed u, are sufficiently defined by the averages in the individual pipe cross sections. Furthermore, it is assumed that there is no pulse loss because of the internal friction in the flow. Only the friction of the flow against the pipe wall is considered. This means that the processes in a pipe element, for example, the conversion of pressure energy into movement energy, are irreversible only as a result of the included wall friction. A nonlinear inhomogeneous differential equation system is accordingly created for a 1-D transient pipe flow within the flow plane (xt lane) based on the conservation equations for mass, pulses, and energy: (10.32)



∂( r ⋅ u) 1 dA ∂r =− − r⋅u⋅ ⋅ ∂x A dx ∂t



∂( r ⋅ u2 + p ) ∂( r ⋅ u) 1 ∂A FR =− − r ⋅ u2 ⋅ ⋅ − ∂t ∂x A ∂x V

(10.33)

(10.34)



∂ ⎡ u ⋅ ( E + p ) ⎤⎦ 1 dA q w ∂E =− ⎣ − u ⋅ (E + p ) ⋅ ⋅ + ∂x A dx V ∂t

438 | Internal Combustion Engine Handbook

6606_Book.indb 438

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.3 The Charge Cycle in Two-Stroke Engines

SB1

R5

6

R6

1

2

R1

R4

5

4

R2 R3

3

I1

7 R7

R13

R9 10

8 R8 9

18

14 R14

R10 15

11 R11 12

SB2 51

23 I3 28

R19 20

24 I4

29

R20

R16

P13

27

R18

R15

17

26 22 I2

19

16

R12 13

R17

25

21 50

P12

C4 C3 C2 C1

R21 R25

38

30 34

R22 R26

J1 42 R29 44 39

32 36

J2 43 R30

41

46

R24 R28 37

33

49

R31 R32

where FR is the wall friction, V is the volume, qw is the flow of heat, and E is the total energy. To solve this problem regarding initial values and boundary values, information is needed concerning the status at the pipe edges. This status vector is determined by the flow in the components that connect the pipe ends with each other. Stated simply, the filling and emptying method assumes that the flow is quasi-stationary through these throttling sites.

10.3 The Charge Cycle in Two-Stroke Engines 10.3.1 Scavenging

The characteristic feature of different two-stroke engine designs is the respective type of cylinder scavenging and the related type of scavenging air supply. The selected scavenging approach

Displacement x Density of the ambient air

45 J3

31 35 R23 R27 40

48

Cat1

47 Figure 10.54 Schematic representation of a four-cylinder gasoline engine.

greatly influences the complexity of the design, component load, operating behavior, air/gas mixing conditions, fuel consumption, and emissions of the engine. When the cylinder is scavenged, the combusted mixture is displaced from the cylinder by fresh gas, without mutual mixing in the ideal exception of displacement scavenging. In contrast, when a cylinder is scavenged in a real engine, a mixture of fresh gas and exhaust occurs in addition to the displacement of the exhaust. As schematically illustrated in Figure 10.55, especially when there is a great deal of scavenging air, such as at high load map points, a part of the scavenging gas mixed with the exhaust is expelled from the cylinder (loss of fresh gas). To evaluate the results or the efficiency of the scavenging procedure in two-stroke engines, the retention rate or air expenditure is used as an index in addition to volumetric efficiency (also refer to [10-12] and [10-13]). Figure 10.56 shows an overview of the most important two-cycle scavenging procedures with methodically related advantages and disadvantages.

Cylinder volumes x Density of the ambient air

Charge enclosed in the cylinder Excess air

Retained fresh gas in the cylinder

Combustion products

Supplied fresh gas in the cylinder

Exhaust

Surplus product

Fresh gas loss

Residual exhaust gas

Figure 10.55 Mass balance of the two-cycle scavenging process according to [10-14].

Internal Combustion Engine Handbook | 439

6606_Book.indb 439

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Scavenging Concept

Advantages

Disadvantages

1. Loop scavenging

• Compact design dimensioning • High rpms are possible • The combustion chamber recess can be located well cooled in the cylinder head • Simple design method without a piston valve

• Asymmetrical control timing diagram is only possible with additional components (piston valve) • Asymmetrical thermal load on the piston • The piston rings are especially endangered by the scavenging and exhaust ports • Comparatively difficult to generate charge turbulence flow

2. Uniflow scavenging with exhaust valves

• Effective scavenging/low air expenditure • Larger overall design height in comparison to 1 • Simple generation and influencing the • A more involved and optimized valve gear combustion chamber turbulence is required for large effective cylinder • The combustion procedure can largely be strokes and low consumption transferred to four-stroke engines • Asymmetrical control timing diagram is possible without additional components

3. Uniflow scavenging with opposed pistons

• Minimization of the combustion chamber surfaces heated in the high-pressure phase • Asymmetrical timing diagram can be achieved only by controlling the piston edges • Effective scavenging/low air expenditure

• More involved constructional effort • Larger overall design height (design width) • Extreme thermal load on the piston controlling the exhaust ports • A conventional combustion method cannot be used because of the arrangement of the nozzle holder/spark plug

4. Reversed head scavenging

• Engine-transmission unit is very similar to that of four-stroke engines • The piston rings are not endangered from scavenging and exhaust ports

• Low scavenging effect/large air expenditure • Because of the restricted opening time cross section, excessive increases in the charge cycle work and consumption at higher rpms

Figure 10.56 Comparison of different scavenging approaches.

440 | Internal Combustion Engine Handbook

6606_Book.indb 440

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.3 The Charge Cycle in Two-Stroke Engines

10.3.1.1 Loop Scavenging In loop scavenging (according to Schnürle), the fresh gas passes into the cylinder through two-to-six scavenging channels that symmetrically mirror one another across the midaxis of the exhaust port and run in the opposite direction of the exiting exhaust (excess flow channels) in the cylinder. The fresh gas streams are directed to one another and form, at the exhaust port opposite side of the cylinder, an ascending stream of fresh gas that reverses its direction in the area of the cylinder head and pushes the exhaust gas out of the cylinder. This widespread type of scavenging that is particularly used in small engines is suitable for high rpms, and results in a simple design and compact engine dimensions and features in DI diesel engines, the option that the combustion bowl can be arranged as a well-cooled unit in the cylinder head. Downsides are the asymmetrical thermal load on the piston, a hazard for the piston rings from the rinsing and exhaust ports and the fact that the use of a pressure lubrication means that oil consumption is difficult to control. In addition, there are technical measures required for the generation of a conventional charge turbulence for DI diesel engines and for creating an asymmetrical control diagram. 10.3.1.2 Uniflow Scavenging With uniflow scavenging, the fresh gas passes into the cylinder through intake ports in the perimeter of the cylinder and displaces the exhaust through several exhaust valves in the cylinder head that are controlled with the crankshaft speed. A tangential arrangement of the scavenging channels makes it comparatively easy to generate and/or influence the turbulence flow that supports the mixture formation. This turbulence generally lasts for the entire work cycle while attenuating, and it does not have to be completely regenerated in the following scavenging cycle. The advantages of uniflow scavenging are: (1) its comparatively effective scavenging efficiency (low air expenditure); (2) the optional implementation of asymmetrical timing can be achieved without additional constructive measures; and (3) tried-and-tested DI diesel combustion methods for four-stroke engines can be transferred largely unchanged to two-stroke engines. In contrast to loop scavenging, it is comparatively easy for the piston rings to freely rotate a given corresponding scavenging port design that increases their life. The overall height of a cylinder head with valves yields a taller engine compared with similar four-stroke engines, especially with over-square stroke-to-bore ratios, because the scavenging ports are covered by the piston shaft and a collision of the connecting rod with the piston shaft must be excluded in the design. In addition, there are substantial demands on the design of the exhaust valve drive because of the double valve actuation frequency and the limited valve opening (crankshaft) angle with the simultaneous requirement for large opening time cross section. 10.3.1.3 Opposed Piston Uniflow Scavenging With opposed piston uniflow scavenging, two pistons move in the opposite direction in one cylinder, and their inner-end position encloses the combustion chamber (TDC position).

In their outer-end position (BDC position), one of the pistons opens the intake ports, and the other piston opens the exhaust ports, so that the incoming flow of fresh gas expels the exhaust from the cylinder with the main direction of flow along the cylinder axis. The advantages are effective scavenging, minimization of the combustion chamber surface heated in the high pressure phase, and easily realizable asymmetrical timing. Serious disadvantages of this approach result mainly from the complex design and construction, the bulky engine dimensions, the extreme thermal load on the exhaust side piston (see also [10-15]), and the limited transferability of the combustion process to modern four-stroke engines. 10.3.1.4 Reversed Head Scavenging With reversed head scavenging, the fresh gas generally flows through at least two or three valves actuated at crankshaft speed at BDC into the cylinder, and displaces the exhaust from the cylinder through the simultaneously opened exhaust valves, supported by a reversal of direction at the piston head. The advantage of this type of scavenging is the design of the engine-transmission unit, which largely corresponds to that of a comparable four-stroke engine. Furthermore, the absence of scavenging in the exhaust ports reduces the hazard to the piston rings. These advantages contrast with the serious disadvantage that intake and exhaust valves must be located on the limited combustion chamber surface of the cylinder head. In contrast to a comparable two-stroke engine with uniflow scavenging and, for example, with four exhaust valves, a basic approximation indicates that the available opening time cross sections are cut in half. At the same time, a great deal more scavenging air is required to introduce the same amount of fresh gas in the cylinder because of the less effective scavenging (mixture of fresh gas and exhaust from turbulence and contact of a large surface area of the gas stream) in reversed head scavenging. For this reason, the required charge cycle work and the resulting specific fuel consumption lie only within an acceptable range at low rpms. These restrictions on the level of the nominal speed and/or in the consumption run counter to the requirements of designs of drives for future passenger cars. Apart from that, short-stroke, reversed head two-cycle diesel engines and possibly two-cycle spark ignition engines hold promise for low-speed airplanes with lower nominal rpms (disregarding an intermediate transmission and using high propeller efficiency) as possible successes for this concept. We will, hereby, refrain from discussing other types of scavenging, such as cross scavenging, fountain scavenging, reverse MAN scavenging, and the various dual-piston scavenging approaches (see also [10-16] and [10-17]) because of their limited scavenging efficiency, complicated design, or other disadvantages.

10.3.2 Gas Exchange Organs

As previously noted, the fresh gas stream entering the cylinder in the case of inflow scavenging, and the exhaust stream leaving the cylinder in the case of loop scavenging, are controlled by ports in the cylinder wall and the ascending and descending

Internal Combustion Engine Handbook | 441

6606_Book.indb 441

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

piston. A characteristic feature of port control is that a large flow cross section can be opened and closed within comparatively small CA ranges in comparison to the conventional valve actuation in the cylinder head. High nominal rpms can, therefore, be obtained with port-controlled two-stroke engines. A characteristic quantity used in designing and determining the gas flow rate through a port is the (opening) time cross section (see also [10-17] and [10-18]). This defines the time integral over the respective port cross-sectional area from the opening to the closing of the respective port. Without additional measures, symmetrical timing results for port-controlled two-stroke engines at the dead centers of the crankshaft. With the set objective of improving the charging of the crank chamber with an asymmetrical intake timing diagram, a series of two-stroke gasoline engines were first equipped with tubular and roller rotary disk valves in the past, and then later with disk type rotary disk valves. With asymmetrical timing of the rotary-disk valves, the start of intake is substantially earlier than with port control. Because the vacuum in the crank chamber is comparatively low now, the air column in the intake tract is excited to form gas column oscillations comparatively less at low and average speeds. This produces more continuous torque characteristics and favorable conditions for the formation of a fuel-air mixture with a very constant air-fuel ratio in the carburetor. Instead of rotary-disk valves, modern two-stroke gasoline engines have frequently used lamellae valves (reed valves) in recent years (see also [10-18] and [10-19]). These act as a check valve and automatically open given a specific pressure gradient toward the crank chamber, and they independently close given an opposite pressure gradient. Figure 10.57 shows the construction of a reed valve for two-stroke engines. By reducing the flowthrough resistance in the form of a gable-type roofformed basic body (made of die-cast aluminum or plastic), the contact areas of the reed valve are generally sprayed with a thin elastomer coating to reduce mechanical load and improve the sealing effect and acoustics. The lamellae fixed on one side of the base body (mechanical replacement model: replacement model: cantilever with a surface load) are made of either 0.15–0.2-mm-thick Cr–Ni sheet steel or more recently 0.4–0.6-mm-thick fiberglass-reinforced epoxy resin plates. Given the same length and width, the natural frequencies of steel and epoxy resin reeds are approximately the same, because the quotients of their elasticity modulus and density are about the same.

Because the reeds open more as the pressure differential increases, a linear relationship between the pressure differential and the mass flow results in a first approximation. To prevent the reeds from moving in an undefined manner (opening too wide with the subsequent premature closing of the reeds, vibration in the second original form, and so on), reed valves with arched stops of sheet steel are provided, so that the reeds contact as they execute a rolling-off movement when they open. The natural frequency of the reeds should be at least 1.3 times that of the opening frequency (intake frequency of the engine). Reed valves are placed either directly on the crank chamber or as shown in Figure 10.58, and are used with the piston intake control.

Figure 10.57 Illustration of the construction of a reed valve for use in an intake system of a two-stroke engine.

Figure 10.59 Partial section of a loop-scavenged cylinder with an exhaust port pivot valve according to [10-20].

Figure 10.58 Intake system with combined piston edge/reed valve control.

With the goal of compensating for the disadvantages of symmetrical timing of port-controlled loop scavenging, some modern high-performance gasoline engines partially use flatseat valves, pivot valves, or rotary-disk valves. This can improve the fresh gas charging, the torque and performance curve, or, as is the case with the Honda-activated radical combustion method, improve the ignition of the air-fuel mixture. Figure 10.59 shows a partial section of such a cylinder.

442 | Internal Combustion Engine Handbook

6606_Book.indb 442

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.3 The Charge Cycle in Two-Stroke Engines

10.3.3 Scavenging Air Supply

Whereas the pressure gradient for the charge cycle arises from the expulsion and intake process of the engine transmission unit itself in four-stroke engines, the required scavenging pressure gradient for the charge cycle in two-stroke engines must be generated by a separate scavenging blower (compressor). The cylinder can be only be scavenged when the intake and exhaust organs are open simultaneously. The flowthrough intake and exhaust organs can be described in simplified terms as a flowthrough two series-connected throttles (see also [10-21] and [10-22]), which can themselves be replaced by an equivalent cross section. Because, apart from influences such as pressure pulsation, gas temperature, and exhaust counterpressure, it does not matter whether the ports or valves open and close a few times slowly, or many times quickly within a given period of time, the air flow rate through a two-cycle engine to produce a respective scavenging pressure gradient is independent of the engine rpm. In contrast to them, there is a quadratic relationship in the first approximation between the scavenging pressure gradient and the scavenging air quantity. At higher engine rpms, a much higher scavenging pressure is, thereby, needed to attain the same result. The amount of scavenging air can, in principle, be varied—assuming that the scavenging blower is correspondingly flexible—over wide ranges for a corresponding mapping point depending on the required engine temperature, exhaust temperature, emissions, consumption, and engine performance (supercharging). A displacement-type compressor (reciprocating piston compressor and rotary piston compressor) and flow compressors can be used for scavenging and/or possibly supercharging two-stroke engines (see also [10-21], [10-23], and [10-24]). Figure 10.60 shows an overview of different blowers and/or supercharging designs.

a)

d)

b)

e)

10.3.3.1 Reciprocating Piston Compressor The simplest type of reciprocating piston compressor for two-stroke engines uses the crankcase housing and the bottom of the piston to enclose the working volume. With this design, which is particularly widespread among small two-stroke gasoline ignition engines (the advantages are a compact design, low additional costs, steep compression curve, and low additional drive power), the working gas generally flows through the holes in the cylinder wall or piston shaft into the crankcase housing when the piston moves upward. When the piston subsequently executes a downward movement, the fresh gas is compressed and flows through overflow ports and from scavenging ports exposed by the piston head into the crankcase housing. Using reed valves or rotary-disk valves, or by transferring to a cross-head charging pump, the amount of scavenging air can be increased that is limited by the stroke-to-bore ratio and the dead space. In particular, given the limited scavenging efficiency of two-stroke engines and the fact that operating at a full load generally requires a substantial amount of excess air even in modern diesel combustion systems because of the smoke limit, the low volumetric efficiency of the crank housing scavenging pump is a profound disadvantage, apart from the complicated stepped piston design. Assuming that a highly effective, flow enhancing oil separator with low-pressure loss cannot be used for the scavenging air, the necessity of minimizing the lubrication oil in the scavenging air (problem: hydrocarbon and particle emissions, piston ring deposits, and racing engine) means that the engine-transmission unit generally cannot be born on tried-and-true, low-noise, economical, and reliable plain bearings with oil spray cooling of the pistons. Another substantial disadvantage of crankcase housing scavenging pumps is that the crank chambers need to be sealed from each

c)

f)

Figure 10.60 Overview of the different designs of blowers and/or superchargers. (a) Vane-type supercharger. (b) Roots supercharger. (c) Rotary piston supercharger. (d) Screw compressor. (e) Spiral supercharger (G-supercharger). (f) Turbocharger.

Internal Combustion Engine Handbook | 443

6606_Book.indb 443

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

other in multicylinder engines. Using a separate, mechanically driven reciprocating piston compressor avoids some of the cited disadvantages, but, thereby, requires, apart from the limited flexibility in adjusting the fuel delivery, substantial additional installation space, and major additional costs are involved. 10.3.3.2 Rotary Compressor Under the general term of rotary compressor (rotary piston compressor), there can be a series of compressors covered whose delivery or compression is determined by the compressing effect of rotating elements or pistons. The drive shaft is mechanically coupled to the crankshaft of the engine to scavenge or supercharge internal combustion engines. Belonging to this group of superchargers are roots superchargers, vane-type superchargers (encapsulated blowers), rotary-piston superchargers, spiral-type superchargers (G-superchargers), and screw compressors. Similar to reciprocating piston compressors, the delivered mass flow is approximately proportional to the drive speed and decreases slightly at higher pressures because of increasing leakage, whereby the average compressor efficiency is generally attained. Compared as an equivalent delivery rate, the dimensions of reciprocating piston compressors and radial compressors are approximately the same. 10.3.3.3 Flow-Type Superchargers Of the flow-type superchargers, primarily radial compressors (turbocompressors) are used for vehicle engines. The delivered flow of radial compressors is approximately linear, and the pressure is approximately the square of the drive rpm. Modern radial superchargers can achieve highly efficient compression. Because, in contrast to a four-stroke engine, the two-stroke engine has a mass flow rate characteristic that is only more or less independent of the engine rpm that can be defined as an opening (throttle) with a constant cross section, a radial blower mechanically coupled to the engine is a suitable scavenging blower. Corresponding to the objective of limiting the size of the radial supercharger, it is useful to drive the supercharger with a high-speed transmission. To optimally adapt the air mass flow delivered by the supercharger, largely independent from the crankshaft speed for each mapping point, to the desired scavenging or supercharging level of a two-stroke engine, it is desirable to drive the supercharger with a variable transmission ratio like the previously discussed displacement supercharger. Such a solution was, for example, used for the ZF-Turmat (see also [10-25]). Apart from high design and construction costs, problems with vibration, and the useful life of variable drive transmissions, a general disadvantage of mechanically driven superchargers is that a substantial amount of the effective output must be sent to the crankshaft to drive the supercharger, whereby this correspondingly increases the specific fuel consumption. 10.3.3.4 Exhaust Turbochargers The exhaust turbocharger, which has been successfully used for decades in four-stroke engines, can also be used in principal for two-stroke engines for passenger cars and trucks as a scavenging and supercharging blower. The advantage of turbocharging is that the exhaust energy converted in the

turbine is used, which would otherwise largely be underused. According to Schieferdecker [10-26], a requirement for the use of freewheeling turbochargers in two-stroke engines is that the joint efficiency of the turbine and compressor must be at least 60%, which is more or less achievable with modern turbochargers used in passenger cars and trucks. To use as much of the exhaust energy as possible in the turbine, it is also essential that the exhaust lines from the respective cylinder to the spiral housing of the supercharger to be optimized not only for good flow but also rather more so optimized for minimal heat loss. In addition to a short, cramped port design, the air gap insulation, and possibly even the use of port liners, needs to be considered. To ensure a positive scavenging pressure gradient over as wide a mapping range as possible, superchargers should be used with variable turbine geometry [adjustable blades, sliding supercharger, and double helix supercharger (Twin skroll and Aisin)] (see also [10-25]). An advantageous side effect of turbocharging and supercharging with superchargers, which have an adjustable turbine geometry, is that the backup of exhaust in front of the turbines allows highly effective charging even for scavenging approaches with symmetrical timing (such as loop scavenging). Such an approach—although in an extreme form—was used for the turbo-compound airplane engine “The Napier Normad” (see also [10-27]). To generate a positive scavenging pressure gradient when accelerating from a low load and low rpm, and when starting an engine, there is a requirement for a series-connected additional mechanically or electrically driven supercharger or a mechanical auxiliary turbocharger drive. An interesting alternative is, hereby, an electrically supported turbocharger. With these types of superchargers, a proportion of the propulsion power for the compressor is supplied as needed by, for example, an asynchronous electrical motor integrated in the supercharger (see also [10-28]). It has been, hereby, noted that with the thermodynamic conditions when coupling with two-stroke engines for the pressure wave supercharger (Comprex supercharger), the same observations apply that were made for turbochargers. A fundamental disadvantage is that the fresh gas is heated when it briefly and directly comes into contact with the exhaust, and that mechanical or electrical support of the compressor output is impossible given the functional principle of the supercharger.

10.4 Variable Valve Actuation VVA can be used to positively influence the desired quantities for the combustion engine, such as specific consumption, the emission behavior, torque, and maximum output. Depending on their physical functional principle, VVA systems are divided into systems that are mechanically, hydraulically, electrically, and pneumatically actuated. Many such systems are known, and there has been extensive research on both simple systems in which the control time can be varied between two positions, and on more complex systems in which even the engine load can be controlled by variable control times. Figure 10.61 shows a detailed categorization of VVA. This fine

444 | Internal Combustion Engine Handbook

6606_Book.indb 444

1/19/16 8:40 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

categorization commences with the component of the CS. The CS criterion creates the first of three selected categorization levels. Systems whose energy is provided for valve actuation without a CS are categorized according to their physical functional principle. This accordingly yields electrically, pneumatically, hydraulically, and mechanically actuated systems. With systems that need to use a CS for control, a distinction is drawn between the use of conventional and special CSs. CSs are, hereby, termed conventional when they illustrate a conventional cam geometry, use common materials, and are created using familiar manufacturing procedures. The second categorization level deals with the location where variability takes effect. One can divide the third categorization level that describes the operational and functional principle of VVA into seventeen groups. Type of valve actuation

Target location of the variability

Only the individual systems will be considered in this section. The series-connected systems are, thereby, of particular interest. In the categorization in Figure 10.61, the groups that use series solutions have a gray background. Many types of VVA make it difficult for a developer to select a suitable type of control for their application. The wide variety of systems affect the cylinder head concept to such an extent that substantial adaptations are required when variable valve control is used. A new cylinder head generation has to be developed simultaneously for VVA in serial engines for utilization in generally used stock engines. Usually a more complex design is required for variable control times in contrast to conventional engines, and this is expressed in higher costs. This environment creates particular challenges for engine designers. Target principle and functional principle Electrical system

System without camshaft

Pneumatic system Hydraulic system Mechanical system

System with camshaft

Variability on camshaft drive unit

Mechanical and hydraulic NM-Adjustment Mechanical CS Drive unit with irregular movement Mechanically

Utilizing conventional camshafts

Variability on transfer element between tappets and valve

Variability with additional camshaft Variability on the valve spring Variability on valve seat

Utilizing specialized camshafts

sealed hydraulic system Hydraulic with constant discharge Hydraulic with staged discharged Mechanical modulation with double CS Electro-magnetic Mechanical

Variability on tappets

Mechanical alignment from tappet parts

Variability with axially adjustable CS

Mechanical variability with space tappets

Variation between tappets and valve

Mechanical, self-closing valve Miscellaneous systems

Figure 10.61 Categories of variable valve control [10-29].

Internal Combustion Engine Handbook | 445

6606_Book.indb 445

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

In the future, the option of using VVA to control engine load will gain in importance. An essential objective for varying the valve lifting curves is to lower charge cycle loss under partial loads and, hence, reduce the fuel consumption. The goal of many developmental activities is to dispense with the throttle valve in gasoline engines to control load solely by varying valve lifting. In comparison to pure TC with the conventional throttle valves, Figure 10.62 shows four load control methods that vary intake valve lifting.

DR

FES

Valve stroke [mm]

SES

SEÖ

VME

Crankshaft angle [°KW]

Figure 10.62 Adjustment possibilities for valve lifting curves with VVA.

The load control method “early inlet closure (EIC)” limits the amount of fresh gas by the early closure of the intake valve, as long as the filling has been achieved using the load to be set. When the engine idles, the intake valve opening time corresponds approximately to a 60° CA. With the control mode “late inlet closure,” the part of the charge that is not needed for the specified output is expelled from the cylinder during the uplift displacement of the piston because of the longer opening time. This charge quantity passes through the throttle site of the valves twice with the corresponding loss. When the load is controlled using the late IO method, the intake valve is only opened when the remaining opening time corresponds to the required amount of inflowing mixture. At the start of induction, a strong vacuum is in the cylinder, which promotes mixture through turbulence. The cylinder charge is influenced by the control mode “variable maximum intake valve stroke” by reducing the valve stroke with equivalent opening angles. Instead of the throttle valve, the valve acts as a throttle location, whereby there is no reduction in the amount of charge cycle work. The valve friction can, however, be lowered, because the valve springs are only partially compressed. The effects of the influencing parameters for altering the valve lifting curve are well known. An ideal valve gear would be one that allows the valve lifting curves to be changed as freely as possible. It would also make sense to combine different load control procedures. Depending on the system, however,

only a limited degree of freedom can be attained using the different types of VVA. In addition, a substantial amount of system engineering is required for VVA before the desired complete variability levels can begin to be realized. Just using systems that turn the CS relative to the crankshaft position, the attainable improvements to the engine are substantial. These systems are widely used in the standard produced engines, so that they can be discussed in greater detail in Chapter 10.4.1. At this point, it is important to make a calculated guess about the degree to which VVA can improve consumption or emissions. In one takes a look at the professional literature, such as [10-29], one can see that improvements to consumption of an average between 5% and 15% within some engine mapping ranges could be achieved. Frequently, however, the engines in the literature are optimized in other ways in addition to variable valve control, so that it is difficult to directly identify to which degree the utilization of VVA has actually affected the application. Compared with gasoline engines, the potential improvement to diesel engines from implementing VVA is somewhat limited. Relatively few investigations have been made into this. The future will show to which extent the systems in the standard production engines will develop.

10.4.1 Camshaft Timing Devices 10.4.1.1 Overview of the Functional Principles of Camshaft Timing Devices As early as September 29, 1918, a patent was issued for rotating a gasoline engine CS [10-29]. The desired variation during the engine operation was attained with a sleeve, interior and exterior teeth, and straight and helical teeth that moved axially between the CS and the drive sprocket (Figure 10.63). Therefore, the angular position of the cam and/or CS was adjusted relative to the crankshaft. 1

2

3

4

Figure 10.63 Patent of a CS adjuster dated 1918 [10-30].

The inventor of this patent, Samuel Haltenberger, intended the timing device to be for an airplane engine to adapt output to different flight heights. The helical-toothed sleeve (2) is moved in an axial direction by air pressure using a timing device linkage (4). The relative angular position of the CS (1), thereby, changes in relation to the driving bevel gear (3) that is linked to the crankshaft. Based on the same functional

446 | Internal Combustion Engine Handbook

6606_Book.indb 446

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

principle of a straight and helical-toothed sleeve, in 1983, Alfa Romeo started mass producing a CS timing device for a two-valve engine with dual CSs (Figure 10.64). The timing device is seated on the intake CS and enables the control times to be adjusted between two positions. While idling, the late control time position is retained by a return spring (10), and an early control time is set depending on the oil pressure and the rpm. A solenoid (6) that actuates the control valve (5) applies the engine oil pressure to the helical-toothed piston (9). The adjusting element is the helical-toothed piston (9), which is moved by the oil pressure against the spring force. The helical toothing (3) on the piston and the CS are used to rotate the CS relative to the driving sprocket (4) and, hence, to the crankshaft when the piston shifts axially. The systems shown in Figure 10.63 and Figure 10.64 are all designs where a mechanical functional principle is used. This, thereby, means that the force to actuate the valves flows only through components that are engaged by friction or are positively engaged. The movement and retention of the adjusting elements, such as the piston in the Alfa Romeo timing device in Figure 10.64, can, however, be moved and held by

1: 2: 3: 4: 5:

Camshaft Oil supply flute Helical toothing Sprocket Control valve

9

8

7

oil pressure. For a CS timing device that operates based on hydraulics, a hydraulic component lies in the flow of force to actuate the valve. This is then executed using a quantity of oil that must be at a correspondingly high pressure to keep the positions of the adjusting elements stable. The position of the CS timing device should logically be directly located next to the CS drive. The flow of force to drive the CS can be most easily interrupted here, and selecting the suitable CS adjustment element can easily create the variation for the CS rotation. If one researches the literature and known patent applications, then one can find many different functional principles for CS shifter devices. Based on executed patent research, the author has identified approximately 3000 different applications. When one plots these registration dates, for example, over a period of the last 25 years following their registration dates, then one can see a strong increase over time in activities in this field. After the Alfa Romeo adjusting devices started being mass produced, the number of patent applications began to increase drastically. In Figure 10.65, which illustrates the situation up to January 2004, the number of applications

10

1

2

6: 7: 8: 9: 10:

3 6

4

5

Displacement magnet Gear wheel hub Straight interlocking Adjustment piston Return spring

Figure 10.64 CS adjuster from Alfa Romeo dated 1983 [10-31].

301

300 248

250 200

170

150 100

118 113

101

188

128 79

2002

2003

2000

2001

1998

1997

1999

1995

1996

1993

1992

1994

1990

1991

1988

1989

1986

1987

1984

1982

214

41

26 25 16 24 21

1983

1980

0

29

8

3

1985

50

year

150

267

259

187

100

1981

Number of patent registrations for camshaft shifters

350

Figure 10.65 Number of recorded patent applications and unexamined applications for CS shifters from 1980 to 2004.

Internal Combustion Engine Handbook | 447

6606_Book.indb 447

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Camshaft shifter

Mechanical target principles

Axial alignment of a piston or pinion in, or on, the camshaft Planetary gearbox on the camshaft drive unit Differential gearbox on the camshaft

Hydraulic target principles

the transmission geometry between the cam and the valve. These systems are used for many different engines (also refer to Chapter 10.4.2). 10.4.1.2 The Effects of Camshaft Timing Devices on Engines The setting of objectives for the application of CS timing devices can vary widely. The maximum output, the torque curve through the rpm, and the exhaust behavior can be positively influenced in passenger car gasoline engines by altering the relative angle of the CS to the crankshaft. Standard CS timing devices offer two angle positions and a variably change in the angular position. Figure 10.67 shows the fundamental options for adjusting valve lifting curves from using two continuously variable CS timing devices. The curves in dashed lines represent the possible end positions of the control time systems.

Valve stroke (mm)

is counted from 1980 to 2003 from a compiled database. For 2002 and 2003, not all applications could be entered, because there are eighteen months between the application date and the date of publication. The known adjusting devices can be categorized according to their different functional principles. Figure 10.66 shows these principles. Essentially, the adjusting devices are systems that are based on either a mechanical or a hydraulic functional principle. The most frequent solution is to axially shift a piston to change the angle by using helical toothing, similar to the Alfa Romeo adjuster. Basically, only three principles are used for production engines—with gray backgrounds in Figure 10.66. Belonging to the first group are systems that use helical toothing, like Alfa Romeo’s approach, based on a mechanical functional principle. A second solution is the hydraulically actuated chain timing device, where the CS is rotated to the desired extent by adjusting the chain sag. Belonging to a more current group are systems with hydraulically actuated swing motors on the CS drive. Individual descriptions of the systems will be covered in more detail in Chapter 10.4.1.3.

Discharge valve displacement UT

Intake valve displacement OT

UT

Crankshaft angle [°KW]

Figure 10.67 Changeable cam contours using continuously variable timing devices on intake and exhaust CS. Belt or chain shifters on the camshaft drive unit Swing motor shifter on the camshaft

Belt or chain shifters on the camshaft drive unit Centrifugal force on the camshaft

Figure 10.66 Categorization of CS shifters according to their functional principles.

All CS timing devices for production engines are on the CS drive. The timing devices do not affect the valve stroke or valve opening time. In addition, there are many other known systems in existence. The working location at which the valve stroke and the valve opening time are adjusted is usually between the cam and the valve. This allows the CS timing devices to be combined with these systems. An example of a system to change the valve stroke and/or opening duration has been determined as the so-called VTEC System from Honda [10-32]. This system enables different valve strokes and opening durations by using the change in

Since CS timing devices are only used to change the position of the control times, and not the valve lifting curves, the effects on the drive are limited. However, the potentially attainable improvement in engines is easier to estimate during development than, for example, with infinite VVA. To estimate the potential improvement, the charge cycles are calculated with numeric programs. The overall charge cycle of the engine can be estimated in reference to the torque and output behavior, and the residual exhaust gas. In this case, all the components participating in the charge cycle, such as the intake manifold or exhaust system, are parameterized and depicted in the calculation model [10-29]. The valve lifting curves are determined and included with the possible control times in the charge cycle calculations. This, therefore, enables a reliable prediction of the engine output and torque characteristic. The parameters required to adjust the CS are roughly estimated and then refined in more detail in experiments. Subsequently, the maximum torque or the maximum output can be positively influenced using a CS timing device on the intake valve side, depending on the cam contour. Only a compromise solution is possible for output and torque for engines with fixed control times and/or cam contour positions. The position at which the intake is closed on the intake valve lifting curve has a decisive influence on maximum engine output. At higher rpms, the inlet is closed at later control

448 | Internal Combustion Engine Handbook

6606_Book.indb 448

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

times. The time is selected to optimize the cylinder charge and, therefore, attain higher volumetric efficiency. A return flow of the charge from the cylinder to the intake port can be avoided by adapting the speed of the inlet closing. With CS timing devices, the valve overlap can be varied, so that the residual gas in the engine exhaust can be controlled. Normally, the residual gas is supplied to the cylinder through an external exhaust return device. The temperature of combustion is restricted by the retention of residual gas in the cylinder. This has a positive influence on the NOx emissions. With continuously variable CS timing devices, internal exhaust-gas recirculation can be achieved by changing the valve overlap. This allows the exhaust to overflow from the exhaust port to the inlet port during the overlap phase at dead center during the charge cycle. The advantage of internal recirculation is attained with a short system dead time, and more even distribution of the recirculated exhaust. Compromises always have to be made when designing the valve overlap. For example, the maximal possible valve overlap is limited by the position of the valves that collide with the piston when the overlap is too great. An example is given at this point with the control strategy for a double CS timing device of an engine from VW (Figure 10.68) [10-33], [10-34]. Four basic positions are shown for an naturally aspirated engine with controlled intake manifold and adjusting the inlet and exhaust CSs, with the relevant short and long intake manifold position. This representation also shows the influence of different intake manifold pipe lengths in combination with CS timing devices on the intake and exhaust valve side. Given the number of possible degrees of freedom that this enables, it is logical to work out a correspondingly appropriate adjustment strategy. The strategy can differ depending on the engine design. For example, a longer intake manifold pipe is necessary to attain a high torque at average speeds. As rpm increases, the intake control time is switched from “early” to “late” depending on the rpm. At a higher rpm, a short intake manifold pipe channel is selected, and the intake CS is shifted in the direction of “late” to attain maximum output. Figure 10.69 shows examples of control times of valve lifting curves for the individual CS and induction pipe positions of six-cylinder engines.

Early Position

Late Position

IOs

 26° before TDC

 26° after TDC

ICs (long channel)

179° after TDC

231° after TDC

ICs (short channel)

184° after TDC

236° after TDC

Outlet opens (short channel)

236° before TDC

214° before TDC

Outlet opens (long channel)

231° before TDC

209° before TDC

Outlet closes

 26° before TDC

 4° before TDC

Figure 10.69 Control times for dual cam adjustment of a VW-V-6-Gasoline engine with 1-mm valve stroke [10-34].

Although the first mass-produced CS timing devices, with only two control time positions, primarily sought to improve the output and/or the torque behavior, today the goal is also to control the inner exhaust-gas recirculation by using continuously variable timing devices [10-33]. The intake CS is shifted for increasing torque, especially at a low rpm, and for internal exhaust-gas recirculation, where the CA is offset from the “IO” output position toward “early” with a maximum of 52° CA. The exhaust shaft can be adjusted from the output position “outlet close” toward “early” to optimize idling or toward “late” to attain maximum exhaust recirculation rates, whereby a maximal 22° CA is sufficient for each respective angle range. Compared with a conventional two-valve engine without CS adjustment against the four-valve engine described in [10-33] with CS adjustment, there are achievable savings in consumption of 15.5% while idling and of 5.5% in the partial load range at 2000 rpm and 2 bar. When using intake and exhaust valve time offsets, the specific consumption reduction is approximately 10%. 10.4.1.3 Camshaft Shifter for Production Engines Following the start of mass production of the Alfa Romeo CS shifter, other serial production designs were introduced from, for example, Mercedes Benz [10-35], Nissan, or other companies [10-36]. Most of these systems used similar solutions as those from Alfa Romeo as the functional principle with straight/helical toothing. A system that adjusts the control times by using a chain side length change is the CS chain timing device from the company Hydraulik-Ring [10-37]. In this case, the adjusting element is located between both the dual CS drive wheels,

Utilized middle pressure

2 1

Intake Full-load manifold setting

2 3

1

4

2 1 2 1

long long short short

3 Partial load 4 Idling Rotational speed

DisInlet-CS chargeCS early late early late

late late late late

early late

late early

Figure 10.68 Control strategy for dual cam adjustment of a VW-V-6 engine [10-34].

Internal Combustion Engine Handbook | 449

6606_Book.indb 449

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Setting “late“

Setting “early“

Hydraulic cylinder

ANW

Chain tensioner

ENW

Angle alignment

whereby the intake CS (IC) is driven by the exhaust CS (EC). The timing device system combines a chain tensioner, which is commonly used for such a short drive with a hydraulic cylinder to change the chain side length. The hydraulic cylinder under oil pressure on both sides is moved, depending on the desired control time position. In this manner, one chain side is lengthened, and the other is simultaneously shortened. This timing device, thereby, implements two control time positions for the intake CS (Figure 10.70). During adjustment, the chain drive also remains taut between the two drive wheels of the CS because of the chain tensioner integrated into the system. The adjustment cylinder of the timing device is controlled by an electronically controlled hydraulic four/two-way valve. The timing device solution, as shown here, uses hydraulically VVA, because the end positions are held only by oil pressure (Figure 10.71). The design is formed in such a way, so that the adjustment is made with the available engine oil pressure even under difficult conditions. A costly additional oil pump can, thereby, be dispensed with. This adjuster principle is a mass production for different engines from Audi, Porsche, and the rest of the Volkswagen concern [10-37], [10-38], [10-39].

Figure 10.70 Functional principle of the CS chain timing device [10-38].

Developments of the continuously variable adjustment of intake CSs have enabled the execution of a retention of more than two CS positions. This solution has not been implemented in mass production engines. BMW was the first company to also use continuously variable adjustment of the CS in mass production (Figure 10.72). Initially, this was only used for the intake CS; later, it was followed by the continuously variable adjustment of the intake and the exhaust CSs [10-40]. Straight/Helical toothing

Adjustment piston

Chain tensioner

Figure 10.72 Continuously adjustable CS shifter of a BMW six-cylinder engine [10-40].

Figure 10.71 CS shifter as a chain timing device for the Porsche Boxter [10-39]. See color section page 1081.

A new generation of CS shifters is represented by systems designed according to the swing motor principle [10-41] (Figure 10.73). In this system, both the intake and the exhaust CSs can be simply adapted to the existing cylinder heads. Inside the timing device is a pivotable rotor, which is firmly fixed to the CS. The outer part is either driven by a chain or a synchronous belt. The connection between the outer and inner parts is formed by the oil space that is filled with engine oil pressure and that contains the pivotable rotor. The blades of the rotor are supplied with oil pressure through an electronically controlled four/ two-way proportional valve. The relative angular position of the CS is changed depending on the change in oil pressure on both sides of the rotor. The angular position of the CS will

450 | Internal Combustion Engine Handbook

6606_Book.indb 450

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

Camshaft sensor

Camshaft CPU

Swing motor shifter Crankshaft sensor

Temperature sensor

4/2-Way proportional valve

Figure 10.74 CS shifter arrangement in a six-cylinder engine according to the swing motor principle [10-34]. See color section page 1082.

Engine oil pump

Throttle Valve

Figure 10.73 Functional principle and control loop of a continuously variable CS shifter designed according to the swing motor principle [10-41].

be measured with a sensor compared with the position set by the engine electronic system. The desired position of the CS is permanently readjusted by controlling the proportional valve to hold stable intermediate positions of the rotor and, thereby, of the CS. The oil is only supplied by engine oil without an additional pump. The system is controlled depending on the rpm, load, and engine temperature. Compare with conventional, toothed, and continuously variable CS shifters, these systems represent a much more economical solution, so that one can expect their increasing use in series gasoline engines. The time and the money spent on manufacturing the components can also be reduced, when the parts of the components are sintered and the seal of the oil chamber has a simple design. Timing devices of this type can be even more economical than toothed, two-step timing adjuster devices. A more precise description of this system follows in [10-42]. The design of the dual CS shifters according to the swing motor principle from Hydraulik-Ring for a six-cylinder engine is illustrated in Figure 10.74 [10-33]. Figure 10.75 shows the arrangement of two CS shifters according to the swing motor principle for the left cylinder bank of the 3.0 IAudi V-6 engine. A two-step timing device is used on the exhaust valve side on this engine, and a continuously variable timing device is used on the intake valve side. With this design of a synchronous belt drive, the CS timing device housing needs to be encapsulated as an oil-tight component. In addition to the mass production implemented systems from Hydraulik-Ring for Audi and VW, similar systems with swing motors according to the swing motor principle are also implemented by Renault, Toyota, and Volvo [10-35].

Rotor setting „late“

Figure 10.75 Dual cam adjustment system in a 3.0 I -V-6 gasoline engine from Audi [10-43].

A wide variety of hydraulic valves are used for the hydraulic control of the CS shifter [10-42]. Directional controlled valves are generally used to control the oil flow. These can be, hereby, subdivided into proportional and switching valves. CS shifters that only hold two end positions and, hence, can have only two different control times are equipped with four/two-way valves. Today, four/three-way proportional valves are primarily used for continuously variable systems (Figure 10.76). The actual know-how for hydraulic valve engineering has less to do with

Internal Combustion Engine Handbook | 451

6606_Book.indb 451

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

6

Q in l/min

5

B

P

A

4 3

Oil flow from P to A

Oil flow from P to B

2 1 0

0

T

100

200

300

400

I in mA

manufacturing individual valves for small series, rather more for implementing the technical requirements for economical large series production. Therefore, the difficult boundary conditions of series production need to be dealt with, such as dirty oil, engine vibration, high temperature fluctuations, or fluctuations in the vehicle power supply. Usually a special valve is used to adapt the valves to the individual engine. A well-thought-out modular system is useful to meet the main demands for economical mass production. The close collaboration between the developers of the VVA system and the developers of the engine is essential for successful mass production. 10.4.1.4 Perspectives for Camshaft Shifters An overview relating to the patent applications for CS shifters, and the number of different systems in production engines clearly indicates that, in the future, probably all modern manufacturers of gasoline engines will implement CS shifters. The author does not know of any serial systems of production engines with only one CS, in which the intake and exhaust cams rotate against each other in opposite directions. Perhaps, it might make sense, however, to offset the entire CS through a CS shifter, if only at narrow adjusting angles. There are many reasons for implementing timing devices that allow the continuous variable offsetting of the CS. It is recommendable to also increase the use of these systems for multivalve engines with dual CSs in which one system is affixed to a CS. In particular, the control of the internal exhaust-gas recirculation with continuously variable systems can have a positive influence on the direct exhaust emissions. To implement the functional principle, designers are constantly preferring timing devices with swing motors. The primary focus of these elements is on the development, and, thereby, the use of light components and weight reduction. Aluminum is increasingly being used as the preferred material for serial production of CS shifters. A current example is, hereby, an adjuster made of sintered aluminum from Hilite International/ Hydraulik Ring in a six-cylinder engine from BMW (Figure 10.77). This timing device is also designed according to the swing motor principle, and is equipped with

500

600

700

800

Figure 10.76 Cross-sectional illustration, Q–I Characteristic curve, technical data, and the hydraulic symbol of a four/three-way proportional valve.

rotor blades that are individually sintered with the boss of the timing device. To lock the timing device in its original end position, it is equipped with locking pins and the respective hydraulic control system.

Figure 10.77 CS shifter made of aluminum for a six-cylinder engine.

Using two CS shifters in this engine creates a total weight saving of 1.3 kg for each engine, as an adjuster without a valve now only weighs 450 g instead of approximately 1000 g. To compensate for the unequal torque on the CS, and for quicker adjustment of the timing device in the direction of the previous CS setting, a spiral spring is used. The silicon proportion of the aluminum alloy is approximately 15% for durability and wear reasons, a criterion that can be especially achieved with sintering. For the first time, this innovative design enables the implementation of a CS chain wheel made of aluminum in a production engine. The know-how for this timing device is founded on the control of the tight gap between the rotor and the stator housing, as well as the ingenious assembly of the individual components. Further reductions in weight and single parts can be expected when the CS shifter is constructed from synthetic materials.

452 | Internal Combustion Engine Handbook

6606_Book.indb 452

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

Initial engine trials have clearly demonstrated the potential for this (Figure 10.78). This illustrates a CS shifter from Hilite/ Hydraulik-Ring, which is made of a thermoset material that eliminates a sintered design by using a special design and injection technology for two components.

for application areas for implementing CS shifters, whereby there is the possibility for creating additional optimization potential for combustion engines. Infinite VVA systems must focus on the potential improvement that can be achieved with these measures.

10.4.2 Systems With Stepped Variation of the Valve Stroke or Opening Time

Figure 10.78 CS shifter made of a synthetic material [10-44].

The swing motors can be economically and easily controlled by hydraulic multidirectional control valves. The harmonization of the valves with the timing device is one of the essential know-how factors during development. Here as well, there is potential to lower costs. In contrast to the rest of the other known systems for changing the control times during operation, CS shifters have a simple design and are correspondingly economical. These systems should be integrated in the cylinder head design early on in the initial development of a cylinder head. Additional cost-saving possibilities can result when the CS shifter for the adjustment of the required proportional valve and the CS could be realized as one system unit. The adjuster and the valve could be securely connected with the CS as one unit, and the OEMs could then be supplied so that additional cost reductions would be achievable. The oil circulation supply required for regulating the system can then be more easily harmonized with the hydraulic directional control valves. Because of the attainable improvement that is achievable, new engines will probably increasingly use continuous variable timing devices. The control of the inner exhaust-gas recirculation requires cylinder head designs with at least two CSs. The continuously variable CS timing device works similarly in gasoline engines with DI. In this case as well, the internal exhaust-gas recirculation can be controlled by implementing this system. CS shifting will also be used for this still-new combustion process. CS shifters can be combined with VVA that allows the valve stroke, or opening period, to be varied. Porsche has implemented this in a production six-cylinder engine among other implementations [10-45], [10-46]. This enables many possibilities

Honda has used VVA, for the first time, in the mass production of gasoline engines in its so-called VTEC-System, which influences the valve stroke or valve opening time [10-32]. The principle is based on a rocker arm solution in which, by moving smaller hydraulically actuated pistons inside the rocker arm, different coupling statuses can be achieved to, thereby, enable switching back and forth between different cam contours. Figure 10.79 shows a principle outline of the system used as an application in a four-valve engine with dual CSs. Figure 10.79 (right) shows an isometric representation of the valve and CS arrangement. For each cylinder, the CS has a central cam with a larger valve stroke and opening time geometry. On each respective side, there is a cam profile with smaller cam contours. Inside the rocker arm module, a two-part piston is shifted by oil pressure parallel to the axis of the CS bearing. This is executed depending on the engine mapping as a function of the engine rpm, the induction pipe pressure, the vehicle speed, or the coolant temperature. The oil for switching the cam contour is supplied through openings and channels in the bearing shaft on which the rocker arm module pivots. When operating at lower rpms, the smaller cam contours act on the gliding surface of the rocker arm. The separation of the rocker arms are executed by precisely harmonizing the geometry of the two-part adjustment piston with the rocker arm width. A relative stroke is created in this case between the central rocker arm and the individual rocker arms on the side. The central rocker arm is, hereby, supported on a spring element. The space for this must, therefore, be created in the cylinder head. In cylinder head concepts with more than four valves, this is a particularly demanding challenge for the developer. When coupled in a status—as illustrated in Figure 10.79—the central cam acts on the rocker arm module, and all the components are moved simultaneously without a relative stroke. Resetting the two-part adjustment piston is executed by a small spring. The adjusting oil pressure is established by the engine oil circuit, without an additional oil pump. The VTEC System is implemented on the intake and exhaust valve side. Honda has developed many patent applications for this and similar solutions. The number of different inventors involved in these patent applications is enough to indicate how enormous the interest is in this development. Four-valve solutions with one or two CSs have been created for production engines. Rocker arms, as well as rocker rollers, with sliding surfaces can be used for such applications [10-35]. The current Honda program uses the VTEC-System in almost every vehicle with different engines. Up to three differently acting cam contours have been implemented in this context.

Internal Combustion Engine Handbook | 453

6606_Book.indb 453

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Central tappets

Minimum tappet contours

Central cam follower

Separate cam follower

Oil pump

Two-part piston

Spring element

Figure 10.79 Honda VTEC system [10-32].

Mitsubishi has also developed a similar series system for four- and six-cylinder engines based on the same functional principle [10-35]. This solution implements three different cam contours, whereby one cam contour consists of a pure base circle to, thereby, achieve the valve stop. Both engines have two and/or three cylinders, which are stopped using this valve actuation system. To achieve this task, Mitsubishi requires a small oil pump in the cylinder head. Daimler Chrysler has implemented a VVA to stop the cylinders in its serial production V-8 and V-12 engines. The solution used is based on a valve rocker module, which is used with a central CS in a three-valve approach. Figure 10.80 shows the valve rocker module of this system without a CS. The functional principle is the same as the described Honda solution. Inside the rocker roller module, a two-part adjustment piston is moved electrohydraulically against a spring force.

Depending on the coupling status, different cam contours are selected between valve strokes, except that one valve stroke is a zero stroke, and this stops the valves to shut off the cylinders. The primary objective of the system used in this instance is to reduce fuel consumption during partial load operation by stopping the cylinders. This is particularly effective with high-displacement engines with higher numbers of cylinders. There is little effect on the smooth running of these engines. These measures can achieve the reductions in consumption of approximately 15% in contrast to the conventional engines. Similar to Mitsubishi and Honda, Toyota has also pursued solutions for mass production involving switching the valve contour on the intake and exhaust valve side. In this case as well, an adjustment piston in a rocker arm module is electrohydraulically pushed against a spring force (Figure 10.81).

Roller

Two-part piston

Hydraulic play compensation

Lost-MotionElement

Slide support

Adjustment piston

Figure 10.81 Toyota valve actuation VVTL-i for different valve strokes [10-45]. Lost-MotionElement Roller support

Figure 10.80 Rocker roller module for stopping the valves from Daimler [10-35].

The interesting fact about this solution is, therefore, that a rocker arm module is used, in which a roller is the contact surface that faces the cam at a low rpm, and a sliding surface is used at a high rpm. At a high rpm, the rocker arm module is pivoted by the sliding contact, whereby a stop below the lost-motion element provides for the coupling. The stop is retained by oil pressure, and is moved by spring force at a low rpm toward the bearing center of the rocker arm module.

454 | Internal Combustion Engine Handbook

6606_Book.indb 454

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

At a high rpm, the sliding surface lowers with the lost-motion element into the rocker arm module. The spring force of the lost-motion element can be minimal, because the moved mass of the element is also low. For this solution, Toyota also uses a continuously variable CS shifter on the intake valve side. This combination allows the valve lifting curve to be widely varied in contrast to engines with fixed control times. Porsche has traditionally implemented bucket tappet solutions for their four-valve engines. In 2000, Porsche presented a turbo engine for the first time with a VVA, which used different valve strokes on a switching bucket tappet [10-46]. Furthermore, an additional CS shifter was implemented on the intake valve side to provide two control time settings. As with Toyota, a combination of two independently functioning systems of VVA is, thereby, used. The switching bucket tappet can execute two valve strokes and consists of an inner and an outer tappet (Figure 10.82). Its rotational position is oriented by a special guide in the cylinder head. This enables the surface to be designed as ball-shaped for correspondingly strong maximum strokes. Inside the pushrod are small hydraulically actuated pistons that activate the inner or outer tappet for valve actuation depending on the position. In this case as well, the term“mechanically VVA” is appropriate, because only the adjustment piston is electrohydraulically controlled, and the valves are actuated by the mechanical positive engagement of the components. Internal stem External stem Adjustment piston

Rotating bearing fixing

Figure 10.82 Switching bucket tappet from Porsche [10-46].

An additional solution for valve switching can be seen in the V-8 engine with push rod drive from Daimler (Figure 10.83). In this instance, there is a hydraulically controlled, and mechanically lockable, element located between the bottom positioned CS and the push rod, which leads to the cylinder head. The switching element is actuated with oil pressure, whereby a smaller locking piston is moved which, thereby, assumes the unlocking and/or locking for the switching element. The advantage with this solution is that there is no

requirement for altering the basic engine in the cylinder head area, and that the design adjustments on the cylinder block are rather minimal. The cylinder switching on a V-8 engine has the primary objective of reducing consumption which, in testing cycles for higher cubic capacity engines, can result in savings of up to 15%, and which were already realized in production in the USA in 1980 on an eight-cylinder from Cadillac [10-35].

Figure 10.83 Valve switching mechanism from Daimler for a push rod V-8 engine.

Usually, a new generation of cylinder heads is used with the application of this type of valve actuation. The geometry of the cam contours is conventional, that is, they are smooth and can be manufactured by normal cam production systems. Corresponding to the categorization in Chapter 10.4, these solutions represent systems with a variable transmission element between the cam and the valve. The functional principle is mechanical, because only mechanical actuators and contact elements are used in the flow of force toward valve actuation. The adjustment piston is controlled hydraulically through an electrically actuated directional control valve. Because of the number of patent applications in this area, it is difficult for the designer to retain an overview of the patent situation. It is anticipated that these, or similar systems, will become more widespread in production engines. All the switching on or switching off systems for production engines for different valve stroke contours are not based on roller cam follower designs. Roller cam followers have been increasingly favored, however, in modern gasoline and diesel engines in recent years because of their lower frictional ratios. Figure 10.84 illustrates a schematic image for a roller cam follower design with a possible functional location for arranging the valve contouring or valve switching. Current publications or patent applications clearly indicate that there are some associated developments underway. Only a few ideas exist for implementing a mechanism for a functional location A directly on the cam. Lowering the hydraulic playcompensation element (functional location B) is the only way to implement pure valve switching. There has been a lot of

Internal Combustion Engine Handbook | 455

6606_Book.indb 455

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

A

D

B

E

C

Target location of the variability A B C D E

On tappets On hydraulic support element On the roller bearing On the lever frame Above the valve

effort invested for solutions that provide switching on and switching off mechanisms directly on the lever (functional location D). Many systems are, hereby, noted in patent literature such as the patent application DE 19510106 from BMW with a so-called kink lever from 1995.

8,8

16

2b

10

7

13

15 4

6 11

12

14

simultaneous controllable residual gas volume are the only way to achieve lower consumption improvements. A positive influencing on the consumption is only initially possible by using the load control for the engine with an infinite VVA. In this conjunction, the valve contour switching system is in direct competition with the infinitely variable mechanically actuated valve actuation. These are, however, more expensive and complicated to develop and to apply for engines, but, however, provide considerably larger potential with regard to lowering the fuel consumption.

10.4.3 Infinitely Variable Valve Actuation

2a

5

11′

Figure 10.84 Possible functional location for valve contour switching on a roller cam follower.

3

Figure 10.85 Patent application DE 19510106 for valve switching on a roller cam follower.

The achievable potential for the engine with valve contour switching is limited. The system is, however, suitable for increasing the torque and the performance through a higher rpm range in the engine. There are no noticeable advantages to be gained by lowering the consumption. The simultaneous application of continuously active CS shifting and the

In the following section, systems will be considered, which enable an infinitely variable valve lifting curve. These are, therefore, considered to be systems with mechanical, hydraulic, and electromechanical functional principles that use a CS or a CS-free system. 10.4.3.1 Review of the Development of Infinitely Variable Mechanical Valve Actuation Experiments to realize an infinitely VVA with a mechanical functional principle have been around as long as engine designing itself. The first ideas were formulated by Louis Renault in 1902 (Figure 10.86). Even back in the days of his German patent application DE 145662, he envisaged two contacting together rocker arms, whose position was controlled with eccentric controlling, so that when the eccentric shaft rotated, on which the rocker arm was bearing, the valve stroke would be variable. A CS-free system using a crankshaft for valve operation was proposed by Torazza (Figure 10.87). A rotating lever is located on the crankshaft drive whose working curve takes an oscillating movement and which has an effect on the valve for the valve stroke by transmitting motion through a rocker arm. Through the change in the position of the rocker arm, the valve stroke and the opening duration can be varied. The

456 | Internal Combustion Engine Handbook

6606_Book.indb 456

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

system could not establish itself because of the high construction cost, the higher friction, and the large space required in a production engine.

utilization in a four-valve engine is unlikely. In addition, this system can only alter the valve stroke, and the valve opening time remains constant.

Figure 10.86 VVA according to Louis Renault, 1902. Figure 10.88 VVA according to Titolo [10-48].

Figure 10.87 VVA according to Torazza [10-47].

Titolo [10-48] used a system with axial displacement of conical cams for the stroke variation (Figure 10.88). According to the categorization in Chapter 10.4, this is a system that implements the variability by moving space cams. This principle is known from old controller units of marine engines, which used axially movable space cams for reversing. Because of the required width of the steep cam and the, therefore, required space, a

The variation in the opening period with a conical space cam leads to cam geometries that are no longer economical to manufacture, and wherein a contact point is formed between the cam and the consumer. Therefore, this type of solution is not applicable for high-valve accelerations and, therefore, for production models. At the University of Vienna, a system was developed by Wichart, in which the CS actuates a rocker arm [10-49]. This rocker arm is not mounted on an axis, but relies on a crank from. A needle roller bearing that is located on this crank rocker acts during the valve stroke with a rotating control shaft for the CS rpm, whose profile comprises an arc and a speed regulation curve. If the control shaft is rotated relative to the CS, this affects not only the arc but also the activating curve on the roller and the valve closes earlier. The solution could also not prevail on production engines, because the intake valve strikes the valve seat very hard during the activating process. To remedy this, a hydraulically acting valve brake would be necessary, which would significantly reduce the valve touch-down speed. The “Delta” controlling described by Kuhn and Schön uses an intermediate member with a working curve as an infinitely variable valve lift actuator between the cam and the pushrod. This intermediate member has at the side at which it is supported on the housing, when resting and working curve the control section.

Internal Combustion Engine Handbook | 457

6606_Book.indb 457

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

L

K

B C

F

Öl-Zufuhr J H I

A

G

E

D

B A C D E F G

Displacement camshaft H Hydraulic silencer Control camshaft I Oil discharge borehole Rocker arm J Oil supply borehole Valve K Oil discharge gap Valve spring L Rocker arm spring Crank Needle rollers, mounted roller

The load controlling for the engine can, therefore, be controlled solely by variable valve lifting. The system is equipped with a special force transmission mechanism with an intermediate lever between the cam and the valve and, corresponding to the categorization in Chapter 10.4, is classified as a mechanical VVA system. Figure 10.91 illustrates the Valvetronic system with the intake CS and the intake valve module. In the flow of force between the CS and the valve, there is a transmission mechanism located that swings the roller cam follower to actuate the valves. A setting shaft designed as an eccentric shaft, and driven by a dc electrical control motor, changes the lever geometry of the transmission mechanism. Valve strokes between 0.3 and 9.7 mm [10-51] and/or 0.18 and 9.9 mm can be set for the second generation of the system [10-52]. The entire adjustment process, through the complete stroke range, is executed within 0.3 s. Therefore, the conventional throttle valve can be dispensed with. The friction loss of the valve gear can be reduced during operation by the variable valve strokes, compared with conventional valve trains, because the valve springs are compressed less with smaller valve strokes. Actuator

Shift shaft Transfer mechanism Camshaft

Figure 10.89 VVA according to Wichart [10-49].

Cams Intermediate element Control section Housing Housing

Rest section

Output element Valve

Figure 10.90 VVA according to Louis Renault [10-50].

A valve lift only occurs when the intermediate member is supported during the movement to the control section on the housing. As long as the locking portion of the working curve has contact with the housing, the valve remains closed. The main disadvantage of this solution is the high friction, because the intermediate member slides both on the cam and on the driven member. 10.4.3.2 Mechanical Systems in Production With its so-called Valvetronic system, BMW has created a continuously VVA system on the intake manifold valve side.

Figure 10.91 “Valvetronic” system from BMW as a module with the valve train components [10-51].

The fundamental principle of the system can be best explained by the sketch in Figure 10.92. An intermediate lever is implemented between the cam and the roller cam follower with a working curve, which is altered with an eccentric shaft in its bearing in its upper linkage. With a fixed positioning of the intermediate lever [Figure 10.92 (left)], the resulting cam rotation through the working curve produces a constant valve stroke with a constant valve opening duration. Depending on the linkage of the intermediate lever, there is a resulting differential valve stroke process. By altering the bearing point position of the intermediate lever along an adjustment path, the transmission ration also alters and, therefore, the valve

458 | Internal Combustion Engine Handbook

6606_Book.indb 458

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

Fixed rotation point equals fixed control time

Variable rotation point equals variable cam stroke

Rotation point

Adjustment route

Eccentric shaft varies the rotation point and the valve flow Eccentric shaft as a solution

Work curve

Figure 10.92 Principal illustration of the Valvetronic system.

stroke with simultaneous alteration of the valve opening duration. For the infinitely variable valve stroke movement, it is necessary to find a design-manageable solution for moving the bearing point position of the intermediate lever and to guide this along the adjustment path. The Valvetronic system uses the bearing point position of the intermediate lever and, therefore, the adjustment of the desired control times through an eccentric shaft. A spring supports the play-free positioning of the intermediate lever on its contact position. Figure 10.93 shows an installed situation for a VVA system in a cross section of a cylinder head for a four-cylinder engine. The exhaust valve side remains conventionally actuated with a roller cam follower. The required space for valve actuation is kept within limits. The space in the vehicle is only required for the control motor. The eccentric shaft, the transmission Actuator

Shift shaft Transfer mechanism Camshaft

Figure 10.93 Arrangement of the “Valvetronic” system in a cylinder head [10-51].

mechanism, the CS, and the control motor are preassembled in a separate cast holder and attached as a module to the cylinder head. The overall optimization of cylinder head and valve train is an iterative process in which CAD models, finite element models for the structural strength of interpretation, and multibody simulation models for the interpretation of the dynamic behavior can be effectively used nowadays. Figure 10.94, hereby, shows an example for a complicated model for a multibody simulation of the valve train. This enables accurate checking of the fundamental kinematic interpretation of the valve train components. The resulting dynamic loads at the contact points are important design parameters for the dimensioning of bearings, or individual components of the valve train. In the validation of the simulation results, there is the special expertise from the development of such systems such as the Valvetronic. With the variation of the transmission mechanism, and the simultaneous change in the spread of the inlet valve stroke by a CS shifter, the intake CS enables a load-dependent and rpm-dependent valve-stroke variation (Figure 10.95). The load control method used in this case is referred to early intake closing (EIC). This system aims to achieve a reduction in the charge cycle losses in a wide engine load range. The process is, hereby, interrupted in such a way, so that the still existing throttle valve remains as a fully open throttle valve during aspiration. The intake valves are closed exactly time when the desired mixture composition is in the cylinder. The charge cycle benefits considerably from full load. Therefore, acting with small loads means that the system is particularly fuel reducing. With small loads, and correspondingly short valve strokes, the valve seat region acts as a throttle point, and the inflow velocities increase from approximately 50 to 300 m/s.

Internal Combustion Engine Handbook | 459

6606_Book.indb 459

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Force element A

Bearing 2D with play, stiffness and damping

Single valve drive model B A A A

B

A B

Contact 2D Circuit with work curve (tappets, excenter…)

B

Friction Stiffness Damping

C

Contact 2D Circuit with circuit

D

Contact 2D Circuit with level

E

A

C D

HVA Model with subsequent setting and sinking behavior Spring model discretized in approx. 20 partial bars with winding contact D

F

Helical toothing with play, stiffness, damping F

B Single valve drive model

– Bending and torsion elastic shafts – Screw mode – Consideration for bearing misalignments – Consideration for various bearing stiffnesses along the engine axle

This effect promotes the mixture preparation for combustion to a particular degree. BMW state fuel savings with some load points at low loads of approximately 20%, which with a stoichiometric air ratio of λ = 1 at the mean point, equates to approximately 10% [10-52]. In the European test cycle, the new six-cylinder with the second generation of the system is operated with valve strokes of less than 1.5 mm, and up to 4 mm was only achieved on urban trips for a short period. Compliance with the EU4 emission standard is achieved with Valvetronic engines, without the need for sulfur-free

Figure 10.94 Single and total valve train model for dynamic interpretation [10-51].

fuel. The engines can, therefore, be used worldwide using conventional gasolines. The basic principle of the eccentric adjustment is executed using a dc motor through a spiral-gear actuator. To adjust from minimum to maximum stroke in the less than 0.3 s, the control structure was redesigned (Figure 10.96). Only the integration of all mechanical, electrical, and control engineering elements of the Valvetronic system, and the use of functional options with a new combustion control, can tap the potential of this fully VVA.

460 | Internal Combustion Engine Handbook

6606_Book.indb 460

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

Throttled engine

p

VALVETRONIC-Engine

p

Expansion

Compression EÖ

AS

Return spring

Background AÖ

EÖ AS

ES V

AÖ

Center of rotation

Eccentric shaft

ES

V

Intermediate lever

Stroke

Work curve with ramping function HVA AÖ

EÖ AS

ES

Crankshaft alignment EÖ Inlet opens AÖ Outlet opens ES Inlet closes AS Outlet closes

(180°) (540°)

Figure 10.97 Execution of the second generation of the Valvetronic system [10-52].

Figure 10.95 Principle of previous EIC.

Vanos KW-Signal ............

Valvetronic functions Combustion regulating

are needed. The opening and closing ramps are incorporated in the work curve of the intermediate lever. In motion imaging of the ramp functions is realized in that the intermediate lever clearly moves in front of the roller cam followers, or significantly slows after this. The support of the intermediate lever upwards through a cylindrical cam attached to the cylinder head. The backdrop has a circular path around the center of rotation of the roller cam follower. The manufacture and assembly of the individual components of the valve train requires particular attention for component tolerances. The lateral fixation of the intermediate lever is assumed, on the one hand, through the cam, and, on the other hand, by the eccentric shaft. All the contact points could be executed as roller contacts (Figure 10.98).

Combustion process 3-rowed roller bearing

Figure 10.96 Valvetronic system with adjuster mode [10-52].

The design of the system for the six-cylinder engine is shown in Figure 10.97. The main differences of this variant compared with the four-, eight-, and twelve-cylinder engines are in the altered basic kinematics. Here, a fixed center of rotation is provided, by which the intermediate lever performs a pure rotation per working cycle. The so-called working curve on the intermediate lever acts on the hydraulic play-compensation element that supports the roller cam follower. For a trouble-free kinematic functioning of the valve train, clearly defined ramp functions

Housing in MIM production process

Work curve milled, classified, exactness in µm-range

Figure 10.98 Structure of intermediate lever.

Internal Combustion Engine Handbook | 461

6606_Book.indb 461

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

10 9

Full load 9.9 mm

Stroke characteristic diagram, kinematic

8

Stroke [mm]

7 6 Partial stroke 1 mm 2 bar 2000 1/min

5 4 3

Early „inlet closes“

Idling 0.2 mm

2 1 0 90

135

180

225 Degree [°NW)

The achievable generation for intake valve lift curve is shown in Figure 10.99. The design of the reduced valve accelerations could be made up to the level of previous tappet engine with 80 mm/rad2. For medium and low valve strokes, shorter valve opening times can be achieved with the new Valvetronic interpretation (Figure 10.99). In the second generation of the Valvetronic system, adjacent intake valves are opened at different later times to increase the charge movement, which is, hereby, described as “Phasing.” This is implemented with the valve strokes of up to 5 mm, a parallel opening occurs above this point. Structurally differ in the eccentric contours the adjacent valves. In addition to the phasing, masking the valve seat (masking) is used at higher stroke valve. This, therefore, enables a higher turbulence to be achieved for the incoming mixture (Figure 10.100). In summary, the overall system of Valvetronic in the second generation, with the additional utilization of CS shifters on six-cylinder engines including the mounting position in the cylinder head, is illustrated in this chapter in Figure 10.101. The complete bearing of the intake CS with lateral fixation and the required housing peripheries for the valve train components are molded in one piece on the cylinder head lower component to create a rigid and very innovative solution.

Valve stroke

Phasing (unequal valve stroke per cylinder) Range: Lower partial load, Catalytic heating

E1

270

Figure 10.99 Producible swarm for intake valve stroke curves.

Figure 10.101 Installation situation of the second generation of Valvetronic in the cylinder head.

The Valvetronic system will be combined with DI in a V-12 engine. BMW, therefore, also showed the way for, in addition to the infinite variability of the valve train, the special potential of the direct-injection for utilization in production engines in this combination [10-53].

Masking

5 mm stroke

Stroke difference

E2 Idling

Full-load

Figure 10.100 Phasing and masking in the second generation of Valvetronic.

462 | Internal Combustion Engine Handbook

6606_Book.indb 462

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

10.4.3.3 Mechanical Systems in Development The following will concentrate on two more infinitely variable valve control systems with mechanical operating principles. These systems have been developed along the wide variety of systems available on the market, after the launch of the Valvetronic system, for application in large-scale production [10-54]. Initially, the system VARIOVALVE from IAV is described (Figure 10.102). This is also an infinitely variable solution based on a roller cam follower drive, which can alter the valve stroke and simultaneously the valve opening period up to the zero stroke. The system uses an additional moving gear member with a return spring between the cams and roller cam followers. All contact points are feasible with roller contacts. The control curve is mounted stationary on an adjusting shaft. By rotating the controlling shaft, and the resulting change in transmission geometry, the moving gear element rolls differentially on the control curve, so that, thereby, the desired valve alteration occurs. The system can take full control of the engine load so that, even in this case, the conventional throttle valve can be omitted. The compact transmission member creates the possibility for high system rigidity. The recoverable lift curve swarm is symmetrical with respect to the standard maxima. When using a CS shifter, the internal residual gas control can also be realized using the valve opening time shifting.

Figure 10.103 Adjusting gear and valve train components of the VARIOVALVE system.

central roller on an axis. This intermediate lever and/or rocker arm are driven by the cam of a CS and move the fork lever with the central roller in a fixed location gallery. The fork lever of the system performs a pure tilting movement about an axis. The working curve is integrated in the intermediate levers, which then is executed on the roller of a roller cam follower, and thus produces the valve stroke.

Camshaft

Forked lever

Background

Eccentric shaft

Work curve

Roller cam follower

Fork lever in detail

Valve with different strokes

Figure 10.102 VARIOVALVE system from IAV.

The gear required for rotating the control shaft, as well as the valve train components, for a possible cylinder head installation situation is shown in Figure 10.103. The second very interesting upcoming development was submitted by the Hilite International/Hydraulic Ring and enTec CONSULTING companies with the UniValve system. This is also deals with an infinitely variable mechanical valve actuation, which can also open the valve strokes of individual adjacent valves of a cylinder differently (phasing). Figure 10.104 illustrates the main valve components of a cylinder. A fork lever located between the cams and the roller cam follower consists of two intermediate levers that are arranged with a

Figure 10.104 Design of the UniValve system [10-55]. See color section page 1082.

The background curve is defined by a circular path with the center of the roller of the roller cam follower and a radius that is defined by the roller diameter of the roller of the rocker arm. To set a valve stroke, the fulcrum of the intermediate lever is shifted toward the CS with an eccentric shaft. The displacement path is approximately 3.5 mm, so that the valve stroke is adjusted from 0 to 10 mm. The cam stroke is designed with approximately 5 mm, so that a very compact arrangement results. The return spring is supported in sliding bearing rollers on the connecting axis of the rocker arm. As a complement to the valve stroke variation, a continuously operating CS

Internal Combustion Engine Handbook | 463

6606_Book.indb 463

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

shifter is used on the intake CS, which, thereby, guarantees the optimum selection for the inlet circuit. The mass distribution of the tilting lever is compensated, respectively, for its rotation about the eccentric shaft, so that the height of the contact force does not increase to the eccentric shaft with the rotational speed. The valve train has been tested up to an rpm of 8000 L/min. The adjustment from zero stroke to full stroke occurs within 120° by rotating the eccentric shaft. The holding and adjusting torque on the eccentric shaft for two valves per cylinder is in the range of approximately 4 Nm (Figure 10.107). In the UniValve valve train the CS center is shifted by 5–10 mm higher and 10–15 mm for the intake system or, depending on the cylinder head geometry, down to the center of the cylinder head when compared with a roller cam follower valve train. The additional restoring spring, which is guided in low-friction rollers on the axis of the fork lever, lies in the shadow of the CS shifter, and does not lead to an increase in the package-relevant engine height. The installation of the system, including adjustment actuator in a four-cylinder engine prototype, is shown in Figure 10.105.

Figure 10.105 Installation of the UniValve system in a four-cylinder head.

Depending on the cylinder head design, the system can alternatively be mounted, including CSs, valve train components, and CS shifters as a module with a CS carrier on the cylinder head lower part (Figure 10.106). This scope can be offered as a complete system by the system supplier. The UniValve fork lever has been revised several times. The lever has been significantly reduced, and the roller for support on the eccentric shaft has been replaced by a flat contact surface (see Figure 10.107). Through these flat contact surface to the eccentric shaft, there is a reduction in tolerances, in particular, the tolerance play of the roller, the roller bore hole, and the roller axis are eliminated. The contact surface is ground in one clamping with the working curve. Therefore, and only using one additional transmission link, the system offers the best conditions for achieving the stroke tolerance for the intake valves of ±5% of a multicylinder engine without adjustment

Figure 10.106 CS carrier with valve train components in modular design.

in a large-scale production. The friction in the contact surface because of sliding is, thereby, slightly increased (Figure 10.107). The eccentric shaft is located as a floating axle directly in the cylinder head, and the outer diameter can be ground centerless after hardening, resulting in a very cost-effective and very rigid solution. The high rigidity of the eccentric shaft and the stationary-located setting, which is preferably made of steel, are a prerequisite for high acceleration and high speeds. The eccentrics of two intake valves can be turned to each other, whereby the so-called phasing of the valves results (see also Figure 10.104). The eccentric shaft illustrates a very small holding torque for two positions. In these two positions, the contact force vector from the intermediate lever goes right through the center of rotation of the eccentric. Thus, such an eccentric for a two-stage valve contour switch is the first development stage of a new cylinder head. Without major changes to the cylinder head geometry, a gradual development over a two-stage switching would be possible to an infinitely variable stroke system. For a switching system between two lift curves, it is particularly important that the energy expenditure for retention of the position is as low as possible. Thus, the system is also suitable for a gradual expansion with high future security. One can, thus, represent an infinitely variable and throttle-free load control without a large change in the cylinder head through a relatively inexpensive two-stage system by expanding with a sensor and a customized actuator system. In a throttle-free load control, the full load is not achieved with full stroke over the entire rpm range. Because the full load in the lower speed range is achieved with partial strokes, it is beneficial when the intake closing is variable with the valve stroke. Thus, the torque on the full load in the lower rpm range without CS adjustment can be significantly increased, that is, without additional regulation of the position of the CS relative to the crankshaft. With such a type of control, no compromise of timing in the idling range or in the upper

464 | Internal Combustion Engine Handbook

6606_Book.indb 464

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

Stopping torque / Cylinder [Nm]

1

Roller contact Level slip contact

0 Sliding surface

–1

–2

–3

–4

Valve height Roller contact: Level slip contact: 0

60

Level slip Roller contact contact

8,457 mm 8,559 mm 120

180

240

300

360

Cam angle [°CS]

speed range must be entered. An opening time with 320 or 340° CA at maximum stroke, as with a pure motor racing design, does not lead to irregular running at low revs or no loss of idle quality. This results in a connection between the opening time and the valve stroke, eventually the opening cross section, which is shown in Figure 10.108. So that the best possible filling of the engine can be achieved, it is advantageous when the acceleration of the intake valve in all speed ranges, and, therefore, in all partial strokes, is as high as possible. Current VVT systems, today, have a maximum valve-related acceleration of approximately 55–80 mm/rad2. In the partial strokes, the related acceleration, thereby, decreases greatly and, therefore, the filling of the full load. Advantageous

Figure 10.107 Adjustment torques on the eccentric shaft.

for the full load and the throttle losses in the partial load range would be a related valve acceleration at partial strokes, which is higher than at full stroke. High-specific and absolute accelerations can only be realized when the acceleration ramps have been clearly defined in all strokes, that is, high acceleration and high engine speeds can be realized only when the acceleration ramp, and, in particular, the closing ramp, are not assumed by the play or contact stiffness for all the valve strokes. Tappet valve train and roller cam valve trains are now designed with a maximum valve-related acceleration of approximately 85 mm/rad2. Because the effective valve acceleration grows with the square of the engine speed, the partial acceleration can be possibly made eventually higher.

10 Target curve

9

Dimensioning

Valve stroke [mm]

8

Measurement

7 6 5 4 3 2 1 0

0

20 40 60 80 100 120 140 150 180 200 220 240 260 280 300 320 340 360 Opening time [°CS]

Figure 10.108 Opening characteristic of the valve. See color section page 1082.

Internal Combustion Engine Handbook | 465

6606_Book.indb 465

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Thus, an advantage would be at the full load in the lower speed range and adjust the consumption in the partial load range because of the low tendency to improve throttle losses. The target for the VVT should be a maximum acceleration at full stroke of approximately 85 mm/rad2 at an rpm of 8500 L/min. The UniValve system is also usable on engines with a bottom CS (Figure 10.109). From the bottom CS, a rocker arm is driven by a roller tappet and a push rod that transmits its movement over a roller on the upper roller of the intermediate lever. With this upper roller, the intermediate lever proceeds simultaneously in a stationary-located gallery. The working curve is integrated in the intermediate lever, which runs on the roller of a roller cam follower arm. The backdrop curve is defined by a circular path with the center of the roller of the roller cam follower and a radius that is defined by the roller diameter of the roller of the rocker arm. To set a valve stroke, the fulcrum of the intermediate lever is shifted toward the CS with an eccentric shaft. The displacement is approximately 3.5 mm, to adjust the valve stroke from 0 to 10 mm. This provides the advantage that in this type of engines, the engine block can continue to be used unchanged, and only modifications to the cylinder head are needed.

Background

In the systems with CS, the CS is used for pressure buildup for a push rod that transmits the reciprocating motion through the medium of oil to the valve. Therefore, the start of opening of the valves is always constant. Depending on the interruption of the oil pressure buildup in the hydraulic tappet, the closing time of the valves may now be changed. The EIC is mainly used for the load-controlling process. Depending on the type of interruption of the pressure buildup by using a control edge or an electromagnetic valve, there is a differentiation according to the categorization in Chapter 10.4. A CS adjuster is recommended for controlling the residual gas content. An example of a design to change the control times, based on a hydraulic functional principle with use of an electromagnetic discharge control, is the system from Fiat shown in Figure 10.110. Developments toward similar solutions were executed by many companies.

3

2

Clarification:

Intermediate lever Roller cam follower

Valve

Camshaft

Figure 10.109 UniValve system for engines with bottom CS.

10.4.3.4 Hydraulically Actuated Systems In the eighties of the previous decade, there was a whole series of research efforts executed, which dealt with hydraulically VVA. The objective of the development was the freely designable actuation of the valve through the medium of oil and setting the, thereby, associated improvement potential for engines. As previously, therefore, solutions with and without a CS were investigated.

1

1 2 3 4 5 6 7

Intake valve Brake pistons Cams Tappet Stem chamber Solenoid valve Pressure tank

Figure 10.110 Hydraulic VVA from Fiat [10-56].

The intake valve is actuated through the CS and a hydraulic transmission mechanism. Therefore, pressure builds up with the movement of the tappet in the tappet chamber, which, thereby, moves the piston above the valve and, hence, moves the valve. The oil pressure in the tappet chamber can be interrupted by a solenoid valve. This, therefore, limits the valve stroke, and the engine load can be controlled without a throttle valve. Oil can be conveyed to the tappet chamber through a small pressure tank. The solenoid valve must be designed to switch extremely quickly. A problem with this type of valve actuation is the operating behavior at low temperatures and the related strongly differing oil viscosities. A reproducible valve lifting curve is also difficult to obtain. A functioning valve brake must be present for the selective delaying of the valve when closing the valve seat. A current example of this design principle is shown in a test engine of the company INA in Figure 10.111. Tappets, tappet chambers, and brake piston from the schematic diagram of Figure 10.110 are arranged in extension of the valve axis superimposed.

466 | Internal Combustion Engine Handbook

6606_Book.indb 466

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

levels [10-57]. The lower part contains the gas exchange and combustion chamber associated functions (thermodynamics module), and the upper part contains the valve actuators (hydraulic module). The actuator is essentially a dual-sided piston module actuated with high pressure oil, whereby, depending on requirements and desired valve stroke on-off valves, the piston module is controlled. The free type of control, and the associated operation of the valves, is the great advantage of this type of valve actuation. Now, there are few activities in the area of hydraulic VVA. The authors recognize the developments from the companies Fiat, Bosch and/or AVL, Lotus, INA, and several other companies where there are intensive works going on. Whether these systems have a chance of being realized in mass production engines or not is difficult to evaluate, and remains exciting to watch.

Figure 10.111 Hydraulic valve actuation system UNIAIR from the company INA on a test engine.

The companies Bosch and AVL are working on a cam-less solution for the implementation of the fully variable valve stroke. An oil pump is required to provide the hydraulic energy, which will be integrated in the auxiliary drive of the engine and provides the pressure required to actuate the valves. The pump conventionally receives prepressurized supplied oil from the normal engine oil pump from the engine oil circuit. A rail is fed through pipelines, from which the individual actuators are supplied in the cylinder head (Figure 10.112). The high pressure for the actuators is in a range of between 50 and 200 bar. The cylinder head is built in two modular

Pressure and temperature sensors

10.4.3.5 Electro-Mechanical Systems The idea to operate gas exchange valves with electrical energy, and to make the designs free from any constraints of a cam drive for the valve stroke movement, is certainly the dream of every motor engineer. For approaches to accomplish this with a purely electric operation, such as by piezoactuators, there has been no shortage. It often occurs as the utilization of a structural element to reduce the driving energy for the valve actuation of a spring as an energy accumulator or valve closing element. Thereby, the first developments for electromechanical cam-free valve actuators began more than 20 years ago. With this system, there is the greatest potential for varying the valve lifting curve. The mostly identified designs use actuators, which each used gas exchange valves. This enables a setting for each individual valve control time. The biggest advantage of these systems is seen as being, in addition to the freely selectable

EHVS-Regulator

Overpressure valve

Signal for: Rotational speed Accelerator pedal position Temperatures PS-Position Viscosity sensor High-pressure pump

Oil Filter (Supply from engine circuit)

High pressure Engine pressure Return flow Leakages

Figure 10.112 Hydraulic valve actuation from Bosch and/or AVL.

Internal Combustion Engine Handbook | 467

6606_Book.indb 467

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

valve stroke strategy through load and rpm, the possibility of additional cylinder deactivation. An armature between two coils, alternately supplied with power, is connected to the charge cycle valve through the armature guide. In addition, springs are used that actuate the armature and/or the valve. The armature is excited to vibrate depending on the power system in the bottom or top coil. The valve stroke can, thereby, be set from 0 mm to a maximum stroke, and the load of the engine can be controlled by a wide variation of the valve strokes. Figure 10.113 shows the basic construction of this type of control. To open the valves, the opener magnet is excited with current, and the closer magnet is excited to close it. When the coils are not excited with current, the armature and, hence, the valve remain in the middle position between the coils. This position is held by the spring. In case of a system failure or engine stoppage, a corresponding clearance in the piston has to be provided.

4 1

Figure 10.114 Four-cylinder test engine of the company BMW with electromechanical valve actuation.

2 3 4

Clarification: 1 2 3 4

Neutral position

Valve closed

Closing magnet Anchor Opening magnet Valve springs

Valve opened

Connector Control device

Anchor plate

Figure 10.113 Schematic representation of an electromagnetic VVT [10-58].

The actuator is provided with a sensor for detecting the current operating state [10-59], [10-60]. The control of the actuator is performed with a special electronic valve actuation, in which the necessary current curves are formed for the control units. The demand for energy is applied by a 42-V generator. For this purpose, use is, hereby, made of a crankshaft start-generator that is integrated on the clutch side of the motor (Figure 10.114). The requirements for actuators and sensors of the electromechanical valve actuation are essentially: •• the lowest possible power consumption •• high durability compared to conventional valve trains

Electro-magnet

Valve spring

Valve

•• realizing the shortest possible switching time for flexible timing strategies •• compliance with reproducible exact control timing. The electromechanical valve actuation is arranged directly on the motor, and is subject to high mechanical and thermal loads. The primary function of the valve actuation is to supply the coils of the actuators with voltage, so that a specific current waveform will be generated. Furthermore, the thermal stability and a high electromechanical compatibility must be ensured.

Figure 10.115 Single actuator of the electromechanical valve train. See color section page 1082.

468 | Internal Combustion Engine Handbook

6606_Book.indb 468

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.4 Variable Valve Actuation

The power consumption of the actuators can only be realized with increased on-board voltage, otherwise the electrical losses and the volumes required per actuator would be too high. A high generator efficiency of about 75% is partly responsible for achievable fuel efficiency. The typical course of movement during an opening operation starts at a certain time, with the command of the engine control to open the corresponding valve. This command is implemented in the activation of the two electromagnets in the electronics:

for such a type of controls on the intake valve side. The field of conflict between criteria and facts will show to what extent this type of variable valve controllers can replace the hitherto purely mechanically acting control systems on production engines in the future. The infinitely variable mechanical valve control systems will continue to demonstrate their potential in production engines. The, thereby, remaining thermodynamic additional improvement potential for electromechanical valve control systems is limited.

•• The holding voltage to the upper coil is switched off.

10.4.4 Perspectives for the Variable Valve Train

•• The tensioned spring accelerates the armature. •• The valve opens. When the armature begins to get close to the maximum opened position, then the armature will be retained by the current flowing in the lower coil and the valve will be jointly retained in the opened status. At the end of the desired timing, the closing command is triggered and the valve returned to its seat. The gentle placement of the valve into its seat must be carried out for reasons of acoustics and wear with a valve inserter speed of below 0.05 m/s. This touchdown behavior is called soft landing. The current strength must be drastically reduced during the flight phase of the armature and valve to reduce the force and soft-landing speed. Compliance with this process, and the reproducibility of the valve closing time to less than 1° of the CA, makes high demands on the regulation of these complex mechanization systems. The company BMW has, hereby, produced a regulatory approach, based on three pillars [10-59]: •• a target value generator, which assumes the role for the timings for displacement, velocity and acceleration, and, in particular, the end position of the armature •• an observer for considering the estimated values for the three timing processes •• the controller that calculates and regulates the required coil voltage from deviation between the targeted and estimated values. Electromechanical valve control of this type with individual actuators has the highest variability in the design of valve stroke curves. All conceivable processes and load control methods, such as EIC, can be realized. There can be a change to an altered valve stroke function from one cycle to another. Moreover, it is possible to operate only so many valves as needed. In a four valve-engine, the opening of both valves is necessary only at full load and high speeds. In partial load operation, individual valves can be used for the shutdown of individual cylinders or up to the different control of adjacent cylinders. Because of the large variability of the valve lift curves, controlling the residual gas portion through the overlapping area between the exhaust and intake valve lifting curve can be done simply. The series production of the electromechanical valve control systems leads to considerable extra costs, to a larger engine height and higher engine weight. Therefore, further developments are being worked on in which only a provision is made

Since the 1980s, significant developments to improve consumption and torque increase at a low and higher rpm, and to reduce the raw emissions of gasoline engines for gasoline engines in the valve train, have been executed. The first valve contour transfer systems were introduced in 1983 on a motorcycle, and later on in automobile engines to achieve higher performance. This, thereby, involved switching switched between a valve lift curve with short timing and a valve lift of approximately 4 mm, to a valve lift curve with a very long timing and a valve lift of 10 mm, and, thereby, solved the conflict among idle quality, acceleration, and higher performance. Significant consumption improvements could only be first achieved, when it became possible for the introduction of a CS shifter on the intake CS, which represented an effective residual gas control by controlling the overlap area in the partial load of gasoline engines. This enabled the achievement of significant improvements in torque at a low and high rpm through, thereby, possible speed-dependent control of the closing timing of the intake valve at full load. A CS timing on the exhaust CS allows the targeted heating of the catalytic converter system in the warm-up by early opening the exhaust valve. The achievable fuel economy improvements in continued throttled operation are in the range of approximately 4% with a CS adjustment on the intake and exhaust CSs with a displacement angle of 60° CA. A further improvement in fuel consumption of 6%–8% was achieved by the introduction of the throttle-free load control of the gasoline engine with load-dependent control of the height of the valve opening cross section of the closing point and timing of the intake valves. The load-dependent regulation of the opening cross section is carried out in electromechanical and servohydraulic valve trains through a variable opening time, that is, the intake valve is always fully opened and adjusted to the length of the opening time of the load. With fully variable mechanical valve trains, however, the opening cross section is controlled by the height of the valve and in some systems by the amount of valve lift and the length of the opening period. The reduction of the valve lift leads to higher gas velocities in the valve gap, and supersonic velocity is achieved at the valve seat with the smallest valve lifts and, thereby, very finely divided fuel droplets. The, thus, improved mixture preparation reduces fuel consumption and untreated emissions. As an alternative, the throttle-free load control with DI was developed for series production. The DI through the stratified

Internal Combustion Engine Handbook | 469

6606_Book.indb 469

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

lean operation of the gasoline engine also provided significant fuel savings. However, the exhaust gas after treatment is significantly more complex and expensive. When using NOx storage catalytic converters, part of the consumption improvements must be abandoned again, and only lower sulfur fuels can be used. This limits the marketing possibilities. The development engineers are faced on the one hand with the situation of selecting from multiple systems to meet the CO2 commitments in 2008 and, on the other hand, they need to decide on a technology that has the largest possible future potential in the customer-related motor functions. Decision-relevant will be the functional advantages, the total cost including the cost of exhaust systems, sensors, and control units, as well as the investment costs, package requirements, and weight accumulations. Today electromechanical valve trains cannot be excelled will their variability; however, the engine speed is limited to approximately 6000 rpm, the power consumption at low speeds is relatively high, and the increase in the engine size and the weight gain must not be neglected. The shutdown of individual valves, tacts, or cylinders contrast to the highest potential for the future. In the mechanically infinite VVTs, a variety of solutions are provided (Figure 10.116), which differ in important functions and make it difficult for the engine developers to select a system with a high future potential. Figure 10.116 lists the patent applications for mechanical VVAs, which implement a transmission member between the cam and the valve. The systems of this group are described with valve contoured or shut-off systems, as well as infinitely variable systems. In particular, the introduction of the mass production of systems of Valvetronic has led to a sharp rise in the activities in the field of fully variable systems.

Speed [m/s] 978,71 87,74 76,77 65,80 54,84 43,67 32,90 21,94 10,97 0,00

Figure 10.117 Simulation of the turbulence with different valve lifts of 1 and 10 mm. See color section page 1083.

36 / 3 / 24

1990

58 / 5 / 25

1991

118 / 5 / 28

1992

97 / 4 / 4

1993

94 / 0 / 3

1994

Year of registration

The new developments have made it possible to switch off individual valves or cylinder groups, or adjust the valve lift and valve opening time differently. The phasing of the valve lift generates a cylinder flow with a distinct turbulence (Figure 10.117). The flow velocities in the turbulence can be

Valve contouring switch /

82 / 3 / 1

1995

Fully variable systems /

91 / 1 / 3

1996

Other systems

64 / 1 / 5

1997

42 / 26 / 20

1998

45 / 24 / 9

1999

27 / 49 / 7

2000

63 / 91 / 18

2001 41 / 62 / 26

2002 2003

30 / 29 / 10 0

20

40

60

80

100

120

140

160

180

Patent registrations for mechanically variable valve controlling

Figure 10.116 Patent applications of systems with valve contour infinite variability.

470 | Internal Combustion Engine Handbook

6606_Book.indb 470

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

adjusted by the valve and by the difference in the valve lift of the two intake valves to the engine rpm and load. This enables additional consumption benefits in association with the masking measures [10-52]. The consumption benefits may not be reduced again by energy losses while setting and holding the valve lift. Even if all the consumption potential of VVTs are used, one will still not achieve the consumption potential of DI with spray-guided combustion process in gasoline engines from today’s perspective. Is therefore the fully VVT only an intermediate step on the way to DI? This question will probably be answered by the development costs. If the fully mechanical valve train can be anticipated as a gain because of the torque advantages, by emissions improvements in starting and in the warm-up, improved throttle response and increased idling quality and fuel consumption improvements in λ = 1 – Operation, then the future will consist of a combination of DI and fully VVT. A variable turbulence intensity by phasing of the valve, and a residual gas control by a second variable opening of the intake valve combined with open exhaust valves, leads to increased discussion about the use of infinitely VVTs in the diesel engine.

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve 10.5.1 Introduction

In many markets, the combustion engine will continue to represent the dominant drive concept for mobile applications in the coming years as it has to date. It, therefore, applies that new

engines will have to fulfill stricter emissions requirements and, because of rising oil prices, comply with the increasing demands for low fuel consumption. In addition, customer requests continue to rise steadily with regard to engine performance and vehicle dynamics. To justify these sometimes contradictory demands, it is essential to pursue new technologies in the development of charge cycles and combustion processes. An interesting and promising development for fulfilling the aforementioned, and sometimes contradictory demands, is the application of an air stroke valve (ASV) for load controlling. This will enable further improvement of performance and fuel consumption, in particular, in gasoline engines with DI. For diesel engines using ASVs, there is, in addition to performance improvements, also the possibility of reducing emissions through the Miller process, analog to the Miller process with early closing of the intake valve.

10.5.2 Technology Description

The ASV has been a well-known process for many years [10-62], which allows a largely free and flexible influence of the charge cycle of internal combustion engines. In this process, every suction line of the intake manifold is equipped with a fast-switching valve between collector and intake valve, as shown schematically in Figure 10.118. This can largely be achieved with the characteristics from electromechanical valve train systems with regard to the free selection of the regarding the free choice of drive time points and, thereby, the thermodynamic potential, without the need for completely redesigning the cylinder head mechanics and combustion chamber geometry. In recent times, many publications and research were publicized with regard to ASVs in which the potential of this technology was assessed from various different angles, for example [10-63], [10-64], [10-65], [10-66], [10-67], [10-68], [10-69].

Cylinder head

Intake manifold Intake valve

Exhaust valve

Spark plug

Collector

Air cycle temperature

Combustion chamber

Injection nozzle

Piston

Figure 10.118 Arrangement of the ASV on the engine.

Internal Combustion Engine Handbook | 471

6606_Book.indb 471

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

The ASV opens among others, the possibilities for influencing the gas exchange in the combustion engine with DI: •• Throttle-free load control in gasoline engines: Reducing the intake air mass without throttling losses by metering the air impulse valve, analog to the reduction of the valve opening period with mechanically VVTs. •• Suction-side cylinder deactivation: Flexible suction-side cylinder deactivation to reduce the gas exchange work through load point shift; alternating shutdown prevents the cooling of individual cylinders. •• Pulse turbocharging: Increasing the engine power in the lower rpm range by inducing pressure waves in the intake manifold and their inclusion in the combustion chamber. •• Temperature control: Heating and cooling of the aspirated air through compression and expansion in the combustion chamber, and in the interstitial volume between the air valve and intake valve stroke by the piston movement. The Miller process should be emphasized here for the reduction of the exhaust emissions of diesel engines.

Figure 10.119 ASV from MAHLE. Principle: magnetic-controlled springmass oscillator with a swivel angle of 45°. Switching time: 2.7 ms. Leakage: <1 m3/h at ∆p = 600 mbar. Channel cross section: 1300 mm2. Self-adaptive sensor-free current control. On-board voltage of 12 V.

10.5.3 Construction Principle and Boundary Conditions

Some of the possible process methods using an ASV are explained in more detail below. The underlying theory is described and demonstrated by test results.

Basically, a plurality of valve concepts (rotary valve, poppet valve, and rotary-disk valve) is conceivable for the ASV, because it is spatially separated from the combustion processes in the combustion chamber, as an add-on solution for existing basic motors (refer to Figure 10.118). The functional assurance and the quality assurance to the ASV to be adapted can, thus, be the responsibility of your supplier. The aim has to be fulfilled the following functional requirements: •• sufficiently fast switching time (typically <3 ms), within which the intake cross section will be opened and/or closed •• tightness requirement (low leakage and closed cross section) •• necessary flow cross section to maintain the rated motor power •• adaptation within the installation space available within the intake system and/or in the engine compartment •• flexible and precise switching drive. Furthermore, these characteristics and the functions involved must have a guaranteed service life of approximately 200–300 million switching operations. The temperature range is set as a basis of Tambient = −40 to +130°C. The electrical supply is to be realized from the 12-V power supply and, thereby, with the lowest possible power consumption. To switch safely and quietly at all the flow conditions in the intake cross section, a control with soft-touch function is required. These high demands for fast and frequent switching of valves with low-power dissipation make a necessity for an extensive development of new drives for this application [10-65], [10-66], [10-70]. The data of the ASV executed by MAHLE as a rotary damper are illustrated in Figure 10.119.

10.5.4 Thermodynamic Potential

10.5.4.1 Dethrottling for Fuel Savings for Gasoline Engines 10.5.4.1.1 Functionality of dethrottling During dethrottling in gasoline engines, the ASV provides the controlling system for metering the required air mass. The function of the throttle valve is in this case taken over by the ASV, with which the metering of fresh air is now executed. The ASV enables, in connection with a conventional valve train control strategy, an analog intake closing (Figure 10.120), and late intake opening can be realization. A combination of early and late introduction of air is possible through a double tact of the ASV in the engine cycle. The minimum opening duration of the ASV (that is, a very short response time) only plays a subordinate role for dethrottling, because it can already have been given an appropriate positioning of the ASV-opening window for the range, in which the ASV and intake valve are opened simultaneously. In addition, fresh air will only be aspirated during this time. However, a short response time (<3 ms) allows greater latitude when choosing the location of the ASV-opening window, relative to the intake phase. Significant points in dethrottling are the tightness of the valve and the dead space between the intake valve and ASV. Especially at low loads, and consequently small air masses, these factors play an important role in the control technology. By leakage air and air in the void volume, a minimum amount of air is already dictated that arrives in any case into the combustion chamber. If an even smaller amount of air is required, then it can no longer be regulated by the ASV. This then must be throttled back so as to take in less air in the dead space and, thereby, into the combustion chamber.

472 | Internal Combustion Engine Handbook

6606_Book.indb 472

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

CS

CS

open

CS

CS

CS

CS

CS

close

CS

CS

Valve opening curves

Figure 10.120 Throttle free engine operation with ASV analogous to EIC.

Figure 10.121 shows the test results for dethrottling with the ASV. When the ASV is opened as long as the intake valve, then the load point is conventionally retained by closing the throttle valve. If the opening duration of the ASV is shortened, by a constantly retained intake valve lifting, then the throttle valve can be gradually opened, and the pressure in the intake manifold increases to almost that of the surrounding pressure. Figure 10.122 shows the pressure process for two differential opening times of the ASV. With a test with the hatched illustration for the pressure process, the opening time is selected to the largest, so that the ASV has no influence on the charge cycle. This corresponds to the conventional partial-load controlling with a throttle valve. A test with a solid line has reduced

opening times for the ASV. The charge cycle work (the area contained by the pressure process in Figure 10.122) could, therefore, be significantly reduced. As clearly shown in Figure 10.123, the indexed mean pressure of the charge cycle loop reduces, with constant increasing of the closing time point when using an ASV, toward the direction of “early.” By reducing the opening time of the ASV, there can, therefore, be a reduction in the charge cycle work, and thus reductions in the fuel consumption. For the four-cylinder gasoline engine of Figure 10.121 of this reduction is in the order of 6% compared with the base engine in the throttle operation.

Closed

Savings in consumption by early-closing of the inlet with the air cycle valve

Fuel consumption [-]

100%

-0.1 -0.2

98%

Consumption

-6%

96%

Intake manifold pressure + 500 mbar

-0.3 -0.4 -0.5

94% 300

GOT

420 480 540 600 Closing moment of the air cycle valve [°CS]

-0.6 660

Vacuum in intake manifold [bar]

opened

Figure 10.121 Consumption reduction by dethrottling with ASV on a four-cylinder 2-L gasoline engine with DI (homogeneous) with n = 2000 rpm and pme = 2 bar.

Internal Combustion Engine Handbook | 473

6606_Book.indb 473

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

1.6 Throttling with ACV Influencing the charging change by means of a throttle valve

1.2 LTV ö 1.0

EÖ 0.8

AS

LTV s

ES

0.6 ES

0.4 0.2

AS 0

50

100

150

200

250 300 350 Volumes [cm3]

400

450

500

550

Figure 10.122 Partial load controlling with a throttle valve and/or ASV at n = 2000 rpm and pme = 2 bar.

-0.20 -0.25 -0.30 -0.35 -0.40

De-throttled partial load mode

Throttle mode

-0.45 -0.50 -0.55 350

400

450 500 600 550 Closing moment of the air cycle valve [°kW]

650

0.0

Basis With ACV

-0.2

bi [g/kWh]

-0.4 38 36 34 32 2

3

4 5 pmi [bar]

6

7 1

2

3

4 5 pmi [bar]

10.5.4.1.2 Results on a full engine (Gasoline V-8) Engines with large displacement benefit from fuel savings by dethrottling even more, because the motor is much more frequently operated in real driving conditions in partial load. To show the full potential, measurements were carried out with the air impulse valve for unthrottled load control on

6

7

30

Figure 10.123 Reducing the charge cycle work by dethrottling with ASV on a four-cylinder 2-L gasoline engine with DI (homogeneous) with n = 2000 rpm and pme = 2 bar.

pmi,LW [bar]

Indexed mean pressure LW [bar]

EÖ

ηHD [%]

Cylinder pressure [bar]

1.4

Figure 10.124 Load average measurements of fuel consumption with dethrottling through the air impulse valve to a V-8 gasoline engine with DI with n = 2000 min–1.

an eight-cylinder gasoline engine with DI, whose results are shown in Figure 10.124 and Figure 10.125. Figure 10.124 shows a load average for the speed of 2000 min–1. Here, the consumption of the dethrottled air impulse stroke valve test engine is applied compared with the base engine. For operation with the air cycling valve, additional measures to increase the cargo movement were used in the

474 | Internal Combustion Engine Handbook

6606_Book.indb 474

1/19/16 8:41 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

10.5.4.2 Intake-Side Cylinder Cutout The idling at n = 550 min–1 on eight-cylinder gasoline engine considered (see Figure 10.125) can be reached by the air impulse valve to reduce the fuel consumption by 23% compared with the base engine. Four air cycling valves were there permanently closed for realizing a suction cylinder deactivation, while the four operating cylinders were dethrottled with the air cycling valves. The reduction in the consumption resulting from out here, on the one hand, the reduction in the charge exchange work and, on the other hand, from the strong increase in the high-pressure efficiency of the working cylinder through the load point shift to higher load. 10.5.4.3 Pulse Charging at Low Speeds (Low-End Torque)

Cylinder pressure [bar]

10.5.4.3.1 Functionality of pulse turbocharging Both in supercharged and naturally aspirated engines, the air impulse valve can be used for pulse charging. An improved torque response is the result, especially in the lower speed range. Raising the torque at low engine speeds with no increase in the displacement is an important target for new engine concepts. It is necessary to guarantee a good acceleration behavior of vehicles with continuously increasing total weight,

Bh [kg/h]

0.9 0.8 0.7 0.6 -0.40 -0.30 -0.20 -0.10 -0.00 30 25 20 15

- 23%

ηi,HD [%]

- 38%

+ 62%

Basis

pmi,LW [bar]

combustion chamber. The measurements show that the amount of consumption reductions depends on the load between 3% and 9%. Even with medium loads there are still savings achievable in the range of 5% 6%. In the investigated operating points, the gas exchange work has been reduced by approximately 50%. For high-pressure efficiency, the air impulse valve operation is also low. It could be improved by 0.5%–1.5% points. With this information, a fuel saving on the NEDC cycle of more than 5% can be expected. It seems realistic that additional potential can be tapped by optimizing the air impulse valve operation litigation and by a favorable combustion chamber geometry.

With ACV

Figure 10.125 Consumption savings due dethrottling the air impulse valve in combination with the suction-side cylinder deactivation (4/8) of a V-8 gasoline engine with DI at n = 550 min–1.

as well as their desired elasticity. High torque at low engine speeds continues to be the basic prerequisite for future downsizing concepts. Fuel-efficient vehicles with sufficient performance can then enable the implementation of new consumption strategies. The selective control of flow and pressure waves is generated by the vibration processes in the intake using the air cycling valve that realize the pulse charging. Thus, at any engine speed can be an ideal resonance charging can be realized, whereby an increase in the high-pressure working is achieved at full load (Figure 10.126). In addition to these actual charging of the increased pressure is also present in the hall between air impulse valve and intake valve. It can be used in the next charge cycle to flush the residual gas through the open exhaust valve from the cylinder.

ACV active, pmi = 9.89 bar ACV open, pmi = 8.13 bar

Stroke volumes normalized

Figure 10.126 Increasing highpressure working by pulse charging by LTV to a V-8 gasoline engine with n = 1000 min–1 and full load.

Internal Combustion Engine Handbook | 475

6606_Book.indb 475

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Crankshaft angle [°CS]

Chapter 10 Charge Cycle

ACV active, 2x Strokes

LTV ö

LTV s

LTV s LTV ö

2. Takten 2nd. Stroke

Cylinder pressure [bar]

1. Takten 1st. Stroke

ACV open

LWOT Kurbelwinkel [°KW]

Figure 10.127 shows the result of the generated intake pressure waves. When operating with LTV open, a drop in pressure is only briefly seen at the beginning of the intake process. There are virtually no usable pulsations available, and an effective suction pipe is not possible charging. If the LTV is active, a significantly higher pressure level over an open LTV can be achieved. At low speeds, for example, twice clocking the LTV is useful. In the present case 1000 min–1 with 2× cycles achieved a higher increase in the torque. This is because of the increased volumetric efficiency and a resulting larger mass of air in the cylinder. The pulse charging with the air impulse valve can be used in gasoline and diesel engines, and, in particular, may contribute to significant improvement in the dynamics of vehicles with exhaust gas turbocharging, as it is described in detail in [10-67]. 10.5.4.3.2 Results on a full engine (gasoline V-8) Measurement results of an eight-cylinder gasoline engine, which are shown in Figure 10.128, demonstrate the potential of the pulse charging with air impulse valve for increasing the torque in the lower speed range. In this way, the torque relative to the base without ASV operation could be increased significantly. When n = 1000 rpm, there is even an increase of 23% achieved. The mentioned values were obtained with the same fixed short intake manifold length. This torque can be built up by leaps and bounds from one cycle to the next, because no filling operation of the suction system delays the torque buildup.

Figure 10.127 Pressure curve in the cylinder during the intake process at n = 1000 rpm and full load.

One problem with the pulsed turbocharging is that the increased volumetric efficiency often does not result in the same degree to an increase in the torque. The pulse turbocharging creates an increase in the temperature of the fresh air in the combustion chamber, because of its procedure (analogous to a compression without intercooling), which increases the tendency to knock. As a result, the motor with pulse turbocharging at very low speeds cannot be operated as optimal filling. Advantageous is shown to be the double opening of the ASV in the intake phase, whereby, respectively, only minimum pressure differences have to be created through the ASV for achieving the turbocharging effect with every gear change. 10.5.4.3.3 Dynamics with load changes As described, the ASV can increase the torque by means of turbocharging in lower rpm ranges at full load and, thereby, eliminate weakness in the low-end torque. Another advantage is the improved dynamics during a load step. Based on the throttle-free test with ASV as partial-load mode, the load-step demand reacts significantly quicker as with the conventional operation of the throttle. The atmospheric pressure is already created in the intake manifold system, and a relatively time-consuming filling of the intake system is not required here. Figure 10.129 shows the measured conversion of a load step from partial load (pme = 2 bar) at full load at n = 2000 rpm and open ASV valve. Using the ASV, it

476 | Internal Combustion Engine Handbook

6606_Book.indb 476

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

Torque [Nm]

Torque with ACV Torque basis

Figure 10.128 Torque increase by pulse turbocharging with the ASV on an eight-cylinder gasoline engine with DI and a fixed short intake manifold.

rpm [min–1]

is possible to achieve significantly quicker maximum torques of 90%, which is not the case with use of the throttle valve. Especially in internal combustion engines, exhaust gas turbocharging in combination with air-stroke valves is shown to be advantageous. The faster response to load change is, because of the increased mass-flowthrough volume, also an improvement for the demand behavior of the turbocharger. The adverse effect of a turbo-lag can, thus, be reduced. 10.5.4.4 Miller Process The Miller process is a process with which the very early closing of the intake valve analogously leads to a targeted reduction of the temperature of the aspirated air in the combustion chamber of turbocharged engines. The air required for combustion mass is included with the ASV in the combustion chamber, while the piston is still located in the intake phase.

The air precooled by a heat exchanger is cooled further using expansion through the piston movement, and then trapped by the intake valve in the combustion chamber. This temperature reduction can be used to reduce the knock tendency of the supercharged engine in gasoline engines. A higher compression ratio in the charger, with the resulting higher discharge temperature, can, therefore, be realized, which, in turn, can be used to further increase the torque and power of the engine without additional thermal combustion chamber loading. For diesel engines, the target size of the Miller process is the reduction of NOx emissions. While retaining the existing valve train, an EIC can be realized through the ASV. The Miller process, thus, reduces the overall pressure and/or temperature level of the working process. The lower combustion temperatures result in fewer oxides of nitrogen. The degree

t1 s 2,6 t2 s 1,9

Torque

90% of des the max. max. Momentes torque

Lastsprung Load jump with mit LTV ACV Lastsprung Load jump with mit DK DK

time

Figure 10.129 Comparison of loadstep reaction with ASV and with conventional throttle valve.

Internal Combustion Engine Handbook | 477

6606_Book.indb 477

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

of reduction of NOxs can be adjusted using the ASV in a simple manner, when the closing time is set to the operating point-specific, as well as for each cylinder and cycle selectively. The Miller process may be differently implemented in the diesel engine with exhaust gas turbocharging, depending on the operating point:

De -th r

g tlin ot

•• with air mass compensation to compensate for the reduced intake opening period by increasing the boost pressure •• without compensation of the reduced air mass (at low load points)

Phase regulator

GDI homogeneous

10.5.5 Evaluation Summary

In conjunction with a conventional valve train, the air cycling valve can affect the charge cycle of a gasoline engine manifold. This opens up the possibility of combining the ASV with different current and future development plans. Figure 10.131 shows two possible technology combinations. A goal-oriented approach is the use of the ASV for dethrottling primarily in large-volume gasoline engines.

na Dy

rq To

Figure 10.131 Engine concepts that can be effectively supported by the air impulse valve.

The gas exchange work can be greatly reduced, resulting in significant fuel savings at partial load and, therefore, in the customary test cycles. In conjunction with the homogeneous operation, a worldwide usable DI complete system can be realized for displaying more efficient engines. The system is economically very interesting compared with competing technologies for dethrottling. In particular, only the proven and cost-effective three-way catalytic converter technology is required for exhaust gas aftertreatment. The possibilities for cold start assistance and pulse turbocharging during acceleration can then be used with.

2.5

1.2

2.0

0.8

AGR

rel. spec. PM [-]

rel. spec. NOx [-]

1.0

0.6

0% AGR 40% AGR

0.4

0% AGR 40% AGR

1.5

1.0

0.5

0.2 0.0 80

ue

+

Tests on a 0.5-L single-cylinder structure show that significant improvements, with respect to emissions, can be achieved especially with the Miller process without air mass compensation, and in combination with an external exhaust-gas recirculation. Contrary to the typical soot-NOx trade-off this enables—at only slightly higher consumption—a significant reduction in the NOx emissions to be realized while reducing soot emissions. Figure 10.130 shows the normalized results of this, hereby, executed measurements performed at n = 2000 rpm and pmi = 4 bar (see [10-71], [10-72], [10-73], [10-74]). The results show that the Miller process without air mass compensation in connection with an external EGR can primarily reduce the NOx reduction through the exhaust-gas recirculation, while the Miller process can reduce the soot emissions [10-71]. Because of this reduced mass flow at the same injected fuel quantity, there occurs additional higher exhaust gas temperature with regard to the exhaust aftertreatment. [10-71], [10-72], [10-74].

m ic

•• in combination with an additional external exhaust-gas recirculation.

100

120 140 160 Close ACV [°KW n. LW-OT]

180

200

0.0 80

100

120 140 160 Close ACV [°KW n. LW-OT]

180

200

Figure 10.130 NOx and soot emissions at different ASV closing times and EGR rates.

478 | Internal Combustion Engine Handbook

6606_Book.indb 478

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

Another approach is shown in Figure 10.131 as the effective support of downsizing strategies for gasoline engines. The dethrottling, thereby, plays a lesser role, because the motor is generally driven by the lower stroke volume in higher load points. To ensure the rated output, a relatively largesized turbocharger is usefully employed, but which has the known disadvantages of torque weakness in the lower speed range and during transient engine operation. In this case, the required low-end torque can be prepared by the ASV. Furthermore, it is possible to create a much quicker torque buildup in comparison to throttle guided systems by the immediate adjustment (already in the next cycle) of the ASV switching times. In experiments, it was also possible to use the application of the ASV and prove it for active temperature control. It is suitable for the implementation of the Miller process wherein significant improvements in NOx and soot emissions can be achieved in combination with an external exhaust-gas recirculation. The high development costs, which must be paid for the mechanics and control of the air cycling valve, as well as its application for the engine represent the only significant drawback. In contrast, for all the processes and applications with the air impulse valve, the high variability can be used because of the freely controllable switching times in the engine cycle (Figure 10.132). In particular, here is the possibility of a cycle and cylinder-selective circuit mentioned. Tolerancerelated unequal distribution of the load on the individual cylinders can, thus, be compensated for by the control unit. Likewise, a change of operating mode because of changed boundary conditions from one cycle to the next can be realized, for example, by switching from an early closing of the ASV timing on its late opening.

Advantages: Complete variability (cycle-selective and cylinder-selective switching) Regulating the air-side loads Switching over from e.g. FES to SEÖ from one working example to the next

10.5.6 Pulse Turbocharging With Controllable Aspiration Valves 10.5.6.1 Introduction In the pulse charging, a higher compression ratio of the combustion required for fresh gas mixture is achieved by the air mass flowing upstream of the intake valves, by means of a fast-switching solenoid valves is selectively controlled. These freely controllable impulse charger valves are arranged in the intake manifold pipe between the intake valve and the air collector (Figure 10.133). At the beginning of the intake stroke when the piston moves down, the impulse charger-valve is kept closed. This creates a vacuum in the combustion chamber. Only when the impulse turbocharger valve is opened before the lower turning point of the piston, the air can be suddenly released and flows through the pressure gradient generated in the cylinder. At the same time, there is a creation of a pulsed pressure wave that reaches the speed of sound in the combustion chamber after back reflection on the collector. The impulse turbocharger valve closes before the pressure wave escapes again, resulting in the increased cylinder filling. This charging effect is available without delay within a power stroke. A valve is required for each cylinder. The control of the valves is assumed by the electronics, the phase-accurate signals of the crankshaft sensor, and the CS position sensor. A standardized interface to the engine controller transmits inter alia, the information when the pulse turbocharging is to be activated. The pulse turbocharging creates an increase in the volumetric efficiency λ a and the indexed mean pressure pmi in the lower rpm range by up to 40%. Without pulse turbocharging,

Disadvantages: Increased development resources in the following working areas: Adjustment of the piston geometry and injection nozzle characteristics to air-cycle valve operation basis Charge movement measures

No effects on performance outputs

Acoustics

Balancing functional tolerances

Technology for manufacturing shut-off mechanisms and seals

Emergency operation without immobilization (compare to EMVT) Known process for component production in valve drive range, but still with complete variability Function diagnostic in operation

Preventing double impact, Anchor / Valve seal Electrics (sensor-free controlling with "Soft Touch" function) Function developments for engine control devices

Good effort/utilization ratio Versatility (de-throttling, impulse charging, temperature controlling, Miller process,.) Figure 10.132 Functionality and economical aspects.

Internal Combustion Engine Handbook | 479

6606_Book.indb 479

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

Throttle Valve

Collector

Injector

Impulse charger Valve

Figure 10.133 Position of the collector, injector pulse turbo loader valve in the intake module. See color section page 1083.

especially at low speeds, the cylinder charge in the full load is very unsatisfactory. In addition, exhaust-driven turbochargers deliver only improved filling after a time delay time and a minimum rpm. The pulse turbocharging reacts spontaneously in these situations and provides the desired torque boost in the next work cycle. The development trend toward downsizing bears fruit when, without further necessary engine measures, the resulting loss of torque in the lower rpm range can be conditionally compensated by the smaller displacement, with pulse charging at full load. It could, with the herein described type and configuration of the pulse turbocharging, already indicate the increases in the measurements for the volumetric efficiency and the indicated mean effective pressure of up to 40% at low speeds [10-75]. In addition, measurable improvements in volumetric efficiency and torque could realized by pulse turbocharging, as shown with differently designed valves [10-76]. The advantage of pressure wave turbocharging to the desired increase in the volumetric efficiency at the end of the intake process, with the smallest possible intake work, have been published

4 Impulse artery valve

Impulse artery control device

CAN

Battery 12 V

Crankshaft sensor (rpm/phase)

[10-78]. It also addresses the problem of knock limit and high cylinder pressure. With the help of fast switching, and very dense impulse turbocharger valves, a higher gas pressure is saved by closing the mechanical intake valves. This side effect in the intake path leads to an improved residual gas scavenging during the valve overlap period, and thus, a reduction in the temperature of the combustion chamber is realized, which has an advantageous effect on the knocking behavior of the engine. The improved residual gas scavenging is made possible, in that the overpressure created during the pulse drive charge in the short suction path between the intake valve and the impulse charger valve is cached, and, during the overlap period of intake and exhaust valve, is used to flush out the remaining residual gas in the combustion chamber. Residual gas shares of up to 2.5% can be realized. Another positive side effect is caused by the higher gas inflow into the cylinder, which affects the mixture formation positively. The intense flow movement especially creates a positive effect in the cold start phase. The resulting stable

Valve triggering – Diagnosis – Status Information

Engine control device Impulse artery addressing – Phase – Continuously opened position

Figure 10.134 Pulse turbocharger controlling, electrical system integration and interfaces.

480 | Internal Combustion Engine Handbook

6606_Book.indb 480

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

combustion can be used to noticeably reduce the hydrocarbon emissions during this cold-running phase when the catalytic converter is not yet effective. The increase in the engine torque achieved by the improved cylinder charging can be used by longer transmission ratio to directly reduce the consumption. The potential for reducing consumption by pulse turbocharging and cylinder deactivation is estimated in gasoline engines at 7%–10% [10-79]. A combination of a simple exhaust-driven turbocharger and pulse turbocharging, with the advantages of spontaneity in positive load change changes down to the idle speed, provides optimal approaches [10-84]. The design of the exhaust-driven turbocharger is, thereby, focused on the upper speed range, the pulse turbocharger on the lower speed range. The use of the pulsed turbocharging is also suitable for DI gasoline engines with channel and DI, and for diesel engines. 10.5.6.2 Requirements for Components for Series Production The use of systems with electrical impulse turbocharging for reciprocating engines must not lead, of course, to the restrictions in the system life and the availability in series production. Today’s demands on the service life of the powertrain of 5000 h mean more than 108 on–off switching operations for the pulse turbocharger valves. It was, thereby, assumed that the pulse turbocharging would be used by an on–off switching operation per engine cycle for full-load improvement at speeds up to 4000 rpm, for residual gas scavenging as well as for cyclical cylinder deactivation during partial load operation. In the other operating modes, the valves are kept in the open position electrically, thereby affecting the effective flow resistance of the suction only minimally. The entire intake module is to be construed as a possible application for ambient temperatures from −40°C to over 125°C. In earlier prototypes of purely electric valve train actuators, the noise behavior caused by the hard impacting of the valves on the valve seat was viewed as very comfort reducing. These findings were considered in the mechatronic integration of the impulse turbocharger valves through the targeted absorption of structure-borne noise and, in particular, by special electrical control algorithms. Compatibility with today’s system architecture for the vehicle is a prerequisite for the application of pulse turbocharging. Design studies have shown that the information necessary for an efficient pulse turbocharging intake path can be represented by a corresponding intake module in the available space. 10.5.6.2.1 Valve concept Fast-switching electrically operated linear valves are used, whose highly dynamic characteristics are achieved by means of two integrated springs and a customized electronic control as a spring-mass resonator. The patent text [10-80] describes the embodiments of such valves for use in the pulse turbocharging. Two coils come into utilization in the stationary part of the valve. One of the coils is energized to hold the movable part of the valve in the closed position as energized, the other in

the open position of the valve. The movable part is formed as a flow body. The, thus, formed poppet is subsequently part of the magnetic circuit of the valve. By dispensing with a separate movable magnetic circuit of the valve, the translational accelerated mass and the mass to be delayed on closing can be kept small. The, thereby, initially possible moderate spring stiffness, which is necessary to achieve the resonance time constant T/2 (which is proportional to the square root of the quotient of mass and spring rigidity) of less than 3 ms (T100), permits an acceptable amount of energy expenditure while retaining the valve. The advantages of this valve shape are, in addition to the transient flight phase of the valve as always flow-economic properties, the excellent tightness in the closed state. This is also a contributing factor for the implementation of residual gas scavenging. The tightness of the linear valve is a significant advantage over flap systems for pulse charging, as described in [10-76], [10-83]. Crucial for an adequate pulse turbocharging effect are the switching times for opening the valve fewer than 3 ms, with consecutive low impact-time velocities. The intelligent control algorithms for the so-called soft landing, hereby, come into utilization. The soft landing also contributes significantly to fulfilling the service life requirements. In addition, noise emissions during operation can be significantly reduced. Many years of experience in the system development of electromagnetic valve control systems [10-79] has allowed the realization of switching times of fewer than 3 ms at low impact speeds far below 1 m/s. The patent text [10-81] describes control algorithms for soft landing. 10.5.6.2.1.1 Implementing the valve Figure 10.135 shows a valve as it would be used in the integrated intake module. The electromagnetic design of the valve was executed with the help of modern FEM (finite element method) calculation programs. The forces occurring during operation, and its compensation, will, thereby, be analyzed by the electromagnets of the valve. Saturation effects in the holding phase (valve open or closed) could be avoided in view of the moderate power consumption.

Figure 10.135 Pulse turbocharger valve.

Figure 10.136 represents the average power consumption over an engine cycle of a pulse turbocharger valve, in dependence

Internal Combustion Engine Handbook | 481

6606_Book.indb 481

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

of engine speed when constantly driven with active pulse turbocharger operation.

Energyconsumption / W

Leistungsaufnahme vs. Motordrehzahl Energy Consumption per EIC Valve

During Impulse Charging Mode

30 25 20 15 10 5 0

plate is determined through the measuring of the scavenging flow and a model-based computation of the valve speed. The submitted control strategy for the existing system is described in more detail in [10-82]. A timed-division of the controlling into four phases, with applied algorithms for determining the voltage required for shutting of the scavenging, realizes a stable, rigid energy-optimized solution. The phases for the controlling are comprised of •• a triggering phase in which the valve moves away from a de-energized, switched scavenging flow •• a proximity phase in which the collected scavenging flow is charged

0

1000

1500

2000 2500

3000

3500 4000

Engine Speed / rpm Figure 10.136 Averaged power consumption for pulse turbocharger valve in full load mode.

Therefore, the pulse turbocharger subsystem, consisting of four valves and a control unit, requires less than 130 W including the dissipation of the electronic control system for a four-cylinder engine during the pulse turbocharger operation. The valve assembly also allows air leakage rates of less than 0.3 kg/h to be achieved at a differential pressure of 600 hPa. This is not reliably reachable even with high-precision executed flaps. The fluid-dynamic optimization of the impulse turbocharger valve in the open state indicated an average pressure drop of only 10 hPa at the rated output point for the maximum air flow rate of a reference engine. Therefore, the performance loss of the combustion engine is very low at higher rpm because of the installation of a pulse turbocharger. 10.5.6.2.2 Control system electronics The pulse turbocharger control device was concepted with regard to low switching losses and good thermal heat removal. The used technologies comply with those for the standard engine-control devices for installation in close proximity to an engine. 10.5.6.2.2.1 Hardware The performance-electrical control device is equipped with Power MOSFETs with very low R DS, on. The losses in pulse turbocharging mode can be limited to 30 W at a maximum engine rpm. Furthermore, the Ohm resistance in the control device can be kept to a minimum by using a careful design for the conducting paths and contact positions to the connector strips. The pulse turbocharging subsystem comprises the control device and the valves compatible with a network supply voltage of 14 V. This creates an essential simplification compared with concepts with 42-V architecture, which have been described in connection with electromagnetic valve actuation to date [10-79], [10-83]. 10.5.6.2.2.2 Triggering The electrical controlling of the pulse turbocharger valve is executed without position sensors. The position of the valve

•• a landing phase in which the proximity speed can be calculated using a simple relationship from known dimensions such as the measured flow, the timely diversion of the flow, and a machine constant •• a retaining phase in which a retriggering of the valve is prevented, and in which the valve can be retained with the least possible energy expenditure. The so-called soft landing can be realized on the basis of this control structure calculation. This, therefore, enables a theoretically achievable valve landing speed of down to 0.1 m/s. The patent text [10-81] describes the regulating structure. Using the soft landing can, on the one hand, reduce the noise in the operation of the pulse turbocharger valve in such a way that there are no longer any noticeable metallic sounding noise buildups. On the other hand, lower impacting speeds lead to light controllable impact tensions on the sealing surface of the pulse turbocharging valve. This enables to permanent rigidity to be ensured. 10.5.6.3 Electrical System Integration Figure 10.134 illustrates the system integration method, in this case, for a four-cylinder engine. The interface of the pulse turbocharger control unit to the existing system architecture of the engine management system is executed through a CAN (controller area metwork)Bus to the engine control device, through the existing control lines to the motor sensors (crankshaft signal and, where applicable, the CS signal), as well as the power supply. The engine control unit transfers the information through the opening and closing time points for the pulse turbocharger valve, which have to be set, to the pulse turbocharger control unit. This, therefore, ensures that, in addition, the complete engine torque structure exclusively remains in the engine control system, and that these only use the existing intake manifold model for computing the pulse turbocharger valve control times from the torque demands. The pulse turbocharger control device works subsequently in a master–slave architecture as a subordinated control member, which, thereby, ensures the correct implementation of the predetermined control time point. Furthermore, diagnostic information about the status of the pulse turbocharger valve, and about the correct processing of the set tasks for the engine controlling, can be transferred through CAN. This concept, therefore, has an additional advantage in that no application requirement for the pulse turbocharger control

482 | Internal Combustion Engine Handbook

6606_Book.indb 482

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

device exists. All the adjustment parameters are located, as previously with concepts without pulse turbocharging, in the engine control again. 10.5.6.4 Mechanical System Integration Because of the thermodynamic requirements, an arrangement of the pulse turbocharger valve is required as a defined spacing to the combustion chamber in close proximity to the intake valve. The connection for the valve to the combustion engine will be selected as somewhat rigid. The valve will be preassembled in a valve mounting (created in aluminum) and electrically contacted (Figure 10.137).

Valve mount Impulse artery valve Intake manifolds Collector

Figure 10.137 Pulse turbocharger valve with aspiration pipe and valve mounting. See color section page 1083.

The guide for the aspiration air to the flow-body designed internal area of the valve will be simulated using CFD computations allowing for all the operating conditions. The constructive design of the aspiration channels will be executed according to the computational results. 10.5.6.5 Integrated Pulse Turbocharger-Aspiration Module Figure 10.138 shows a highly integrated aspiration module for a gasoline engine, which comprises the valve mounting with the pulse turbocharger valve, the synthetic aspiration pipe, the air filter with the applicable pulse turbocharger control unit, the fuel rail with the injectors, and the throttle valve-setting member and the sensors of the latest generation.

The selected design enables the application with not only charged but also with noncharged engines. This will be replaced on applications for diesel engines with an exhaust valve position and diesel prethrottle, and the preassembly of the injection system can be disregarded. 10.5.6.5.1 Throttle valve positioning member as a combination The installation of the pulse turbocharger control unit is executed in the illustrated example on an air filter casing. A cooling of the electronic is executed by the fresh air flow aspirated by the engine. 10.5.6.6 Prospects for the Future Additional works concern themselves with the possibilities for throttle-free load controlling by using EIC with the use of the freely controllable pulse turbocharger valve, whereby partial load mode for additional fuel savings in gasoline engines can be provided. Further examinations are being made for diesel engines where, along with torque increases, the additional potential for further emission and fuel reductions with pulse turbocharging are being considered.

Bibliography

10-1. Spicher, U. Verbrennungsmotoren A und B, Vorlesungsumdruck Universtät Karlsruhe (TH). 10-2. Schwelk et. al. 1989. Fachkunde Fahrzeugtechnik. Jansen Verlag Stuttgart: Holland. 10-3. Stoffregen, J. 2010. Motorradtechnik, 7 Aufl. Vieweg+Teubner: Wiesbaden. 10-4. Marquard, R. 1992. Konzeption von Ladungswechselsystemen für Pkw-Vierventilmotoren unter Fahrzeugrandbedingungen, Dissertation TH Aachen. 10-5. Pischinger, S. Verbrennungsmotoren I und II, Vorlesungsumdruck, FH Aachen. 10-6. Jungbluth, G. et. al. 1983. Bau und Berechnung von Verbrennungsmotoren, Springer Verlag: Berlin. 10-7. Shell Lexikon Verbrenungsmotoren, Supplement der ATZ und MTZ 10-8. Köhler, E. and Flierl, R. 2008. Verbrennungsmotoren, 5 Aufl. Vieweg+Teubner: Wiesbaden. 10-9. Aoi, K., Nomura, K., and Matsuzaka, H. “Optimization of Multi-Valve, Four Cycle Engine Design: The Benefit of Five-Valve Technology.” SAE Technical Paper 860032. 10-10. Brüggemann, H., Schäfer, M., and Gobien, E. 1985. “Die neuen Mercedes-Benz 2,6 und 3,0-Liter-Sechszylinder-Ottomotoren für die neue Baureihe W 124.” MTZ 46. 10-11. Duelli, H. 1987. “Berechnungen und Versuche zur Optimierung von Ansaugsystemen für Mehrzylindermotoren und Einzylinder-Einspritzung.” VDI-Fortschrittberichte, Reihe 12: p. 85. 10-12. Küntscher, V. (Hrsg.). 1995. Kraftfahrzeugmotoren—Auslegung und Konstruktion, 3rd Edition. Verlag Technik: Berlin. 10-13. List, H. 1950. Der Ladungswechsel der Verbrennungskraftmaschine, Teil II, Der Zweitakt. Springer-Verlag: Wien. 10-14. Schweitzer, P. H. 1949. Scavenging of Two-Stroke Cycle Engines. Macmilian: New York, NY.

Figure 10.138 Integrated aspiration module. See color section page 1083.

10-15. Gerecke, W. 1953. “Entwicklung und Betriebsverhalten des Feuerrings als Dichtelement hoch beanspruchter Kolben.” MTZ Jahrgang 14, 6: pp. 182–186.

Internal Combustion Engine Handbook | 483

6606_Book.indb 483

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 10 Charge Cycle

10-16. Venediger, H. J. 1947. Zweitaktspülung insbesondere Umkehrspülung. Franckh’sche Verlagshandlung: Stuttgart. 10-17. Bönsch, H. W. 1983. Der schnelllaufende Zweitaktmotor, 2nd Edition. Motorbuch Verlag: Stuttgart. 10-18. Kuhnt, H.-W., Budihartono, H., and Schneider, M. 2001. Auslegungsrichtlinien für Hochleistungs-2-Takt-Motoren; Vortrag bei der 4. Internationalen Jahrestagung für die Entwicklung von Kleinmotoren: Offenburg. pp. 16 und 17. 10-19. Blair, G. P. 1996. Design and Simulation of Two-Stroke Engines. SAE-Verlag: Warrendale, PA. ISBN 1-56091-685-0. 10-20. Bartsch, Ch. 1996. “Ein neuer Weg für den einfachen Zweitakter, Honda EXP-2 als Versuchsobjekt.” Automobil Revue 5, 1. 10-21. Zinner, K. 1985. Aufladung von Verbrennungsmotoren, Grundlagen— Berechnung—Ausführung, 3rd Aufl. Springer-Verlag: Berlin; Heidelberg; New York; Tokyo. 10-22. Wanscheid, W. A. 1968. Theorie der Dieselmotoren, 2nd Edition. VEB Verlag Technik: Berlin. 10-23. Küntscher, V. (Hrsg.). 1995. Kraftfahrzeugmotoren—Auslegung und Konstruktion, 3rd Edition. Verlag Technik: Berlin. 10-24. Zeman, J. 1935. Zweitaktdieselmaschinen. Springer-Verlag: Wien.

10-41. Knecht, A. 1998. Nockenwellenverstellsystem “Double-V-Cam”: Ein neues System für variable Steuerzeiten. Sonderdruck aus Systems Partners 98. Vieweg Verlagsgesellschaft mbH Wiesbaden. 10-42. Wenzel, C., Stephan, W., and Hannibal, W. 2000. Hydraulische Komponenten für variable Ventilsteuerungen. Vortrag Haus der Technik Essen. 10-43. Endres, H., Erdmann, H.-D., Eiser, A., Leitner, P., Kaulen, W., and Böhme, J. 2000. “Der neue Audi A4; Der neue 3,0-l-V6-Ottomotor.” Sonderausgabe der ATZ und MTZ. 10-44. Hannibal, W, Knecht, A., and Stephan, W. 2002. Nockenwellenversteller für Ottomotoren. Verlag Moderne Industrie: Band 247, Landsberg am Lech. 10-45. N. N. 2001. Die variable Ventilsteuerung VVTL-i der Fa, Toyota. Presseinformation der Fa. Toyota Köln. 10-46. Schwarzenthal, D., Hofstetter, M., Deeg, H.-P., Kerkau, M., and Lanz, H.-W. 2000. VarioCam Plus, die innovative Ventilsteuerung des neue 911 Turbo, Vortrag. 9. Aachener Kolloquium: pp. 4–6. 10-47. Torazza, G. 1972. A Variable Lift and Event Control Device for Piston Engine Valve Operation, 14. FISITA Kongress: London. 10-48. Titolo, A. 1986. “Die variable Ventilsteuerung von Fiat.” MTZ 47, 5.

10-25. Hack/Langkabel. 1999. Turbo- und Kompressormotoren; Entwicklung, Technik, Typen. Motorbuchverlag: Stuttgart.

10-49. Wichart, K. 1987. “Möglichkeiten und Maßnahmen zur Verminderung der Ladungswechselverluste beim Ottomotor.” VDI-Fortschrittsberichte Reihe 12, 91.

10-26. N.N. 1993. Fahrzeugmotoren im Vergleich: Tagung Dresden 3.–4. Juni 1993, VDI Gesellschaft Fahrzeugtechnik, VDI Berichte 1066. VDI-Verlag: Düsseldorf.

10-50. Schön, H. 1992. Untersuchungen an einem viergliedrigen Kurvenrastgetriebe zur variablen Betätigung der Ladungswechselventile in Hubkolbenmotoren. Dissertation, TH Karlsruhe.

10-27. N.N. 1954. “Der Napier-Diesel-Flugmotor Normad.” MTZ Jahrgang 15, 8: pp. 236–239.

10-51. Unger, H. 2004. Valvetronic, Der Beitrag des Ventiltriebs zur Reduzierung der CO2-Emission des Ottomotors. Verlag Moderne Industrie: Band 263, Landsberg am Lech.

10-28. Huber, G. 1997. Elektrisch unterstützte ATL-Aufladung (euATL)— Schaffung eines neuen Freiheitsgrades bei der motorischenVerbrennung, 6. Aufladetechnische Konferenz: Dresden. 10-29. Hagen, W. W. 1993. Vergleichende Untersuchung verschiedener variabler Ventilsteuerungen für Serien-Ottomotoren. Dissertation, Universität Stuttgart. 10-30. Haltenberger, S. 1918. Vorrichtung zur Ventilverstellung. Patent DE PS 368775. 10-31. Bassi, A., Arcari, F., and Perrone, F. 1985. C.E.M.—The Alfa Romeo Engine Management System-Design Concepts-Trends for the future. SAE-Paper 85 0290. 10-32. Inoue, K., Nagahiro, R., and Ajiki, Y. 1989. A High Power, Wide Torque Range, Efficient Engine with a Newly Developed Variable Valve-Lift and -Timing Mechanism. SAE-Paper 89 0675. 10-33. Metzner, F.-T. and Flebbe, H. 1999. Doppelnockenwellenverstellung an V-Motoren. 8. Aachener Kolloquium Fahrzeug- und Motorentechnik. 10-34. Ebel, B. and Metzner, F.-T. 2000. “Die neuen V-Motoren von Volkswagen mit Doppelnockenwellenverstellung.” MTZ 61, 12. 10-35. Hannibal, W. and Meyer, K. 2000. Patentrecherche und Überblick zu variablen Ventilsteuerungen. Vortrag Haus der Technik. 10-36. Ulrich, J. and Fiedler, O. 1991. “Der Motor des neuen Porsche 968.” MTZ 52, 12. 10-37. Knirsch, S., Mann, M., Dillig, H., and Reichert, H.-J. 1996. “Bartholmeß, T.: Der neue Sechszylinder-V-Motor von Audi mit Fünfventiltechnik.” MTZ 57. 10-38. Metzner, F.-T. and Keiser, P. 1999. Der neue V6-4V-Motor von Volkswagen, 20. Internationales Wiener Motorensymposium. 10-39. Batzill, M., Kirchner, W., Körkemeier, H., and Ulrich, H. J. 1997. “Der Antrieb für den neuen Porsche Boxter.” Sonderausgabe der ATZ und MTZ. 10-40. Braun, H. S., Flierl, R., Kramer, Marder, R., Schlerf, G., and Schopp, J. 1998. “Die neuen BMW Sechszylindermotoren.” Sonderausgabe ATZ und MTZ.

10-52. Klaus, B. 2004. Die Valvetronic der 2. Generation im neuen BMW-Reihen-Sechszylindermotor. VDI-Tagung: Stuttgart. pp. 15 und 16. 10-53. Jägerbauer, E., Fröhlich, K., and Fischer, H. 2003. “Der neue 6,0 l-Zwölfzylindermotor von BMW.” MTZ 64, 7/8. 10-54. Hannibal, W., Flierl, R., Stiegler, L., and Meyer, R. 2004. Overview of current continuously variable valve lift systems for fourstroke spark-ignition engines and the criteria for their design ratings. SAE-Paper 2004-01-1263, Detroit, MI, USA. 10-55. Hannibal, W., Flierl, R., Meyer, R., Knecht, A., and Gollasch, D. 2004. Aktueller Überblick über mechanisch variable Ventilsteuerungen und erste Versuchsergebnisse einer neuen mechanischen variablen Ventilsteuerung für hohe Drehzahlen, Variable Ventilsteuerung II. Expert Verlag, Band 32. 10-56. Hack, G. 1999. “Freie Wahl.” Auto Motor Sport 17: pp. 48–50. 10-57. Mischker, K. and Denger, D. 2003. Anforderungen an einen vollvariablen Ventiltrieb und Realisierung durch elektrohydraulische Ventilsteuerung EHVS, 24. Wiener Motoren Symposium. 10-58. Koch, A., Kramer, W., and Warnecke, V. 1999. Die Systemkomponenten eines elektromagnetischen Ventiltriebs, 20. Wiener Motoren Symposium. 10-59. Cosfeld, R., Klüting, M., and Grudno, A. 1999. Technologische Ansätze zur Darstellung eines elektromagnetischen Ventiltriebs. 8. Aachener Kolloquium Fahrzeug- und Motorentechnik. 10-60. Langen, P., Cosfeld, R., Grudno, A., and Reif, K. 2000. Der elektromagnetische Ventiltrieb als Basis zukünftiger Ottomotorenkonzepte. Vortrag Wiener Motoren Symposium. 10-61. Kirsten K. 2011. Variabler Ventiltrieb im Spannungsfeld von Downsizing und Hybridantrieb, 32. Internationales Wiener Motorensymposium. pp 5 und 6. 10-62. Elsäßer, A., Schilling, W., Schmidt, J., Brodesser, K., and Schatz, O. 2001. “Impulsaufladung und Laststeuerung von Hubkolbenmotoren durch ein Lufttaktventil.” MTZ 62.

484 | Internal Combustion Engine Handbook

6606_Book.indb 484

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

10.5 Pulse Turbocharging and Load Control of Reciprocating Piston Engines Using an Air Stroke Valve

10-63. Vogel, O. D. 2000. Load Control of SI Engines using Secondary Valves, Dissertation ETH Nr. 13633. Swiss Federal Institut of Technology: Zürich. 10-64. Schneider, E., Scholten, L., Wallrafen, W., and Zentgraf, M. Ein integriertes Saugmodul zur Anwendung der Impulsaufladung in der Großserie, Tagung 12. Aachener Kolloquium Fahrzeug- und Motorentechnik.

10-74. Broda, A., Eilts, P., Elsäßer, A., and Genieser, P. 2008. Intake Air Management with Air Pulse Valve—New Degrees of Freedom for DI-DieselEngines. THIESEL 2008 Conference on Thermo- and Fluid Dynamic Processes in Diesel Engines. 10-75. Kreuter, P., Bey, R., and Wensing, M. 2001. Impulslader für Otto- und Dieselmotoren, 22. Internationales Wiener Motorensymposium, Wien.

10-65. Findeisen, H., Linhart, J., and Wild, S. 2003. Development of an Actuator for a Fast Moving Flap for Impulse Charging. SAE-Paper 2003-01-0402.

10-76. Elsässer, A., Schilling, W., Schmidt, J., Brodesser, K., and Schatz, O. 2001. “Impulsaufladung und Laststeuerung von Hubkolbenmotoren durch ein Lufttaktventil.” MTZ 62.

10-66. Elsäßer, A., Schilling, W., Schmidt, J., Kallenbach, E., Beyer, F., and Baumbach, J. 2004. Entwurf schnellwirkender magnetischer Aktoren für Lufttaktventile. Tagung Innovative Klein- und Mikroantriebstechnik. 6. EGT-/ GMM-Fachtagung: Darmstadt.

10-77. Salber, W., Duesmann, M., and Schwaderlapp, M. 2002. Der Weg zur drosselfreien Laststeuerung. Fachtagung “Entwicklungstendenzen beim Ottomotor,” Ostfildern.

10-67. Koch, A., Kreuter, P., Wensing, M., Metzger, O., Bey, R., and Nüsser, J. 2004. Impulsaufladung: Umsetzbarkeit unter realen Fahrbedingungen und in Großserie am Beispiel eines Ottomotors mit Abgasturboaufladung, Variable Ventilsteuerung II—Ein Verfahren zur Reduzierung von Kraftstoffverbrauch und Emissionen. Expert Verlag: Haus der Technik Fachbuch, Band 32. 10-68. Sarikoc, F., Kettner, M., Velji, A., Spicher, U., Krause, A., and Elsäßer, A. 2006. Potential of Reducing the NOX Emissions in a Spray Guided DI Gasoline Engine by Stratified Exhaust Gas Recirculation (EGR). SAE-Paper 2006-01-1261. 10-69. Takahashi, H., Harada, T., Yamaki, T., and Oikawa, T. 2006. Study on Impulse Charger for Enhancement of Volumetric Efficiency of SI Engine. SAE-Paper 2006-01-0191. 10-70. Dingelstadt, R., Elsäßer, A., Schilling, W., Schmidt, J., Kallenbach, E., Beyer, F., Baumbach, J., Otto, R., and Kucera, U. 2005. Modellbasierte Optimierung von Magnetantrieben für Lufttaktventile. Tagung Mechatronik ’05, Wiesloch. 10-71. Eilts, P. and Broda, A. 2008. Potentiale des variablen Ventiltriebs am Dieselmotor. MTZ-Konferenz—Ladungswechsel im Verbrennungsmotor, Stuttgart. 10-72. Broda, A., Rieping, M., Eilts, P., Elsäßer, A., and Genieser, P. 2008. Active Air Management with High Speed Flap for DI-Diesel-Engines. SAE-Paper 2008-01-1345. 10-73. Broda, A., Eilts, P., Elsäßer, A., and Genieser, P. 2008. Active Air Management with Air Pulse Valve—Strategies to overcome Emission Legislations for DI-Diesel-Engines. Fisita F2008-06-053.

10-78. Jahrens, H.-U., Krebs, R., Lieske, S., Middendorf, H., Breuer, M., and Wedowski, S. 2001. Untersuchungen zum Saugrohreinfluss auf die Klopfbegrenzung eines Ottomotors. 10. Aachener Kolloquium Fahrzeugund Motorentechnik, 2001, Aachen. 10-79. Koch, A., Kramer, W., and Warnecke, V. 1999. Die Systemkomponenten eines elektromagnetischen Ventiltriebs. 20. Internationales Wiener Motorensymposium, Wien. 10-80. META Motoren Und Energietechnik GmbH. 2001. In einem Einlasskanal einer Kolbenbrennkraftmaschine angeordnete Zusatzventileinrichtung. Patentschrift DE 10137828A1. 10-81. Siemens AG. 1999. Apparatus for Controlling an Electromechanical Actuator. Patentschrift WO 0079548A2. 10-82. Gunselmann, C. and Melbert, J. 2003. Improved Robustness and Energy Consumption for Sensorless Electromagnetic Valve Train. SAE 2003 World Congress, Detroit, MI, USA. 10-83. Findeisen, H., Linhart, J., and Wild, S. 2003. Development of an Actuator for a Fast Moving Flap for Impulse Charging. SAE 2003 World Congress, Detroit, MI, USA. 10-84. Kreuter, P., Wensing, M., Bey, R., Peter, U., and Böcker, O. 2002. Kombination vom Abgasturbolader und Impulsaufladung. 11. Aachener Motorensymposium, Aachen. 10-85. Franz, A., Wild, S., and Katsivelos, H. 2004. Der Wettbewerb von Strömungsmaschine und Impulslader für ein optimales transientes Verhalten und geringste Abgasemissionen des Verbrennungsmotors. 25. Int. Wiener Motorensymposium 2004, Band 2. Fortschritt-Berichte VDI, Reihe 12, Nr. 566.

Internal Combustion Engine Handbook | 485

6606_Book.indb 485

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 486

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11 Supercharging of Internal Combustion Engines The most important goals in the development of internal combustion engines, namely high efficiency, that is, low fuel consumption and reduced emissions, have been explicitly illustrated in the previous chapters. An additional important point here is the increase in performance concentration of an internal combustion engine. This should therefore concentrate on achieving the highest possible power from a defined engine volume and/or from a predetermined engine weight. Increasing the performance concentration is, under certain conditions, additionally linked with improved efficiency. The power output of an internal combustion engine is directly proportional to the effective mean pressure curve pme, the rpm n and the total piston displacement volumes VH. 1 Pe = pme ⋅ n⋅VH ⋅ Z



p me = r2 ⋅ lL ⋅ he ⋅

Hu l ⋅ Lmin

where Z

= 2   4 stroke, Z = 1   2 stroke

(11.0)

(11.1)

An increase in the piston displacement will, alongside an increased performance, result in a respective increase in the dimensions for the engine and the required construction space, as well as a reduction in the efficiency due to the increased friction levels involved. The friction losses increase disproportionately to the increase in revolutions which, however, can also achieve an increase in performance. Net calorific value Hu and minimum air requirement Lmin are fuel characteristic values and are taken to be predetermined. pme ~ r2 ⋅



1 he ⋅ lL l

Therefore, the effective mean pressure curve is proportional to the air density, the effective efficiency and the degree of supply as well as, conversely proportional to the air ratio. The air density is dependent on the charging pressure and the charge air temperature. r2 =



p2 R ⋅T2

ρ 2 = density behind the charger

pme = effective mean pressure

p2 = charge pressure

= rpm

(11.3)

where

Pe = effective performance n

(11.2)

R = gas constant

VH = displacement volumes

T2 = temperature behind the compressor.

ρ 2 = density according to charge

Increasing the air density will also significantly increase the effective performance. Especially with today’s larger engines, up to 31 bar mean pressure can be achieved and private vehicle engines can already achieve 23 bar (Otto) and/or 29 bar (Diesel). In addition to increasing the performance density, supercharging is of vital importance for minimizing engine exhaust

λ L = degree of supply η e = effective efficiency Hu = lower net calorific value

λ

= air ratio

Lmin = minimum air ratio.

Internal Combustion Engine Handbook | 487

6606_Book.indb 487

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

emissions and has today become an integral component for relevant concepts accordingly.

11.1 Mechanical Supercharging During mechanical supercharging, the compressor is driven mechanically by the crankshaft (Figure 11.1). The compression work must therefore be performed by the engine, whereby a proportion is fed back via the piston during the intake stroke.

Figure 11.2a Compressor map, roots supercharger [11-1].

Figure 11.1 Schematic illustration of mechanical supercharging.

The process now functions at an increased pressure level. This then leads to a respective increase in the mean pressure, on the assumption that the air–fuel ratio remains constant. The mechanical supercharging initially results in a deterioration in efficiency despite the increase in performance. If one compares the mechanically supercharged engine with a naturally aspirated engine of the same output, then the mechanically supercharged engine produces a higher efficiency due to the lower mechanical and thermal losses. A radial compressor is seldom utilized as a compressor (with transfer gear), roots blowers (Figure 11.2 and Figure11.2a), screw-type compressor (Figure 11.3) or spiral-type turbocharger (Figure 11.4) are mostly utilized. Mechanical supercharging is mainly utilized today for private gasoline engines. In these applications, it has the benefit that, during cold starting, no heat is taken from the exhaust gas stream, which is of great importance for the starting of the catalytic converter during the warm-up phase.

Figure 11.2 Roots supercharger [11-1].

Figure 11.3 Screw supercharger [11-2].

Figure 11.4 Spiral supercharger [11-3].

488 | Internal Combustion Engine Handbook

6606_Book.indb 488

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.2 Exhaust Gas Turbocharging

11.2 Exhaust Gas Turbocharging During exhaust gas turbocharging, the engine and the turbocharger (Figure 11.5) are linked thermodynamically and not mechanically. The compressor is driven by the turbine. The turbine receives the exhaust gas stream from the engine and, thereby, covers the power requirement of the compressor.

Figure 11.6 Pressure waves in shock or pulse charging.

Figure 11.5 Schematic illustration of the exhaust gas turbocharging.

11.2.1 Constant Pressure Turbocharging

an increase in the residual gas content. With turbocharged spark ignition (SI) engines, the increased residual gas content results in a greater knock tendency; this, in turn, leads to a delayed ignition angle and, hence, to a loss of torque and increased fuel consumption. The exhaust-driven turbocharger consists of a compressor and a turbine Figure 11.7. The internals are shown in Figure 11.8.

With constant pressure turbocharging, a large exhaust gas line volume is provided between the turbine and compressor with the aim of reducing the pressure hammer of the individual cylinders and for pressurizing the turbine as continuously as possible, that is, with a constant status p3, T3. If we assume in an initial approximation that pressure p3 is equal pressure p2, then the engine will be operated at a high-pressure level without any change in the thermal efficiency. If we look more closely, however, then one can determine that a larger volume is relieved in the turbine so that a slight gain is possible. When p2 > p3, then a part of the turbocharger work will be output again to the crankshaft via the positive charge cycle loop.

11.2.2 Ram Induction Turbocharging or Impulse Turbocharging

During ram induction or impulse turbocharging, the kinetic energy of the exhaust gas is additionally used in the form of pressure waves. Figure 11.6 shows the pressure curve of a turbine. Compared with the ram induction, this offers a gain as an isentropic expansion to the ambient state takes place instead of the irreversible throttling from the cylinder pressure to the exhaust gas counter pressure p3. In fact, this gain cannot be completely exploited as a throttling takes place at the exhaust valves anyway and because the turbine efficiencies with nonstatic charging are lower than with static charging. Compared with ram induction, pulse turbocharging has advantages especially in part-load operation and in the acceleration behavior. Appropriate grouping of the cylinders with the given ignition firing sequence prevents exhaust gas from being pressed into a cylinder during the valve overlap, as this would result in

Figure 11.7 Exhaust turbocharger [11-4].

Figure 11.8 Turbocharger internals [11-4].

The operation of the compressor is imaged in a compressor map Figure 11.9.

Internal Combustion Engine Handbook | 489

6606_Book.indb 489

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

11.3 Intercooling If one considers an isentropic compression process from 1 to 2s (Figure 11.11), this results in a temperature increase. Due to an isentropic compression according to the Eq. (11.4), T2 ⎛ p2 ⎞ = T1 ⎜⎝ p1 ⎠⎟



k −1 k



(11.4)

where T1 = temperature up line of the compressor T2 = temperature behind the compressor p1 = pressure up line of the compressor p2 = charge pressure Figure 11.9 Compressor map radial compressor—determined on hot-gas test bench.

Above the volumetric flow V!1 (reduced) and pressure ratio p2/ p1 (total pressure), the compressor rpm and isentropic compressor efficiency are applied. If one follows a constant surface speed (rpm) to the left, that is, one throttles the compressor on the pressure side increasingly, one will reach the pump limit. This must not be reached during operation as otherwise the compressor would be destroyed. For the presentation of the turbine behavior, the isentropic turbine efficiency and the throughflow coefficient are plotted against the turbine pressure ratio p3/p4 (total/static) (Figure 11.10) with the peripheral speed (turbine rpm) as a parameter. If the behavior of the turbine is determined on a hot gas test bench, then the mechanical efficiency of the exhaust-driven turbocharger is included in the turbine efficiency.

κ = isentropic exponent.

p2

T

p1 2

2s

1

s

Figure 11.11 Isentropic and polytropic compression.

Since the compression is performed polytropically instead of isentropically, a further increase in temperature occurs according to the following equation: T2 − T1 =

(T2 − T1 )s hsV ⋅ t K



(11.5)

where T1 = temperature up line of the compressor T2 = temperature behind the compressor

η sV = isentropic compression efficiency τ K = cooling coefficient of the compressor. The isentropic compression efficiency hsV is calculated as follows: hsV =

c p ⋅ (T2 − T1 ) h2 − h1 ≈ h2 s − h1 c p ⋅ (T2 s − T1 )

(11.6)

where h1 = enthalpy up line of the compressor Figure 11.10 Turbine map radial turbine—determined on hot-gas test bench.

h2 = enthalpy downline of the compressor

490 | Internal Combustion Engine Handbook

6606_Book.indb 490

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.4 Interaction Between Engine and Compressor

11.4 Interaction Between Engine and Compressor

h2s = enthalpy downline of the compressor, isentropic cp = specific thermal capacity for p = constant. The cooling coefficient τ K mentioned in Eq. (11.5) for turbochargers makes allowance for the heat dissipation via the compressor housing (particularly with large compressors) to the environment and lies in the range between 1.04 and 1.1. The temperature increase associated with an increase in pressure leads to a reduction in the density as shown in Eq. (11.3). An intercooler allows the charge density, and also the output, to be increased as shown in Eq. (11.2).

11.4.1 Four-Stroke Engine in Compressor Map

Figure 11.13 shows the displacement lines of a four-stroke internal combustion engine. If the engine speed n is held constant, then the volumetric flow V!1 shows only a slight linear increase with increasing compression ratio p2/p1. The engine then operates as a volumetric displacement machine, and its throughput increases in relation to the increase in engine speed. With increasing valve overlap and constant engine speed, the volumetric flow V!1 increases less sharply with increasing compression ratio p2/p1.

Example p1 = 1 bar;

T1 = 293°K (20°C)

Compressor: π V = p2/p1 = 2.5

η sV = 0.85



T2 = 313 K (40°C)

11.4.1.1 Positive Displacement Superchargers Some of the examples for positive displacement superchargers are piston compressors, reciprocating pistons and rotary pistons, roots blowers, and screw-type superchargers. From Figure 11.14, it can be seen that the throughput increases with increasing compressor speed and drops slightly with increasing counter pressure. At constant speed, we obtain the working points 1, 2, or 3, depending on the counter pressure.

Figure 11.12 shows a comparison between a naturally aspirated engine, a turbocharged engine, and a turbocharged engine with intercooling to 40°C. The same air–fuel ratio has been assumed for all three cases. This shows a direct relationship between the density and the output. The ambient state 1 bar and 20°C is assumed for the naturally aspirated engine. Comparison of the naturally aspirated engine with the turbocharged engine and a turbocharger compressor ratio of 2.5 shows an increase in mean pressure to 187%, and for the turbocharged engine with intercooling to 40°C, an increase in mean pressure to 234% is shown. r2

Engine Naturally aspirated engine

kg m3

n1 pL

Mean pressure

1.19

Turbocharged engine

2.23

187%

2.78

234%

n2

<

n3

1

2

100%

Turbocharged engine with intercooling

<

3

Figure 11.12 Density and mean pressure of engines. V1

Figure 11.14 Displacements lines positive displacement superchargers. n1

<

n2

<

n3

P2 P1 Without valve overlap With valve overlap

V (m3/s) Sip lines 4-stroke internal combustion engine (= piston engine)

Figure 11.13 Displacements line 4-stroke motor.

Internal Combustion Engine Handbook | 491

6606_Book.indb 491

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

11.4.1.2 Radial Compressor The radial compressor operates on the centrifugal principle. The increase in pressure is created by the difference in circumferential speed between the inlet and outlet at the impeller. The kinetic energy thereby admitted is converted into pressure in the diffuser. The compressor map shown in Figure 11.15 is limited by the pump limit.

A stable compressor operation is not possible to the left of the pump limit. In this area, there results a breakaway of the flow which commences at the inside of the compressor blades and results in extreme fluctuations that, under certain circumstances, may destroy the compressor. The speed lines drop slightly to the right of the pump limit; they drop more sharply towards the displacement limit. Depending on the counterpressure, this results in working points 1, 2, or 3 at constant compressor speed.

Surge limit

11.4.2 Mechanical Supercharging

pL

11.4.2.1 Positive Displacement Supercharger—Mechanically Linked to the Four-Stroke Engine (Figure 11.16) With a given transmission ratio, the operating curve 1–2–3–4 will result. By changing the transmission ratio, we can also create the operating curve 1′–2′–3′–4′ that leads to an increase in the mean working pressure.

n3

1

n2

2 3

n1

11.4.2.2 Radial Compressor—Mechanically Linked to the Four-Stroke Engine As shown in Figure 11.17, air throughput and charge pressure increase at roughly the square of the rise in speed. This results in the mean pressure curve over speed illustrated in Figure 11.18.

V1

Figure 11.15 Compressor map radial compressor. 1/4 nm

1/2 nm

1/4 nL

3/4 nm

3/4 nL

pL

1′ 1

nm

1/2 nL

2′

nL

3′

2

4′

3

4

Figure 11.16 Mechanical inking of positive displacement turbocharger and four-stroke engine.

V1

Su

rg e

lim

it

pL

r

pe

O

g

in at

lin

e

nL

3/4 nL 1/4 nL

1/2 nL

V1

Figure 11.17 Mechanical inking of radial compressor and four-stroke engine.

492 | Internal Combustion Engine Handbook

6606_Book.indb 492

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.4 Interaction Between Engine and Compressor

mechanically charged with centrifugal compressor

Full load line



pme

k −1 ⎡ ⎤ k 1 ⎢ ⎛ p2 ⎞ k 1 ∆hsV = R1 ⋅T1 ⋅ ⋅ ⎜ ⎟ − 1⎥ ⎥ k 1 − 1 ⎢⎝ p1 ⎠ ⎢⎣ ⎥⎦

(11.12)

R1 = gas constant upline of compressor

Naturally aspirated engine

T1 = temperature upline of compressor

κ 1 = isentropic exponent upline of compressor p1 = pressure upline of compressor p2 = charge pressure

nM

Figure 11.18 Mean pressure curve over speed (rpm).

∆hsT = R3 ⋅T3 ⋅

11.4.3 Exhaust Gas Turbocharging

During exhaust gas turbocharging, the engine and exhaustdriven turbocharger are linked thermodynamically. The respective turbocharger speed is set depending on the power balance between compressor and turbine. If we consider the power balance at the turbocharger shaft, then the change in the angular velocity will result as follows: dwTL ⋅ JTL ⋅ wTL = PV + PT dt



(11.7)

where dwTL = change in the angle velocity ATL dt JTL = turbocharger polar mass moment of inertia

ω TL

= compressor power

PT

= turbine power.

ΔhsT = isentropic enthalpy gradient in the turbine T3 = temperature upline of turbine

κ 3 = isentropic exponent of exhaust gas p3 = exhaust gas counterpressure p4 = pressure downline of turbine. The group efficiency η TL is defined as the overall efficiency of the charge group:

PV + PT = 0

(11.8)



!V +m !B = m !T m

(11.9)

where ṁV = compressor mass flow The operating point lies on the engine displacement line. The power balance can thus be developed further. 1 hsV ⋅ hmV

(11.10)

π V = compression power ratio and with K L= 1.4, the turbocharger main equation will be

∆hsV = isentropic enthalpy gradient in the compressor



(11.16)

cp1 = specific thermal capacity upline of the compressor cp3 = specific thermal capacity upline of the turbine

η TL = group efficiency.



η mV = mechanical compressor efficiency

η mT = mechanical turbine efficiency

3,5

where

η sV = isentropic compressor efficiency

∆hsT = isentropic enthalpy gradient in the turbine

k 3 −1 ⎡ ⎤ ⎛ ! c p3 T3 m p ⎞ k3 ⎥ pV = ⎢1 + T ⋅ ⋅ ⋅ hTL ⋅ ⎜ 1 − 4 ⎟ ⎢ m ⎥ ! V c p1 T1 p3 ⎠ ⎝ ⎢⎣ ⎥⎦

! T ≡ 1.03 − 1.07, then the compressor ratio is If we assume m a function of the following factors:

where

! T ⋅ ∆hsT ⋅ hsT ⋅ hmT PT = m

(11.15)

where



ṁB = fuel mass flow.

p V = p2 /p1



ṁT = turbine mass flow



(11.14)

Utilizing Eqs. (11.9 to 11.13), the power balance to be solved for π V and calculated as follows:





hTL = hmV ⋅ hsV ⋅ hmT ⋅ hsT

where η sT = isentropic turbine efficiency.

The left-hand side of the equation is 0 for the static state:

! V ⋅ ∆hsV ⋅ PV = m

(11.13)

R3 = gas constant upline of turbine



= angular velocity ATL

PV

k 3 −1 ⎡ ⎤ k 3 ⎢ ⎛ p4 ⎞ k 3 ⎥ ⋅ 1− ⎜ ⎟ ⎥ k 3 − 1 ⎢ ⎝ p3 ⎠ ⎢⎣ ⎥⎦

(11.11)

⎛T p ⎞ p V = p V ⎜ 3 ; hTL ; 4 ⎟ p3 ⎠ ⎝ T1

(11.17)

The charge pressure p2 thus increases with increasing exhaust gas temperature T3 and increasing pressure upline of the turbine p3 (whereby the change in the group efficiency dependent on T3 and p3 has still been neglected).

Internal Combustion Engine Handbook | 493

6606_Book.indb 493

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

The pressure p3 is obtained with a given turbine as a function of the mass throughput and gas status and can be calculated for the ram induction mode from ! T = AT red ⋅y T 2 ⋅ p3 ⋅ r3 m



2

with y T =

⎛ p ⎞ k3 ⎛ p ⎞ k3 ⋅ ⎜ 4⎟ −⎜ 4⎟ k 3 − 1 ⎝ p3 ⎠ ⎝ p3 ⎠

(11.18) k 3 +1 k3



1. Generator operation mode In the so-called “generator mode,” the speed has to be kept as constant as possible in view of the high demands on the constant rotational frequency of the generator nM (Figure 11.19).

nm = constant pme = variable

(11.19)

where

pme

AT red = turbine equivalent cross section

ψ T = throughflow function κ3

Full load line

= isentropic exponent of the exhaust gas.

If one considers the turbine as a throttle point (with p3 upline and p4 downline of the throttle point), then this results in the following relationship: !2 r r3 2 m ( n ⋅V ⋅ r ) ⋅ v3 ~ 2T ⋅ 2 3 ~ M H 2 2 r3 AT red r3

Generator operation

2

p3 − p 4 =

(11.20)

where

nm

Figure 11.19 Generator mode.

ρ2

= density downline of the turbocharger

ρ3

= density upline of the turbocharger

v3

= flow velocity, turbine

For the engine with a mechanical turbocharger, one stays at one operating point, nM = constant (Figure 11.20).

AT red = turbine equivalent cross section nM

= engine speed (rpm)

VH = piston displacement volumes. The mass throughput ṁT through the turbine depends, in a first approximation, on the gas status at the intake organs (p2, T2), on the engine speed nM (displacement line), and the density ρ 3. The reduced turbine cross-sectional area AT red is assumed to be constant in this consideration. The following relationship thereby exists:

p3 p3 = ( p2 ,T2 , nM ,T3 , AT red ) p4 p4

pL

nL

nM Operating point in generator mode

(11.21)

where T2 = temperature behind the compressor. Whereby in the case of an engine with a mechanically driven turbocharger and constant transmission ratio the charge pressure and, hence, the maximum torque are only a question of the engine speed, as the Eq. (11.20) illustrates, it is possible via a further reduction in the reduced turbine cross section AT red to increase the exhaust gas counterpressure p3. This results in an increase in the enthalpy gradient at the turbine. The turbocharger output and speed are therefore increased and, consequently, the charge pressure also increases. In principle, different operating points for the same AT red result in a different enthalpy gradient at the turbine and therefore also a different charge pressure. This thermodynamic interaction of the engine and exhaust-driven turbocharger should now be discussed with regard to three borderline cases.

V1

Figure 11.20 Working point in generator mode.

For the exhaust-driven gas turbocharged engine, the change in load results in a different p3 and T3 and, hence, in a different turbine power and different charge pressure. The operating points 1, 2, and 3 all lie on the engine displacement line that belongs to the generator speed (Figure 11.21). With an increase in load (increase in fuel injection), p3 and T3 increases and, thereby, the turbine power. The turbocharger speed increases, as do charge pressure p2 and mass throughput.

494 | Internal Combustion Engine Handbook

6606_Book.indb 494

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.4 Interaction Between Engine and Compressor

3 pL

Operating line in generator mode 2

nTL = const.

1

 . Propeller mode nM = variable, pme ~ nM2 3 For ship drives with a fixed propeller, the propeller torque taken up depends on the square of the propeller speed. In the compressor map, Figure 11.23, the operating line lies between generator mode and speed reduction. Figure 11.24 illustrates a superimposition of all lines of constant load and constant speed. Therefore, this enables complete coverage of the whole area in vehicle mode. This requires wider compression mapping. Figure 11.25 illustrates the mean pressure curve for the full-load line of naturally aspirated, mechanically turbocharged and exhaust-driven gas turbocharged engines.

V1 Surge limit

Figure 11.21 Engine displacement line and generator mode.

nL = const.

 . Speed reduction pme = constant, nM = variable 2 As illustrated in Figure 11.22, the mean pressure moves along a horizontal line for different engine speeds. This results in a flatter operating line (a) in the compressor map (Figure 11.23); that is, with decreasing speed, the operating point moves toward the pump limit (Danger!). This mode of speed reduction also occurs roughly in in vehicle mode along the full-load line and makes the highest demands on exhaust gas turbocharging.

pL

const. Rotational speed

V1

Full load line

pme

ηL = const.

const. Load

Figure 11.24 Superimpositions of characteristic curves. pme = const.

Propeller line

pme n

Figure 11.22 Speed reduction.

Surge limit pL

a b

c

a Constant mean pressure b Propeller operation mode c Generator operation mode nTL = const.

V1

Full-load

Exhaust gas turbocharging (unregulated) Mechanically charged Naturally aspirated engine

n

Figure 11.25 Full-load curves for various engine variants.

This latter line shows a highly unfavorable behavior as the torque also drops with decreasing speed. For good acceleration behavior in vehicle mode, however, a rise in the mean pressure curve is required with decreasing speed. This can be achieved by external controlling of the turbocharger.

Figure 11.23 Operating line between generator mode and speed reduction.

Internal Combustion Engine Handbook | 495

6606_Book.indb 495

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

11.4.3.1 Optimum Torque Curve by Adaption of the Charge Pressure In order to achieve a high charge pressure, even at low engine speeds (private vehicles < 1750 rpm, commercial vehicles < 1000 rpm), a small turbine neck cross section AT red should be selected; this therefore increases the pressure upline of the turbine. With increasing speed, however, the charge pressure also increases because of the increasing exhaust gas enthalpy stream, so that the maximum pressure in the cylinder also rises. To limit the associated component load, the charge pressure will be controlled to a constant value so that the excessive exhaust gas enthalpy stream bypasses the turbine (Waste Gate) and thereby escapes unused into the exhaust pipe (Figure 11.26), which represents a loss for the engine.

exhaust gas counterpressures, and a correspondingly higher charge pressure is achieved.

Figure 11.28 Variable turbine geometry, blade position open [11-4].

With higher speed and, thereby, increasing mass throughput, the blades are turned in the direction of maximum contact cross section (blade position is illustrated in Figure 11.28). Figure 11.29 illustrates the turbocharger group with two exhaust-driven turbochargers for six-cylinder gasoline engine. Figure 11.30 illustrates a turbocharged medium-speed ship diesel engine and Figure 11.31 illustrates a corresponding exhaust-driven turbocharger with axial turbine.

Figure 11.26 Schematic illustration of the exhaust gas turbocharging with waste gate.

The charge pressure curve along the full-load line and the effective mean pressure are illustrated in Figure 11.27 for an Audi 2.7 I Biturbo engine. 4,0

22,0 20,0

3,5

Boost pressure [bar]

Mean pressure [bar]

18,0

3,0

16,0

2,5

14,0 12,0

2,0

10,0

1,5

8,0 6,0

0

1000

1,0 2000 3000 4000 5000 1/min 7000 Rotational speed

Figure 11.27 Mean pressure and charge pressure curves of an AUDI V6 2,7 l Biturbo [11-5]

By utilizing the variable turbine geometry (Figure 11.28), it is possible to set the reduced turbine cross section as very small even at low speeds. This therefore generates higher

Figure 11.29 Biturbo turbocharging group [11-6].

For the dual-staged turbocharging (Figure 11.32), two exhaust-driven turbochargers are connected in series, where the compressed air is postcooled downline of the first compressor and cooled again downline of the high-pressure compressor.

496 | Internal Combustion Engine Handbook

6606_Book.indb 496

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.4 Interaction Between Engine and Compressor

Figure 11.30 Diesel-electric plant of the Queen Elizabeth 2, 9 × 9 L 58/64, 95.5 MW [11-7].

Figure 11.32 Schematic illustration of two-staged supercharging. Figure 11.31 MAN-exhaust-driven turbocharger with axial turbine [11-7].

This dual-stage compression with after cooling produces a good compression efficiency and, with a compressor pressure ratio >5, also produces a correspondingly high mean pressure of up to 30 bar. In applications with higher demands for dynamic and throughput ranges (automobile applications), the concept will

be expanded via the utilization of one (only high pressure, commercial vehicles) and/or two waste gates (private vehicles) for controlled two-stage turbocharging. The high degree of integration of the turbocharger group with multiple exhaust-driven turbochargers is clearly illustrated in Figure 11.33 (boxed detail) and Figure 11.33a.

Internal Combustion Engine Handbook | 497

6606_Book.indb 497

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

Figure 11.33 20 V 4000 M93, 3900 kW @ 2100 1/min [11-8].

Figure 11.33a Two-staged turbocharger group, private vehicle [11-9].

11.5 Dynamic Behavior The internal combustion engine forms part of a drive system from which a rapid response behavior is demanded. This applies for all applications. Emergency power generators have to be able to assume the full power from a standstill within the shortest possible time (<15 s). In vehicle operation, the internal combustion engine also has to spontaneously (<2 s) develop its maximum torque even under extreme load conditions (e.g., a car with a trailer starting in the mountains). Naturally aspirated engines control the torque more or less directly with the throttle plate angle (SI) or via the fuel injection volume (diesel engine). If we calculate the twist equation for a roughly torsionally rigid drive system (Eq. 11.22), then one can see that with a given consumer torque MV (= Load) the effective engine

torque M Me and the polar inertia moment of the whole drive system Jges A = JM + JA significantly influence the gradient of the crankshaft angular velocity.

( JM + JA ) ⋅

d wM = M Me + MV dt

(11.22)

where JM

= polar mass moment of inertia of the engine

JA = polar mass moment of inertia of the drive unit d wM = change in crankshaft angular velocity dt M Me = effective engine torque MV

= consumer torque.

Figure 11.34 illustrates an elasticity test for a vehicle with a turbocharged SI engine for acceleration from 60 to 100 km/h in 5th gear on the highly dynamic test rig.

498 | Internal Combustion Engine Handbook

6606_Book.indb 498

1/19/16 8:42 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.5 Dynamic Behavior

0

1

2

3

4

5 6 Time [s]

7

8

9

10

11

0

1

2

3

4

5 6 Time [s]

7

8

9

10

11

0

1

2

3

4

5 6 Time [s]

7

8

9

10

11

1.8 1.6 1.4 1.2 1 0.8 0.6 0.4

18 16 14 12 10 8 6 4 2 0

It takes almost 3.5 s for the intake manifold pressure, and, hence, the mean pressure, to reach its static value. Figure 11.35 illustrates further measurements for a load shift in a dynamic SI engine at constant rpm n = 2,000 1/min = constant on the highly dynamic engine test rig. The mean

pressure is hereby normalized to the static maximum value. The measured load signal rises rectangularly at 1 s to 100%. After a dead time, the naturally aspirated engine produces an equally spontaneous rise. The exhaust gas turbocharged SI engine rises with the same spontaneity up to approximately

110

110

90

90

Actual load signal [% ]

100

100

80

80

70

70 60

60

Actual load signal

50

50

40

Turbo engine

40

20

Naturally aspirated engine

20

30

30 10

10 0

Figure 11.34 Elasticity test (60–100 km/h 5th gear) highly dynamic test rig, SI engine with exhaust gas turbocharging.

Torque [%]

Mean pressure [bar]

Intake manifold pressure [bar]

Speed [km/h]

105 100 95 90 85 80 75 70 65 60 55

0

1

2

3

4

5 Time [s]

6

7

8

9

10

0

Figure 11.35 Comparison of naturally aspirated engine and turbocharged SI engine, load shift at n = 2000 1/min = constant.

Internal Combustion Engine Handbook | 499

6606_Book.indb 499

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

18

Mean pressure [bar)

16 14 12 10

Turbocharger

8 6

Mechanical supercharging

4 2 0

0

1

2

3

4

5

6 7 Time [s]

8

55% of the achievable static mean pressure. The subsequent slow rise of 13%/s is attributable to the acceleration of the turbocharger internals. The engine reaches its maximum mean pressure after approximately 3 s. Before we proceed to discuss measures for improving the torque development in the exhaust-driven turbocharger internal combustion engine, we see in Figure 11.36 the acceleration behavior of a mechanically supercharged engine, which achieves a significantly faster buildup of mean pressure compared with the exhaust gas turbocharger.

11.5.1 Improvement Measures

Adjustment devices, such as exhaust-driven turbocharger with waste gate or variable turbine geometry, enable the charge pressure to be built up significantly faster during an acceleration phase. In addition, the dynamic charge pressure buildup during transient processes can be improved by using smaller impellers for turbine and compressor. The influence of the polar mass moment of inertia J TL of the internals is illustrated in the twist equation (Eq. 11.23) for the exhaust gas turbocharger shaft. In V engines, for example, the dynamic behavior can be improved by grouping the cylinders on the exhaust gas side into a bank feeding two smaller turbines; on the air intake side, the two compressors are connected to a common intake pipe.



1 d wTL = ⋅ ( PT + Pν ) dj wTL ⋅ JTL

(11.23)

where d wTL = change in the angle velocity ATL dj ω TL = angular velocity ATL JTL

= polar mass moment of inertia for the turbocharger

PT

= turbine output

PV

= compressor output.

11.5.2 Active Residual Gas Discharge

The output torque of a combustion engine is basically proportional to the filling of the cylinder with fresh charges. A reduced residual gas proportion does not correspondingly lead to an

9

10

11

12

Figure 11.36 Comparison of mechanical supercharging–exhaust gas turbocharging, vehicle acceleration process (elasticity test) in the SI gasoline engine.

increased engine torque with exhaust-driven turbocharger engines, rather more it also immediately contributes to a quicker start-up of the exhaust gas turbocharger via the increased exhaust gas enthalpy flow. A prerequisite for an active residual gas discharge is a sufficiently high flushing gradient via the cylinder during the valve overlap. To prevent malfunctions from exhaust-side return effects from the blow-down from other cylinders, the cylinders will be combined with sufficiently high ignition spacing in a joint exhaust channel. The relevant groups will be arranged either with their own turbines (e.g., Biturbo on 6 cylinder), or differing channels of a joint turbine (e.g., twin-flow on 4 cylinder). Brief injection of additional air into the compressor means, on the one hand, that the internal combustion engine is adequately supplied with air immediately after a load demand, and that the increased fuel injection volume corresponding to the limit air ratio provides a rapid increase in torque. On the other hand, the blow compressor wheel is accelerated so that the compressor delivers correspondingly more air with the increasing speed. The air injection is terminated when the turbine takes over the compressor power and the additional acceleration work required.

11.5.3 Electric Support for Exhaust Gas Turbocharging

Since the internal combustion engine does not spontaneously provide sufficient acceleration power for the turbocharged internals in response to a demand for torque, it is expedient to utilized stored electrical energy to accelerate the turbocharger internals by utilizing an electric motor, which is connected between the compressor and turbine (“euATL,” ) (Figure 11.37) [11-10]. The electric motor must also withstand the high turbocharger speeds when it is switched off, and have sufficient torque for acceleration of the internals (compressor and turbine wheel). If an electrically driven compressor (“eBooster,” [11-10]) is connected in series (Figure 11.38) that briefly takes over the air supply to the internal combustion engine, then the electric motor has to accelerate only the compressor wheel whose polar mass moment of inertia is only one-third of that of the turbine wheel. With an appropriate design of the

500 | Internal Combustion Engine Handbook

6606_Book.indb 500

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.6 Additional Measures for Supercharged Internal Combustion Engines

the exhaust-driven turbocharger, especially the turbine, on the upper revolution range/power range without losses for dynamics in torque development.

Figure 11.37 Schematic representation of the electric support for exhaustdriven turbocharger.

eBooster compressor, the maximum speed is lower than with the euATL, which therefore provides benefits for the design of the eBooster. The wider compressor map of this two-stage controlled supercharging also offers the possibility of correspondingly raising the charge pressure and, hence, the torque of the internal combustion engine in the lower speed range, provided that sufficient electrical energy is available.

Figure 11.38b Schematic illustration of a supercharger system with mechanical auxiliary compressor.

11.6 Additional Measures for Supercharged Internal Combustion Engines 11.6.1 SI Engines

Figure 11.38a Schematic illustration of the eBooster supercharging system.

11.5.4 Mechanical Auxiliary Compressor

The combination of a mechanical drive unit and a suppression supercharger (Figure 11.38a) differentiates itself when compared to an electrically driven radial compressor in two fundamental points. The first point is that the mechanical linking to the crankshaft means that the start-up for the required compression power is not so critical and available without time limitations, secondly the suppression compressor demonstrates a significantly more economical supply characteristic. The possibility for an unlimited time period drive unit thereby additionally permits an optimization of

With the turbocharged SI engines, the higher charge pressure results in higher ultimate compression temperatures. This increases the risk of autoignition and of knocking. For this reason, it can be necessary to lower the compression ratio. In any case, the start of ignition of the SI engine must be shifted toward “retard,” in order to avoid impermissible high ignition pressures and knocking combustion (Figure 11.39). High exhaust gas recirculation rates increase the risk of knocking, particularly if an unfavorable exhaust pipe design exists upline of the turbine inlet. In part-load operation, the mass flow from turbocharged SI engines is bypassed by the compressor through the opened venting flap so that the mass flow, which is not required by the engine (part load), is returned upline of the compressor. As a result, no pressure is built up downline of the compressor. A recirculation air valve will be opened for engines that are supercharged with exhaust-driven turbochargers to prevent damage to compressor pumps when rapidly closing the throttle valve. The high-heat resistant turbine materials that are standardized today (NiCr-Steels) create the possibilities for exhaust gas temperature T3 of up to 950°C. It is also possible today to utilize turbine materials that can withstand temperatures of up to 1050°C. The strength and rigidity of the utilized

Internal Combustion Engine Handbook | 501

6606_Book.indb 501

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

70 60

Pressure [bar]

50 40 30 20 10 0 –90

–45

0

45

Crankshaft angle [°kW ]

90

135

Figure 11.39 Pressure curve of a turbocharged SI engine with retarded ignition angle.

materials can, however, weaken drastically above the permissible temperatures. As supercharged gasoline engines at full load can exceed the entered limiting values for exhaust temperatures, an active limitation is required. This is achieved by enriching the fresh charge via the engine control device.

11.6.2 Diesel Engines

In diesel engines, the high-charge pressure also results in very high ultimate compression pressures with compression ratios of ε > 14. Depending on the mechanical strength, the start of injection must therefore be set very late in diesel engines so that, under certain circumstances, the compression pressure can be equal to or higher than the ignition pressure. With medium-speed diesel engines, high-charge pressures are used in conjunction with large valve overlaps (of up to 120° on the crankshaft angle) also to reduce the thermal load on the engine. The medium-speed engine is operated with high excess air factors (λ ~= 2). With turbocharged diesel engines, the external exhaust gas recirculation demands additional measures in the form of a clocked regulation valve and software to control the charge pressure and the exhaust-gas recirculation rate. Measures must also be taken to ensure that a negative scavenging gradient (high pressure: p2 − p3 < 0; Reduced pressure p 0 − p4 < 0) is provided.

11.7 Performance Explosion by Register and Two-Stage Supercharging for Private Vehicles The last two decades of automobile development have continuously increased the demands for the introduction of sufficient power from internal combustion engines. This is due to the constantly increasing demands for comfort, although the

vehicles continuously become heavier. This resulting lost on driving dynamics that hereby occur, in this case “driving pleasure,” are however unacceptable. The opposite appears to be the case in that the driving dynamics may not be retained at the existing levels, rather more they must be actually improved. A component that has a major part in this development in the area of private vehicles is, and will always be, the turbocharger. While diesel engines without turbochargers are a rarity today, turbochargers for gasoline engine concepts are continuing to experience increasing demand. The specific performance of diesel engines has almost quadrupled within the last two decades due to the increased implementation of turbochargers. The utilization in gasoline engines can be set at almost double (Figure 11.40). As single-staged turbocharger concepts for both diesel and gasoline engines have now almost reached their limitations, despite high levels of development, the sequential turbocharger and/or two-stage turbocharger are experiencing a boom in attention for the area of private car drive units. This idea is not fundamentally new, however it has only been primarily implemented to date in static engines, marine applications, and commercial vehicle drive units.

11.7.1 The History and Evolution of Two-Staged Turbocharger Processes

The development of turbochargers for internal combustion engines has been characterized by an interesting past. MAN was a pioneer for the first application of a mechanical turbocharger system for increasing power (30%) in ship diesel engines in their works in Augsburg. The exhaust gas turbocharging, which was patented by Alfred Büchi (1905), was initially implemented in larger engines due to its technical production tolerances. Until 1938 before exhaust-driven turbochargers were able to be constructed to such small dimensions that they could be implemented for serial production HGVs. After the Second World War, mechanical supercharging established its role in motor racing, and could also play a major role in other racing areas (e.g., Drag Racing) and some other serial applications in the market place. The development of exhaust gas turbocharging was constantly confronted with poor response behavior and increased susceptibility to malfunctions and faults in its early years. It took as long as 1978 before Mercedes-Benz managed to effectively utilize exhaust-gas turbochargers for diesel engines in commercial vehicles. The increase in development budgets, and the efforts introduced by Porsche and BMW, resulted in the first serial model in 1973 (BMW 2002 Turbo) and the first supercharging concept with sequential turbocharger for gasoline engines. The Study 917 CanAm (1971/1972), which surprised the technical experts with previously unknown performance values (5.4 l, 1200 PS), formed the basis for a two-staged turbocharger system for gasoline engines with a phenomenal development of the performance density (Figure 11.41). Based on the sixcylinder boxer from the 911, Porsche developed a sequential turbocharger for the 959. Excellent transient behavior and

502 | Internal Combustion Engine Handbook

6606_Book.indb 502

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.7 Performance Explosion by Register and Two-Stage Supercharging for Private Vehicles

120 110 100

Specific engine power in KW/I

90 80 70

Year 1973

Combustion Application process

spec . performance

Gasoline

First automobile with ATL (BMW 2002 Turbo)

1985

Gasoline

Mech. Additional charge (Lancia Delta S4).

63 kW/I

1989

Diesel

First turbo diesel (Audi 2.5TDI 100)

36 kW/I

1994

Wankel

Wankel with turbo charging (Mazda RX-7)

2003

Diesel

Two-stage turbo charging (Opel OPC)

*1 82 kW/I

2004

Diesel

Two-stage turbo charging (BMW 535d)

67 kW/I

104 kW/I

*1 206 kW/chamber volume 2*654 cc

to

Ot

b

tur

es

gin

n oe

60 50 40

Natu

30

soline

ted ga

spira rally a

s

engine

Naturally as

20

el engines

pirated dies

urbo

el t Dies

es

engin

10 0 1930

1940

1950

1960

1970

1980

1990

2000

year

2010

Figure 11.40 The evolution of the specific performance from gasoline and diesel engines.

start-up torque quickly made the sequential turbocharger a favorite. The application of two charges in this style remained unique until 1994 and was then modified into a bi-turbo concept. The sequential turbocharger was then implemented into the so-called “Mazda Sequential Twin Turbo System” for Mazda’s Wankel Engine Sports Car RX-7 in 1994 with an all-electronic regulated exhaust-driven turbocharger. Subsequent variations from other manufacturers like Fiat with their experimental system “ECV,” or the Subaru Impreza 2.0 l-Boxer with 206 kW, continued to develop the sequential turbocharger for gasoline engines further. However, combinations with mechanical supercharging still continued to find applications in motor racing, for example, Lancia Delta S4 from 1985 (Figure 11.42).

It was not until 1996 that Volvo introduced a compressor turbomotor for HGVs (5.5 I, 184 kW). The two-staged supercharger process has also been utilized in recent times for diesel engines in private vehicles. So OPEL 2003 presented the two-staged concept in the “Vectra OPC,” which achieved a liter performance of 82 kW/I with the 1.9 l-TDCI and 156 kW. BMW had already implemented the first serial application in the 535d by the end of 2004 (Figure 11.43). The 6-Cylinder with 3.0 l cubic capacity delivers 200 kW due to its “variable-twin-turbo“ systems. That computes to a liter performance of 67 kW/l. The diesel drive unit provides the perfect basis for the high-charge level for the two-staged supercharger, although gasoline engines with their higher utilizable revolution range can also benefit enormously from the characteristics from sequential turbochargers.

11.7.2 Thermodynamics of Two-Staged Superchargers

The fundamental principle of two-staged supercharger systems is increasing the provided charging pressure as a result of sequential compression. Multiplying the separate pressure ratios by two- or multiple-staged compression enables significantly larger pressure ratios over a wider operating range than with single-stage superchargers.

Figure 11.41 Porsche 917 CanAm-Motor, 1200PS, n = 7800 min–1 [11-12]

11.7.2.1 Compression Efficiency of Two-Staged Superchargers If two-staged compression is considered to be a purely thermodynamic, then the following will be evident. If the same

Internal Combustion Engine Handbook | 503

6606_Book.indb 503

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

Figure 11.42 Lancia Delta S4, Roots/ATL [11-12].

more benefits than single-stage systems when considering the isentropic compression efficiency. If the intermediate cooling is disregarded, then the isentropic compression efficiency reduces in cases of sequential compression compared to the single-stage supercharging. But the increased achievable pressure ratios and the pronounced broadening of the operative range still remain available.

T

Single-stage charging p2

Two-stage turbo charging

2one

2two

2s

p1

Figure 11.43 BMW, two-staged supercharger 535d [BMW AG].

charging pressure is to be achieved, and one compares the one-stage with the two-stage supercharger including intermediate cooling, then the latter is the more effective method for compression as the isentropic compression efficiency is higher (Figure 11.44). Moreover, the model representation has more derivations which, in cases of joint switching with innumerous charging stages with the relevant intermediate cooling, creates a theoretical isothermal compression. hsCompressor, Single-stage = =

T2, s − T1 T2,Single-stage − T1

< hs Compressor, Two-staged

T2, s − T1 T2,Two-staged − T1

This ratio is independent of whether the consecutive subsequent polytropic compressions result from a flow charge or a mechanical supercharging unit, for example, a roots blower. In this respect, all combinations are differential, although twostage supercharger systems utilizing intermediate cooling have

p2

T

2s

pbetween 2one

p1

1z,s 1

s

s

Figure 11.44 Single-stage versus two-stage (with intermediate cooling) compression in T-s-diagram.

The previously described advantages of the two-staged/ multiple-staged supercharging process for stationary operating behavior reverse when considering the transient behavior of two-stage exhaust gas turbocharging. If the exhaust gas side is considered here, then it is evident that the classical two-stage supercharger as a typically slower ramp-up time and/or charge pressure buildup, as a result of the distribution of the exhaust gas enthalpy stream on both constantly driven turbines. The transient behavior can still be considerably improved by utilizing a special combination and/or switching of the individual stages.

504 | Internal Combustion Engine Handbook

6606_Book.indb 504

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.7 Performance Explosion by Register and Two-Stage Supercharging for Private Vehicles

11.7.3 Sequential Turbocharger and Two-Stage Supercharger Concept/Supercharger Systems

The combination possibilities for single-stage supercharging units, which have established themselves during the development process for superchargers and/or are auspicious with regard to their illustrated characteristics for higher specific performance with increased efficiency will be listed subsequently. 11.7.3.1 Sequential Turbocharger Sequential turbochargers are considered to be a combination of two exhaust-driven turbochargers, in which a turbocharger is completely switched off in the lower revolution range. In this case, the complete exhaust gas enthalpy flow is directed over the second exhaust-driven turbocharger. This therefore guarantees increased representable mean pressures, as well as improved response behavior with reduced flow rates. The second exhaust-driven turbocharger can be subsequently switched on with increasing charge pressure requirements. In principle, single-stage and multiple-stage sequential turbochargers can

480

360 kW 340

440

320

PS

320 280 240 200 160 120 80 40

•• Package requirements are controllable.  isadvantages D •• Increase in fuel consumption when operating both stages •• Increasing the nominal power is somewhat limited with single-stage sequential turbochargers. The sequential turbocharger creates an effective possibility for improving the response behavior for all bi-turbo engines, therefore combustion engines based on two-bed concepts (Figure 11.45). A sharp expansion of the characteristic diagram to very high nominal powers is therefore not achievable. The realistic representable charge pressures are therefore over 3 bar.

Nm 500

280

Torque

360

 dvantages A •• Significant improvement in the response behavior compared to bi-turbo engines (Two-bed concept)

300

Output ( kW /PS)

400

be differentiated. The multiple-stage sequential turbocharger differs from the single-stage sequential turbocharger by the number of jointly switched exhaust-driven turbochargers within a group.

260 240

400

220 200

300

180 160

200

140 120 100 80 60 40 20 0 1000

2000

3000

4000

5000

6000

7000

8000

Engine speed (rpm) Power and torque with Porsche turbo charging Power and torque with conventional dual boot in parallel operation Power and torque with loader turbo charging Increase in power and torque with Porsche turbo charging for conventional double charging in parallel operation

Figure 11.45 Sequential turbocharger versus two-bed turbocharger in the Porsche 959 [11-12].

Internal Combustion Engine Handbook | 505

6606_Book.indb 505

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

11.7.3.2 Two-Stage Turbocharger In order to achieve an expansion of the characteristic diagram range to larger air throughputs, a combination from two charges with differential throughput ranges can be utilized. The most common processes can be divided in combinations of electrical, mechanical, and pure exhaust gas turbine-driven turbocharger units. 11.7.3.2.1 Electrical compressor/Exhaustdriven turbocharger The combination of an electrical compressor and a combustion engine is almost as old as the history of the turbocharger itself. The first flow compressor for ship drives and transient engines was powered by high-performance electrical machines. This idea was taken up, and with today’s technically feasible components, was implemented for private vehicle applications in two variants. The electrically driven flow compressor as an auxiliary turbocharger, which is arranged independently from the exhaust-driven turbocharger in the air pathway of the combustion engine, represents the auspicious variant for an electrical supercharger system [11-13]. The electrical compressor is mostly located close to the intake side for quicker charge pressure build up. The combination with an exhaust-driven turbocharger means that the electrical auxiliary compressor only has to be optimized for operating with smaller throughputs. Significant nominal power increases are achievable with the expanded design of the exhaust-driven turbocharger to higher throughput values, whereby the response behaviors are not worse when compared to the output variant.  dvantages A •• Expanding the characteristic diagram area by utilizing additional flow compressors •• Excellent response behavior compared to single-stage turbocharger systems •• Selecting the desired positioning in the air pathway (simplified package).  isadvantages D •• A bypass is necessary as only smaller throughput areas of the electrical compressor will be covered •• Higher electrical performances are required for the electrical compressor (Board network loading) •• Regeneration is not possible •• High costs due to the electrical components. An electrical auxiliary compressor makes higher demands on the board network with regard to the electrical service provision (input current of up to approximately 200 Ampere, depending on the voltage of the board network). The electrical performance balance is dependent on the driving style and the energy storage must therefore be designed for a reliable functionality. This is the main reason why the electrical systems have not managed to establish themselves to date despite intensive development.

11.7.3.2.2 Mechanical Chargers/Exhaustdriven turbochargers The combination of a mechanical supercharger with an exhaust-driven turbocharger contributes represent a significant broadening of the compressor-side operating range. Displayable boost pressure values are also in the range of about 3.0 bar, depending on the choice of the mechanical supercharger. The direct connection of the mechanical supercharger to the engine, which is designed by an appropriate translation for low speeds, the transient response is compared with an exhaust gas turbo-driven design improves. The direct connection of the mechanical supercharger to the engine, which is designed by on appropriate translation for low speeds, the transient response is compared with an combustion engine design improves. If a consumption concepts are shown, the mechanical supercharger may only be used in acceleration phases.  dvantages A •• High potential boost pressure at low engine speeds •• Very good response at low speeds •• Increasing the maximum pressure ratio by connecting the compressor •• Expansion of the map area to large throughputs (power rating).  isadvantages D •• Noise potential is very high •• Huge driving power necessary when relative boost pressures must be shown as greater than 1 bar with a mechanical supercharger •• High friction losses result in high fuel consumption during operation of the mechanical supercharger •• Costs compared to single-stage systems are significantly higher. There can be a significant improvement in the transient behavior with an appropriate design, similar to that of naturally aspirated represent. The represented total operating range charge pressure increases depend on the design. Two significant disadvantages are the shift shock of the engagement of the magnetic coupling at the request of the mechanical supercharger and the still complex reduction of the noise of mechanical charging. 11.7.3.2.3 Two-stage exhaust gas turbocharging With a two-stage regulated charging, the big drawbacks to simple two-stage exhaust gas turbocharger can be significantly improved, and the sequential charging partially improved. To this end, two exhaust-driven turbochargers are connected together, so that either one of the two stages alone, or both can work together. An actuator distributes the exhaust gas mass flow onto the two turbines (Figure 11.46). The design of each exhaust-driven turbocharger is carried out for different throughput ranges.

506 | Internal Combustion Engine Handbook

6606_Book.indb 506

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs

11.7.4 Applications Areas EGR-valves

Engine LLK Switching flap

Bypass

ZK Large ATL

Small ATL 2-stage turbo charging

Figure 11.46 Block diagram of the two-stage sequential exhaust gas turbocharging.

The compressors are connected in series and the highpressure turbocharger is provided with a bypass. Depending on the application case, the usable performance range can be adapted to the requirements [11-14]. Between a high starting torque and consequent rated power, the variants are all possible. In addition, emission concepts with special design for high exhaust gas recirculation rates are also possible for diesel engines. Even with this method, charge pressure values from over 3 bar are representable depending on the design of both stages [11-15]. The current best compromise from fuel economy savings potential and maximum representable boost pressure is represented by the two-stage exhaust gas turbocharging [11-16]. A special feature of engines with two-stage exhaust gas turbocharging is the process for the specific fuel consumption, which is characterized throughout the full load range with two local minimums. This results from the respective efficiency maximums of both singularly driven exhaust-driven turbochargers. In the area where both chargers are supplied together, the fuel consumption increases slightly, which is partly due to the fluid-dynamic losses in the division upstream of the two turbines.  dvantages A •• Expansion of the characteristic map area up to large throughputs (power rating) •• Shifting the pump limit to smaller throughputs •• Increasing the maximum pressure ratio by row switching •• Reducing the inertia when operating with smaller turbocharger stages.  isadvantages D •• Extra space requirements •• Complex routing of flue and air ducting •• Very large space requirements where intercooling is required •• Costs compared to single-stage systems are higher.

A further increase in the nominal capacity or extension of the active map region, and improving the response behavior, can be only be achieved in modern private vehicle applications by charging using the two-stage turbocharger systems. Two-stage turbocharger systems enable an advance to completely new nominal capacity areas for diesel engines, while maintaining the diesel-typical response behavior, or alternatively, a further improvement in the response behavior while maintaining the now achievable variable turbine power ratings. Specific power of approximately 100 kW/l also seem possible for diesel engines as well as specific torques of 200–250 Nm/l. Corresponding values of approximately 25–30 bar of the mean effective pressure can be represented in the foreseeable future. For gasoline engines, the two-stage turbocharger opens further scope for representing downsizing concepts for reducing fuel consumption or performance concepts. With two-stage systems, it seems possible that due to the large spread of the airflow rate of the gasoline engine, the problematic response behavior can be improved considerably and, therefore, provide an increase in the acceptance of these concepts. Should this succeed, fuel consumption savings of about 15–20% from downsizing concepts in comparison to the same power from naturally aspirated engines could be a realistic aim by utilizing use two-stage turbocharger systems. Further potential for reducing fuel consumption through downsizing with two-stage supercharging concepts could be implemented with a combination with variable compression in gasoline engines [11-17].

11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs Turbocharger test rigs have been mainly used for a long time by turbocharger manufacturers and a few universities as a research and development tool. Several factors led to the fact that turbocharger test rigs are now also utilized by engine developers. Among the decisive factors for this increasing use of turbocharger test rigs include the following: 1. The increasing use of exhaust gas turbocharging as a technology for reducing fuel consumption 2. The increasing shift of the development work on the entire charging system from the turbocharger manufacturer to the engine developers 3. The increasing use of simulation and model-based control and, consequently, the increased need for comparable and other areas for measured turbocharger maps. The stationary determined maps on turbocharger test rigs from the compressor and turbine provide information about the performance of an exhaust-driven turbocharger, and are often used as input parameters for the engine process simulation or for the so-called “matching” of exhaust-driven turbocharger and

Internal Combustion Engine Handbook | 507

6606_Book.indb 507

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

engine. To determine the maps, an exhaust-driven turbocharger must be built and measured on the turbocharger test rig. A general view of the measuring methodology on a turbocharger test rig is described in SAE J1826. Information on possible impact factors are illustrated in [11-28], [11-29]. Moreover, as part of the framework of the VFI Project “TC-mapping,” detailed recommendations were drawn up, such as for the optimal measuring points position, for the measuring devices to be utilized, for measuring points execution, for the test setup and operation and for the information supplied to a map. The roles for a turbocharger test rig can be very diverse, but consist principally of being able to operate a turbocharger independently of an engine in the widest possible performance range. Usually, the turbocharger is then operated stationary, but some applications also require a transient operation. For both stationary, and for the transient operation, the engine environment must be simulated. To do this, •• hot gas must be generated and supplied to the turbine •• the consumer linked downstream of the compressor, in principle an engine with its intake control organs modeled in the form of an adjustable throttle •• the oil supply for the hydrodynamic plain bearings are provided •• for water-cooled turbochargers, the cooling water supply is to be provided as required.

11.8.1 Basic Design of a Turbocharger Test Rig

A turbocharger test rig consists of the following main components: Hot gas generator, throttle upstream from the compressor, oil supply, cooling water where applicable, control and regulating electronic systems and the measuring technology. The arrangement of the individual components is illustrated in Figure 11.47 and will be clarified here briefly. 11.8.1.1 Hot Gas Generator The hot gas generator has the role of heating the gases, which are supplied to the turbine, to a constant increased temperature which is set between 150 and normally 1000°C, but which can

vary dependent on the design. The temperature increase is mainly executed by the combustion of fuel, such as natural gas, diesel fuel, or kerosene. Hot gas generators of this type are called combustion chambers and are utilized in many various applications [11-30], [11-31]. Electrical heating devices are primarily utilized for lower temperatures, that is. <400°C, due to the instable behavior of combustion chambersin these temperature ranges. The air required is prepared in an electrically driven compressor and then fed through the combustion chamber for increasing the temperature. The exhaust gas power that is prepared for the turbine can be printed out according to the following equation: ! A ⋅ hA = m ! A ⋅ c pA ⋅ (TA − T0 ) PA = m



(11.24)

11.8.1.2 Throttle Upstream from the Compressor The exhaust gas power will be converted into drive power for the compressor, with an applicable efficiency which is set by the operating point of the turbine. Dependent on the subsequently switched consumer displacement behavior on the compressor, there is now a set constant turbocharger speed under transient conditions. The subsequent switched consumer on the compressor is usually an electrically adjustable valve. The throughput on the compressor and, therefore conversely, also the turbocharger speed can be set depending on the valve setting (Figure 11.48). 11.8.1.3 Oil Supply Hydrodynamic lubricated plain bearings are generally utilized today for the bearings of the turbocharger internals as they are more cost effective. Lubricating oil is utilized for lubricating the plain bearings. An oil circuit, which is constructed for this purpose, consists of at least an oil filter, an oil pump, an oil conditioner and the turbocharger bearing housing. The oil conditioner can cool the oil, or heat it, as necessary.

11.8.2 Compressor and Turbine Maps

An important area of application for a turbocharger test rig is the measurement of compressor and turbine maps. In principle,

Turbocharger turbine Combustion chamber

Fuel

4 Throttle

3 2

Turbocharger compressor

1 Oil supply

Electrically driven compressor

Combustio chamber air

Figure 11.47 Schematic design of a turbocharger test rig. See color section page 1084.

508 | Internal Combustion Engine Handbook

6606_Book.indb 508

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs

Quenching behavior in different valve positions ΠV

As a parameter for the pressure increase in the compressor, the total pressure ratio is often used.

Closing the throttle valve nTL

Opening the throttle valve PV=PT=const.

the following assumptions are made for determining the characteristic field relevant variables:

u2max = 520 m/s

and

u2min = 150 m/s ⋅ (≈ 0.28 ⋅ u2max)

(11.26)

The throughput of a compressor can be specified as volumetric flow or mass flow, in both cases, a normalized representation is chosen. For standardization, there are pressure and temperature, among other things, the following values pair variants:

2. The compressor and/or the turbine efficiency can reach at an isentropic compression and/or relaxation process the value 1.



1 pt = ps + pd = ps + rc 2 2



1. Compressor and turbine are adiabatic machines

11.8.2.1 Compressor Map In principle, the increase in pressure is applied to the flow in a compressor map for constant turbocharger speeds. This so-called speed characteristic, hereinafter referred to as characteristic curve, is limited to high throughputs by the resistance of the pipeline downstream of the compressor pipeline system, and to low throughputs by the so-called pump limit. The pumping of the compressor is an effect that forms in particular at relatively high-pressure conditions, and at the same time low mass flows. During a pumping cycle, there is initially an increased release of the flow from the blades, whereby the mass flow continues to decrease until reverse flow occurs through the compressor. The return flow results in an immediate pressure drop downstream of the compressor, as the flow can be applied on the blades again and the normal supply is taken up again. If the amount of the consumer demand remains unchanged as low, then the pressure downstream of the compressor increases rapidly again and the pumping cycle repeats itself [11-32]. The position of the pump limit is, on the one hand, dependent on the behavior of the turbocharger compressor, on the other hand on the geometry of the inlet and outlet lines [11-34], [11-35]. Down to low speeds, the compressor map will be limited by the combustion chamber operating range among others, up to high speeds the compressor map is limited by the maximum permissible rigidity of the rotor assembly and, consequently, by the maximum permissible turbocharger speed. Conventional peripheral speed limits at the compressor outlet for characteristic curves illustrated in manufacturer’s maps are

(11.25)

Here, the total pressure can be a measurement value or an computed value. In the latter case, use is made of the Bernoulli equation to calculate the dynamic proportion of the pressure from the measured compressor mass flow, air density and the flow cross section at the static pressure measuring point:

. m VN Figure 11.48 Schematic illustration of a compressor mapwith combustion displacement lines and a line-constant combustion power.

p2t p1t

ΠV =



TN

pN

Application with

in K

in mbar

–

273.15

1,013.25

DIN 1342, Sulzer

298

1,000

SAE J 1826, SAE J922, BorgWarner Turbosystems

293

981

BorgWarner Turbosystems

The equations for standardizing the measured dimensions are as follows:



! VN = m !V Standardized Mass Flow m

T1t pN ⋅ TN p1t

(11.27)

T Standardized Volumetric Flow V!VN = V!V 1t (11.28) TN

Even with the registered in a compressor map turbocharger speed values is normalized values. Hereby one uses a similarity measure of the fluid dynamics, the Mach number: u1 uN = a1 aN u1 uN ⇒ = kRT1 kRTN

Ma1 = Ma N ⇒

⇒



u1 u = N T1 TN

⇒ uN = u1



(11.29)

TN T1

and with

n=

u 2p ⋅ r

nN = n1

TN T1

(11.30)

(11.31)

The turbocharger speed is usually determined with a sensor whose operation is based on the eddy current measurement

Internal Combustion Engine Handbook | 509

6606_Book.indb 509

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

principle. The sensor is arranged in such as way that all the blades of the impeller pass the sensor. When calculating the compressor efficiency,the ratio is formed between the enthalpy differential ∆hsV for an isentropic compression and the enthalpy differential ∆hV for a real compression (Figure 11.49). Thus, the calculation of the compressor efficiency is executed according to the following equation: hsV =



T2 s − T1 T2 − T1

(11.32)

with k L −1



T2 s = T1 ⋅Π vk L

(11.33)

∆h = c p ∆T

(11.34)

and

Maximum compressor efficiency values are in the range 70–80%. 11.8.2.2 Turbine Map In a turbine map, the throughput is plotted against the pressure drop for the respective constant turbocharger speeds, illustrated in Figure 11.50. A typical size for the throughput is the reduced turbine mass flow. Taking into account the conditions on the turbine, the reduced turbine mass flow is calculated as follows: ! T red m



! ⋅ T3 m = T p3t

(11.35)

For the pressure drop, the turbine pressure ratio is used, here the quotient of the total pressure and static pressure upstream of the turbine and downstream of the turbine: ΠT =



Surge limit



1 2 c2 2

Resistance characteristic curve

ηsV

∆hsV



hsT ⋅ hm =

nTL



(11.37)

hTL hsV

(11.38)

⎛ k L −1 ⎞ !V c pL ⋅T1 ⋅ ⎜ ΠVk L − 1⎟ ⋅ m ⎠ ⎝ = k A −1 ⎛ p4s k A ⎞ ! ⋅h c pA ⋅T3 ⋅ ⎜ 1 − ⎟ ⋅m ⎜⎝ ⎟⎠ T sV p3t

(11.39)

p2t p2

2

∆hV p1t

. m VN

1t

Figure 11.49 Compressor map with contour lines for compressor efficiency, pumping limit and resistance characteristic (left), isentropic and real compression of the air, illustrated in h-s diagram (right).

p1

1 s

p3t

. mTred

h

p3 3t 3

nTL

! V ⋅ ∆hsV m ! T ⋅ ∆hsT ⋅ hsV m

2s 1 2 c1 2

hTL = hsV ⋅ hsT ⋅ hm ⇒ hsT ⋅ hm =

1 2 c2 2

h

(11.36)

The total pressure upstream of the turbine can be calculated according to Eq. (11.26). The temperature upstream of the turbine is adjusted, depending on the application range of the turbocharger, to a constant value of, for example, 600°C for diesel engines and 950°C for gasoline engine turbochargers, a uniform regulation for this is still not available. In particular, temperature stratification in the measuring pipe downstream of the turbine does not allow for the application of the conventional method in the compressor side for calculating the isentropic efficiency in accordance with Eq. (11.32). The turbine efficiency is therefore determined from the turbocharger efficiency, whereby the mechanical losses from the entire turbocharger of the turbine must be attributed.

2t ΠV

p3t p4s

ηsT ◊ηm

1 2 c3 2

∆hT

∆hsT 4

p4

4s ΠT

s

Figure 11.50 Turbine map with speed and efficiency lines (left), isentropic and real relaxation of the exhaust gas represented in h-s diagram (right)

510 | Internal Combustion Engine Handbook

6606_Book.indb 510

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs

Maximum turbine efficiency values are in the range 60–70%. The turbine efficiency is applied for constant turbocharger speeds via the turbine pressure ratio Π T (Gl. 11.36), or so-called tip speed ratio. The tip speed ratio is the ratio of peripheral speed u3 at the turbine inlet and the flow velocity c 0. The theoretically achievable flow velocity c 0 is achieved if the available isentropic enthalpy ∆hsT is converted into kinetic energy without any losses i.e.: (11.40)



1 2 c0 = ∆hsT 2



⇒ c0 = 2 ⋅ ∆hsT

(11.41)

This allows the tip speed ratio to be calculated as follows: 1

u3 u3 = = c0 2 ⋅ ∆hsT =



2 ⋅ ∆hsT u32



1 k A −1 4s k A

⎛ p 2 ⋅ c pA ⋅T3 ⋅ ⎜ 1 − ⎜⎝ p3t u32

(11.42)

⎞ ⎟ ⎟⎠

The next chapter describes the special features that must be considered when using turbocharger engine maps in the engine process simulation.

11.8.3 Special Features When Using Turbocharger Engine Maps in the Engine Process Simulation

The compressor and/or the turbine are modeled as points in the piping system of an engine, in which a pressure increase or a pressure reduction takes place. Furthermore, the compressor and turbine in the process simulation engine usually have no spatial expansion; therefore, they can neither be considered for gas dynamic effects nor heat transfers that may occur, for those that occur in an acceleration process in a turbocharger. The operational behavior of a turbocharger is therefore represented

only by the stationary measured characteristic fields from the test rig and the thermal transfers which occur there. 11.8.3.1 Map Range for Smaller Turbocharger Speeds Section 11.8.2 illustrates the operating range shown in manufacturer’s maps. Especially with smaller turbocharger speeds (<0.3 * η TLmax), it has been found that heat transfer effects from the turbocharger efficiency can assume relatively large impact and that the assumed conclusion reached in Section 11.8.2 that the compressor and the turbine are adiabatic machines, with declining turbocharger speed occurs less often [11-33], [11-36]. Indicators for this are, for example, turbine efficiency, where the values are larger than one. In [11-37], a special test is described which makes it possible to determine turbine efficiencies for small turbocharger speeds. The turbine power is not determined in the conventional way by means of the applied compressor capacity, rather the torque on the shaft of the electrically braked turbine is measured. Furthermore, the so-called burn-through speed will be ascertained, one for the extrapolation of useful value on the x-axis in the turbine efficiency map (u/c0-representation). Numerical methods can thus assist to extrapolate measured maps meaningfully in the region of small turbocharger speeds. Relevant information is provided in [11-33], [11-34] among others. 11.8.3.2 Pump Limit and Resistance Characteristic Curve As explained already in 11.8.2, a characteristic line of a compressor is limited by the pump limit and the resistance curve (Figure 11.51). The slope of a characteristic curve is based on the resistance characteristic curve up to low throughputs negatively, and can also accept a zero value and/or also values greater than zero to the pump limit. Problematical is such a curvature of the engine process simulation in as much as there are two different throughflow rates for a turbocharger speed and at a pressure ratio of Π V1 (Figure 11.51). A few years ago, it was therefore common to modify the characteristic curves so that their slope in the entire region was negative. Nowadays, the customary engine process simulation programs can also process characteristic curves with curvatures that have values greater than zero [11-38].

Figure 11.51 Theoretical and effective process of the characteristic curve of a centrifugal compressor.

Internal Combustion Engine Handbook | 511

6606_Book.indb 511

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 11 Supercharging of Internal Combustion Engines

It must be assumed that the losses which occur in the compressor will increase for the area below the resistance characteristic curve. If one wants to ascertain the point on the x-axis for a characteristic curve, then the resistance should be minimized downstream of the compressor, ideally to zero. For this purpose, the compressor air is not fed to a throttle as illustrated Figure 11.47, rather it is blown into the environment. The thereby measured value for the compressor mass flow is then the searched for point on the x-axis. 11.8.3.3 Quadrant In acceleration processes of supercharged engines, there could be a formation of inverse pressure conditions on the compressor. For technical-measuring ascertainment of this operating range, the compressor inlet can be supplied with compressed air. With an appropriate amount of air, and at constant turbocharger speed, there will therefore be a higher pressure existing on the compressor inlet. 11.8.3.4 Pulsating Turbines Impingement A major difference between the operation of a turbocharger on a turbocharger test rig, and the utilization on an engine, is the presence on the engine of existing pulsating action on the turbine with exhaust gas as a result of the charge cycle. For the calculation of exhaust gas pulsations in the engine process simulation, this transient outflow is regarded as an inertia-free succession of short spells of stationary operating conditions, that is, at each calculation step the transient prevailing status dimensions will be transferred to the turbine in order to read out the stationary measured mapped data. This modeling style is insufficient for certain application cases and may cause relatively large deviations from experimentally obtained results. Therefore, the pulse charging of the turbocharger turbine has been examined for the last 20 years on the turbocharger test rig [11-39], [11-40], [11-41]. In [11-42], a cylinder head has been added in addition to the standard design, as illustrated schematically in Figure 11.47, between the combustion chamber and the turbine in such an arrangement so that actually the combustion chamber of an engine-facing side of the cylinder head was provided by the combustion chamber with the exhaust gas. The camshaft of the exhaust valves was driven by a variable speed electric motor, so that the pulse rate could be set free. This enables the formation of realistic pressure pulses to be simulated at engine-like temperatures.

Bibliography

11-1. EATON Corporation, USA. 11-2. IHI Corporation, Japan. 11-3. SIG Schweiz-Industrie-Gesellschaft

11-7. MAN Diesel SE, Augsburg. 11-8. MTU Friedrichshafen GmbH. 11-9. Leweux, J., Schommers, J., Betz, T., Huter, J., Jutz, B., Knauel, P., Renner, G., Sass, H. 2008. Der neue 4-Zylinder Pkw-Dieselmotor von Mercedes-Benz für weltweiten Einsatz. 17. Aachener Kolloquium Fahrzeugund Motorentechnik. 11-10. Hoecker, P., Jaisle, JW., Münz, S. 2001. Der eBooster—Schlüsselkomponente eines neuen Aufladesystems von BorgWarner Turbo Systems für Personenkraftwagen. 22. Internationales Wiener Motorensymposium. 11-11. Zinner, K. 1980. Aufladung von Verbrennungsmotoren. Berlin; Heidelberg; New York: Springer-Verlag. 11-12. Hack, G., Langkabel. 2003. Turbo- und Kompressormotoren; Entwicklung und Technik; 3. Auflage; Motorbuch Verlag. 11-13. Sens, M., Offer, T., Bals, R., Kahrstedt, J. 2003. Auslegung der Aufladegruppe eines hochaufgeladenen Ottomotors durch den Einsatz der Motorprozesssimulation; 6. Tagung Motorische Verbrennung; Haus der Technik e.V.; München. 11-14. Tomm, U., Schmitt, F. Optimierung von Hoch- und Niederdruckverdichter für die 2-stufig geregelte Aufladung (R2STM). 9. Aufladetechnische Konferenz 23./24. September 2004, Dresden. TU Dresden, Lehrstuhl Verbrennungsmotoren (Veranst.). 11-15. Schorn, N. et al. Die Aufladung zukünftiger Pkw-Dieselmotoren. Welche Systeme ergänzen die VNT Technologie? 9. Aufladetechnische Konferenz 23./24. September 2004, Dresden. TU Dresden, Lehrstuhl Verbrennungsmotoren (Veranst.). 11-16. Sens, M., Sauerstein, R., Dingel, O., Kahrstedt, J. Möglichkeiten und Besonderheiten bei der Darstellung eines Downsizing-Konzeptes Benzindirekteinspritzung. 9. Aufladetechnische Konferenz 23./24. September 2004, Dresden. TU Dresden, Lehrstuhl Verbrennungsmotoren (Veranst.). 11-17. Drangel, H., Bergsten, L. 2000. Der neue SAAB SVC Motor—Ein Zusammenspiel zur Verbrauchsreduzierung von variabler Verdichtung, Hochaufladung und Downsizing; 9. Aachener Kollo-quium Fahrzeug- und Motorentechnik; Aachen. 11-18. Lehmann, HG. et al. Potenzialbetrachtung zum Instationärverhalten von hochaufgeladenen Ottomotoren. 9. Aufladetechnische Konferenz 23./24. September 2004, Dresden. TU Dresden, Lehrstuhl Verbrennungsmotoren (Veranst.). 11-19. Müller, W. et al 2011. High sophisticated boost-pressure control: the MAHLE electrical waste gate actuator. 11th Stuttgart International Symposium, Automotive and Engine Technology. 11-20. Sauerstein, R. et al. 2010. Die geregelte zweistufige Abgasturboaufladung am Ottomotor—Auslegung, Regelung und Betriebsverhalten. 31. Internationales Wiener Motorensymposium, 29.–30. 11.21. Sauerstein, R. et al. 2010. Downsizing und Hochaufladung: Herausforderungen und Lösungen. TAE, 9. Symposium Ottomotorentechnik, 2. und 3. 11.22. Schernus, C. et al. 2010. Aufladekonzepte für kleine Ottomotoren. TAE, 9. Symposium Ottomotorentechnik, 2. und 3. 11.23. Schmitt, S. et al. 2010. Vergleich von Direkteinspritzung und Saugrohreinspritzung am Otto-Turbo-Motor mit mechanisch vollvariablem Ventiltrieb. TAE, 9. Symposium Ottomotorentechnik, 2. und 3. 11-24. Schmuck-Soldan et al. 2011. S. 2-stufige Aufladung von Ottomotoren 32. Internationales Wiener Motorensymposium 5. und 6. 11-25. Steinparzer F. et al. 2011. Der neue BMW 3,0l 4-Zylinder Ottomotor mit Twin Power Turbo Technologie 32. Internationales Wiener Motorensymposium 5. und 6.

11-4. BorgWarner Turbo Systems GmbH, Kirchheimbolanden

11-26. Zinner, K. 1985. Aufladung von Verbrennungsmotoren; 3. Auflage; Berlin, Heidelberg, New York: Springer Verlag.

11-5. Eiser, A., Grabow, J., Königstedt, J., Werner, A. 1997. Moderne Aufladekonzepte der Turbomotoren von AUDI. 6. Aufladetechnische Konferenz.

11-27. Hiereth, H., Prenninger, P. 2003. Aufladung der Verbrennungskraftmaschine; 1. Auflage; Springer Wien New York.

11-6. Welter, A., Unger, H., Hoyer, U., Brüner, T., Kiefer, W. 2006. Der neue aufgeladene BMW Reihensechszylinder Ottomotor. 15. Aachener Kolloquium „Fahrzeug- und Motorentechnik.

11-28. Nickel, J., Sens, M., Grigoriadis, P., Pucher, H. 2005. Einfluss der Sensorik und der Messstellenanordnung bei der Kennfeldvermessung und im Fahrzeugeinsatz von Turboladern, 10. Aufladetechnische Konferenz. Dresden.

512 | Internal Combustion Engine Handbook

6606_Book.indb 512

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

11.8 Ascertaining Turbocharger CharacteristicMaps on Turbocharger Test Rigs

11-29. Lechmann, A. Simulation und Aufladung von Verbrennungsmotoren, Springer-Verlag, 2008, Berlin, Heidelberg, S. 195–210. 11-30. Pucher, H., Eggert, T., Schenk, B. 1997. Experimentelle Entwicklungswerkzeuge für Turbolader von Fahrzeugmotoren. 6. Aufladetechnische Konferenz, Dresden. 11-31. Forcada, I., Bolz, H., Mandel, M. 2003. Neue Prüfstandstechnik für die Entwicklung moderner Turbolader, MTZ Motortechnische Zeitschrift 64, Nr. 1, S. 38. 11-32. Pischinger, R.; Kraßnig, G.; Taučar, G.; Sams, Th. 1989. Thermodynamik der Verbrennungskraftmaschine. Springer, Berlin, Wien, New York. 11-33. Pucher, H., Berndt, R., Grigoriadis, P., Nickel, J., Abdelhamid, S., Hagelstein, D., Seume, J. 2003. Erweiterte Darstellung und Extrapolation von Turbolader-Kennfeldern als Randbedingung der Motorprozesssimulation. Abschlussbericht über Vorhaben Nr. 754, FVV, Frankfurt/M. 11-34. Grigoriadis, P. 2008. Experimentelle Erfassung und Simulation instationärer Verdichterphänomene bei Turboladern von Fahrzeugmotoren, Dissertation TU Berlin, Berlin.

11-40. Capobianco, M., Gambarotta, A. 1990. Unsteady flow performance of turbocharger radial turbines. 4th International Conference on turbocharger and turbochargers. IMechE Conference Publications, C405/017, S. 123–132. 11-41. Westin, F. 2002. Accuracy of turbocharged SI-engine simulations. Licentiate Thesis, Dept. of Combustion Engines at KTH Stockholm, Stockholm. 11-42. Grigoriadis, P., Nickel, J., Pucher, H. 2004. Experimentelle Untersuchungen instationärer Phänomene in Fahrzeug-Turboladern, 9. Aufladetechnische Konferenz, Dresden. 11-43. Koch, A., Claus, H., Frankenstein, D., Herfurth, R. 2010. Turboladerdesign für elektrische Waste-Gate-Betätigung minimiert Leckage, MTZ 10.2010, Wiesbaden. 11-44. Schmid, A., Grill, M., Berner, HJ., Bargende, M. 2010. Transiente Simulation mit Scavengine beim Turbo-Ottomotor, MTZ 11.2010, Wiesbaden. 11-45. Sailer, T., Bucher, S., Durst, B., Schwarz, C. 2011. Simulation des Verdichterverhaltens von Abgasturboladern, MTZ 01.2011, Wiesbaden.

11-35. Schorn, N., Kindl, H., Späder, U., Casey, MV. 2007. Der Turboladerverdichter als Randbedingung in der Ladungswechselrechnung, MTZ-Konferenz, Ladungswechsel im Verbrennungsmotor Stuttgart.

11-46 Adomeit, P., Sehr, A., Glück, S., Wedowski, S. 2010. Zweistufige Turboaufladung-Konzept für hochaufgeladene Ottomotoren, MTZ 05.2010, Wiesbaden.

11-36. Shaaban, S., Seume, J., Berndt, R., Pucher, H., Linnhoff, HJ. 2006. Part-load Performance Predicion of Turbocharged Engines. 8th IMechE Conference, London.

11-47. Schicker, J., Sievert, R., Fedelic, B., Klingelhöffer, H., Skrotzki, B. 2010. Versagensabschätzung thermomechanisch belasteter Heißteile in Turboladern, MTZ 06.2010, Wiesbaden.

11-37. Schorn, N., Smiljanowski, V., Späder, U., Stalman, R., Kindl, H. 2008. Turbocharger Turbines in Engine Cycle Simulation, 13. Aufladetechnische Konferenz, Dresden.

11-48. N.N: Ladungswechsel bei Turbomotoren—Titelthema, MTZ 11.2010, Wiesbaden, November 2010.

11-38. Linnhoff, HJ. 1989. Berechnung instationärer Ladungswechselvorgänge aufgeladener Otto- und Dieselmotoren, Abschlussbericht über Vorhaben Nr. 366, Forschungsvereingung Verbrennungskraftmaschinen, Frankfurt am Main. 11-39. Dale, A., Watson, N. 1986. Vaneless diffuser turbocharger turbine performance. 3rd International Conference on turbocharger and Turbochargers. IMechE Conference Publications, C110/86, S. 65–76.

11-49. Schernus, R. et.al. 2011. Turboaufladung für kleine Ottomotoren mit weniger als 4 Zylindern, 3rd IAV Conferenze Engine Process Simulation and Turbocharging, Berlin. 11-50. Ross, T., Zellbeck, H. 2010. Neues ATL-Konzept von VierzylinderOttomotoren, MTZ 12.2010, Wiesbaden. 11-51. Getzlaff, U., Hensel, S., Reichl, S. 2010. Simulation des Thermomanagements eines wassergekühlten Turboladers, MTZ 09.2010, Wiesbaden.0

Internal Combustion Engine Handbook | 513

6606_Book.indb 513

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 514

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12 Mixture Format Ion and Related Systems In combustion, from a chemical perspective, the oxidation of fuel molecules requires that the oxidating agent (oxygen) has sufficient access to the fuel molecule. It is, therefore, necessary to prepare the fuel, that is, transform it into a gaseous phase and mix it with air. This is normally done using mixture formation systems. A distinction is drawn between internal and external mixture formation during engine operation.

represent a large share of all engines produced, since their potential for reducing fuel consumption appears greater than with diesel engines with direct injection. Whereas conventionally operating diesel engines utilize an internal mixture formation that produces an inhomogeneous distribution of air and fuel within the cylinder, additional advantages may be achieved in the future with homogeneous diesel combustion that allows the reduction of emissions and fuel consumption.

12.1 Internal Mixture Formation Internal mixture formation takes place in the cylinder of an internal combustion engine. The air is inducted through the piston and compressed; the fuel is then injected into the compressed air at a suitable time. The air–fuel mixture achieves an ignitable composition within certain ranges that leads to the ignition of the mixture at a corresponding temperature. Pronounced in homogeneities arise with this type of mixture formation, and local air–fuel ratios of λ = 0 (pure fuel) to λ = ∞ (pure air) arise. A diffusion flame causes combustion. The utilized fuel must meet certain criteria to achieve ignition quality. The reaction occurs with prepared droplets, that is, droplets surrounded by an ignitable mixture. Diesel engines have provided a typical example of internal mixture formation up to now. Recently, there has been increasing development of spark-ignition engines that also utilize internal mixture formation, so-called spark-ignition engines with direct injection. The basic difference from the diesel engine is the use of gasoline and an external ignition source. In future, it can be expected that gasoline engines with internal mixture formation will

12.2 External Mixture Formation External mixture formation is characteristic in conventional spark-ignition engines. The air and fuel are mixed prior to entering the engine cylinder. Therefore, a more or less homogeneous mixture consisting of air and fuel vapor is generated. This method was found predominantly in engines that had a carburetor or single-point injection as the mixture forming mechanism. There was enough time available to mix air and fuel and then transport this mixture to the intake valve. The danger of these mixture formers was, however, that the fuel present in a vapor phase would condense on cold intake manifold walls, and the mixture would be unevenly distributed to the individual cylinders. The type of intake manifold injection used today eliminates these disadvantages as the fuel is injected directly before the intake valve and partially toward the open intake valve and into the cylinder. In this case, there is also sufficient time to homogenize the mixture over the intake and compression phases.

Internal Combustion Engine Handbook | 515

6606_Book.indb 515

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

12.3 Mixture Formation in Gasoline Engines (Carburetor/ Gasoline Injection) Apart from a few exceptions, mixture formation using a carburetor is no longer the preferred state-of-the-art approach in passenger car engines. Large numbers of carburetors are used only for varieties in certain countries and for two-wheeled vehicle drives. Therefore, this section will only concentrate on the basic details of mixture formation in relation to carburetors. The task of a carburetor is to provide the required amount of fuel to the inducted air for the desired mixture ratio, depending on the operating status of the engine. The throttle valve that governs the air and/or mixture flow is integrated into the carburetor. The energy required for metering the fuel, and conveying it within the carburetor, is taken from the air stream. The carburetor and the connected intake manifold, with its branches to the individual cylinders that distribute the mixture generated by the carburetor, are to be viewed as a functional unit. The operating behavior of the engine is greatly influenced by the precision with which the intake manifold was engineered to produce an even mixture distribution under all operating conditions.

12.3.1 Mode of Operation of the Carburetor

The functional principle of the carburetor is based on the fact that by reducing a cross section in an air-conducting channel, there is less pressure than in the larger cross section or than in the atmosphere due to the greater flow rate in the narrow cross section. This pressure differential is used to supply fuel to the air through suitable cross sections (Figure 12.1). A characteristic feature of carburetors is the generation of a differential pressure signal from an air stream, and its direct conversion into a fuel stream. In principle, the air and fuel sides are designed identically and can be described with the Bernoulli equation for fluid mechanics. Given a simplified assumption of an incompressible flow, the following equation results for the air mass flow:

Input

Pressure difference, air

Air filter side

! L = AL ⋅a L ⋅ e ⋅ 2 ⋅ ∆pL ⋅ rL m



(12.1)

In this formula AL = cross section of the venturi, α L = flow coefficient, ε = factor for air compressibility, ∆pL = differential pressure venturi to environmental, ρ L = air density in the air funnel. For the valid fuel mass flow, ! Kr = AKr ⋅ α Kr ⋅ 2 ⋅ ∆pKr ⋅ rKr m



(12.2)

In this formula, AKr = cross-section fuel nozzle, α Kr = flow speed of the nozzle, ∆pKr = pressure differential across the nozzle, ρ Kr = fuel density. A carburetor has a fuel reservoir (float chamber) with a free fuel surface whose level is kept constant. One must differentiate between constant air funnel cross section and variable air funnel cross section. Constant air funnel cross section (fixed air funnel carburetor) (Figure 12.1): Most carburetors are essentially constructed according to this principle. There is a venturi-shaped air funnel located in the intake air duct with a fixed cross section. It is assigned with at least one main nozzle. The pressure differential generated with the air funnel remains small in smaller airflows. One must therefore calculate and consider the existing pressure difference between the intake and intake manifold for metering the fuel. Carburetors with a constant air funnel cross section require several nozzle systems and an accelerator pump for an appropriate fuel supply in the engine map. One mixes compensating air into the fuel to compensate for the influence of the different Reynold’s numbers in the fuel and air flows. Variable air funnel cross section: The intake air duct cross section is normally changed with a movable element. Commonly used are •• an air damper •• a piston that penetrates the channel •• a swiveling lever that constricts the channel. This enables a wide range of air streams to be controlled using a differential pressure that changes only slightly. For

Fuel inlet Fuel nozzle (main nozzle)

Ventilation

Float

Restriction (venturi pipe)

Float chamber Mixing chamber

Pressure difference, fuel

Throttle valve Intake manifold side

Fuel level

Figure 12.1 Principle of the carburetor.

516 | Internal Combustion Engine Handbook

6606_Book.indb 516

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

A throttle valve stop screw serves to set the opening of the throttle valve and, hence, control the air and mixture flow for idling. In a further development, an idling fuel shutoff valve was added. The accelerator pump with a diaphragm, as well as an intake and pressure valve, is actuated via a linkage together with the throttle valve. The outlet orifice of the injection tube is calibrated. A pressure-tapping hole for the spark advance ends in the narrowest cross section of the air funnel. The single-barrel carburetor can be equipped with an automatic choke.

symmetrical reasons, a conical nozzle needle that extends into a needle jet is connected to a movable element to meter the fuel. If the movable element also works while the engine is idling, the fuel can be dosed with the needle jet for the entire range of the air streams in an idling engine warm from operation. One refers here to a constant vacuum carburetor. When the movable element does not work while the engine is idling, rather it rests on a stop, one refers to a constant pressure stage. Constant pressure stages are also frequently used as a second stage in multistage carburetors.

12.3.2.1.2 Dual carburetor A dual carburetor is the combination of two single-barrel carburetors in one housing. Each of the intake air ducts is assigned with an equivalent set of systems. Generally, there is only one float chamber and one accelerator pump. The required control organs for the starting control are only provided singularly. The dual carburetor has two parallel intake air ducts in a common carburetor housing, each equipped with a throttle valve that supplies the two separate intake manifolds. The throttle valves are activated simultaneously and can be either on a common shaft or on two parallel shafts. The same holds true for the chokes.

12.3.2 Designs

The designs can be categorized according to the number of intake air ducts and their spatial position. 12.3.2.1 Number of Intake Air Ducts 12.3.2.1.1 Single-barrel carburetor This type of carburetor has an intake air duct with a throttle valve and is the most frequently utilized design. Figure 12.2 shows a design example of a downdraft carburetoron which several dimensions and expansion stages of the carburetor are based on the VW Beetle. This is considered to be a fixed air funnel carburetor. The system consists of a float chamber with a float and float needle valve, as well as an internal float chamber ventilation in the inlet. It is equipped with a main system with an air funnel, a discharge arm, a compensating air nozzle with a venturi tube and a main nozzle, as well as a dependent idling system with an idling nozzle,idling air nozzle,idle mixture adjusting screw, and transition holes.

12.3.2.1.3 Triple-barrel carburetor This is a combination of three fixed air funnels and has three in-line parallel intake air ducts, each with a throttle valve in a common carburetor housing. The classic application area was two triple-barrel carburetors on a six-cylinder boxer engine.

Compensating air nozzle with mixing pipe Float chamber vent pipe Restricting throttle Single pull down Idle air adjusting nozzle Bimetallic spring

Float artery valve

Injection pipe

Cone pulley Throttle valve setting screw

Fuel inlet

Venturi pipe Connection pipe for ignition control Pressure valve Suction valve

Outlet arm Throttle valve

Pump lever

Stud screw Idling mix/ Regulating screw

Pump spring Main nozzle carrier with main nozzle

Pump transfer lever Transition boreholes Idling duct

Figure 12.2 Downdraft carburetor as a fixed venturi carburetor.

Internal Combustion Engine Handbook | 517

6606_Book.indb 517

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

12.3.2.1.4 Two-stage carburetor One divides the possible airflow into two stages with a larger range of air flows. The first is used for the smaller air throughputs, including idling and partial load, and the second stage, frequently with a larger cross section, is opened only to attain maximum output. The two-stage carburetor contains two parallel intake air ducts in a common housing that are both connected to an intake manifold. Both throttle valves in the carburetor are opened sequentially. The first stage of a two-stage carburetor is equipped with all the additional necessary systems. The accelerator pump and starting control are only contrastingly required for the first stage. This is generally opened directly with the gas pedal; the second stage is linked to the opening of the first stage in the carburetor. Similar to single-barrel carburetors, a two-stage carburetor can be combined to form a double-barrel two-stage carburetor. Usually, double-barrel two-stage carburetors are supplied by means of a float chamber. Such carburetors are used in large volume six- and eight-cylinder engines. 12.3.2.2 Location of the Intake Air Duct One can differentiate between the respective locations for the air ducts: 12.3.2.2.1 Downdraft carburetor The intake air duct is arranged at a right angle; the throttle valve is located in the bottom part of the carburetor. The air flows from top to bottom. 12.3.2.2.2 Horizontal draft carburetor The intake air duct lies horizontally in a flat stream or horizontal draft carburetor.

12.3.2.2.3 Semi-downdraft carburetor The intake air duct lies at an angle; the air stream can be directed upward or downward. 12.3.2.2.4 Updraft carburetor In this case, the air stream rises from bottom to top. The throttle valve is in the top of the carburetor. 12.3.2.3 Designs for Special Applications 12.3.2.3.1 Pressurized carburetor A pressurized carburetor is utilized on the pressure side of the supercharger of a supercharged engine sealed to the outside. 12.3.2.3.2 Carburetor for two-stroke engines Given the limited intake capabilities of the generally used crankcase scavenging, and the strong pulsations in the intake air duct, the cross section of the airflow is bigger than in carburetors for four-stroke engines.

12.3.3 Important Systems on Carburetors

Additional systems on the basic carburetor allow the carburetor to be used within the entire operating range of the engine with minimized emissions and fuel consumption, as well as improved driving behavior. The following briefly lists the most important systems. It merely represents a selection and no claim is made for completeness. 12.3.3.1 Acceleration Enrichment The preparation of the mixture, which initially begins in the carburetor, primarily occurs in the intake manifold. The boiling point of the fuel, and the heating of the intake air and

Pre-atomizer Supplementary air nozzle Restricting throttle Injection nozzle

Adjusting air nozzle Basic idle nozzle

Pump lever

Float needle

Pump tappet

Fuel inlet Accelerating pump

Float artery valve Float chamber Fuel level

Pump piston

Float Basic idle nozzle Idling duct

Mixing pipe

Pump chamber

Main nozzle

Suction valve

Additive mixture

Throttle valve

Ground idle mix-adjusting screw

Idle mix-shut off valve

Transition boreholes

Pressure valve

Venturi pipe

Figure 12.3 Carburettor with a mechanically actuated accelerator pump.

518 | Internal Combustion Engine Handbook

6606_Book.indb 518

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

the intake manifold, must be harmonized with each other to achieve the most homogeneous mixture in hot running engines. However, high-boiling components of the fuel are still liquid in the intake manifold for warm operating engines, respectively, as film on the wall that is entrained by the air. The air thereby enters the cylinder faster than the fuel. When the throttle valve is opened, the mixture becomes lean so that a temporary enrichment is necessary. This is done by acceleration enrichment using an accelerator pump. Accelerator pumps are volume-displacing pumps, with an intake and pressure valve. In the intake stroke, fuel streams from the float chamber through the intake valve into the pump interior. Upon the delivery stroke, fuel is generally delivered by the pressure valve to an injection device. This is calibrated and leads into the intake air duct. The delivery stroke is executed by a pump spring that is pretensioned in various ways. One makes a distinction between •• mechanically actuated plunger pump (Figure 12.3) •• mechanically actuated diaphragm pump •• pneumatically actuated accelerator pump. 12.3.3.2 Actuating and Constructing a Second Carburetor Stage In multistage carburetors, there are several systems for actuating the second stage, which have similar constructions. It is important for the second stage to always be reliably closed with the first stage. One has to make a distinction is drawn between mechanically and pneumatically actuated systems. Mechanically actuated second stage: In mechanically actuated systems, the second carburetor stage is primarily actuated by a trailing linkage, whereby an opening damper is often installed to prevent a temporary lean mixture. The second stage initially begins to pen when the first stage is already half open and a catch of the throttle control lever is correspondingly positioned. Both valves reach the full opening position simultaneously. The design of the second stage largely corresponds to that of the first stage: There is a main system and transition system similar to the idling system in the first stage with a large fuel reserve. The latter system should ensure a smooth torque increase when the second stage opens quickly at low rpm, with a correspondingly high pressure in the intake manifold. For carburetors whose second stage is a constant pressure stage, the throttle valves on a common shaft actuate the second stage with a trailing linkage. Pneumatically actuated second stage: In this design, both stages are constructed as fixed air funnel carburetors. A diaphragm unit actuates the second stage with a connecting rod articulated to the throttle control lever of the second stage. A corresponding design ensures that the second stage can only open when the first is almost completely open, and the second is closed with the first stage.

12.3.3.3 Nozzles These meter fuel, compensating air, and the premixture. To form an optimum flow and protect the actual calibrating section, they have an inlet cone and sometimes an outlet cone. Figure 12.4 shows a typical main nozzle. They can be used in both flow directions. Calibrating route

Figure 12.4 Compression temperature increase with intake air.

12.3.3.4 Float Chamber A float chamber serves to control the fuel level in the carburetor as well as fuel accumulator and contains a guided float which immerses into the fuel. The float actuates a needle, thereby closing the body of the float needle valve that blocks the inflow of fuel when a set level has been attained. There are carburetor designs that do not control the fuel level by utilizing a float chamber. The level in the fuel accumulator is controlled by means of the pressure in the fuel accumulator (diaphragm carburetor). The float chambers must be ventilated to fulfill their functionality. One has to make a distinction between external and internal ventilation. With external ventilation, the gas area of the float chamber is directly connected to the environment surrounding the carburetor. This prevents problems when the engine is running hot, which could otherwise occur from the fuel evaporating out of the float chamber. A pipe leads into the inlet and/or pure air side of the air filter with internal ventilation. 12.3.3.5 Starting Controls In carburetor engines, special attention must be given to controlling the engine from a cold start through warm-up to a “hot” engine. In particular, the legal limits for exhaust emissions, problems associated with mixture distribution, fuel deposits on the wall, costs, and operating and driving comfort are requirements that must be taken into account. The requirements, in particular the enrichment of the mixture, result from •• increased friction from unattained operating temperature •• insufficient mixture preparation •• increased power demand for auxiliary systems. In a stoichiometric mixture of air and conventional fuel (boiling curve), the dew point under environmental pressure is approximately 35°C. The homogenization of the mixture can generally only occur in the cylinder when the dew point

Internal Combustion Engine Handbook | 519

6606_Book.indb 519

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

temperature of the mixture in the cylinder is exceeded during the compression stroke. Consequently, there is always a large amount of liquid fuel in the intake system in cold engines. The mixture must therefore be enriched in comparison to a warm operating state, even when the engine is idling, to compensate for deficient homogenization. The fuel must also be enriched during acceleration. The amount of enrichment must therefore be greater, depending on how low the temperature of the air on the intake manifold wall is. The main reference variable for enrichment is the temperature of the intake manifold. The mixture formation must correspond to the following phases: •• Initial poststart phase: The engine starts as soon as there is an ignitable mixture. This process occurs faster at higher temperatures than at lower temperatures. Moreover, a stationary mixture enrichment is also additionally required

depending on the initial temperature and the time after the start. •• Run-up to a stable idling speed: An initial, very strong enrichment must be attenuated so that the mixture remains flammable. •• Warm-up to operating temperature: The mixture stream and the enriching can be adjusted depending on the temperature of the intake manifold corresponding to the engine heat. In principle, one differentiates between three systems for start control: •• The manual starter, which will not be discussed here since it is now defunct •• Automatic choke (Figure 12.5). The mixture stream for starting and idling the cold engine is ensured by opening the throttle valve; the mixture is enriched with the choke.

Restricting throttle closed Bimetallic spring

Starter

Restricting throttle shaft Restricting throttle Cone pulley

Cone pulley in cold start position

Tappet Throttle valve set screw

Throttle valve Throttle valve lever Interaction between Bimetallic spring, Restricting throttle, Cone pulley and Throttle valve set screw

Restricting throttle fully open Cone pulley in idle position

Automatic choke off

Automatic choke in cold start position

Restricting throttle partially open Cone pulley in intermediate position

Automatic choke in intermediate position

Figure 12.5 Automatic choke.

520 | Internal Combustion Engine Handbook

6606_Book.indb 520

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

The temperature-dependent control element is a bimetal spring against which a choke can be drawn by the air stream. The functional relationship between the position of the choke and the position of the throttle valve is established by a cone pulley. In contrast to a fully automatic start, the functional process before the start must be triggered by a single depression of the gas pedal. The bimetal spring is heated to raise its temperature to that of the engine. As the temperature of the bimetal spring rises, the choke opens and mixture enrichment is attenuated. •• Fully automatic start. The essential difference from an automatic choke is that this system is not triggered before starting. On the one hand, the control of the fuel mixture stream for idling a cold engine is separate from the mixture enrichment and, on the other hand, depending on the inducted air stream. There are a whole series of devices and additional functions for carburetors that improve their operating behavior, such as the throttle valve actuator, pressure tapping holes, systems for compensating air, idling transition, additional mixture, and circulating air, to name only a few, which, however, are not discussed here.

12.3.4 Electronically Controlled Carburetors

Electronically controlled carburetors were developed to improve the adaptation of the mixture to all engine-operating ranges and, thereby, to achieve the demands for minimizing untreated emissions and lowering fuel consumption. A lambda control was later adapted to the system. The mechanical design of the electronically controlled carburetor is basically identical to conventional carburetors. Additional features are an actuation of the throttle valve in the near-idling range, an influencing of the mixture enrichment, sensors and an electronic control unit (Figure 12.6). The electronically controlled carburetor is in principle a fixed air funnel carburetor with a two-stage design with a pneumatically actuated second stage. The throttle valve for the first stage is actuated in the range close to idling with a continuously variable, position-regulated throttle valve actuator. The continuously variable mixture enrichment throughout the entire range is executed by means of the choke in the first stage that is actuated to restrict the airflow to enrich the mixture. A throttle valve potentiometer is located on the throttle valve shaft of the first stage that determines the throttle valve position or the change of the throttle valve position. Signals from

Restricting throttle regulator Restricting throttle valve (RT)

Throttle valve regulator (TVR)

Otto engine Throttle valve (TV) Throttle valve potentiometer Temperature sensor Electronic control device Temperature

Idling rpm regulating

TV-angle

Acceleration enrichment End stage restricting throttle

Output

Lambda-control

Reworking

End stage throttle valve

Start up and warm run control

Characteristic diagram correction

Lambda-Signal Rotational speed

Overrun cut-off, engine stop

Converting

TVR-Position

Input

Figure 12.6 Block diagram of an electronically controlled carburetor.

Internal Combustion Engine Handbook | 521

6606_Book.indb 521

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

the temperature sensors for the coolant and intake manifold, the idle switch on the throttle valve actuator and a speed tap are input variables for the electronic control that influence the throttle valve of both the first stage and the choke. The following functions, among others, are therefore illustrated: •• control of engine start and warm-up •• acceleration enrichment •• lambda control •• influencing the air–fuel ratio in the program map •• idling rpm regulating •• overrun fuel cut-off •• catalytic-converter protection function by shutting off the fuel. When starting with an electronic carburetor, the throttle valve of the first stage is actuated, and the choke is fully closed. During the run-up phase, the idling speed control takes over the control of the throttle valve with a set point depending on the coolant temperature of the engine. The mixture enrichment is executed with the choke, and a basic enrichment is dictated as a function of the intake manifold temperature in the program map while stationary. This is an addition to acceleration enrichment when the load of the engine increases.

12.3.5 Constant Vacuum Carburetor

The functional principle of the constant vacuum carburetors is illustrated in Figure 12.7 with the example of a Zenith– Stromberg CD carburetor. The system works with a variable air cross section in the form of a horizontal draft carburetor. The alteration in the cross section in the intake air duct is caused by a piston inserted from the top. In addition to pressure from its weight, a spring presses the piston against a bridge, which then fills the bottom section of the intake air duct. An

Carburetor cover

area is sealed above the piston with a diaphragm that abuts to the bottom of the plunger and accordingly creates a pressure compensation. The bottom of the diaphragm is ventilated toward the inlet. The spring creates the necessity for a linear rise in the pressure differential for raising the piston. Due to this pressure differential, the inducted air is accelerated between the inlet and the narrowest place between the piston and the bridge. The position of the piston is therefore a measurement for the inducted air stream. The float chamber is located below the bridge. A nozzle needle is fixed in the center of the piston, which extends into a needle nozzle and meters the fuel as a function of the piston displacement. The needle nozzle enables the fuel measurement on a characteristic curve depending on the air stream. A constant vacuum carburetor is adapted so that the piston is completely opened at a full load and approximately half-maximum speed. With larger air streams, the constant vacuum carburetor then operates like a fixed air funnel carburetor. By exploiting the pressure oscillations in the intake system, one can set the mixture ratio of the air and fuel in the entire engine-operating range.

12.3.6 Operating Behavior 12.3.6.1 Hot Operation The carburetor is cooled by the inducted air and the initial evaporation of the fuel during the engine operation. High environmental temperatures, and subsequent heat after the engine is turned off, can lead to operational problems. The main problems are high component temperatures in the carburetor, which lead to fuel evaporation, engines with cylinder heads in a continuous flow arrangement in which the intake manifold and exhaust manifold are stacked, and the boiling behavior of the fuel.

Sealed

Idle mix, shut-off valve Damper oil Damper piston

Connection for early ignition shifting

Piston spring Diaphragm

Idle mix adjusting screw

Carburetor inlet Fuel inlet

Needle nozzle

Fuel return

Throttle valve

Intake manifold

Mixing chamber

Fuel return valve Bimetallic spring Nozzle needle Float Float chamber Fuel level

Bypassing the mixture pipe

Figure 12.7 Constant vacuum carburetor.

522 | Internal Combustion Engine Handbook

6606_Book.indb 522

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

Improvements in relation to hot operation are possible with the following: •• A hot idling valve to open a bypass for air to circumvent the carburetor. This is primarily utilized especially in idling to counter mixture enrichment. •• Fuel return. A valve connected parallel to the float needle valve returns unneeded fuel to the tank. This removes heat from the carburetor. •• Heat is reduced by shielding plates and an insulating flange. 12.3.6.2 Altitude Compensation The mixture enrichment is executed in carburetors, in a first consideration, as proportional to the root of the environmental pressure. For example, the enrichment at 1600 m above sea level is approximately 10%. The following corrective options are possible: •• influencing the pressure in the float chamber •• influencing the differential pressure signal for fuel metering •• influencing the air correction and the metering cross section for fuel. 12.3.6.3 Icing Because of the evaporation heat of the fuel, heat energy is drawn from the inducted air and the components of the carburetor; they cool down. This can then cause the water vapor in the air to freeze and, hence, cause malfunctions in operation. The tendency toward ice formation is greatest at air temperatures of 5°C with high humidity levels, especially in fog. Icing can be controlled very effectively by introducing additives in the fuel (such as alcohols), which then lower the freezing point. Another possibility is created by preheating the intake air and, hence, increasing the air temperature. One makes a differential between the following: Idling icing: This can occur when ice from the cooling mixture forms on the edge of the throttle valve at low engine loads. In addition, ice can form on the holes through which the premixture for idling exits as well as on the transition holes, which can distort the differential pressure signal. Additives can also provide an increase in the intake air preheating as well as a change to using fuel with additives. Icing during idling disappears as the heat increases. Full-load icing: Full-load icing, or air funnel icing, occurs primarily in older carburetors in which the venturi tube of the main system is located in the middle of the air funnel, and the premixture enters the air over a short path. This can be dealt with by increased preheating of the intake air, or by fuel additives. Carburetor icing rarely occurs in carburetors in which the venturi tube is not in the middle of the air funnel. Idling rpm control: The idling speed can be indirectly controlled with mechanical means utilizing an intake manifold pressure regulator. A separate idle controller is integrated in electronically controlled carburetors.

Engine after running: To prevent the engine from “after running” after being shut off due to the self-ignition of the mixture at hot areas in the combustion chamber, the supply of fuel is shut off along with the ignition. Overrun: The transition from load operation to overrun must, in certain cases, be supported in two ways. Adding mixture in overrun prevents an undesirable lean mixture. This is executed with an additional mixture system. The other possibility is to shut off the fuel in overrun, which is normally speed controlled, and additionally serves to lower the fuel consumption. The transition to overrun mode is accomplished with a throttle valve dashpot or an overrun air valve to minimize load change reactions.

12.3.7 Lambda Closed-Loop Control

Two solutions are implemented for effective lambda closedloop control: •• In a fixed air funnel carburetor, an additional intervention possibility to achieve lambda closed-loop control is to provide valves that are in the main system and/or idling system and that influence the air–fuel ratio. Alternately, one can utilize an arrangement in the form of a partial load control in which a channel to the intake manifold is opened and closed with a solenoid valve. •• With electronic carburetors, the lambda closed-loop control is one of the features of an electronic control unit. The intervention relating to the air-fuel mixture ratio is executed with a choke.

12.3.8 Mixture Formation by Utilizing Gasoline Injection 12.3.8.1 Intake Manifold Injection Systems The demands for low vehicle emissions and reduced fuel consumption, whereby fuel is injected via electronically controlled fuel injectors for individual cylinders into the intake arms of the gasoline engines, are the primary influences on the design of modern intake manifold injection systems. A typical configuration for fulfilling standards for minimum emissions is shown in Figure 12.8. The measures to reduce emissions beyond the basic functions of the engine control systems—injection and ignition—are mainly based on the required emissions standards, the untreated emissions of the combustion engine, and the vehicle weight class in the smog test. For example, after treatment of exhaust, measures, such as secondary air injection in combination with retarded ignition, are required to quickly heat the catalytic converter. These measures, as well as their diagnosis, increase the complexity of the engine control systems in the form of sensors, actuators, cabling, and computer programming. Typical functional characteristics of modern engine control systems are as follows: •• torque-based load control with an electronically controlled throttle valve (ETC = electronic throttle control)

Internal Combustion Engine Handbook | 523

6606_Book.indb 523

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

Phase variable intake camshaft

Improved spray injector valve Camshaft position sensor

Pin ignition coils

Engine control device

Intake manifold pressure sensor

Cooling water temperature sensor Linear lambda sensor pre-catalytic converter

Knock sensor

2nd Lambda Sensor jump probe after catalytic converter

Electronically controlled throttle valve

Tank ventilation valve

Crankshaft position sensor

Muffler Three-way catalytic converter close to the engine

Figure 12.8 Intake manifold injection system.

•• model-based functions such as a model-based intake manifold filling with load detection via a hot film air mass meter or an intake manifold pressure sensor •• control of the position of a continuously adjustable camshaft on the inlet and/or outlet side •• control of various relays for switching on and/or off components (main relay, fuel pump relay, fan relay, starter relay, and compressor relay) •• active camshaft position sensor for rapid detection of the camshaft position and, thereby, for fast synchronization of motor control during engine start •• cylinder-selective knock control based on a crankcase vibration sensor for optimum power and consumption control of the ignition timing point •• control of the tank ventilation valves to regenerate the active carbon canister while the engine is running •• special catalytic converter heating function with secondary air system, retarded ignition, and transmission shifting point control •• precise control of the mixture composition via an oxygen sensor (“lambda sensor”) upstream from the catalytic converter and a so-called “trim control” via a second oxygen sensor downstream from the catalytic converter

In principle, one differentiates between two basic fuel system designs: 1. Fuel system with a return flow system (Figure 12.9): A characteristic of this fuel system is that the pressure regulator is directly on the fuel distribution rail. The pressure diaphragm receives the intake manifold pressure on one side, so that a constant differential pressure exists between the fuel in the fuel rail and the intake manifold. Given a constant fuel injector control time, this makes the injected fuel quantity independent of the intake manifold pressure. The advantages of the fuel system with a return flow system are as follows: •• favorable fuel pressure control dynamics •• good hot start behavior by rinsing the fuel distribution rail with cool fuel from the tank •• an injected fuel quantity which is independent of the intake manifold pressure. An essential disadvantage is that the fuel is heated in the tank (up to 10°K in contrast to systems without a return flow). This increases the fuel evaporation in the tank and results in an increased load on the active carbon canister.

•• “on-board” diagnosis (OBD) of all exhaust-relevant components and functions.

2. For this reason, as well as to reduce system costs, returnfree fuel systems were developed (Figure 12.10). They are characterized by the integration of the fuel pump and pressure control valve in the tank, or close to the tank.

In the design of the fuel system, special care is taken to ensure that the fuel in the fuel distribution rail is only slightly heated. If the fuel is heated in the phase after shutting off the engine (so-called “hot soak”), then vapor bubbles can occur in the fuel distribution rail that can lead to start problems in a subsequent hot start.

The advantage of this design is that the excess fuel does not have to be first pumped into the engine compartment and then flows via the pressure regulator back into the tank. The injection times are correspondingly corrected by the engine control system as a function of the constant fuel pressure of approximately 350 kPa (3.5 bar ± 0.5 bar).

524 | Internal Combustion Engine Handbook

6606_Book.indb 524

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

Vacuum intake manifold

Fuel distribution rail

Pressure control valve

Engine compartment Injectors

Check valve Fuel pump

Fuel return line

Fuel tank

Figure 12.9 Design of a fuel distribution system with return flow.

Fuel distribution rail

Damper element

Engine compartment Check valve

Pressure control valve

Injectors

Fuel tank

Fuel pump

Fuel return line

Figure 12.10 Layout of a return-free fuel distribution system.

Internal Combustion Engine Handbook | 525

6606_Book.indb 525

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

reflection/ deflection

wall guided

swirl tumble

charge motion guided

spray guided

Figure 12.11 Combustion.

To avoid large pressure fluctuations in the fuel distribution rail that can lead to fluctuations in the injected fuel quantity, pressure pulsation dampers are used in returnfree fuel systems. 12.3.8.2 Systems for Direct Injection In addition to the described possibility for injecting the fuel as a liquid into the intake manifold of the spark-ignition engine, direct injection systems have been developed in recent years. The fuel is hereby directly injected into the combustion chamber from a central fuel distribution rail under high pressure by electronically controlled fuel injectors (“the common rail” principle). By layering the fuel and air by injecting during the compression phase, a relatively lean mixture is therefore produced with a relatively rich mixture cloud close to the spark plug that ensures reliable ignition. Due to the excess air in this mode of operation, there is a reduction in the charge cycle work and in the loss of wall heat during the high-pressure phase of combustion, which together substantially lower the specific fuel consumption. In addition, the direct injection cools the interior of the cylinder from the evaporation of the fuel, which then reduces knocking at full loads. This therefore makes it possible to increase the compression by approximately one unit. This additionally creates a lower specific fuel consumption. The stratified operation only makes sense only within a limited range of operation under a partial load in the gasoline engines. In the other ranges, the engine is operated homogeneously lean, stoichiometric, or rich under a full load. Depending on the installation site and position of the fuel injector, and the arrangement of the air intake into the cylinder, one differentiates between wall-directed, air-directed, and jet-directed combustion (Figure 12.11): •• HPDI* with wall-directed combustion (Figure 12.11 links): The fuel injector is arranged on the side, and the fuel is sprayed onto the piston head. The shape of the piston recess and the type of airflow dictate how the injected fuel is directed toward the spark plug. One differentiates with the air streams, depending on the geometry of the intake ducts, between a so-called reverse tumble procedure (“reversion role”) and swirl procedure with air-directed combustion.

*HPDI means: High Pressure Direct Injection.

•• HPDI with air-directed combustion (Figure 12.11, middle). •• The fuel injector is also located on the side, but the fuel is injected into the air in the center of the combustion chamber in contrast to wall-directed combustion. Therefore, a substantial amount of air movement is necessary, which is generated by a variable tumble. HPDI with jet-directed combustion (Figure 12.11, right). This type of combustion has the greatest potential for lean engine operation and, therefore, the greatest potential savings in fuel consumption of more than 30% at low partial load. The fuel injector is located in the center of the combustion chamber and the spark plug is close to it at a lateral angle. This therefore prevents the fuel from contacting the piston or combustion chamber walls. This type of combustion places increased demand on the preparation of the jet by the fuel injector. To ensure reliable ignition, and low fouling of the spark plug, the fuel in the area of the spark plug must be finely atomized and the spray pattern may not change substantially when the combustion chamber pressure changes. In addition, there exists air-supported direct injection (OCPTM)†, in which the central injector position close to the spark plug has proven to be particularly advantageous. The overall lean mixture during combustion in stratified operation creates a problem for exhaust after treatment as conventional three-way catalytic converters cannot reduce NOx emissions. Despite a reduced NOx pipe emission of up to 30% by utilizing the exhaust gas recirculation rate, a specialized after treatment is necessary for fulfilling the exhaust gas-limiting values for the NOx emissions. Along with the selective NOx reduction catalytic converters, which have a low-thermal stability, so-called NOx storage catalytic converters are utilized today. Storage catalytic converters absorb the NOx emissions in lean operation and then convert them in substoichiometric operation into N2 and CO2. A complex function in the engine management system controls this process. Storage catalytic converters tend to experience “sulfur poisoning” and are, therefore, reliant on fuels with low sulfur content. The exhaust after treatment measures reduce the savings in effective fuel consumption that result from direct injection.

† OCPTM means: “Orbital Combustions System” is a registered trade mark of Orbital Corporation.

526 | Internal Combustion Engine Handbook

6606_Book.indb 526

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

fuel injector

air injector

Figure 12.12 Air and fuel distribution rail with fuel injector and air injection valve.

12.3.8.2.1 Air-supported direct injection In addition to liquid high-pressure direct injection systems, there are air-supported direct injection systems, such as OCPTM. Air-supported direct injection systems enable stable combustion with good stratification, that is, compatible with large amounts of recycled exhaust gas, because of the high quality of the mixture preparation. The main component of air-supported direct injection is an arrangement of an electromagnetic fuel injector and an electromagnetically actuated air injection valve (Figure 12.12) that injects finely atomized fuel into the combustion chamber. Figure 12.13 provides an overview of the air-supported direct injection system. The injection system is divided into two subsystems, the compressed air path and the fuel path. A compressor driven by gears, or a belt, generates the required compressed air. The pressure level is set by a mechanical pressure regulator to a set point and the fuel is then conveyed

by an electrical fuel pump. The fuel pressure is regulated to be at a constant differential pressure from the compressed air (approximately 0.7–1.5 bar). The fuel is metered is executed with a conventional fuel injector for intake manifold injection. The fuel is injected into a ventur in the air injection valve. By utilizing the air injector, a finely atomized mixture cloud is introduced into the combustion chamber that can then be directly ignited. In stratified-charge operation, the jet-directed combustion is, therefore, possible with low untreated emissions. By synchronizing the timed arrangement for the phase angle of injection (= the actuation of the air injector) and ignition, an optimum air–fuel ratio in the mixture can be maintained at the spark plug for stable combustion under all operating conditions. As Figure 12.13 illustrates, the system consists of an air filter, an air mass meter with an integrated air temperature sensor, an electrical throttle valve actuator, and the intake manifold.

Figure 12.13 System overview of the air-supported direct injection system.

Internal Combustion Engine Handbook | 527

6606_Book.indb 527

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

The externally recycled exhaust can be introduced into the individual intake tubes via the intake manifold, or via a line in the cylinder head. In any case, a position-controlled EGR valve is necessary to provide precise metering. Optionally, an internal exhaust-gas recirculation can be provided by adjusting the phase angle of the intake and exhaust camshaft, allowing the torque characteristic and output of the engine to be improved. The exhaust after treatment subsystem consists of a threeway catalytic converter close to the engine and an underfloor NOx storage reduction catalytic converter. A broadband oxygen sensor enables the air–fuel ratio to be controlled in lean operation, including the regeneration phase and the stoichiometric operation. The exhaust temperature sensor optimally controls the NOx storage catalytic converter as well as triggering measures for catalytic converter protection. A binary lambda sensor downstream from the NOx storage catalytic converter is required for the control of the regeneration phase to work properly. Alternately, an NOx sensor can also be used. The compressor in the system offers a new and interesting solution for scavenging the active carbon canister. The air inducted by the compressor is guided through the active carbon canister. This therefore allows stratified-charge operation to also be used to attain a sufficient scavenging rate without the disadvantages of untreated emissions. The process of fuel preparation in the OCPTM injector differs from that in a high-pressure direct fuel injector in the following ways: With a high-pressure injector, the jet decomposes primarily from turbulence and inertia in the liquid jet itself. Up to the end of the decomposition process, a distance of approximately 10–50 times of that of the orifice diameter is necessary. In the case of an air-supported injector, the jet dissipation is executed when the aerodynamic forces exceed the surface tension in the liquid. The pressure level in the air injector is selected in such a way that the critical pressure ratio at the valve orifice is exceeded during injection. The thereby resulting speed of sound of the airflow causes strong aerodynamic forces to be exerted on the fuel jet. The

SMD = 10.3 µm D10 = 6 µm

12.3.8.2.2 High-pressure injection A further possibility for directly injecting fuel into the cylinder is liquid fuel injection using the common rail principle (fuel injection from a common pressure line). Figure 12.15 provides a system overview of high-pressure gasoline injection divided into an engine control system and a fuel system. A gasoline engine with direct injection requires the use of an electronically controlled throttle valve for the various modes of operation with homogeneous charging and stratified charging. The mixture control for the lean mixture with stratified charging requires a linear λ -sensor, which can also ensure this function for homogeneous operation with λ = 1.

D50 = 13 µm D90 = 34 µm

SMD [µm]

Volume fraction [%]

8.0

essential part of the atomization process is already completed directly at the valve outlet. Other thermodynamic effects, such as evaporation, play a special role particularly when injecting fuel into a medium with a high temperature. This interaction of the fuel spray with the cylinder charge represents the interface between the fuel system and the combustion system. The atomization quality can be seen in Figure 12.14. The average Sauter mean diameter (SMD) is 10.3 μ m; only an insignificant number of droplets have a diameter over 40 μm. The required air mass flow for the OCPTM-air injector in reference to the entire amount of air inducted by the engine varies from 15% in homogeneous idling to 1.5% in full-load operation. As an absolute quantity, this results in approximately 5–9 mg air per injected fuel per pulse for a 1.5-liter four-cylinder engine. The absolute pressure in the air rail is preferably set at 6.5 bar. Compressed air is generated by means of a water-cooled piston compressor, which is driven by the engine via gears or a belt. Due to the high turbulence and stratification in the airsupported OCPTM-direct injection systems, normally no additional measures are required to move the charge, such as swirl and tumble plates. Because of the low sensitivity compared to internal cylinder flow, this injection system is particularly suitable for engines with different valve configurations without active or passive measures to move the charge.

4.0

0.0

Fuel pressure: Air pressure: 40

20

0 0

40

80

Particle diameter [microns]

120

720 kPa 650 kPa

0

5

10

Time [msec]

15

20

Figure 12.14 Drop size distribution in relation to volume and time in air-supported direct injection.

528 | Internal Combustion Engine Handbook

6606_Book.indb 528

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

Figure 12.15 Gasoline direct injection—System overview.

The high-pressure pump is fed from the low-pressure system that has a pressure of 1–4 bar. In the mechanically driven high-pressure pump, the fuel pressure is increased to up to 120 bar. The high pressure is controlled via an electrically controlled pressure control valve. The return flow from the high-pressure line exits directly from the pressure control valve into the feed for the high-pressure pump. A pressure sensorserves to detect the pressure. An overpressure valve is integrated in the high-pressure circuit for safety reasons, which then limits the maximum fuel pressure. If the fuel pressure in the fuel distribution rail is kept constant over the entire speed and load ranges, then the electrical pressure regulator can, in principle, be replaced by a mechanical pressure regulator. The fuel injector valve is located directly in the cylinder head. Due to the high fuel pressure, the magnetic force required to open the valve needle must be much higher than with lowpressure fuel injectors. In addition, the valve needle must be opened and closed extremely quickly for the reasons of charge stratification and metering.

The fuel atomization quality is extremely dependent on the fuel pressure, the counter pressure, the flow calibration, and the spray dispersal angle. Figure 12.16 shows the atomization quality of a high-pressure fuel injector in comparison to a low-pressure fuel injector and an air-supported fuel injector. Different types of combustion and combustion chambers require different flow calibrations and jet shapes. The high-pressure pump provides the pressure (50 to 120 bar) for the fuel distribution rail. The injection pressure has, however, been recently developed toward 200 bar, and no upper limit is yet in sight for jet-directed procedures. In connection with a multiple-hole nozzle, one can therefore achieve a stable spray dispersal angle and a good evaporation and/ or mixture preparation. The high-pressure pump is driven directly by the engine camshaft and is, therefore, mounted on the engine. One makes a differentiation between radial and axial high-pressure pumps. Figure 12.17 illustrates an axial plunger pump. A swash plate is rotated by a mechanically driven shaft and is, therefore,

Sauter diameter [microns]

100 Intake manifold-injection

80 60 40

Air-content injection valve

20 0

0

10

20

Direct-injection

30

40 50 60 Injection pressure [bar]

70

80

90

Figure 12.16 Comparison of the atomization quality of high-pressure injection valves and intake manifold fuel injection valves.

Internal Combustion Engine Handbook | 529

6606_Book.indb 529

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

responsible for the alternating stroke movement of the three pistons. Fuel passes into the cylinder through a groove in the swash plate into the cylinder and is then ejected via a check valve in the outlet. Each piston is mounted in a ball joint on the swash plate. The bearing and pump chambers are separated by a shaft-sealing ring. The combined materials and coatings are adapted to the wear and lubrication requirements of the gasoline operation. The extraordinarily narrow tolerances require that the fuel has to be finely filtered.

•• The high-pressure fuel injectorsrequire a control system that is adapted to the special requirements of this technology. The high fuel pressure and the increased demands for linearity and reproducibility from injection to injection make it necessary to alter the fuel injector control system. For the fuel injectors to open quickly, an increase in voltage and current to 80 V/10 A is required. When integrating the driver stages in the electronic control unit, the greater power loss of the driver stages must be taken into consideration. •• For lean engines, the use of an electrical throttle valve cannot be discounted. The control of this engine-operated throttle valve ensures that the pedal and throttle valve positions are completely independent. •• In un-throttled engine operation, there is no pressure differential for scavenging the active carbon canister. To attain the necessary scavenging rates, a pump is required for scavenging the active carbon canister in engine designs with a high-stratified charge component.

Figure 12.17 High-pressure fuel pump.

There are numerous new requirements on the engine control system for high-pressure direct injection: •• The pressure in the high-pressure fuel systems must be regulated. •• The lean operationfor engines with direct injection requires a linear lambda sensor, which covers the lean range and operation range where λ = 1.

homogeneous

In contrast to conventional engines with intake manifold injection systems, which work under nearly all operating conditions with a homogeneous stoichiometric mixture (i.e., the mixture is enriched only in special engine states, such as cold start, warm-up, and full load), the gasoline engine with direct injection is operated using different injection and combustion strategies. The fuel preparation strategies that produce different homogeneous operating conditions, and stratified charging, are clarified below (Figure. 12.18). The objective of the stratified charging is to concentrate a well-prepared fuel–air mixture at the spark plug location so that a locally limited, ignitable mixture arises (λ ≈ 1) which, despite the overall lean mixture, still creates favorable conditions for combustion. Due to the localized concentration of the mixture in the center of the combustion chamber, stratified charge operation also allows high exhaust gas recycling rates. The stratification of the mixture around the spark plug is achieved by late injection during the compression cycle.

stratified

• throttled operation

• non-throttled operation

• early injection during the intake stroke

• Late injection in the compression stroke

• homogeneous mixture distribution

• Charge stratified on the spark plug

Figure 12.18 Engine operation with homogeneous and stratified charging.

530 | Internal Combustion Engine Handbook

6606_Book.indb 530

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.3 Mixture Formation in Gasoline Engines (Carburetor/Gasoline Injection)

The throttle valve is opened completely for maximum air induction into the cylinder. The jet direction, jet shape, jet penetration depth and air flow in the cylinder are the decisive parameters for a successful stratification of the injected fuel near the spark plug. The jet direction cannot be changed during engine operation. The penetration depth depends on the difference between the fuel jet speed and the airflow speed in the cylinder. The jet speed can be influenced by the injection pressure. To attain a specific objective-related movement of air flow in the combustion chamber with turbulence near the spark plug (for good mixture preparation), the design of the intake duct and the combustion chamber (a key responsibility of the engine manufacturer) must be adapted and optimized. The current intake systems are constructed with a tumble or swirl design. When the operating state of the engine is to be changed from stratified charging to homogeneous charging with the same engine output, the inducted air mass must be reduced by closing the throttle valve ; and simultaneously, the amount of injected fuel in the cylinder must be increased to compensate for the greater throttling loss. In order to produce a homogeneous mixture in the combustion chamber, the fuel is injected during the intake cycle at the point in time in which the air speed is at its maximum. Problems with mixture preparation, driving ability and, above all, exhaust emissions prevent stratified charging from being used over the entire operating range of the engine. The utilization of the different operation states over the working range of the engine are shown in Figure 12.19. The example also considers the cooling water temperature to illustrate the influence of different environmental states. The following combustion states exist: •• homogeneous rich •• homogeneous λ = 1 with/without exhaust-gas recirculation

hot

•• homogeneous lean with exhaust-gas recirculation •• stratified charging with a high exhaust gas recirculation rate. At low and medium loads and speeds, the engine is operated with stratified charging and a high exhaust gas recirculation rate. This results in lower fuel consumption. The exhaust temperature determines the operating range at low loads in which the engine can be operated without throttling. In order for the catalytic converter to be able to convert the pollutants, the catalytic converter temperature may not fall below 250°C. A completely unthrottled operation in the idling mode is therefore impossible. Even in a cold start and during warm-up, the engine runs with a homogeneous, slightly lean mixture to achieve a quick start for the catalytic converter. A hot engine can function with throttled stratified charging. At a low partial load and high speeds, it is difficult to attain good mixture preparation with stratified charging because of the short mixture preparation time and the danger of soot formation. A homogeneous mixture with EGR should therefore be preferred. The NOx emissions, and the danger of soot formation, pose limits for stratified charging in the upper partial-load range. In this range, operation with homogeneous charging and EGR has only a slightly negative effect on fuel consumption in comparison to stratified charging with EGR, although lower NOx emissions and there is no danger of soot formation. Homogeneous lean operation will be limited by the exhaust temperature. At temperatures over 500°C, the storage catalytic converters are no longer able to store nitrogen oxides, so that the engine is operated with a stoichiometric mixture and high EGR rates to reduce the NOx emissions and fuel consumption. Exhaust-gas recirculation is not possible when operating at full load. The engine is controlled in the same manner as engines with intake manifold injection, that is, with a mixture for maximum performance and optimal catalytic converter protection.

homogeneous, l < 1 homogeneous, l =1,AGR

Engine cooling water temperature

homogeneous lean, EGR Layered, EGR

cold

Last

homogen, l < 1

Layered, throttled EGR

homogen, l = 1, AGR

homogen, λ =1,05 Rotational speed

Figure 12.19 Operating strategies in the engine map.

Internal Combustion Engine Handbook | 531

6606_Book.indb 531

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

In addition to changing the operating states during the transitions between different load states (such as the transition from stratified charging at partial load to a homogeneous enriched mixture at full load during acceleration), a change between two operating conditions can also be necessary in an unchanging load for exhaust after treatment. The main requirement is a transition without a change in the torque that would be perceived by the driver. To regenerate the catalytic converter during lean operation, increased fuel consumption of up to 3% is to be expected. 12.3.8.2.3 Injected fuel metering The fuel for injection is metered via electronically controlled fuel injectors (Figure 12.20). The fuel in the annular orifice between the valve needle and needle seat is metered by the length of time in which the needle is opened. In principle, the design of the fuel injectors is the same for the intake manifold injection and for direct injection. The differences lie in the design of the magnetic circuit and the maximum flow.

•• Pencil stream valve: The injection jet has a small spray dispersal angle of a maximum of 8°. This type of injector is primarily used for applications in which the fuel injector is installed relatively distant from the intake valve. •• Cone spray valve: The injection jet has a larger spray dispersal angle of 10–30°. This fuel injector is mainly used when the distance to the intake valve is relatively small. The droplet sizes are smaller than with the pencil stream jet valve. •• Split stream valve: The injected fuel is divided into two injection streams. The angle between both jet axes is normally 15–35°. This type of fuel injector is mostly used for multiplevalve engines with two intake valves. In addition to the jet geometry, there is a series of other quantities that must be determined for the fuel injection applications for an engine: •• Static flow: This is the maximum flow through a fuel injector, when maximum current is applied to the coil. It depends on the fuel pressure, the diameter of the holes in the orifice plate at the fuel injector exit and the needle stroke. •• Dynamic flow: Provides a flow for a coil operating time of 2.5 ms. •• Linear flow range: “Linear flow range” (LFR) is the ratio of the maximum and minimum flows with a maximum 5% deviation from the linearity lines (lines through the characteristic of the injected fuel quantity over the operating time of the coil). •• Droplet size: This characterizes the atomization of the fuel injector. The droplet size of a cloud of droplets is usually indicated by the Sauter mean diameter that describes the ratio of the average droplet volume to the average droplet surface in a delimited measured volume. In addition to the average droplet size, the droplet size distribution in the injection jet also has a large influence on the emission behavior of the internal combustion engine. Moreover, the droplet speed is important since it characterizes the penetration depth of the fuel jet when injected into the air, and it characterizes the secondary jet decomposition when the droplets contact a surface.

Figure 12.20 Fuel injector.

The needle is lifted by applying current to the solenoid coil, when the magnetic force on the needle is greater than the force from the fuel pressure, the spring, and the friction. As soon as the flow of current is interrupted in the coil, the magnetic field then starts to decay and the needle closes the annular orifice supported by the spring force and the fuel pressure. After the fuel exits the fuel injector, a jet geometry arises that is a function dependent on the geometry of the fuel injector after the metering annular orifice (above all the needle seat and orifice plate geometry). One makes a differential between the following:

•• Tightness (“Leak rate”): There are strict requirements due to the applicable legislation concerning evaporation emissions. Since it is difficult to determine the leak rate with liquid media, it is therefore evaluated with nitrogen. The leakage may not exceed 1.5 cm3/min.

12.4 Mixture Formation in Diesel Engines Diesel engines operate using internal mixture formation, compare with Section 12.1. At the end of the compression cycle, liquid fuel is injected into the area around the ignition top dead center (TDC) into the highly compressed air. Directly after penetration of the fuel droplets whose average Sauter mean diameter is 2 and 10 μ m, the physical and chemical preparation

532 | Internal Combustion Engine Handbook

6606_Book.indb 532

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

While for whirl chamber engines, whose development was essentially discontinued in the 1990s, the injection pressure level has remained constant at approximately 300–400 bar, for engines with direct injection, the direct injection pressure has risen dramatically over the last 10 years (Figure 12.21). This is also essentially related to the development of highspeed diesel engines with direct injection for passenger cars. Since the available time is very short because of the high speeds in these instances, a correspondingly large amount of mixture formation energy must be made available by a high injection pressure. 2500

Maximum injection pressure

of an ignitable air–fuel mixture begins. The processes of fuel evaporation, the mixture with the air, and subsequent ignition and combustion occur simultaneously. The goal of mixture formation is, on the one hand, to achieve the fastest possible ignition of the fuel vapor and, on the other hand, to burn as completely as possible all the injected fuel while avoiding high peak combustion temperatures. When these two basic conditions are fulfilled, the pollution generated by combustion is very low with the avoidance of extreme pressure peaks and, hence, loud combustion and a high mechanical and thermal load, also compare with Sections 14.3 and 15.1. The air–fuel mixture in the combustion chamber is strongly inhomogeneous, in terms of both location and time. The local air ratio in the combustion chamber ranges from 0 (in fuel droplets) to infinite (zones of pure air). The global air ratio, that is, the ratio of the air mass actually in the combustion chamber to the air mass required for complete combustion of the injected fuel ranges from approximately 1.1 to 6 in practically designed diesel engines. There is only a very short time available for mixture formation in diesel engines. Assuming an injection time of approximately 36° of the crankshaft angle, then there is only 1.5 ms available at a speed of 4000rpm. In comparison, the mixture formation time for a conventional gasoline spark-ignition engine, with intake manifold injection at a comparable speed, is 15 ms. The time from the start of injection to the first ignition of an air–fuel mixture is much shorter. This time, termed as the ignition lag, is up to 2 ms. Ignition lag is strongly influenced by the temperature and pressure conditions in the combustion chamber and the atomization of the fuel. After the first ignition, the additional mixture formation of the noncombustible hydrocarbons with the available air oxygen is accelerated by the beginning combustion and the related temperature increase as well as the arising turbulence. The energy required for the mixture formation comes from either the injection system or the air movement and the starting combustion itself. In engines with a divided combustion chamber (antechamber or whirl chamber enginees), the rich combustion starting primarily in the whirl chamber is responsible for the main energy for mixture formation in the main combustion chamber. Minimal requirements are hereby placed on the injection system; depending on the whirl chamber procedure, the air movement participates to different degrees. In the direct injection system normally utilized today, without a divided combustion chamber, the injection system contributes the majority of energy. In engines with a greater speed range, or injection systems with comparatively low injection pressure, the air is guided so that turbulence arises in the combustion chamber that supports the mixture formation. Increasing the proportion of the air movement in mixture formation therefore relates to lower injection pressure. It must therefore be noted that the generation of swirling air is associated with greater losses in the charge cycle. The fuel injection in the combustion chamber is therefore of prime importance for mixture formation in diesel engines. In addition to other functions, the injection pressure particularly plays the central role.

?

2000

1500 HD engines

1000

Car DI engines

500

Car-chamber engines (IDI) 0 1960

1980

year

2000

2020

Figure 12.21 Development of the maximum injection pressurein recent decades.

When injecting the liquid fuel into the combustion chamber, it is important for the fuel to be distributed in many very small droplets and, hence, present a large surface for evaporation, and also to reach the complete air in the combustion chamber to prevent heavy soot formation because of locally insufficient oxygen. This is accomplished by carefully adjusting the injection pressure, the nozzle aperture geometry, and combustion chamber recess, as well as the correct injection time. The fuel droplets must be prevented from reaching the cylinder wall beyond the combustion chamber recess and therefore collecting in the fire web area between the piston and cylinder. The droplets would then bypass the combustion process, then evaporate, and leave the exhaust pipe as noncombustible hydrocarbons. Figure 12.22 shows the penetration depth of the liquid and vapor fuel over the period after the start of injection as a function of the injection pressure [12-7]. One can thereby clearly see that the penetration depth of the liquid jet is independent of the pressure. The speed of the jet tip is, however, much faster at the maximum injection pressure. The stronger pulse ensures greater air entrainment in the injection jet and, hence, faster evaporation. The larger nozzle aperture at the same injection pressure causes the liquid fuel to penetrate

Internal Combustion Engine Handbook | 533

6606_Book.indb 533

1/19/16 8:43 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

less deeply. It should therefore be noted that, as the drop in size increases with the nozzle hole diameter, the aerodynamic resistance (increased by the square of the drop diameter and the speed) can rise so much that the penetration depth falls with larger holes. A balance is therefore required between the nozzle hole cross section, injection pressure, and air movement. 45

5 x 0,165 mm

Penetration depth [mm]

40 35 30 25

300 bar 500 bar 800 bar 1350 bar 1800 bar

20 15 10

mixture

5 0

liquid 200

400

600

800

1000 1200 1400 1600

Time after ASB [microseconds]

Figure 12.22 Penetration depth of the liquid and vapor fuels, measured in an injection chamber, as a function of the injection pressure [12-7].

In addition to this classic type of internal mixture formation in diesel engines with liquid fuels, there are various special types of diesel engine mixture formation such as in diesel or gas engines and homogeneous diesel combustion that have recently been in subject of intense investigation compare to Section 15.1.2.4. A detailed description of mixture formation in connection with diesel combustion follows in Chapter 15.1.1.

12.4.1 Injection Systems—An Overview 12.4.1.1 Tasks The injection system is therefore largely responsible for providing diesel engines with high exhaust quality, low-fuel consumption, fast response behavior, and smooth running with little noise. Depending on the use of the diesel engine, these stated goals can assume different amounts of importance. The injection system and diesel engine must be correspondingly adapted. The main tasks of the injection system are as follows [12-8], [12-20]: Precise metering of the fuel mass per work cycle. The load of diesel engines is controlled by metering and injecting a variable fuel mass (quality control) which, on the one hand, must be as precise as possible to attain maximum performance and, on the other hand, creates a risk of nonpermitted soot emission when the stoichiometric air–fuel ratio is approached. The more precise and stable the metering of the fuel in a

full-load curve is, the smaller the safety margin from the smoke limit is, that is, the engine performance can be maxed out. The fuel quantity tolerances should be as low as possible under a full load and should not exceed approximately ±2.5%. During idling and in the partial load range, in particular in stationary mode, that is, when there is no intentional control applied, the requirements on the stability of the fuel metering are very high from cylinder to cylinder as well as from injection to injection. The deviations should be less than 1 mg/ injection. It may be necessary to adapt the injected amount of fuel to each cylinder to achieve the desired smooth running. Adapting the injection rate to the operation conditions. The injected fuel mass during the injection process per unit time (injection rate and its change: dm/dt = f(t)) decisively influences exhaust emissions, smooth operation, and fuel consumption. In principle, the injection rate can be influenced by altering the injection orifice cross section of the nozzle and by changing the injection pressure. Despite substantial efforts, a reliable nozzle with a variable spiracle cross section has not been successfully created to date, which leaves only pressure modulation as an alternative. The pressure can be modulated comparatively easily with a slight degree of variation by adjusting th​e cam shape and, therefore, via the piston speed or plunger speed in the high-pressure injection pump. Pressure modulation in accumulator injection systems is an intensive, but nevertheless interesting, solution possibility. However, injection rates can also be changed to a certain extent by pressure stages in the nozzle holder. Figure 12.23 shows altered injection rates during the injection time for a main injection [12-9], [12-10]. In general, a high injection rate and the therefore related large amount of injected fuel at the beginning of injection produces a strong burst of combustion with a high local temperature and, hence, high NOx formation. Multiple injections. Controlling the injection rate during an injection is frequently insufficient to fulfill the given requirements. Increasingly, multiple injections are required with different quantities depending on the operation point in the program map. Figure 12.24 shows a selection of conceivable multiple injection systems, some of which are in use. Figure 12.25 shows multiple injections at the optimum operating point and their share of fuel at three places in the program map. A small, reduced pre-injection substantially reduces the ignition lag for the subsequent main injection and can, therefore, soften the combustion characteristic and which will lead to reduced combustion noise. A secondary injection under high pressure and, hence, with a well-prepared jet is able to oxidize the soot generated during preinjection or increase the exhaust temperature given a corresponding exhaust after treatment system, for example, to regenerate the particle filter. In addition, secondary injection can also support the supply of noncombustible hydrocarbons with a “subsequent” exhaust after treatment. Depending on today’s operating conditions, up to eight injections are applied in car engines.

534 | Internal Combustion Engine Handbook

6606_Book.indb 534

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

6 mm3 °KW 5

1-Cylinder unit, V h ≈ 1,0 l n = 1400 min –1 Fullload

Injection rate

4 3

rectangula triangular “Boot“ shapea

2 1 0 –20 –10 [v. O T]

0 10 20 30 Crankshaft angle

Engine speed:

40 50 [n. OT ]

2000 min –1

mm3 °KW

Engine speed: 2000 min -1

10 20 30 0 Injection duration ∆f [°CS]

with pilot injection

40

10 20 30 0 Injection duration ∆f [°CS]

with two pilot injections

With pre- and two post injections

Split-injection

2,5

1,5 VE1+HE 1,0

VE1+HE+NE

HE

0,5

0

VE1+VE2+HE+NE

VE1+VE2+HE 1000

2000

Engine speed

3000

4000

Figure 12.25 Example of optimum operation-point-dependent multiple injections in the engine map; quantity distribution in percent [12-11].

40

Figure 12.23 Different injection curves (Injection rate = f(t)) for the main injection, top [12-9], bottom [12-10].

Pre- and post-injection

Quintuple injection

VE Pre-injection HE Main injection NE post injection

2,0

Spec. work [kJ/dm3]

mm3 °K W

Injection rate

Injection rate

Radial piston pump Pump nozzle Common Rail

Figure 12.24 Qualitative nozzle needle stroke characteristics for multiple injection (Selection).

Minimum injection functionality. In connection with multiple injections, in particular where the preliminary and secondary injections range from approximately 1 to 3 mg per injection, for example, in passenger car engines, the need for precisely metering the very small quantities increases dramatically. The injected fuel quantity tolerances should therefore be less than 0.5 mg per injection. Since the fuel injector needles always move within the so-called ballistic area for these minimum amounts, that is, they do not reach the mechanical stop, all the manufacturing-related tolerances strongly affect the volume quality. The result is a considerable increase in the requirements for component quality and long-term stability for the fuel injectors, in particular affecting the nozzle needle, the needle seat and the injection orifice. Adapting the injection time. The purely speed-related adaptation of the commencement of fuel supply in systems with

Internal Combustion Engine Handbook | 535

6606_Book.indb 535

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

long fuel injection lines is, as known, insufficient today. Even systems without injection lines, or with electronically controlled injectors, require a freely settable injection start from early, for example, in cold starts, to late in certain areas of the program map to reduce nitrogen oxides. The optimum setting for fuel consumption is desirable in other operating areas. In addition, individual injection commencement intervals between these separate injections must be implemented for multiple injections. The precision for implementing the beginning of injection should be < ±1° Crankshaft angle. Flexible adaptation to operating and environmental conditions. In addition to the aforementioned main tasks, a modern injection system should react flexibly to dynamic processes depending on the air mass. For example, when accelerating under a full load, the fuel quantity should only be adapted to the dynamically measured air mass to avoid undesirable smoke emissions. When the engine speed has achieved, the quantity of fuel must be reduced corresponding to the application range so that diesel engine is protected against excess revolutions (full-load speed regulation). Within the lower operating range, the engine is to be operated at the lowest possible speed both stably and nearly load independent. The respective amount of required fuel is to be adapted quickly depending on the environmental and fuel temperatures so that a rapid engine heat up is achieved. The amount of fuel needs to be adapted to geographical elevations as well. Higher elevations above sea level, where the amount of fuel under a full load should be reduced because of the lower air density to prevent exceeding the permissible smoke limit. When the engine is overrunning, a ramp-like increase in the amount of

injected fuel is required to “catch up” the engine when idling with a reducing speed. The injection masses have to be adjusted to the respective operating conditions independently of the charge pressure and exhaust-gas recirculation. These multifaceted, and sometimes interrelated tasks and demands on the injection systems, can only be managed by electronically controlled systems. Mechanically controlled systems with edge-controlled fuel volume metering cannot fulfill these demands at all, or only when large compromises are accepted. In application areas where the aforementioned increased demands, especially on the dynamic behavior, are not of prime importance, these mechanical robust systems can therefore continue to work satisfactorily. The mechanically controlled and edge-controlled systems have already been discarded for car engines which have to fulfill strict exhaust legislation and, at the same time, utilize as little fuel as possible under dynamic behavior.  esign and Parts D The mechanical system, hydraulics, electrical system, and electronics work together in a modern-day fuel injection system. The overall system can be divided into five subsystems (Figure 12.26). Low-pressure system. The low-pressure system ensures that the fuel from the tank is supplied for the actual high-pressure injection. The required pump can either pump the fuel from the tank to the engine as a tank module or, as suck the fuel from the tank as an integrated pump in the high-pressure pump and prepare it to the required supply pressure for the high-pressure unit within the high-pressure pump. This

• Pump • Edge controlled • Time controlled • Injection process

h pre H ig

ssure syste m

Regulating/controlling

pressure system

ro ction p cess Inje

Low -

• Pump • Filter • Pressure regulating • Volume regulating

• Bucking • Sensors • Diagnosis,OBD • Quantity • Injection start • Driving speed • ABS • EGR • Charge pressure• ASR • ESP • Idling • Transmission • Rotational • speed

• Nozzle-holder assembly • Injector Solenoid/Piezo • Injection process • Hole geometry

Inje em ction t iming syst

• SV device • Actuator • Angle-time acquisition

Figure 12.26 The diesel injection system as the sum of different subsystem.

536 | Internal Combustion Engine Handbook

6606_Book.indb 536

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

feed pressure can be between 1 and 15 bar depending on the injection system and rpm. In a common rail system, a rail-pressure-dependent, low-pressure side control is used to reduce the power consumption of the high-pressure pump. When designing low-pressure pumps, it must be considered that in most cases a substantial volumetric flow of fuel is required to cool the injection components and to compensate for leakage. The utilized filters are both large-pore prefilters as well as fine filters; the latter have an average pore size of 5 to 10 μm, depending on the injection system. The filter also has the task of removing any water that is in the fuel, to prevent corrosion of the injection components. To ensure operation at extremely low temperatures, the filter frequently has an additional electrical fuel heater and/or heated fuel from the high-pressure area is recycled to the fuel before the filter, to prevent the filter from being blocked by paraffin deposits from the fuel. The high-pressure system is essentially characterized by the high-pressure pump. The required high-pressure and injection output is exclusively generated by plunger pumps. Internally or externally supported radial piston pumps and single-cylinder axial piston pumps are also used. Only these pumps are capable of generating pressures greater than 1000 bar stably over a long period of time and also metering the required amounts as needed. In conventional, so-called edge-controlled systems, the actual pump element also has the function of metering precise quantities in addition to the tasks of volume feed and increasing pressure. In modern systems with electronically controlled valves, the high-pressure pump exclusively serves to generate injection pressure and supply the fuel. The exactly metering is assumed by an electronically controllable valve (normally a solenoid valve or Piezo actuator). Injection timing device system. To supply the fuel to the engine at the right time, a so-called injection timing device system is necessary. In general, there is a differentiation between two basic systems. In conventional systems, because of mechanical or hydraulic forces, there is a phase displacement between the pump driveshaft and, hence, the camshaft or ignition distributor shaft of the high-pressure pump and the crankshaft of the internal combustion engine. In systems in which the metering is accomplished by electronically controllable valves, the required adaptation of the injection time can be achieved by a change of the operating time of the valves or by a combination of the operating times and phase displacement. Fuel injector (injection nozzle). The fuel conveyed by the high-pressure pump is fed via the fuel injector to the combustion chamber of the engine. In addition to the injection time and precise metering, the main task of these valves is jet preparation for subsequent mixture formation and combustion. Either the fuel injector can be directly pressure controlled or, corresponding to the pressure generation by the high-pressure pump as is the case with standard common rail (CR) systems, it can be hydraulically controlled by an electronically controllable valve. Recent developments of CR injectors allow direct needle movement by the piezo actuator without the detour via the hydraulics.

The connection between the high-pressure system and the fuel injector is executed with in line systems using highpressure lines. These are steel lines with internal diameters of approximately 1.5 to 2.5 mm. To increase the fatigue limit, it is important for the inside of the pipes to be as smooth as possible without rough recesses, overlaps, and defects. Special procedures are required to achieve this, such as so-called autofrettage, which is plastic smoothing and residual stress generation on of the inside of the pipe at extreme high pressure. Regulating/controlling. The aforementioned four subsystems are coordinated by a regulation and control system. While conventional systems are primarily controlled mechanically/ hydraulically and, in certain cases, even pneumatically, modern injection systems use electronic information acquisition and processing systems and electrically actuated control organs. The electrical actuators are most utilized only for control, the displacement energy is normally generated hydraulically or pneumatically. The electronic control and regulation system is incorporated into the complete engine and vehicle management system and is, hence, connected to all the subsystems from which an effect on the engine torque and/or speed must be executed from. Figure 12.27 schematically illustrates the injection systems characterized by pumps that are used today; the edge-controlled systems are in the top row. The regulation and control can therefore be both mechanical and electrical. In the middle and bottom rows, the systems are illustrated in which the feed is controlled by electrically actuated regulation valves. With the exception of computer-regulated systems (CRS), injection can only occur in the other systems while volume is conveyed and pressure is generated, that is, while the pump plunger is moving. In general, the injection systems used today can be divided into three categories: these are the so-called camedge-controlled, cam-time-controlled, and storage controlled systems (Figure 12.28). In cam-edge-controlled systems,the manipulated variable is the cam angle of the camshaft and/or the delivery stroke of the piston. Since the piston speed increases with increased revolutions, and the thereby increasing pressure (also additionally influenced by predelivery and postdelivery effects in the case of edge control), the delivery angle at high speeds is less than at low speeds assuming a constant quantity. In cam-time-controlled systems, the manipulated variable is the control time. In this case as well, the control time is lower at high speeds than at low speeds due to the speed-related pressure. In the case of storage-time-controlled systems, the pressure can be constantly set by means of the speed and, hence, the control time. The delivery angle therefore doubles itself as the speed doubles.

12.4.2 Systems with Injection-Synchronous Pressure Generation

The main feature of injection systems with injection synchronous pressure generation is that the pressure generation and the fuel delivery, or injection, occurs individually at the right

Internal Combustion Engine Handbook | 537

6606_Book.indb 537

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

Series pump mech./ elektr. .

Hub slide pump Electrical

Pumpes-NW

Plug-in pump mech./elektr.

Pumpes-NW

Engine-NW

Axial piston distributor pump Magnetventil

Pump-nozzle unit Magnetventil/Piezo

Axial piston distributor pump mech./elektr.

Radial piston pump Mechanical Solenoid valve

Pump-line nozzle Solenoid valve

Common-Rail Magnetventil/Piezo

Engine-NW

Engine-NW

Conveying duration/Controlling duration t Conveying angle f

Figure 12.27 

m E = const.; t ~ p = f (n )

p = f (n )

f

t

f nM p = const.

t

t

f

f

2000

4000 min –1

Cam, edge controlled t

2000 4000 min –1 Rotational speed n M Cam, time controlled t

time for each engine cylinder, that is, the pressure is generated at the same rhythm as the engine ignition sequence. Individual pump systems, sequential fuel-injection pumps, distributor injection pumps, and line-free pump nozzle systems work according to this principle. The metering can therefore be executed by pure edge control (mechanically or electronically controlled) or by electrically actuated regulation valves, also compare with Section 12.4.1.

2000

4000 min –1

Storage, time controlled t

Figure 12.28 Breakdown of diesel injection systems for pressure generation and the manipulated variable for metering.

12.4.2.1 Individual Pump Systems with a Line In addition to sequential fuel injection pumps, the individual fuel injection pump system shown in Figure 12.29 with mechanical control is one of the oldest diesel injection systems. This system is characterized in as much as that the drive for the pump plunger is provided by special cams that are on the camshaft for the valve actuation of the engine. This design model permits the use of individual injection pump systems

538 | Internal Combustion Engine Handbook

6606_Book.indb 538

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

(also frequently termed as “plug-in pumps”) only for engines with bottom-mounted camshafts. The system is, therefore, not suitable for modern, high-speed passenger car diesel engines, whose valves are controlled exclusively by overhead camshafts. The main areas of use for mechanically controlled individual pump systems are, therefore, small engines, engines for construction machinery, and stationary engines, as well as large engines for diesel locomotives or ship engines.

1 2

3 4 5

6 7

Figure 12.29 Mechanically controlled single-plunger fuel injection, individual injection pump for large engines [12-12] Type Bosch, 1 Pressure valve; 2 Vent screw; 3 Pump cylinder; 4 Pump piston; 5 Control rack; 6 Control sleeve; 7 Guide bushing.

The attainable injection pressures for large engines lie in the range of 2000 bar. For applications in ship engines, special designs for heavy oil operation are available. The long life and reliability required in these cases lead to a very robust construction with sealed cylinders on one side, so-called blind-hole elements. A great deal of engineering is required to freely adapt the start of delivery and, therefore, fuel injection. Subsequent developments have therefore led to the production of the solenoid valve controlled individual pump system, the so-called pump line nozzle (PLN) system, and also the unit pump system (UPS) or electronic unit pump (EUP). As a result of the shorter fuel injection line between the individual pump and the nozzle-holder assembly, the demands are slight regarding the adjustment of the beginning of delivery, and this can be flexibly accomplished by controlling the start for the solenoid valve on the delivery cam. These systems are, therefore, useful for high-speed running commercial vehicle engines Figure 12.30. In addition, there is a greater need for freely adjusting the start of injection in large engines [12-18], [12-21], 12-22].

Unit pump systems for utility vehicles attain maximum injection pressures of approximately 1800 bar with a potential of 2000 bar. The unit pump system is a further development of the edge-controlled plug-in pump for the cam-time-controlled electronic individual pump system. 12.4.2.2 Inline Fuel Injection Pumps The elements of the inline fuel injection pump consisting of the pump barrel and pump plunger, and corresponding to the number of available engine cylinders, are in their own housing in high-speed aluminum engines. The pump plungers are moved by the pump’s own camshaft, which is itself driven by the timing gear drive of the engine. The fuel metering is executed exclusively via edge control by rotating the pump plunger. Each pump plunger has an angled timing edge, so that a different delivery stroke and, hence, a different amount of injected fuel is delivered or can be set in connection with the cylinder-side, fixed spill port as a function of the angular position of the pump plunger. The entire plunger stroke remains constant and corresponds to the cam pitch. The plunger rotation is executed via a so-called control sleeve, which mates with a laterally movable control rack. The control rack itself is moved by the governor connected to the injection pump. The governor can be either a mechanical governor that primarily shifts the control rack depending on the speed and, in particular, provides full-load speed regulation, or an electronic governor that acts on the control rack by using an electromagnetic actuator mechanism. To adapt the injected fuel quantity to the wide variety of operating conditions, add-on modules are required for mechanically controlled pumps such as a charge pressure-dependent full-load stop, and temperature and quantity-dependent adjusters of the injected fuel quantity. To reliably supply the pump elements with fuel, a lowpressure supply pump is usually mounted on the inline fuel-injection pump that is actuated by a special cam on the pump’s camshaft. The supply pump feeds fuel to the fuel gallery of the inline fuel injection pump at a pressure of up to approximately 3 bar. At the high-pressure exit of the inline fuel injection pump, a pressure valve separates the highpressure area in the pump from the fuel injection line and the nozzle-holder assembly so that, after the line and nozzle system injects the fuel, the fuel in the system remains under pressure, that is, a certain amount of static pressure remains. In addition, a return-flow throttle valve is frequently integrated in this pressure valve, which therefore prevents undesirable secondary injections at only a slight injection pressure. Similar to the situation with mechanically controlled individual injection pumps, inline fuel injection pumps require a comparatively great deal of engineering to freely adapt the start of delivery. The revolution-dependent control of the start of delivery can be achieved with a front-end injection-timing device. This functions with the aid of flyweights and suitable kinematics to yield a phase displacement between the pump camshaft and the engine crankshaft. A simple load-dependent control of the start of delivery is occasionally enabled by a top-timing edge

Internal Combustion Engine Handbook | 539

6606_Book.indb 539

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

1

2

3

4

14

15

5

16

6

17 18

7

19

8 9

20

10

21

11

22 23 24 25

12

26 27

13

on the pump plunger. Hydraulic front-end timing devices are also utilized which can alter the start of delivery in relation to both load and speed within certain limits. Despite these various solution possibilities, the absence of a freely adjustable start of delivery is a disadvantage in conventional inline fuel injection pumps. This has led to the design and construction of the control-sleeve pump. Figure 12.31 shows a section through a pump element of a control-sleeve inline fuel injection pump. Its main feature is that it can move within the area of the piston timing edge of the pump barrel (control sleeve). This enables the adjustment of the plunger lift to port closing i.e. the piston travel until the inlet passage for the fuel. A small plunger lift corresponds to an early start of delivery, a large plunger lift corresponds

Figure 12.30 Pump line nozzle (PLN) or unit pump system (UPS) for commercial vehicle engines [12-12] Type Bosch, 1 Injection nozzle holder assembly; 2 pressure connection; 3 high-pressure line; 4 connection; 5 lift stop; 6 solenoid valve needle; 7 Plate; 8 Pump housing; 9 high pressure chamber (element chamber); 10 Pump plunger; 11 Engine block; 12 Roller tappet bolt; 13 Cam; 14 Spring seat; 15 Solenoid valve spring; 16 Valve housing with oil and magnet core; 17 Armature plate; 18 Intermediate plate; 19 Seal; 20 Fuel feed (low pressure); 21 Fuel return; 22 Pump plunger retainer; 23 Tappet spring; 24 Tappet body; 25 Spring seat; 26 Roller tappet; 27 Tappet roller.

to a late start of delivery. The height of the control sleeves of the individual pump elements are adjusted by a common actuator shaft. The actuator shaft, as well as the control rack necessary for fuel metering, are activated by two separate, electromagnetic actuator mechanisms. Fuel metering is compared with standard inline fuel injection pumps, varying the start of delivery by changing the plunger lift to port closing demands higher cams. For this reason, and because two actuator mechanisms are required, this pump design is used only for commercial vehicle engines. Conventional inline fuel injection pumps with mechanical or electromagnetic control are contrastingly utilized in all engine sizes. The larger inline fuel injection pumps for ship engines are presently controlled only mechanically. The pressure range

540 | Internal Combustion Engine Handbook

6606_Book.indb 540

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

Sleeve

∅h1,2

Control borehole In the sleeve Piston control edge

Effective ∅h2 stroke

h2 Upper disk position

h1

Effective ∅h1 stroke

1 2

Total cam stroke/piston stroke

Cam piston stroke

Conveying start Conveying end

6

7 8

3 4 5

9 10

11

Cam angle

Figure 12.31 Control sleeve inline fuel injection pump [12-8], [12-12] Type Bosch, left: Functional principle for varying the start of delivery, right: Pump with electromagnetic actuator mechanisms: 1 Pump cylinder; 2 Control sleeve; 3 Control rack; 4 Pump plunger; 5 Camshaft; 6 Start of delivery-actuator solenoid; 7 Control sleeve setting shaft; 8 Control rack travel-actuator solenoid; 9 Inductive control rack travel sensor; 10 Plug-in connector; 11 Disc to block delivery start and part of the oil return pump

of inline fuel injection pumps extends from approximately 550 bar for the small inline pump (M-type) to approximately 1350 bar for the control-sleeve pump. Because of the ever-increasing demands for lower exhaust emissions and reduced fuel consumption, and the related increased demands on the fuel injection system with regard to maximum injection pressure, multiple injections, and freely variable start of delivery, the inline fuel injection pump system is increasingly being replaced by solenoid-valve-controlled injection systems. 12.4.2.3 Distribution Injection Pump In addition to the inline fuel injection pump, the distributor injection pump is the second compact pump design. It consists of a low-pressure supply pump, a high-pressure supply pump, a timing device unit, a speed/volume governor and various mechanical and electrical components. The high-pressure pump can either be designed as an axial piston pump or be designed as a radial piston pump. Figure 12.32 shows an axial piston pump with edge-controlled fuel metering and an electromagnetic actuator mechanism. In contrast to the inline pump, this pump is distinguished in that only one pump element is required for all engine cylinders. This is therefore possible because the frequency of the stroke of the pump plungers corresponds to the ignition frequency of the internal combustion engine, and not that of an individual engine cylinder. The pump plunger simultaneously rotates at the camshaft speed. The fuel volume is fed to the engine cylinders by the stroke displacement of the plunger. The rotation enables the fuel distribution to the engine cylinders corresponding to the firing sequence. This dual function of the plunger allows the distributor injection pump to be used

for engines with up to six cylinders. The application areas for this pump are, in particular, for high-speed diesel engines for passenger cars and light utility vehicles. In individual cases, they can also be used for medium-duty class engines. The nozzle-side injection pressure reaches values of 1200 to 1300 bar. A particular advantage of the distributor injection pump is the integration of a solenoid-valve-controlled commencementtiming device. This makes it therefore particularly suitable for engines with a large rpm ranges. Instead of electromagnetic actuator mechanisms, an inertiasupported actuator assumes mechanical control of the timing of the control sleeve and, hence, the fuel metering mechanisms in the older conventional design. A further development of the axial piston distributor pump is shown in Figure 12.33. This can be classified as a cam-time-controlled pump that has a solenoid valve in the high-pressure area, which can then control both the charging of the pump element and the start and end of delivery, thereby the delivery volume. Due to the slight amount of dead space in the high-pressure area of the pump, pressures of approximately 1600 bar can be attained with this variation. An electronic control unit is arranged on the pump body that assumes the pump control functions, in particular, the activation of the fuel solenoid valve and the solenoid valve for timing the start of delivery. The information for this is provided from the pump internal incremental pump angle time signal (IAT-Signal). This is generated by the speed and/or angle-of-rotation sensor on the trigger wheel of the drive shaft. The position of the sensor is jointly altered with the adjustment of the roller ring as necessary for timing the start of delivery. With the cam plate speed (camshaft speed), and the assignment of the delivery stroke to the TDC signal of the crankshaft speed sensor, the start of delivery

Internal Combustion Engine Handbook | 541

6606_Book.indb 541

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

7

6

11

5

4

8

3

can be precisely timed without a needle motion sensor in the nozzle-holder assembly having to provide the information at the beginning of injection. At the same time, the IAT signals can set the exact control time of the fuel solenoid valve and, therefore, the fuel metering

1

9

2

10

Figure 12.32 Axial piston distribution injection pump, edgecontrolled with electromagnetic actuator mechanism [12-12], Bosch design, 1 Distributor plunger; 2 Solenoid valve for jet timing; 3 Control collar; 4 Cam plate; 5 Jet timing; 6 Supply pump; 7 Electrical fuel volume actuator mechanism with feedback sensor; 8 Actuator mechanisms; 9 Electrical shutoff device; 10 Pressure valve holder; 11 Roller holder

1

The pump electronic control unit (ECU) is connected to the second ECU, the engine control unit, or it can contain its functions so that only one more electronic control unit is necessary. The drive and the commencement timing devices are similar to the edge-controlled axial piston distributor injection pump.

5

6

7 2 8

3

4 9

Figure 12.33 Axial piston distributor pump, cam-timed with solenoid valve [12-12], Bosch design, 1 Rotating angle sensor 2 Drive shaft; 3 Mounting ring; 4 Roller holder; 5 Control device; 6 Cam plate; 7 Pump plunger; 8 High-pressure solenoid valve; 9 Pressure valve holder.

542 | Internal Combustion Engine Handbook

6606_Book.indb 542

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

For distributor injection pumps with a radial piston highpressure pump, the pressure generation function and the distribution function are separate. As the name indicates, the pressure generation piston is in a radial position. The fuel volume metering can be controlled by either the cam pitch or the cam time via solenoid valves—similar to the solenoidvalve-controlled axial piston distributor pump. Figure 12.34 shows a cam-pitch-controlled radial piston distributor pump. With this type of pump, the delivery stroke (which corresponds to the respective overall stroke) is altered by axially shifting the ignition distributor shaft and by conical surfaces. The axial position of the ignition distributor shaft is detected by an inductive travel sensor and represents a measure for the delivery stroke and/or the delivery quantity. This pump is also primarily used for passenger cars, similar to the axial piston distributor pump. To attain the maximum injection pressure, a radial piston pump was developed with a higher supply rate and solenoid valve control (Figure 12.35). With this pump, injection pressures of almost 2000 bar can be attained. A cam ring bears inner radial cams whose stroke is transferred via rollers and slippers to the radial delivery plungers that take over the supply of fuel and thereby generate high pressure. The number of supply pistons and the diameter of the delivery plungers determine the fuel delivery rate. The short, direct flow of force within the cam drive produces a certain amount of resilience and, therefore, a very rigid system that allows high injection pressure.

3

7

6

8

9

10

5

4

3

2

1

Figure 12.34 Radial piston distributor injection pump, displacement controlled and electronically controlled [12-13] Delphi model, 1 Pressure valve holder; 2 Distribution shaft, 3 Slippers; 4 Return spring; 5 Supply pump; 6 Pressure retaining valve; 7 Jet timing; 8 Solenoid valve (Jet timing, Return flow); 9 Solenoid valve (Shut-off, charging); 10 Sensor for axial rotor position

The fuel is distributed to the engine cylinder is executed via the rotating ignition distributor shaft, in which the controlling solenoid valve needle is centrally located. The force-generating actuator solenoid lies stationary in the distributor head. The needle is, therefore, geometrically separated into two parts. The sealing seat and the sealing part of the needle are located in the

4

2 5 1

6 7

8

Figure 12.35 Radial piston pump, cam time controlled with solenoid valve [12-14] Bosch design, 1 Drive shaft; 2 Vane-type supply pump; 3 Angle-of-rotation sensor; 4 Controldevice; 5 Radial piston, high-pressure pump; 6 Distributor shaft; 7 High-pressure solenoid valve; 8 Pump outlet

Internal Combustion Engine Handbook | 543

6606_Book.indb 543

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

rotating ignition distributor shaft. The magnetic force is thereby transmitted by the external needle part to the concomitantly rotating needle part. In a de-energized state, the high-pressure valve is opened by spring force; the pump area can thereby fill the radial pistons with fuel via the low-pressure circuit. After current is applied, the valve then closes and the high-pressure delivery begins. The volume metering is executed via the closing and opening time points of the solenoid valve. The commencement-timing device basically has a design similar to the axial piston distributor pump, although its dimensions are adapted to the increased requirements. The solenoid valves for fuel quantity and start of delivery are controlled analogously to the axial piston distributor pump via the pump ECU that, consecutively, receives the required information via the IAT signals as described previously. Due to the high fuel delivery rate of this pump, a so-called return-flow throttle valve in the area of the pump outlet is usually sufficient, that is, it is an open system. Therefore, the pump and line do not have to be joined by a hermetic seal between the individual injections into the engine cylinder. The application areas for this radial piston pump range from passenger car engines to heavy-duty engines. Whereas conventional distributor injection pumps with edge control are not suitable for generating preinjections by interrupting the supply phases, this is possible with solenoidvalve-controlled systems. Figure 12.36 shows the nozzle needle stroke of an injection system with a solenoid-valve-controlled radial piston pump when using a two-spring nozzle-holder assembly. At low and high speeds, one can clearly see the preinjection generated by repeatedly switching the solenoid valve and, in the case of low speeds, the rate-of-discharge curve formed by the two-spring nozzle-holder assembly. One thing in common for all distributor injection pumps is that the crank gear is exclusively lubricated by the fuel in contrast to individual injection systems and inline fuel injection pumps. In the latter two systems, the crank gear, that is, the combination of cams/tappets is lubricated by the engine oil and is, therefore, tribologic insensitive in comparison with the fuel-lubricated distributor pump crank gear. This state of affairs leads to the fact that the utilized diesel fuel must, therefore, fulfill a minimum lubrication standard. The new fuel standard DIN EN 590 [12-15] established in the 1990s ensures this.

12.4.2.4 Pump Nozzle System The so-called pump nozzle units (PNU), also termed unit injector systems (UIS) or electronic unit injectors (EUI), form the high-pressure generating pump element and the fuel injector forms a component. The drive for the pump plungers is executed via the engine’s own overhead camshaft on which special injection cams are arranged. Figure 12.37 shows the PNU system in the cylinder head with the example of a passenger car engine and its design. Because of the absence of fuel injection lines, the fuel volume to be compressed during delivery (dead space) is very small. This therefore enables the system to achieve very high injection pressure. At present, the injection pressure is just over 2000 bar with a potential of 2500 bar. The system can be utilized both in passenger car engines and in commercial vehicle engines. A prerequisite hereby is that the engines have an overhead camshaft. The arrangement of the pump nozzle element in the cylinder head requires a completely new cylinder head design with an integrated fuel supply and removal system, as well as a particularly rigid and robust camshaft drive. While pump nozzle elements were also used with hydraulic and mechanical control in individual cases in the past, solenoid-valve-controlled systems predominate today. A piezo-controlled variant is also standard for passenger car applications. The illustrated system in Figure 12.37 shows how preinjection can be implemented over a wide range of the engine program approve a small hydraulically actuated bypass plunger. After the first time, the nozzle needle is opened, pressure continues to rise, and a bypass plunger is actuated that interrupts the injection and simultaneously increases the nozzle needle opening pressure for the main injection. If the delivery generated achieves this increased opening pressure, then the main injection starts. At present, preinjection and secondary injection (in connection with exhaust after treatment) are implemented by repeated triggering of the solenoid valve. An increase in flexibility for the fuel injection point is required in connection with diesel particulate filter regeneration, something which cannot be fulfilled by the PNU. Figure 12.38 illustrates a summary and qualitative evaluation of the attained injection pressure levels for the different injection systems.

QPI = const. QMI = const.

Low rpm

High rpm

Figure 12.36 Preinjection implemented with a radial piston distributor injection pump in connection with a two-spring nozzle holder.

544 | Internal Combustion Engine Handbook

6606_Book.indb 544

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

28

2 cm

27

1

2 3 4 26

6 9 5 7

25

8

24 23

10 11 12

22

13 14

21

19 20

16 18 15 17

Figure 12.37 Pump-nozzle unit for passenger car engines [12-12] Bosch model, 1 Ball pin; 2 Return spring; 3 Pump plunger; 4 Pump body; 5 Plug-in connector; 6 Magnet core; 7 Compensating spring; 8 Solenoid valve needle; 9 Armature; 10 Coil of the electromagnet; 11 Fuel return flow (low-pressure part); 12 Seal; 13 Inlet holes (approx. 350 laser-drilled holes as a filter); 14 Hydraulic stop (Damping unit); 15 Needle seat; 16 Sealing washer; 17 Combustion chamber of the engine; 18 Nozzle needle; 19 Tensioning nut; 20 Integrated injection nozzle; 21 Cylinder head of the engine; 22 Compression spring (nozzle spring); 23 Storage plunger; 24 Storage chamber; 25 High-pressure chamber (Element chamber); 26 Solenoid valve spring; 27 Camshaft; 28 Rocker arm

Maximum injection pressure [bar]

12.4.3 Systems with a Central Pressure Reservoir Radial piston pump Conventional fuel injection system

Motordrehzahl [min-1]

Figure 12.38 Injection pressures for various injection systems.

Injection systems with a central pressure reservoir are presently referred to as common rail (CR = joint lines). The common-rail injection system allows a freely selectable pressure within given limits. This provides the developer with another degree of freedom for optimizing combustion in contrast to cam-controlled injection systems. The flexibility for being practically unrestricted in setting essential injection parameters has always been a desired goal of diesel injection engineering, and opens up new dimensions to combustion process developers. In addition to the freely selectable injection pressure, there is a fundamental principle for the possibility of multiple injections independent of a cam ramp or contour, as in distributor pumps and pump nozzle systems. Physically, the CR system also permits injection at any selected time. The number of injections, and their time

Internal Combustion Engine Handbook | 545

6606_Book.indb 545

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

point, are essentially limited by the manufacturer’s expenses and efforts for creating the required engine control device. Together with the possibilities for an ideal engine design integration for the system, and the clearly reduced load on the pump drive in comparison with all other cam-controlled systems, this means that the common rail system Figure 12.39 will further increase its popularity for injection systems in diesel engines in the future.

High pressure connection

Fuel return High-pressure control valve (PCV) Fuel feed pump (ITP)

High-pressure pump element

High pressure pipes Piezo-Injectoren

Volume flow control valve (VCV)

Rail

Figure 12.40 High-pressure pump, 1 Fuel resupply pump; 2 Volumetric flow control valve; 3 High-pressure pump element; 4 High-pressure control valve; 5a Fuel feed; 5b High-pressure connection; 5c Fuel return flow. See color section page 1084.

ECU Pressure sensor

to transient pressure alterations, a valve is often integrated in the high-pressure range of the system. On the other hand, the volumetric flow regulating displays better energy efficiency in the whole working area compared to high-pressure regulating. It has at maximum delivery over the whole pump revolution range and pressure range of 200–1800 bar an efficiency of between 70% and 90%, whereby there are large areas with over 80% effective energy utilization. The advantages of this approach are especially evident in partial delivery. Over the entire range of the injection pressure and fuel delivery rate, the efficiency down to the lowest pump speeds is over 50%. With a high-pressure blow-off approach this can correspond, with the same boundary conditions, to an energy efficiency of below 20%. In volumetric-flow-regulated high-pressure pumps, the influence of the fuel delivery rate on torque fluctuations (Figure 12.41) of the pump drive is an important factor. This illustrates the rail pressures of 500 and 1500 bar at a pump speed of 1000 rpm. It is representative for the entire pressure range that, as the fuel delivery rate drops, the mean torque falls in connection with a moderate rise in torque fluctuations. The drive unit of the common-rail high-pressure pump is considerably less critical compared with cam-controlled systems.

High-pressure pump

Figure 12.39 Common-rail system.

12.4.3.1 High-Pressure Pump The early stages of development for high-pressure pumps concentrated on the three-piston pump (Figure 12.40). Two- and one-piston pumps are currently also utilized. This enables an optimization for weight and costs. At present, two fuel-metering approaches are used: •• high-pressure blow-off •• the volumetric flow approach. In contrast to a high-pressure blow-off approach, the volumetric flow approach displays two advantages. On the one hand, the volumetric flow control combines has a lower injection of heat into the system by returning the fuel into the tank in the program map. To be able to react quickly

500 bar

1500 bar

Full conveying 25 Nm 20 15 10 5 0 25 20 15 10 5 ms 0 0 20 40 60 80

50 % Conveying

0

20

40

60

25 % Conveying

80

0

20

40

60

80

Figure 12.41 Supply rates and torque fluctuations.

546 | Internal Combustion Engine Handbook

6606_Book.indb 546

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

1,6

Full conveying

50 % Conveying

25 % Conveying

1,2 kbar

0,8 0,4

0 1,52 kbar 1,51

1,50

1,49

1,48

ms

0

30

60

90

0

30

60

90

0

30

Figure 12.42 clearly illustrates the influence of the volumetric flow control at 1500 bar on the pressure fluctuations in the rail, which can impair the injected fuel quantity. Just as with torque fluctuations in the drive, the fuel delivery rate has practically no influence on the pressure pulsations in the rail as there is no real influence. Given a rail pressure of 1500 bar and maximum delivery, pulsations result as a range of ±5 bar. Given the boundary conditions of a very low fuel delivery rate of 25%, the width of the fluctuations increases to approximately 15 bar. An example of this are the common-rail high-pressure pumps that control pressure via a high-pressure bypass, and those that control pressure via a volumetric flow control valve in reference to pressure oscillation behavior. Both control approaches produce similar results. The concern that controlling pressure by means of a volumetric flow valve could induce impermissible pressure oscillations in the rail is, therefore, unfounded. Accordingly, the aforementioned advantages of pressure control can be exploited with a volumetric flow control valve, without disadvantages. A comparison of the pressure pulsations in the rail, with high-pressure bypassing and volumetric flow control is shown in Figure 12.43. Proportional directional-control valves are used as valves for volumetric flow control, and proportional pressure limiting valves are used for high-pressure blow-off applications. The fuel can be predelivered to the high-pressure pump via an electrical presupply pump (e.g., integrated in the fuel tank)

1,6

High pressure bypassing

60

Figure 12.42 Pressure pulsations in the rail with different fuel delivery rates and volumetric flow control.

90

or a mechanical supply pump that is separate or integrated in the high-pressure pump. The advantage of the first solution is that, after the fuel tank has been emptied, the system can be rapidly refilled, whereas the presupply pump integrated in the high-pressure pump housing has the advantage that there are fewer components in the injection system, and the overall fuel system can be cheaper. 12.4.3.2 Rail and Lines The rail (Figure 12.44) serves as a high-pressure reservoir for the fuel which is delivered by the high-pressure pump. Furthermore, it supplies the injectors with the necessary amount of fuel for all operational conditions. The rail is designed in such a way so that it can, on the one hand, be tensioned quickly to the required pressure and, on the other hand, it can be rapidly damped by the triggered vibrations caused by the injection procedure. The length and the diameter of the lines between the rail and injectors are also to be selected accordingly. The goal of such a design is to have a similar pressure for each cylinder at the same engine operation time during each injection, since otherwise the injector-to-injector spread can become too great due to the timing, which can cause problems with emissions and driving dynamics for a vehicle. Rails are either forged or welded designs from drawn steel. Special attention must hereby be given to the high-pressure fatigue strength of the welded connections. The connection to the lines should be created in such a way so that no tension arises in the weld seam.

Volumetric flow control

1,2 kbar 0,8 0,4 0 1,52 1,51 1,50 1,49 1,48

0

30

60

ms 90

0

30

60

90

Figure 12.43 Comparison of pressure pulsations in the rail with high-pressure bypassing (left) and volumetric flow control (right).

Internal Combustion Engine Handbook | 547

6606_Book.indb 547

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

High-pressure pipes to the injectors

High-pressure line to the variable high pressure pump DCP

Figure 12.44 High-pressure rail.

The lines between the pump and rail, and between the rail and injectors, are manufactured from seamless drawn steel. 12.4.3.3 Injector Figure 12.45 shows the layout of the common rail injector using the example of a servo-valve actuated piezo injector.

Figure 12.45 Common-rail piezo injector. See color section page 1084.

The heart of the injector is a piezo actuator that, on the on hand, allows relatively low electrical voltages while, at the same time, satisfies the automotive requirements regarding temperature and vibration. The actuator is able to open or close the servo valve in less than 100 μs. Together with the harmonized input and output throttle combination to the control area above the nozzle needle, the nozzle opening speed can be influenced and therefore the rate-of-discharge curve and also the minimum injected fuel quantity that is determined by the minimum operating time. These processes are triggered with practically no response time. This clearly indicates that piezo engineering enables highly reproducible fuel injection. An executed example of such an injector is shown in the enlarged sectional view in Figure 12.46. The piezo actuator (4) is a multilayer stack in the so-called multilayer technology, where numerous individual ceramic platelets are joined together. These are subjected to initial pressure in housing. A problem that has to be hereby solved is temperature compensation. Because of the large range of temperatures in a vehicle, the expansion of the housing compared to the longitudinal expansion of the ceramic platelets under tension is large. Temperature compensation is executed by selecting a suitable material for the installation housing surrounding the piezo stack, together with the pretension spring, as well as appropriately setting the empty displacement of the actuator. Hereby it may not be that the injector remains open (too little play) which can cause engine damage. Alternatively, the injector cannot remain closed during very small control times (too much play), which will noticeably increase combustion noise in the absence of pilot injection. Another peculiarity is the servo-valve, which opens inwards to the high-pressure area instead of outwards in contrast to a magnet-actuated valve. The reason for this is that the piezo element expands and also exerts a large outward force when a voltage is applied. This makes therefore makes it more functionally appropriate for the piezo element to open the valve against high pressure

548 | Internal Combustion Engine Handbook

6606_Book.indb 548

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

and yields a simpler injector design than if the movement is in the opposite direction when voltage is applied to the piezo element when the servo-valve is being closed. Overall, this actuator design allows a regulation valve lift of approximately 40 μm to be maintained over the entire temperature range of a vehicle engine from −30°C to +140°C. The functionality of this design can be seen in Figure 12.47. If the injector is not controlled (left half of the figure), there is fuel at both the high pressure of the rail and in the injector control area (2) as well as in the high-pressure chamber (3) of the nozzle. The hole for the fuel return flow (5) is sealed by the valve end pad (4) with a spring. The hydraulic force, which is exerted by the high fuel pressure on the nozzle needle (6) in the control area (2) (F1), is greater than the hydraulic force acting on the nozzle tip (F2) since the area of the control plunger in the control area is greater than the free area under the nozzle needle. The nozzle of the injector is closed. If the injector is controlled (right half of the figure), then the piezo actuator (7) presses on the valve plunger (8), and the valve end pad (4) opens the hole that connects the control area (2) with the fuel return. This causes the pressure in the control area to drop and the hydraulic force, which acts on the nozzle needle tip (F2), is greater than the force acting on the control plunger (F1) in the control area. The nozzle needle (6) moves upward, and the fuel passes via the injection orifices into the combustion chamber of the engine. When the engine is at a standstill, the valve that connects the control area with the fuel return is closed by spring force along with the injector nozzle. Figure 12.48 shows the performance ability of the piezo injector. Engine sizes with a cylinder volume of 0.5 l have a beneficial overall adjustment for the injectors,

3

4 1

5 2 6

1 High pressure connection 2 Fuel return 3 Electrical connection to the engine control unit (ECU) 4 Piezo-Actuator 5 Valve piston 6 Valve head 7 Control piston 8 Nozzle needle 9 High-pressure chamber nozzle 10 Nozzle injection holes

7

8

9 10

Figure 12.46 Sectional view of a piezo injector. See color section page 1085.

Piezo actuator not controlled

Piezo actuator controlled 7

4

4

1

2

2 5

8 F1x

F1 3 6

1 2 3 4 5 6

High-pressure inlet Control chamber High-pressure chamber Valve head Fuel-return Nozzle needle

6

F2

F2x

7 Piezo-Actuator 8 Valve piston F1 Force on the control piston F2 Force acting on the nozzle needle Figure 12.47 Injector function. See color section page 1085.

Internal Combustion Engine Handbook | 549

6606_Book.indb 549

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

150

1,6 1,2 0,8 0,4 0,0

Quintuple injection n = 2000 min–1 pRail = 1000 bar

0

300

600

900

1200

Rail pressure [bar]

1500

100 50

350 250 150 50

0 –40

1800

–30

–20

–10

0

10

20

Crankshaft angle [°KW]

30

Nadenhub [μm]

2,0

Zylinderdruck [bar]

Injection quantity [mm3/Hub]

Minimal injection quantity

40

Figure 12.48 Performance of second-generation CR injectors with piezo technology.

that is, with sufficiently fast opening and closing flanks, up to 1800 bar minimum injection volume from under 1.5 mm3, in areas with lower pressure even considerably under 1.0 mm3. At the same time, the distance of the start of injection from pre-injection to the main injection can be very short. Depending on the rail pressure, minimum injection breaks of 0 to 250 μ s are possible. Larger injection commencement distances are not subject to any limitations. Preinjection and postinjection are both possible in the complete program map range, that is, either in the complete pressure range or also in the complete revolution spectrum. 12.4.3.4 Injection Nozzle The task of the injection nozzle is to atomize and distribute the fuel to attain the desired micromixture and macro-mixture.

The injection nozzles that are used in the common-rail injection system are seat-hole nozzles and blind-hole nozzles (Section 12.5.4). 12.4.3.5 Electronics The following system block diagram, Figure 12.49, illustrates the sensors and actuators. This clearly illustrates the complete functional scope of the common-rail injection system. All driver stages, including energy recovery, are integrated within the engine electronics. In contrast to solenoids, piezo technology requires a totally new driver stage approach. Whereas with solenoids current flows during the entire valve-opening phase regulated by peak and hold, the piezo actuator is electrically similar to a capacitor. The piezo element is charged

Crankshaft Camshaft Accelerator pedal Injectors

Rail pressure

Pressure regulating valve

Charge air cooler

Charge pressure Ambient pressure

Intake throttle valve

Air temperature Air Mass Cooling water temperature Terminal 15 Clutch

Rail

+

FGR Keypad

ABS/ASR

Driving speed

exhaust-gas recirculation

EPC

VTG

Relais

Electric fuel pump

Relais

Glow relay +

Automatic transmission Climate compressor

EPC

Fuel tank

M

Diagnostic lamp Service tester

Additional end stages

Figure 12.49 System block diagram for CR.

550 | Internal Combustion Engine Handbook

6606_Book.indb 550

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

Piezo

Solenoid Voltage

Voltage

Current

Current

Figure 12.50 A comparison of the electrical properties of solenoids and piezo elements.

and thereby lengthens itself, it is discharged and returns to its initial length at the end. A comparison of the electrical properties of solenoids and piezo elements is shown in Figure 12.50. With controlling, the trend covers the reversal driver stage with fixed current form back to the so-called CC-final stage, when the current profile can be individually regulated differently depending on the requirement (CC = current control). Another aspect of piezo technology is the electromagnetic compatibility. In principal, one expects strong voltage peaks are expected because of the fast switching times. However, since the piezoelectric actuator can be charged and discharged with a sine (and/or sinusoidal) current form, this technology proves itself to be not more critical with regard to EMC as the solenoid technology with pulsed peak and hold phase. Piezo technology can also be utilized for energy recovery. For example, approximately 50% of the utilized energy can be recovered in piezo technology within the boundary conditions for extremely fast switching. The absence of magnetic remanence allows not only the high degree of repeatability in a piezo actuator from shot to shot, rather more it also allows very close individual injections to be rapidly fired to form an injection series, which allows the targeted control of combustion. The timed intervals between the injections are only limited by the speed of the driver stage. 12.4.3.6 Development Trends General developmental trends, also for the common-rail injection system of the future, are the following: •• increasing the injection pressures •• flexible injection rate control •• interchangeable nozzle injection hole technology •• increased utilization of closed-loop control strategies •• reduction of tolerances. In the first instance, one certainly has to consider a further increase in the injection pressure for an improved fuel preparation and combustion. Piezo technology provides an optimum prerequisite for a flexible injection rate regulation, which will be utilized more and more in the future for fulfilling the emission requirements which will be introduced. Injection nozzles with

variable jet-hole geometry (e.g., two-stage nozzle, TSN) will be developed for ideal mixture preparation in the complete engine program map. Furthermore, closed-loop strategies will also be increasingly implemented in injectors. One example is the piezo actuator that directly actuates injectors with motion reversal, whereby, on the one hand, opening and closing of the nozzle needle can be controlled by feedback signals by a piezo actuator utilized as a sensor and, on the other hand, also the steepness of the flanks and the needle lift in its height can be specifically modified and regulated. Fundamentally, the demands on component tolerances and the smallest possible volume capabilities will have to increase further to fulfill the emission levels (Figure 12.51). With piezo technology, with given adapted injection orifices, preinjections of 0.8 mm3 will be feasible.

12.4.4 Injection Nozzles and NozzleHolder Assemblies

The fuel delivered by the pump element is injected through the injection nozzle at a high pressure into the combustion chamber of the diesel engine and distributed as finely as possible. The nozzle itself is mounted in a nozzle-holder assembly that is screwed or inserted into the cylinder head as a sealed component. In cases of pump-nozzle units, the high-pressure element, nozzle-holder assembly, and nozzle form a constructive unit. With common-rail systems, the injector serves as a control element and also controls the functionality of the nozzle-holder assembly. The primary tasks of the nozzle, in combination with the nozzle-holder assembly, are to form the rate-of-discharge curve, atomize and distribute the fuel in the combustion chamber, and seal the hydraulic system from the combustion chamber. The nozzle construction and design need to be precisely harmonized to the different engine conditions. These are primarily •• combustion processes (DI, IDI) •• geometry of the combustion chamber •• number of injection jets, the spray shape, and spray direction •• injection time •• injection rate.

Internal Combustion Engine Handbook | 551

6606_Book.indb 551

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

250

Needle stroke [μm]

200

150

100

50

0

time[ms] 0

1

2

3

Figure 12.52 shows a few basic designs of pintle nozzles (for IDI Engines) and hole-type nozzles (for DI Engines). In all cases, this concerns nozzles that open inwards. Outwardopening nozzles are no longer utilized as serial, mass-produced solutions. The pintle profile design in pin nozzles can be used to adapt the nozzle stroke-related opening cross section and, hence, the fuel flow or rate-of-discharge curve to engine requirements. In so-called flattened pin nozzles, a larger duct is released so that combustion chamber-side coking of the nozzle needle can be reduced. For the reasons of strength, the design of the nozzle cone shape in hole-type nozzles is very important. In addition, the size of the residual volume between the tip of the nozzle needle

4

Figure 12.51 Jet symmetry over the needle stroke at 1000 bar.

and the inner contour of the nozzle body between the nozzle needle seat and the injection orifice is important because of the fuel volume inside that does not participate in combustion. The smaller the volume, lesser the hydrocarbons evaporate from this volume, which can then be found in the exhaust as un-combusted HC emissions. So-called blind-hole nozzles usually have a greater cone strength and greater residual volume than seat-hole nozzles, in which the injection orifice is found in the area of the nozzle seat. The residual volume is thereby separated from the combustion chamber, and only the fuel remaining in the individual injection orifices can evaporate. Most injection orifices themselves are presently made by electro-erosion and are no longer mechanically drilled.

Throttle pin nozzle Flat pin nozzle with conica polished section

Hole nozzle with a conical blind hole

Hole nozzle with a cylindrical blind hole

Seat-hole nozzle

Figure 12.52 Basic designs of injection nozzles.

552 | Internal Combustion Engine Handbook

6606_Book.indb 552

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

Seat-hole nozzle (VCO)

Micro seat-hole nozzle

Alternative production processes, such as laser applications, are being developed for the future requirements for even smaller nozzle holes (∅ <100 µm). Increasing injection pressure, and allowing for the fact that the volume distribution per injection hole in a blind-hole nozzle is more even by with a seat-hole nozzle (VCO), there is an increasing tendency toward pollutant-volume reduced blind-hole nozzles (microblind-hole nozzles). Figure 12.53 shows a comparison for both design types. The even spray ratio of the microblind-hole nozzle id particularly important for the smallest volumes for preinjection and postinjection. In seat-hole nozzles, an uneven injection pattern of preinjected fuel volumes (<1–3 mm3 per injection) can arise when the strokes are very small due to manufacturing-related tolerances. If the seat is moved back and positioned in the direction of flow, uneven cross sections in the seat area and, hence, uneven pressure do not as strongly affect injections in which the seat throttle predominates (minimum strokes), since the injection orifice does not start directly in the seat area. Figure 12.54 compares the individual injected fuel quantities per injection Hole 1 120 %

Hole 6

nP = 725 min–1

100 % Hole 2 80 %

Hole 5

Hole 3

Hole 4

Qges = 7 mm3/E

Qges = 40 mm3/E

Figure 12.54 Individual injected fuel quantities per injection orifice for a blind-hole nozzle with two different volume levels.

Figure 12.53 Seat-hole nozzle and microblind-hole nozzle.

orifice for small and medium injection volumes, determined with a measuring procedure according to [12-26]. The number of injection orifices strongly depends on the combustion behavior and the air circulation (turbulence) in the combustion chamber. In principal, the following is valid: in general, the greater the turbulence, the fewer injection orifices that are thereby necessary and vice versa. Today, it is normal to implement 6 to 12 injection orifices in engines with direct injection. In passenger car diesel engines, the number is normally 6 to 8 holes. The dimensioning of the injection orifices creates a target conflict for the injection nozzles for diesel engines that are currently mass produced. Relatively low injection volumes are required for partial-load areas and/ or for preinjection and postinjection, and volumes that must be atomized through the smallest possible injection orifices to achieve a reduction in exhaust values. Hereby, the best fuel atomization is achieved when the total pressure reduction, and thus, the conversion into kinetic energy, takes place exclusively as possible in the injection holes. The small injection orifices, however, prevent the required fullload fuel quantity from being injected into a predetermined, optimal time window. Therefore, a compromise has to found for the design of the injection orifice diameter which, on the one hand will fulfill the demands and/or exhaust emissions and, on the other hand, be sufficient to achieve the maximum engine power. The required minimum injection quantities are therefore achieved in certain areas of the program map, as the nozzle needle is not fully opened, and thus, the flow is limited in the needle seating area. However, this also means a reduction, and thus, pressure loss before the spray holes because the complete pressure is no longer available for energy conversion into kinetic energy. In order to resolve this conflict there are nozzle concepts which display variable, or staged, injection cross sections. Various design solution-applications for variable nozzles exist, one of which is the Coaxial-VarioNozzle (CVN), which is also defined as a Two Stage Nozzle (TSN), and which is illustrated in Figure 12.55. Hereby, the nozzle tip is provided with rows of injection orifices, which are controlled separately by an inner and an outer needle. The required minimum quantities for preinjection and postinjection can be optimally designed through less, or smaller, injection orifices in the first injection orifice level. The secondary injection orifice row will be additionally opened

Internal Combustion Engine Handbook | 553

6606_Book.indb 553

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

Outer and inner needle closed

Inner needle opened; outer needle closed

A

B

Cylindrical injection port with 10% HE-rounding (CF = 0; HE = 10 %)

Outer and inner needle opened

Figure 12.55 Coaxialvario-nozzle (CVN).

C

Conical injection port with 10% HE-rounding (CF = 1,5; HE = 10 %)

for the full-load range, which is equipped with multiple or larger injection orifices. The standard minimum hole diameter, which is represented today, is approximately 0.10 mm. More important than the diameter is, however, the flow through the nozzle holes. The nozzle holes have sharp edges at the start of the nozzle hole after eroding, that is, on the internal face of the nozzle. Rounding the inlet edges would achieve a considerable increase of the hydraulic throughflow with the same hole diameter. If the injection drilling could also be executed conically, then an even speed control profile would be created and cavitation zones would be prevented. Figure 12.56 illustrates cylindrical and conical injection orifice designs with different rounding degrees, as they are standardly applied. Another possibility of evening out the injection jets for each nozzle hole is to improve the guidance of the nozzle needle by means of a dual-needle guide (Figure 12.57). 12.4.4.1 Nozzle-Holder Assembly The nozzle is, as previously mentioned, installed in the nozzleholder assembly. The nozzle needle is closed by the initial pressure from the compression spring in the nozzle-holder assembly. If the hydraulic force (proportional pressure and (d2Needle–d2Seat)) exceeds the initial force, then the nozzle opens. In principle, the fuel must overcome two throttling points. Firstly, the stroke-related seat throttle (variable throttle), and subsequently, the fixed restriction characterized by the injection orifice geometry. For smaller strokes, the seat throttle predominates. When the nozzle is completely open, and the nozzle needle lies on the mechanical stop, the injection orifice geometry determines the flow cross section. Figure 12.58

Conical injection port with 20 % HE-rounding (CF = 1,5; HE = 20 %)

Figure 12.56 Nozzle-hole inlet with, and without, hydroerosive rounding.

Nozzle body

Nozzle needle

Double needle guide

Figure 12.57 Dual-needle guide of a seat-hole nozzle

554 | Internal Combustion Engine Handbook

6606_Book.indb 554

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.4 Mixture Formation in Diesel Engines

800

Throughflow [cm 3/30 s]

700

Limited by pump capacity

600 500 400 300 100 bar 500 bar 1000 bar

200 100 0 0.00

0.05

0.10

0.15 0.20 Needle stroke [mm]

0.25

0.30

Figure 12.58 Volumetric flow of a nozzle-holder assembly (seathole nozzle) that depends on the nozzle needle stroke for three different pressures.

shows the characteristic nozzle flow of a hole-type nozzle. The seat throttle determines the through flow in the area of the smallest strokes. This steeply increases with the stroke. In this area, the so-called ballistic area of the nozzle needle movement, the manufacturing tolerances and set tolerances play a particularly large role. To shape the rate-of-discharge curve, particularly in edgecontrolled injection systems, the two-spring nozzle-holder assembly is frequently used (Figure 12.59). In this case, the

15 14 1

4 5

•• smoke limit (especially at a full load)

1

12 11 2 10 3 9 3 4 5

8

12.4.5 Adapting the Injection System to the Engine

So that diesel engines can provide the best results at every working point corresponding to requirements, the entire injection system must be exactly adapted to the engine. One refers to the application of the injection system to the engine. To best resolve the individual tasks defined in Section 12.4.1, numerous geometric parameters of the injection system components and working-point-related input variables of the injection system must be determined and implemented corresponding to the target values. The electronically controlled systems therefore offer many more degrees of freedom and possibilities for optimization when compared to conventional, mechanically regulated systems. For example, the following important engine-related and vehicle-related restrictions must be observed when determining the fuel mass to be injected:

13

2

initial force of the spring is initially overcome in the top part of the nozzle-holder assembly. The opening pressure is approximately 120 to 180 bar and, after passing through a few 100ths of a millimeter plunger lift, the initial force of the second (lower) spring will be additionally overcome. The opening pressure of this second stage is then from 250 to up to over 300 bar. Therefore, this makes it possible, in the lower revolution ranges, to achieve an “adherent preinjection” (Figure 12.36). At higher revolutions, the pressure builds up so strongly and quickly that the first stage is immediately overcome, and a normal needle lift characteristic arises. In the common-rail injectors that are built today, the nozzle needle is opened and closed purely by hydraulic force, or the direct mechanical connection between the actuator and nozzle needle, also compare to Section 12.4.3. Other systems are also utilized (CVA), whereby injection process applications are possible by pressure modulation during the injection process with common-rail injectors. In addition, developments are already recognized in which the nozzle-hole cross section changes depending on the needle stroke, so-called Vario nozzles or register nozzles. The advantage of these different constructions is that only a small cross section is exposed and the jet preparation is also improved while the engine is idling or under a partial load (Figure 12.55).

6

7

•• maximum permissible cylinder pressure •• exhaust gas temperature

6 7

Figure 12.59 Conventional nozzle-holder combination (left) and two-spring nozzle-holder assembly with an integrated needle motion sensor to commence the injection start (right) [12-24] left: 1 Edge filter; 2 Inlet passage; 3 Pressure pin; 4 Intermediate disk; 5 Nozzle retaining nut; 6 Head thickness; 7 Nozzle; 8 Locating pin; 9 Pressure spring; 10 Compression spring; 11 Fuel-leak hole; 12 Fuel-leak connecting thread; 13 Retaining element; 14 Connection thread; 15 Sealing cone right: 1 Retaining element; 2 Needle movement sensor; 3 Pressure spring 1; 4 Guide washer; 5 Pressure spring 2; 6 Pressure spring; 7 Nozzle retaining nut.

•• engine revolutions •• torque and revolution upper limits. The required injection volume per work cycle and cylinder for the four-stroke engine is calculated using the following equation: VK =

Pe ⋅be ⋅ 2 z ⋅ nM ⋅ rK

(12.3)

Pe: Effective performance of the engine be: Specific fuel consumption of the engine (mass/performance and time)

Internal Combustion Engine Handbook | 555

6606_Book.indb 555

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

z: Number of cylinders

VOff = AD ⋅ ∆t ⋅a ⋅

nM: Engine speed



ρ K: Fuel density. The implementation of these requirements by the injection system depends on the fuel metering principle. In conventional edge-controlled or directly lift-controlled pumps, the volume Vstroke per pump stroke depends only on the cross section of the piston and the size of the effective stroke: VHub = Apiston ∗ hutilization



(12.4)

APiston: = Cross-sectional surface of the pump plunger hUtilization: = Effective delivery stroke of the pump plunger. In cam-time-controlled systems, the delivered quantity per injection is dependent on the closing time of the solenoid valve, the plunger cross section, and the piston speed:

Vstroke = Apiston ∗ vpiston ∗ ∆tSD

(12.5)

APiston = Cross-sectional surface of the pump plunger vPiston

= Mean speed of the pump plunger during the delivery period

∆tSD

= Closing duration (≙ delivery time) of the control valve.

The predelivery and postdelivery effects, as well as the delivery during opening and closing of the high-pressure solenoid valves, are hereby not included. The fuel volume exiting the nozzle is contrastingly simplified by the formula.

2 ⋅ ∆p rK

(12.6)

AD = geometrical nozzle-hole cross section ∆t = Injection time

α = Throughflow rate ρ K = Fuel density ∆p = Differential pressure (fuel-side to combustion-side). The injection time can be calculated in a simplified form from the needle lift signal, or the control time of the injector sided, electrically actuated valve in the Common-Rail systems. It must be hereby considered that the metering cross section of the nozzle, the differential pressure between the inside of the nozzle and the combustion chamber during the injection phase are not constant. The same applies for the flow coefficient. In addition, it must also be considered when designing the high-pressure pump, which the fuel at high pressures of over 1000 bar cannot be assumed to be incompressible. The highpressure pump must, therefore, be larger in reference to its delivery capacity, depending on the dead space and available pressure and temperature levels to cover the “storage behavior” of the fuel volume to be compressed. Various powerful tools are available for numerical simulations for injection systems and for determining processes that cannot be measured, or can only be measured with difficulty, in the pump, lines, nozzle, solenoid valve or injector [12-18], [12-28], [12-29], [12-32], [12-33]. Figure 12.60 provides a generalized overview of the hardware side adjustment parameters of the injection system for the

a) Cam-driven injection systems

Injection process, Injection duration, Injection pressure

Droplet size Distribution in the combustion chamber

b) Common Rail System

Injection process Injection duration Injection pressure

Droplet size Distribution in the combustion chamber

Cam speed Element-∆ Time response of the solenoid valve, when necessary Pressure relief valve/degree of relief Pressure valve spring Number Dead volumes (total equipment) Injection ∆ Pressure line length and-∆ hole hydraulic throughflow Needle Length Opening pressure stroke Spring rigidity (nozzle holder) Pressure stage (closing pressure) Injection hole design Injection pressure High-pressure pump Rail volume Rail pressure Inlet throttle in the injector Outlet throttle in the injector Time response of the actuator Pressure fluctuations in the injector Dead volumes (total equipment) Injection Pressure line length and-∆ hole hydraulic throughflow Needle Nozzle pressure stage stroke Injection hole arrangement Blind hole design Injection pressure

Number ∆ Length

Figure 12.60 Important design arrangements and adjustment parameters for injection systems [12-18].

556 | Internal Combustion Engine Handbook

6606_Book.indb 556

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

engine application for cam-driven injection systems and for a common-rail system. In addition, there are the operating parameters such as temperature, charge pressure, air pressure, exhaust-gas recycling, as well as information from other vehicle systems, for example, ESP or ASR as well as the driver’s wishes (gas pedal position, cruise control), and information from the sensors of the exhaust-gas recirculation system. In the past, all of these tasks were only accomplished by “mechanical engine management” in the form of the mechanical control of diesel injection. None of these requirements could be implemented from the vehicle and engine operation, and the driver’s wishes. The mechanical control was restricted to the basic functions of operating the engine, such as idling control, maximum speed governing, full-load torque control, manifold pressure-dependent fuel quantity compensation, atmospheric-pressure-dependent full-load torque control, and temperature-dependent fuel quantity compensation (e.g. while starting). Only with the introduction of electronic diesel control could the aforementioned requirements be comprehensively covered in the application. The electronic diesel control (EDC = electronic diesel control) can be divided into three system blocks: 12.4.5.1 Sensors and Target Value Adjusters They detect the operating conditions of the engine and the target values, and convert physical quantities into electrical signals so that they can be processed in the second block, the electronic control unit. In the electronic control unit, this information is processed according to mathematical guidelines (control and regulating algorithms). The electronic control unit also provides the electrical output signals for the actuators, and is therefore the interface for other systems and for diagnosis. The third block consists of the actuators (positioning elements). They reconvert the electrical output signals of the electronic control unit into mechanical quantities, such as triggering the solenoid valve for metering fuel. Figure 12.61 shows the complete overview of the EDC systemfor a PDE-System for a passenger car. The basic design of an electronic diesel control is illustrated in Figure 12.62. Strictly speaking, diesel control covers not only control but also regulation, since in many cases the actuators are activated based on input variables by predetermined data program maps or characteristics, without the reaction directly being checked. On the other hand, in a series of cases, reactions, such as the speed of the engine in the idling speed regulation system and the nozzle-needle movement in the injection start regulation system are measured and used to activate the actuators. The electronic control unit in the electronic diesel control system is in principle, strictly speaking, a control and regulation unit. For further details concerning electronic engine management, compare with Section 16.

12.5 Fuel Supply Systems Supplying a passenger engine car with fuel requires the provision of a fuel tank, which is normally located in the area of the rear axle of the vehicle. The tank is a component of the fuel supply system with numerous functions. This functions include et alia the tanking of the fuel, filling level limiting, storing the fuel, supplying the fuel for the engine and aerating and venting the fuel tank during tanking as well as during operation. The fuel supply system provides the engine with fuel via the feed line within a predetermined pressure range and in sufficient, defined quantities.

12.5.1 Fuel Tanks

Fuel tanks are mainly manufactured from synthetic and/or plastic materials (PE-HD with various barrier layers) or metal (stainless steel, hot-dip aluminized sheet metal or aluminum). Normally, synthetic fuel tanks are constructed as a six-layer tank. The six layers comprise of, among others, new materials, bonding agents, barrier layers, and reclaimed materials. The barrier layer serves to prevent any diffusion of the carbon dioxide, which is released from the fuel, through the material. In order to comply with the increasing emission regulations, and a service life of 15 years and/or 150,000 miles, stainless steel tanks have been occasionally manufactured in recent years. As a result of the high manufacturing costs and further developments in manufacturing processes for synthetic tanks, the numbers of stainless ste,el tanks is starting to reduce. Fuel supply systems have been designed for various fuels— gasoline, diesel fuel, or also liquid gas. Defending on the target market and the valid emission legislation, the fuel systems differentiate with regard to flow guidance during the fuel tanking and the gases that result in operation. In principal, one differentiates between tank systems for diesel fuels and gasoline. The different characteristics of these fuels create a differentiation for the systems required for the filling system, when tanking and the delivery technology. 12.5.1.1 Diesel Fuel Tank Diesel fuel has almost no tendency to release any fuel gases. The emissions that result during tanking with diesel fuel are essentially the gas content, which the fuel forces out of the tank during filling. This gas is discharged directly to the surroundings via a venting pipe. Ambient air flows into the tank during operation in proportion to the volume of the consumed diesel fuel. A higher filling rate (up to 60 I/min) can be tanked due to the higher density of the diesel fuel, which is mirrored in the larger flow cross section of the diesel-filling pipe in comparison with the gasoline fuel system. 12.5.1.2 Gasoline Tank One can differentiate between numerous variants for gasoline fuel tanks. The first differential characteristic is the type of fuel leaded or unleaded fuel. The difference in the fuel systems is

Internal Combustion Engine Handbook | 557

6606_Book.indb 557

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

Overview of EDC for Unit Ejector System (UES) in cars Engine rpm (CS) (crankshaft)

Unit Injector (max. 5 pro Steuergerät)

Control unit EDC 15 P Signal inputs

Engine rpm (CS) and cylinder recognition (camshaft)

Sensor evaluation Signal processing

Accelerator pedal sensor with empty gas switch and kick-down (2. Sensor)

Atmospheric pressure (ADF) Idle speed regulating Running smoothness regulator

Charge pressure

Start of conveying, duration

Active jerking damper, Cruise control (CC)

Air temperature

Glow relay control

extern. Quantity intervention, quantity control and limitation

Water temperature

Climate shutdown

Exhaust gas recirculation Charge pressure control

Exhaust gas recirculation plate

Air mass

Start of conveying correction (GDP)

Charge pressure plate

Driving speed

Conveying commencement computation and conveying duration computation

Fuel temperature

Additional inputs for analog signals

Climate compressor controlling

6

Actuators

Glow time controlling

ML-Request

ISO K diagnostic request

Immobilizer replacement functions 5

FGR-Keypad

Balance +

Solenoid valve end stages

+

Brake (2)

+

Power end stages Serial interface

Clutch

2

CAN communication

+

Terminal 15

SG

Air input

SG

MIL-lamp

CAN

Main relies

SG

ASR/MSR

System lamp

MUX-Signal (PBM-A)

Diagnosis Glow time relay status Transmission

Additional low-power end stages (1.. 2A)

7

Speed signal (TD-Signal) CAN +

Sensors *optional

–

BOSCH

Consumption signal (TQ-Signal)

Communication

Figure 12.61 System configuration of an injection system utilizing the example of a PNU [12-22] (Bosch).

558 | Internal Combustion Engine Handbook

6606_Book.indb 558

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

Fuel regulating loop 1 (injection component) Fuel regulating loop 2 (injection component) "Detour" via the driver

Air regulating loop Data and signal flow

Driver´s requirements – Driver's wish – Cruise control – Engine brake...

EDC control device Controlling the injection component

CAN

Regulation and control of other actuators

Control of injection

Sensors and set point adjusters – Acc pedal sensor – Speed Sensor – Switches...

Driver´s requirements – Driver's wish – Cruise control – Engine brake...

Engine

Air

Fuel

System for filling control – Supercharging – Exhaust gas recirculation Pos elements (actuators) – Electro-pneu.Converter. – Retarder – Fan – Glow time controlling...

Injection components: – Line injection pump – Distributor injection pump – Unit Injector/Unit Pump – Common rail high-pressure pump and injectors, – Nozzle holder and nozzle

also found again in the filling pipe. Delivery valves for leaded fuel have a larger diameter (∅ 23.6 mm) compared to those for unleaded (∅ 21.3 mm). The delivery valve diameter has been reduced to prevent vehicles operating with unleaded fuel being tanked with leaded fuel. Fuel systems that still operate with leaded fuel do not normally have an active carbon filter. The active carbon filter prevents the unrestricted emission of carbon dioxide into the atmosphere. The active carbon filter is regenerated during the operation of the vehicle, controlled by the engine management system.

Figure 12.62 Basic design of the electronic diesel control system (EDC) [12-27].

The fuel supply system for unleaded fuels differentiates itself depending on the target market and the valid emission legislation for this market. Normally, only infinitesimal quantities of carbon dioxide may be emitted into the atmosphere via the filling pipe during tanking. This is achieved in Europe in that the gases which result during tanking are sucked off in a gas pendulum process via the filling valve. In the USA, conversely, all resulting gases must be cleaned by the vehicle. These different procedures create the differentiation between the European tank system and those in the USA by way of

Internal Combustion Engine Handbook | 559

6606_Book.indb 559

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

the filling pipe, the size of the active carbon filter and the diagnostic processes. One defines the fuel tank utilized in Europe as an ECE-Tank (Economic Commission for Europe), the one in the USA as an ORVR-Tank (onboard refueling vapor recovery). With the ORVR-Tank one differentiates additionally with the procedures, which prevent the emission of carbon dioxide from the filling pipe, as “liquid seal” and “mechanical seal.” With the liquid-seal system (fluid-type sealing) there is the creation of a vacuum with the fuel flow in the filling pipe during tanking. Air from the atmosphere is sucked into the tank system in this procedure. The principal of recirculation is frequently utilized to minimize the gas volume flow through the active carbon filter. Hereby, a proportion of the vapor is sucked from the fuel tank via an additional pipe to the filling pipe and is then resupplied into the fuel tank with the air sucked from the atmosphere. This cycle reduces the fuel vapor quantity that would normally enter the active carbon filter during tanking. With the mechanical-seal system (mechanical sealing), the gap between the filling valve and the cap mounting is sealed. The different gas quantities that arise during tanking determine the volume of the active carbon filter. A “carbon volume” of 0.8–1 l is sufficient for an ECE tank. Conversely, the volume for an ORVR tank can be up to 4 l. Another difference is caused by the diagnostic procedure. The ORVR tanks have an additional implementation of a leak diagnostic function in the diagnostic system for the fuel supply system in the vehicle. The leak diagnostic system should determine whether the tank cap is unscrewed, and whether there is a leak located in the fuel supply system. The leak diagnostic function can be executed in a vacuum status procedure—or excess pressure procedure. During the diagnostic period for vacuum procedures, a vacuum is created in the tank by utilizing the vacuum in the intake manifold, and then the tank pressure is monitored via a pressure sensor. In the excess pressure system, excesspressure is created by a pump in the tank; a leak can then be detected via the flow uptake on the pump.

the filling process must substantially below the venting points. The level limit can via a dip tube—carried out by a float valve in the refueling vent line or by a float-valve system on filling tube—a tube at the end of the refueling vent line, which is closed by the rising fuel.

Figure 12.63 Tank with an external venting system (top, middle).

Compensation reservoirs

12.5.2 The Tank Venting System

The fuel tank must always be aerated and vented in all operating statuses (at standstill, operating mode, filling). Various measures can be implemented to prevent liquid fuel exiting the tank system. The venting point from the tank must be selected in such a way as to ensure that the tank can always be vented, irrespective of which angle it is currently at, whereby no fuel can exit the tank at the same time via the other venting points (Figure 12.63). The venting pipes either can be closed with float valves or must be led to an external, compensation tank. This compensation tank is integrated in the main tank in most modern tank systems in order to optimize the permeation ratios (Figure 12.64). The fill-level limiter also plays a major role in tank systems alongside the venting procedure. Since petrol expands under the influence of temperature, the liquid level when terminating

Fill-level limiting valve

Figure 12.64 Tank with an internal venting system. See color section page 1085.

12.5.3 Requirements for the Fuel Supply System

The fuel delivery system has the task of supplying the engine with enough fuel from the tank during all possible driving situations. These include the vehicle manufacturer’s defined static and dynamic driving conditions such as at standstill, cornering, driving uphill and downhill. Other typical requirements are the initial suction height by initial filling of the

560 | Internal Combustion Engine Handbook

6606_Book.indb 560

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

tank, the tank refilling height as well as the residual suction height at standstill or while driving. The initial suction describes the required fuel height after first filling the tank with fuel, which is necessary for the fuel delivery system to start easily and the engine is supplied with sufficient fuel. The refilling height describes the required level of fuel during filling, which ensures a safe starting of the engine after running the tank dry. The residual suction height indicates how much fuel is allowed to remain after emptying in the tank. With multichamber tank, the fuel supply system must empty all the chambers of the tank up to the required residual suction height. In addition, the level measurement is included in each fuel delivery system. A distinction is made between diesel and petrol fuel delivery systems.

bottom. A coarse fuel filter is connected to the filtering at the end of the intake manifold. For better performance when cornering and driving in mountainous terrain, this design is extended with a surge chamber and a hydraulically operated ejector pump (Figure 12.66). A level transmitter is used for level measurement to the surge chamber.

12.5.3.1 Diesel Fuel Supply System Diesel fuel supply systems make a differentiation between the diesel fuel supply unit and the diesel exhaust unit. The use of a conveyor or suction unit depends on the numerous types of engine types and different requirements that are imposed on the system. The diesel aspiration unit is more cost-effective compared to the diesel fuel supply due to the lack of a diesel tank pump. An application of a diesel aspiration unit makes it is necessary that the high-pressure pump on the motor can build up enough vacuum to suck the diesel fuel from the diesel fuel tank. However, the resulting high vacuum can produce considerable cavitation in the high-pressure pump, which can ultimately lead to increased wear on the high-pressure pump. This is dependent on the design and quality of the high-pressure pump. Newer diesel fuel supply systems, therefore, have the application of a diesel fuel supply unit, which supplies the high-pressure pump on the motor with a slight overpressure (Figure 12.65).

Figure 12.66 Representation of a diesel exhaust unit.

Diesel suction units do not have electrical diesel in-tank pumps in contrast to diesel fuel supply units. The diesel fuel is aspirated from an engine mounted high-pressure pump directly from the surge chamber in the tank. The amount of fuel that is not consumed by the engine flows back into the suction unit in this case.

12.5.3.1.1 Diesel exhaust unit The diesel exhaust unit consists, in the simplest case, of a flange from which an intake manifold extends to the tank

Returns

Fine filter

Radiator

Common Rail

Flow

High-pressure pump

Fuel pump Ejector pump

Engine Tank

Figure 12.65 Principle of a diesel fuel supply system for common rail. See color section page 1086.

Internal Combustion Engine Handbook | 561

6606_Book.indb 561

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

The extraction of the diesel fuel and the pressure drop from a tank to the high-pressure pump occurs at the inlet of the pump and creates a negative pressure, which can lead, in conjunction with high temperatures, to bubble formation. The bubble formation leads to cavitation in the high-pressure pump. This then results in increased wear of the pump. A further disadvantage of diesel suction units are problems when starting the engine after complete emptying of the tank, if the diesel injection system is not self-venting. 12.5.3.1.2 Diesel supply unit A diesel in-tank pump with a filter is integrated upstream of the pump in the surge chamber for the diesel fuel supply unit. The electrical interfaces (plug-in connections to the pump) are available in the flange. Figure 12.67 shows a diesel supply unit. The surge chamber is filled permanently through one or more hydraulically operated ejector pumps and serves as a reservoir to ensure the delivery of fuel to the engine, even at low quantities of fuel in the tank and all driving situations.

A highly accurate pressure regulation as in petrol delivery units is thereby not necessary. 12.5.3.1.3 Diesel tank pumps The diesel tank pump has the task of supplying the engine, under all operating conditions, with sufficient diesel fuel. It is installed upstream as part of the low-pressure region of the diesel injection unit and the (high pressure) injection pump. Customary system pressures for in-tank pumps lie between 0.5 to 1 bar (with distributor pump injection systems) and 1.5 to 5 bar (for Common-Rail- and pump-nozzle injection systems). The flow rates are 100 to 300 l/h at 12 V rated voltage. Today’s diesel in-tank pumps essentially comprise of a connecting unit, the electric motor and the pumping stage. The connecting unit connects the electrical and hydraulic contacts. Normally, a check valve and a pressure-limiting valve are thereby integrated. The pressure retention in the fuel supply system is initially actuated with a switched off pump and thereby also prevents a discharge from the fuel supply system which leads to shorter start-up times for the vehicle. The pressure-limiting valve is a safety valve and opens when there are impermissible, high pressures in the fuel supply system. Furthermore, the connection unit is also equipped with radio interference suppression, which comprises a throttle coil and a condensator, depending on the vehicle manufacturer’s requirements. The in-tank pump is powered by utilizing a DC electric motor. This consists of an anchor, the return surge component with permanent magnet and a commutation system comprising carbon brushes and commutator. The pumping stages comply with the hydrostatic principle (displacement pumps) or the hydro-dynamic principle (flow pumps). In principle, the aspiration of the fuel is executed in displacement pumps in a chamber that can expand itself. Once the fuel has exited the inlet area, then the displacement is executed in a subsequently smaller chamber in the outlet. Currently, displacement pumps are usually implemented as G-Rotor pumps with an eccentrically orientated internal rotor gear against the external gear (compare to Figure 12.68). A roller

Figure 12.67 Representation of a diesel supply unit.

In multichamber tanks, additional hydraulically-driven suction spray pumps ensure the evacuation of all chambers. Diesel supply units have the task of supplying the diesel fuel under all possible load and engine speed conditions, in a given pressure window, from the diesel fuel tank to the injection system. The amount of this which is not hereby consumed by the engine flows as reflux back into the diesel fuel supply unit. Prefiltering of diesel fuel is already executed in the diesel fuel supply unit, the fine filter outside the tank. Unlike petrol supply systems for intake manifold injection, the diesel fuel supply unit does not provide the necessary injection pressure, but rather only supplies a high-pressure pump attached to the engine, which applies the injection pressure according to the relevant injection system.

Figure 12.68 Eccentric to the external rotor-oriented internal rotor gear of a G-Rotor pump.

562 | Internal Combustion Engine Handbook

6606_Book.indb 562

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

cell and/or vane pump can also be implemented, whereby the sealing for the altered chamber volumes is executed via rollers and/or radially moving vanes. Displacement pumps are particularly utilized for high system pressures (>3.5 bar), as this principle has shown itself to be particularly beneficial in this area. This can achieve an efficiency of up to 25%. Flow pumps are mostly utilized in diesel supply systems for lower system pressures (<3.5 bar). A turbine wheel, which is fitted with one or more concentric vane rings, rotates in a stationary pump chamber. Both pump chambers form a housing component, and both have a passage. Between the start and end of the passage there is a so-called scraper for sealing the pressure and intake areas. A buildup of pressure takes place along the passage due to the impulse exchange of the impeller vanes with the fuel particles. Depending on the position of the passage, one refers to a side passage (lateral arrangement of the passages and the vanes) or peripheral impeller principle (radially positioning the passage and vane to the outside). The advantage of the flow pumps is their almost nonexistent pulsation pressure build up and, compared to the displacement pumps, their relatively simple and, thereby, cost-effective design. 12.5.3.2 Gasoline Supply Systems Current gasoline fuel supply systems consist of a surge chamber with a gasoline in-tank pump, a fuel pressure regulator, a pump prefilter and fuel fine filter as well as one, or multiple, intake spray pumps and a level regulator for filling level measuring, which all comprise one component (Figure 12.69). This component closes the service opening of the tank by utilizing a flange. The flange is equipped with all the electrical and hydraulic interfaces to the tank. The surge chamber is filled permanently by the suction spray pump(s), and serves as a reservoir to ensure the delivery of fuel to the engine, even at low quantities of fuel in the tank and all driving situations. In multichamber tanks, additional suction spray pumps ensure the evacuation of all chambers. One makes a differentiation between conventional systems and return-flow-free systems for gasoline supply systems.

Returns

Figure 12.69 Illustration of a gasoline supply system with a fuel fine filter.

Conventional systems have a return flow. The gasoline in-tank pump supplies a constant fuel quantity to the fuel pressure regulator on the engine via a feed line and a fuel filter situated outside the tank (Figure 12.70). As the engine does not always consume the complete quantity, surplus fuel is fed through a return flow pipe to the tank and, thereby, back into the fuel circulation cycle. More recent return-flow-free systems also integrate the fuel fine filter and the fuel pressure regulator in the supply unit and, thereby, in the fuel tank (Figure 12.71). They fulfill the high demands for the emission sealing for tank systems. This allows for the dispensation of the fuel-flow return pipe and, thereby, the hot return flow from engine to tank. The reduced fuel temperature, less sealing points as well as the fact that only the outer surfaces of the fuel pipes create hydrocarbon emissions, enable the higher requirements for emissions to be fulfilled.

Pressure Regulator

Flow Fuel pump Fuel Filters

Engine Ejector pump

Tank

Figure 12.70 Principle of a conventional gasoline supply system.

Internal Combustion Engine Handbook | 563

6606_Book.indb 563

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

Pressure regulator

Flow

Engine Fuel pump

Tank

Ejector pump

Figure 12.71 Principle of a returnflow free gasoline supply system.

12.5.3.2.1 On-demand regulated systems The traditional fuel supply systems in vehicles with fuel injection engines are equipped with an electrical gasoline in-tank pump for supplying the fuel to the engine. The preparation of the necessary system pressure is implemented by a mechanical pressure regulator, which returns the superfluous fuel to the tank, whereby the pump always operates at full load. This means, on the one hand, that there is unnecessary energy consumption, on the other hand, there is a tendency for an additional heat input into the tank, especially in the compact and powerful vehicles, which leads to impermissible pollutant emissions due to the loss-performance of the gasoline in-tank pump. On-demand regulated systems can hereby provide assistance. These consist of a fuel delivery unit, a pressure sensor inserted into the fuel rail and electronics, which so control the fuel pump in their performance that, regardless of consumption, a constant system pressure is achieved (Figure 12.72).

Driving tests indicate an average reduction in performance uptake of approximately 50% compared with other traditional systems. An additional benefit is the considerable noise reductions, especially when idling. The performance reduction also increases the service life of the gasoline in-tank pump. The utilization of electronics in the fuel delivery system also permits the use of a pump with an electronically commutated (EC) DC motor, which is also suitable as a lifetime component for critical fuels, such as ethanol or liquefied gas, since the wear of the mechanical commutator is eliminated. Further benefits arise from utilizing an electronic system, such as the settings for different system pressures, for engines with direct injection or an additional possibility for the immobilizer. 12.5.3.2.2 Filtering The prefiltering is executed by a pump prefilter that filters the intake fuel before entry into the gasoline in-tank pump.

Engine control unit

Electronic Fuel Control Unit

Fuel pump with EC-Motor

Figure 12.72 Principle of an on-demand regulated gasoline supply system. See color section page 1086.

564 | Internal Combustion Engine Handbook

6606_Book.indb 564

1/19/16 8:44 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

In this instance, the contaminated particles of a size range from ≥30–60 μ m are filtered out, so that the gasoline in-tank pump and the fuel pressure regulator are protected against blockages. The filter materials are fleece material or webbing made of thermoplastics. The fuel fine filter is utilized for the fine filtration of the fuel to protect the important components of the engine, such as e.g. injection valves, against dirt particles and the thereby resulting wear. The fuel fine filter is located between the gasoline in-tank pump and the engine. If the filter is integrated in the gasoline supply system, then it must fulfill its functionality throughout the vehicle service life (lifetime component). Alongside the classical filter paper, more and more newly developed high-performance synthetic materials are being implemented, especially for engines with gasoline direct injection. These are multilayered from the original filter paper and ensure the respective high dirt absorption capabilities, with simultaneously high initial separation efficiency of over 90% with particle sizes between 3–5 μ m. 12.5.3.2.3 Pressure regulating To ensure a constant injection pressure in the filling pipe to the engine, a fuel pressure regulator is installed as a bypass valve downstream of the gasoline in-tank pump in returnflow free gasoline supply systems. The working principle correlates to that of a proportional regulator. Depending on the engine consumption, the resulting surplus quantity from the whole supply flow of the gasoline in-tank pump is regulated, and then resupplied to the gasoline supply system. Engines with MPI injection have a pressure between 3.0 bar to 4.3 bar. Engines with direct injection have up to 8.0 bar. 12.5.3.2.4 Gasoline fuel in-tank pump The gasoline in-tank pump has the task of supplying the engine, under all operating conditions, with sufficient fuel. In this case the gasoline fuel injection system provides the necessary system pressure without an additional high-pressure pump, conversely to the diesel fuel injection system. Normal system pressures are hereby, depending on which injection system, between 3–4.3 bar for multi-point systems, and 4.7–5.2 bar with charged gasoline engines, and between 5 and 8 bar for gasoline direct injection systems. The supply volumes are approximately 80 to 200 l/h at 12 V rated voltage. Today’s gasoline in-tank pumps essentially comprise (as with diesel in-tank pumps) of a connecting unit, the electric motor and the pumping stage. The connecting units connect the electrical and hydraulic contacts. Normally, a check valve and a pressure-limiting valve are thereby integrated in the connecting unit. The pressure retention in the fuel supply system is initially actuated with a switched off pump and thereby also prevents a discharge from the fuel supply system which leads to shorter start-up times for the vehicle. The pressure-limiting valve is a safety valve and opens when there are impermissible, high pressures

in the fuel supply system. Furthermore, the connection unit is also equipped with radio interference suppression, which comprises a throttle coil and a condenser, depending on the vehicle manufacturer’s requirements. The drive unit of the in-tank pump is executed with today’s systems by utilizing a DC electric motor. This consists of an anchor, the return surge component with permanent magnet and a commutation system comprising carbon brushes and commutator. Electronic commutation electrical motors are currently being developed, which have service life advantages especially when utilizing critical fuels for the carbon commutator system. The pumping stages either comply with the hydrostatic principle (displacement pumps) or the hydro-dynamic principle (flow pumps). The fuel is sucked into a self-expanding chamber with displacement pumps. Once the fuel has exited the inlet area, then the displacement is executed in a subsequently smaller chamber in the outlet. Currently, displacement pumps are usually implemented as G-Rotor pumps with an eccentrically orientated internal rotor gear against the external gear. A roller cell and/or vane pump can also be implemented, whereby the sealing for the altered chamber volumes is executed via rollers and/or radially moving vanes. The downside with displacement pumps with gasoline utilization is the loss in the supply volume with higher fuel temperatures due to gas bubble build up. Due to this, applications for gasoline fuel operation usually a hydrodynamic precursor whose task, on the one hand, is the separation of the gas bubbles through the passages located in the degassing holes and, on the other hand, to reduce the tendency of gas bubble build up in the displacement stage. This however leads to relatively timeconsuming design processes for solutions. Due to these disadvantages, the flow pumps have mainly succeeded for gasoline fuel applications. A turbine wheel, like the one illustrated in Figure 12.73, which is fitted with one or more concentric vane rings, rotates in a stationary pump chamber. Both pump chambers form a housing component, and both have a passage. Between the start and end of the passage, there is a so-called scraper for sealing the pressure and intake areas. The impulse exchange of the gear vanes

Figure 12.73 Turbine wheel of a gasoline in-tank pump.

Internal Combustion Engine Handbook | 565

6606_Book.indb 565

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

with the fuel particulates, and the formation of a helically formed circulating flow in the passages in the vane area, takes place in the passage as a pressure build up. Depending on the position of the passage, one refers to a side passage (lateral arrangement of the passages and vanes) or peripheral impeller principle (radially positioning the passage and vane to the outside). The exhaust gas hole, which is located in the passage, can hereby assist to achieve a relatively constant supply quantity due to the exit of gas bubbles from the passage area, also with high fuel temperatures. Additional advantages for flow pumps are the almost pulsation-free pressure build up and the cost-effective design. The most commonly utilized version is hereby the lateral passage pump. With today’s compact construction methods, they can achieve system pressures of approximately 4.5 bar per stage and efficiency of over 20%. The sequential switching of two pump stages can also achieve system pressures of approximately 9 bar. 12.5.3.2.5 Electronic on-demand regulated systems The electronic on-demand regulated system (electronic fuel control unit) regulates the gasoline in-tank pump, and creates a constant system pressure by utilizing a pressure sensor integrated in the fuel line. The electronics must be the actual value of the pressure sensor compared with the nominal value and provide the determined therefrom drive power for the pump. This occurs due to the loss-performance reduction by utilizing a pulse-width modulated (PWM) signal. The PWM frequency should be above 15 kHz; otherwise, audible noises will occur. On the other hand, the frequency must be as low as possible to prevent electro-magnetic disruptions—approximately 20 kHz is normal. If a pump with an electronically commutated (EC) DC motor is utilized, then the commutation signals also have to be generated by the electronics. Since the electronics must be placed near the pump for EMC reasons, and the pump must therefore be located so that the pipes can be short, other possibilities have to be implemented, such as the collection and processing of the signals of the level sensor or emission monitoring of the tank. An additional immobilization protection could also be considered. Therefore, additional communication possibilities must be provided.

12.5.4.1 Requirements for the Filling-Level Measuring The first priority is to achieve a reliable filling-level measurement. An exact measurement must be realized, especially with low fuel quantities in the tank. The fill-level indicator in the vehicle is primarily dependent on the measuring accuracy in the tank. A deviation in the indicated level of several liters can, in extreme cases, result in the vehicle coming to a standstill. This is also influenced by the diagnostic capabilities of the sensor. If an error occurs, then it must be ensured that the fill-level indicator returns to “min” or “zero.” As well as the indicator accuracy for the fill level, other particular requirements are placed on the mechanical stability and medium resistance. The fill-level sensor is designed for a vehicle service life. Replacing a defective sensor is, dependent on the manufacturer’s processes, not often possible. The mechanical stress results from vibrations and shocks, as they usually prevail in the tank while driving, and from sloshing of the fuel. A correct measuring of the fill level must be ensured under all conditions. Breaks in the contacts cannot occur at any time. Particular attention must also be taken for medium resistance. In particular in Brazil and the USA, the normal gasoline is also supplemented with other fuels, such as flex fuels, that is, mixtures of ethanol and/or methanol with gasoline. An essential difference arises, among others, from the larger conductivities of the fuels. In addition to conventional diesel fuels, there is also often an increased utilization of fatty acid methyl esters (FAME) in pure form, and up to 5% admixture to diesel fuel. In the case of diesel fuels, it must be therefore considered that these often occur in layers with water. A special corrosion protection is therefore unavoidable. 12.5.4.2 Lever-Type Sensor The mostly widely utilized sensor is a lever-type sensor that is implemented in the supply unit. It normally consists of a thick-film network (TFN) on a ceramic substrate, and a wiper contact spring which, via a lever with float (Figure 12.74) picks up a level-dependent series of switched film resistors.

12.5.4 The Filling-Level Measuring

Monitoring the fuel reserves is not executed for comfort reasons, rather more as a necessity for safety reasons. The filling-level indicator has to be implemented in such a way as to, independent from the dynamic driving situation—such as resulting from curve behavior and acceleration stages—prevent any excessive irregularities in the indicator. This behavior is achieved by appropriate damping algorithms in the electronic’s evaluation. If the filling-level sensor is, however, placed originally at a suitable position in the tank, then these fluctuations caused by sloshing are minimized. This is, however, not always possible due to the complicated tank geometry. Branched geometries, for example in multichamber tanks, require the utilization of several level sensors.

Figure 12.74 Lever-type sensor for the fuel-level indicator.

566 | Internal Combustion Engine Handbook

6606_Book.indb 566

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

12.5 Fuel Supply Systems

The resulting complete resistance is proportional to the fill level. The nonlinear filling volume of the tank can be linearized by utilizing approved designs in the TFN to create an indication. The design of the network is executed in such a way that the resistance value increases as the fill-level sinks. If there are any contact errors or breaks in the cables, then the smallest possible fill level will be indicated. The latest generations of lever-type sensors are equipped with contact springs with more than one contact reed. This therefore results in an improved resistance against fouling in fuels. The redundant pick-up leads to an optimized behavior during oscillation stress and increased abrasion resistance. These open sensor elements are operated with pulsating DC to prevent electrochemical effects. 12.5.4.3 Magnetic Passive Position Sensor The MAPPS (magnetic passive position sensor) also consists of a ceramic substrate with 52 in series, switched layer resistors. Every resistor has its own individual pick-up. A soft magnetic contact spring is positioned very close to the pick-ups. The system is hermetically sealed with a circumferentially soldered cap against the fuel, which surrounds the MAPPS. The contact point of the individual contact tongues to the spring is executed by a magnet, which runs along the back of the MAPPS on the ceramic instead of a strip. The interface to the surrounding medium is also formed here with the lever with a positioned float. The displacement of the magnets thereby corresponds to an orbital radius with a permissible angular range of approximately 90°. The electrical output signal is proportionally varied dependent on the position of the magnet. The resistance range extends from 100% to 0% of the fill level. An additional serial resistor is inserted for diagnostic purposes so that, if an error occurs (for example, the magnet is outside the permissible angle range), a defined total resistance results. The fully closed system ensures that the micro-contacts, even under extreme environmental conditions, are not exposed to contamination or exposure through diverse fuels (Figure 12.75). The contact surfaces are considerably less stressed by mechanical affects as conventional wiper systems. This reduced wear contact system guarantees and increased service life for the sensor element.

Bibliography

12-1. Pierburg. 1970. Vergaser für Kraftfahrzeugmotoren, Vertrieb VDI-Verlag, Düsseldorf. 12-2. Löhner, K., Müller, H. 1967. Gemischbildung und Verbrennung im Ottomotor, Band 6 in H. List: Die Verbrennungskraftmaschine, SpringerVerlag, Wien, New York. 12-3. Lenz, HP. 1990. Gemischbildung bei Ottomotoren, Band 6 in H. List, A. Pischinger: Die Verbrennungskraftmaschine, Neue Folge, Springer-Verlag, Wien, New York. 12-4. Behr, A. 1998. Elektronisches Vergasersystem der Zukunft, MTZ 44, Nr. 9, S. 344. 12-5. Großmann, D. 2002. Lexikon Verbrennungsmotor: Vergaser, Supplement MTZ/ATZ, erscheint. 12-6. Schöppe D. et al. 2011. Anforderungen an zukünftige Otto DI Einspritzsysteme und entsprechende Plattformlösungen 32. Internationales Wiener Motorensymposium 5. und 6. 12-7. Schöppe, D. 1997. Anforderungen an moderne Dieseleinspritzsysteme für das nächste Jahrhundert. Tagung Dieselmotorentechnik. Esslingen. 12-8. Egger, K.; Schöppe, D. 1998. Diesel Common Rail II—Einspritztechnologie für die Herausforderungen der Zukunft. Internationales Wiener Motorensymposium. 12-9. Klügl, W.; Egger, K.; Schöppe, D.; Freudenberg, H. 1998. The Next Generation of Diesel Fuel Injection Systems Using Piezo Technology, FISITA World Automotive Congress, Paris. 12-10. Eichlseder H., Rechberger E., Staub P. 1995. Einfluß des Einspritzsystems auf den Verbrennungsablauf bei DI-Dieselmotoren für PKW. Tagung “Der Arbeitsprozeß des Verbrennungsmotors,” Graz. 12-11. Egger, K., Lingener, U., Schöppe, D., Warga, J. 2001. Die Möglichkeiten der Einspritzung mit einem Piezo-Common-Rail-Einspritzsystem für Pkw, Int. Wiener Motorensymposium. 12-12. Bauer, St., Zhang, H., Pirkl, R., Pfeifer, A., Wenzlawski, K., Wiehoff, H.-J. 2008. Ein neuer Piezo Common Rail Injektor mit Direktantrieb und Mengenregelkreis: Konzept und motorische Vorteile, Int. Wiener Motorensymposium. 12-13. N.N. 2011. Optimierte Gemischbildung—Durch innovative Einspritztechnik—Titelthema, MTZ 01.2011, Wiesbaden. 12-14. N.N. 2010. Einspritztechnik—Der lange Weg zum Druck—Titelthema; MTZ 02.2010, Wiesbaden. 12-15. Schmidt, S. et al. 2010. Einfluss des Hub-Bohrungsverhältnisses und der Einlasskanalgeometrie auf Ladungsbewegung und Gemischbildung bei BDE-Ottomotoren. 9. Internationales Symposium für Verbrennungsdiagnostik, AVL, 8./9. 12-16. Warga J. 2011. Konsequente Weiterentwicklung der Hochdruck-PkwDieseleinspritzsysteme, Int. Wiener Motorensymposium. 12-17. Pauer, T., Wirth, R., Brüggeman, D. 2000. Zeitaufgelöste Analyse der Gemischbildung und Entflammung durch Kombination optischer Messtechniken an DI-Dieseleinspritzdüsen in einer Hochtemperatur-Hochdruckkammer. 4. Internationales Symposium für Verbrennungsdiagnostik, Baden-Baden. 12-18. Mollenhauer, K., Tschöke, H. (Hrsg.) 2007. Handbuch Dieselmotoren, 3. Aufl. Berlin, Heidelberg: Springer. 12-19. Härle, H. 2000. Einfluss des Einspritzverlaufs auf die Emissionen des Nkw-DI-Motors. Tagung Diesel- und Benzindirekteinspritzung. Berlin. 12-20. Eichlseder, H. 1995. Der Einfluss des Einspritzsystems auf den Verbrennungsablauf bei DI-Dieselmotoren für Pkw. 5. Tagung “Der Arbeitsprozess des Verbrennungsmotors” Graz.

Figure 12.75 Magnetic passive position sensor (MAPPS) for the filllevel indicator.

12-21. Chmela, F., Jager, P., Herzog, P., Wirbeleit, F. 1999. Emissionsverbesserung an Dieselmotoren mit Direkteinspritzung mittels Einspritzverlaufsformung. In: MTZ 60 Heft 9, S. 552–558. 12-22. Dieselmotor-Management: 4. Auflage, Vieweg Verlag, 2004. 12-23. Lewis, GR. 1992. Das EPIC-System von Lucas. In: MTZ 53 Heft 5, S. 224–229.

Internal Combustion Engine Handbook | 567

6606_Book.indb 567

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 12 Mixture Format Ion and Related Systems

12-24 Robert Bosch GmbH (Hrsg.). 1998. Technische Unterrichtung Kraftfahrzeugtechnik: Diesel-Radialkolben-Verteilereinspritzpumpen VR. Ausgabe. 12-25. DIN Deutsches Institut für Normung (Hrsg.): DIN EN 590 (Ausgabe 2000-02): Kraftstoffe für Kraftfahrzeuge.—Dieselkraftstoff.—Anforderungen und Prüfverfahren. Berlin: Beuth, 2000.

12-34. Senghaas, C., Schneider, H., Reinhard, S., Jay, D., Ehrström, K. 2011. Neues Schweröl-Common-Rail-Einspritzsystem, MTZ 01.2011, Wiesbaden. 12-35. Simon, C., Will, BC., Dörksen, H., Mengel, C. 2010. Erzeugung und Einspritzung von Diesel-Wasser-Emulsionen, MTZ 07.– 08.2010, Wiesbaden.

12-26. Tschöke, H., Kilic, A., Schulze, L. 2000. Messadapter für Mehrlochdüse. Offenlegungsschrift DE 199 09 164 A1 vom.

12-36. Borchsenius, F., Stegemann, D., Gebhardt, X., Jagni, J., Lyubar, A. 2010. Simulation von Diesel-Common-Rail-Einspritzsystemen, MTZ 06.2010, Wiesbaden.

12-27. Robert Bosch GmbH (Hrsg.). 2001. Technische Unterrichtung Kraftfahrzeugtechnik: Elektronische Dieselregelung EDC. Ausgabe.

12-37. Clever, S., Isermann, R. 2010. Modellgestützte Fehlererkennung und Diagnose für Common-Rail-Einspritzsysteme, MTZ 02.2010, Wiesbaden.

12-28. Tschöke, H., Leyh, B. (Hrsg.). 2001. Diesel- und Benzindirekteinspritzung. Renningen-Malmsheim: expert-Verlag.

12-38. Leonhard, R., Warga, J., Pauer, T., Rückle, M., Schnell, M. 2010. Magnetventil—Common-Rail-Injektor mit 1800 bar; MTZ 02.2010, Wiesbaden.

12-29. Kull, E. 2003. Einfluss der Geometrie des Spritzloches von Dieseleinspritzdüsen auf das Einspritzverhalten, Dissertation. 12-30. Bonse, B., u.a. Innovative Dieseleinspritzdüse—Chancen für Emissionen, Verbrauch und Geräusch; 5. Internationales Stuttgarter Symposium 2003, Kraftfahrwesen und Verbrennungsmotoren. 12-31. Tschöke, H., Legh, B. (Hrsg.). 2003. Diesel- und Benzindirekteinspritzung. Renningen–Malmsheim: expert-Verlag. 12-32. Tschöke, H. (Hrsg.). Diesel- und Benzindirekteinspritzung. Renningen–Malmsheim: expert-Verlag 2005 sowie 2007 und 2009.

12-39. Fürhapter, A., Piock, WF., Fraidl, GK. 2004. Verbrennung: Homogene Selbstzündung—die praktische Umsetzung am transienten Vollmotor. In: MTZ 65, Heft 2, S. 94ff. 12-40. Stegemann, J., Meyer, S., Rölle, T., Merker, GP. 2004. Berechnung und Simulation: Einspritzsystem für eine vollvariable Verlaufsform. In: MTZ 65, Heft 2, S. 114ff. 12-41. Hummel, K., Boecking, F., Groß, J., Stein, JO., Dohle, U. 2004. 3rd Generation Pkw-Common-Rail von Bosch mit Piezo-Inline-Injektoren. In: MTZ 65, Heft 3, S. 180ff.

12-33. Shinohara, Y., Takeuchi, K., Herrmann, OE., Laumen, HJ. 2011. Common-Rail-Einspritzsystem mir 3000 bar, MTZ 01.2011, Wiesbaden.

568 | Internal Combustion Engine Handbook

6606_Book.indb 568

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13 Ignition 13.1 Gasoline Engines 13.1.1 Introduction to Ignition

External ignition systems in combustion engines (gasoline engines) trigger the combustion process by an electrical discharge in the combustion chamber toward the end of the compression cycle. The required components for this process are an ignition coil as the high-voltage source, and a spark plug as the electrode in the combustion chamber. From the spark, a high-temperature plasma channel arises between the spark plug electrodes. An exothermic chemical reaction occurs in a thin reaction layer around this channel. This develops into a self-sustaining and expanding flame front [13-1].

13.1.2 Requirements for the Ignition Systems

The ignition system must ensure a reproducible ignition process throughout all conceivable changes and dynamic fluctuations of the engine’s operating states. For the spark to jump to the spark plug electrodes, the ignition system must have sufficient high voltage. The pressure, temperature, and density of the mixture at and between the ignition electrodes at the ignition timing influence the required voltage. These parameters vary widely over the speed and load. According to the Paschen law, the required ignition voltage increases linearly with the pressure and electrode spacing. The energy transferred to the mixture by the spark must therefore suffice to trigger self-sustaining combustion. The optimum ignition timing plays a central role, and is measured in the engine during the application phase and saved in a program map in the engine control unit as a function of the speed and load.

13.1.3 Minimum Ignition Energy

Homogeneous, stoichiometric joined fuel-air mixtures require energy of less than 1 mJ for ignition while idling. In richer or emaciated mixtures, the energy requirement increases to 3 mJ [13-2]. In real engine conditions, these conditions are

much less favorable. The energy requirement rises sharply because of the inhomogeneous distribution of air, fuel, recycled exhaust gas, etc., between the cylinders, and because of inhomogeneous cylinder charging and transfer and heat losses to feed lines and electrodes. Conventional ignition systems provide approximately 40 mJ with a spark duration of 1 ms at the spark plug to ensure ignition.

13.1.4 Fundamentals of Spark Ignition 13.1.4.1 Phases of the Spark The spark forming at the spark plug can be divided into three sequential types of discharge with very different energy and plasma physical properties (Figure 13.1) [13-3], [13-4], [13-5]. Initially, the voltage at the spark plug grows sharply. As soon as the current charge forming in the field reaches the opposing electrode, (breakdown) occurs within a few nanoseconds. The impedance of the electrode path falls drastically, and the current rises from the discharge of the leakage capacitance of the spark plug. The spark then transitions into the (arc phase) with very small voltages, in which the current is determined by the discharge of the high-voltage-side capacitance. At the cathode, a hot spot (ignition spot) arises because of the strong emission of electrons; cathode material vaporizes and strongly erodes the electrodes. The temperature in the channel drops to approximately 6,000 K. The plasma now expands by heat conduction and diffusion processes and the exothermic reaction that leads to a progressive flame front starts. At currents below 100 mA there is a transition to glow discharge. Multiple transitions can occur between arc and glow discharge in a transition range depending on changes and movements of the mixture between the electrodes. In the glow discharge phase, the voltage rises again (the electron stream is supported by the contacting ions); the temperature in the channel is now only approximately 3,000 K. This is below the melting temperature, and the electrodes are now primarily atomized by contacting charge carriers [13-6].

Internal Combustion Engine Handbook | 569

6606_Book.indb 569

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

Bend, %

Glow, %

<1

5

<1

5

45

70

Total losses

6

50

70

Plasma energy

94

50

30

Voltage, V

10 4 10 3

Radiation loss

Glow discharge

Transition region

Arc phase

Transition phase

Predischarge phase

10

5

Breakdown phase

Breakthrough, %

Heat dissipation at the electrodes

Figure 13.2 Energy balance of the three forms of discharge [13-3]. 15kV 1 mJ

50 V 1 mJ

300 mV 30 mJ

10 2 10 1 10 0 10–9

10–6

10–3

13.1.5 Coil Ignition System (Inductive)

10 3

Coils used in distributor less ignition systems switched with transistors are dry ignition coils cast with epoxied resin that consist of a closed magnetic circuit made of laminated low-loss electrical sheet steel with concentrically superposed primary and secondary windings Figure 13.3. When the primary current is turned on, energy is inductively stored in the air gap of the magnetic circuit. After the primary current is interrupted by the transistor (Figure 13.4), a secondary-side voltage builds

10 2

Current, A

ignition, and the ignition reliability increases with the peak current and the length of the discharge [13-7]. A long spark duration favors reliable ignition. Even with lean mixtures (Aλ = 1.5) and a fast flow (>30 m/s) the long glow discharge of transistorized coil ignition is sufficient by itself to continually ignite a flammable mixture that is transported through the flow field into the electrode area [13-8].

10 1 10 0 10 –1 10 –2 10–9

10–6 Time. s

10–3

Secondary winding

Figure 13.1 Chronological process for current and voltage of a transistor-coil ignition (TSZ) [13-4]. Typical values for occurring voltages and energy transfer in the individual spark phases start quickly. Because of the fast rise in voltage in the ignition coil, arcing does not occur upon reaching static breakdown voltage but upon overvoltage due to the ignition lag. Very high temperatures of 60,000 K arise in the conductive channel from the complete dissociation and ionization of the atoms and molecules. The pressure wave begins to propagate at supersonic speed.

Core lamination stack

Primary winding High-voltage diode

The energy storage coil discharges thereby completely in the discharge channel. When the voltage falls below the threshold voltage necessary for maintaining the channel, the spark terminates. The residual energy decays in the secondary winding of the ignition coil. 13.1.4.2 Energy Transfer Efficiency Figure 13.2 shows the amount of energy that can be sent to the mixture in the described phases of the spark. The breakthrough phase in this case has the highest ignition efficiency and causes a faster conversion of energy in the initial stage of the combustion process. By increasing the spark plasma and its propagation speed, the ignition reliability can be improved [13-4]. Because of the considerable heat loss through the electrodes present in the spark plasma energy is much smaller than the spark plug to the electric energy supplied. With conventional transistorized coil ignition, basically the glow phase stimulates

Suppression

Pressure spring

Figure 13.3 Structure of ignition coil.

570 | Internal Combustion Engine Handbook

6606_Book.indb 570

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.1 Gasoline Engines

Diode

Rprim

Rsec

Lprim

Lsec

RSupp.

Spark plug

UZener

Figure 13.4 Schematic layout of a transistor coil ignition (TZT).

After arcing, energy discharges in the spark via the secondary winding of the coil. During this glow phase (combustion time), the spark gap at the spark plug can be considered from an electrical point of view as being replaced by a Zener diode gap that restricts the secondary voltage to the value of the firing voltage and keeps it constant until the spark breaks contact. The definitions of the properties of such an ignition coil are uniformly governed by ISO 6518. The available voltage is defined as the maximum attainable voltage with substitutionary resistance corresponding to the relevant obstruction. For example 1 MΩ // 25 pF of the electrical load corresponds to an ignition coil directly connected to the spark plug, and 1 MΩ // 50 pF corresponds to an ignition coil that is connected to the spark plug via an ignition cable. The output or internal energy is determined by measuring the discharge duration concluded with a Zener diode circuit with 1,000 V ignition coil. By means of the turn’s ration and the interrupting current of the coil, the maximum spark current (glow current) is defined on the secondary side of the ignition coil. The spark duration can be varied within wide limits by setting the stored inductance and operating point of the magnetic circuit. The coupling between the primary and secondary sides of the ignition coil is more than 90%. Of the electrical energy stored in the primary current circuit, only approximately 50% arrives at the spark plug because of the transmission loss and resistance in the circuit. The conditions in the combustion chamber (pressure, temperature, mixture movement, etc.) determine the firing voltage during spark ignition together with the electrode distance. The influence on the energy and spark duration is shown in Figure 13.5. Double spark ignition coilsare used widely in which both ends of the secondary winding are series connected via ignition cables to the spark plugs that belong to cylinders whose firing sequence is shifted by 360° KW crankshaft angle. When there are four cylinders, then these are the cylinders 1 and 4 as well as the cylinders 2 and 3 which are respectively

60

3,0

50

2,5

40

Energy Combustion duration

30 20 400

600

800

1000

Combustion duration [ms]

2,0 1,5

Operating voltage [V]

UBatt.

connected to a coil. The series connection causes two spark plugs to fire simultaneously—one in a cylinder filled with a fuel-air mixture, the other in a cylinder in the exhaust cycle in which a support spark arises with only a small amount of additionally required voltage due to the pressureless state.

Energy [mJ]

in the coil until breakdown at the spark plug. The maximum attainable voltage essentially depends on the cutoff voltage and the secondary/primary turns ratio in the coil.

1,0 1200

Figure 13.5 Burning voltage influence on energy and spark duration.

Because of the series connection, one of the two spark plugs ignites with a positive high voltage and the other with a negative high voltage. For ignition with a negative high voltage, the required voltage is slightly less (1–2 kV) than with a positive voltage because of the higher temperature of the middle electrode of the spark plug and the subsequently reduced work function of the electrons while the engine operates. At the same time, the electrode erosion at the spark plugs is strongly asymmetrical because of the different polarities of the ignition voltage. Different arrangements are possible for ignition with double spark coils. On the one hand, the double spark coils can be combined into a block or packet, and the spark plugs can be connected via ignition cables; alternately, the ignition coil can be directly placed on or connected to a spark plug, and the connection to the spark plug in the correlating cylinder can be with an ignition cable. In higher-end vehicles, single spark coils are used to better control ignition and problems with valve overlap, etc., where each cylinder is fired with its own ignition coil Figure 13.6. The coils are mounted on the cylinder head and directly contact the spark plug or are combined into blocks with several individual spark coils and ignition cables connected to the spark plugs. In single spark ignition coils, a high-voltage diode is required in the secondary circuit to suppress the voltage pulse that arises at the inductance coil when the current is switched on since an ignitable mixture can be in the cylinder at this time at a low pressure and, hence, low required voltage. By directly connecting these coils to the spark plug and, hence, dispensing with the interference-free ignition cables, the ignition coil itself must have the interference suppression element such as a wound, inductive resistor to suppress high-frequency interference that arises from the flashover at the spark plug.

Internal Combustion Engine Handbook | 571

6606_Book.indb 571

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

the ignition coil, such as the installation of electronic semiconductor switches and/or the integration of diagnostic and self-protection tasks.

Figure 13.6 Directly to the spark plug single-spark ignition with 70 mJ, 35 kV output voltage and 2 ms-burning time.

The use of ignition coils with or without ignition cables (built separately or plugged on directly), determines from the different external capacitive loads the optimum turns’ ratio with which the coil can provide the maximum output voltage (Figure 13.7). 34 32

Usek [kV]

30 28

1MOhm/25pF 1MOhm/50pF

26 24 22 20

50

60

70

80 90 100 110 120 Ratio

Figure 13.7 Influence of the external load of the ignition coil on the optimum turns’ ratio.

Pencil coils are becoming more important(Figure 13.8). Their design with an open, long magnetic circuit allows the size and diameter of the ignition coil to be reduced, so that the coil can be mounted directly in the spark plug shaft. The component requirements for temperature resistance and insulation strength are, hence, greater. The system that is chosen depends on the application, the special requirements, and cost. The same holds true for the integration of other components and intelligent functions in

Figure 13.8 Flush-fitted pencil coil (pencil coil) with diameter 22 mm, 32 kV output voltage and 60 mJ.

13.1.6 Other Ignition Systems

Despite repeated efforts at introducing alternative ignition systems (plasma ignition, laser ignition, and many others), the traditional coil ignition has become generally accepted because of its favorable cost-benefit ratio [13-7]. Only in exceptional cases (e.g., racing car engines) will the high-voltage capacitor ignition (CDI) be utilized. With CDI, the energy is temporarily stored in a capacitor, and the required high voltage is switched via a fast low-loss ignition transformer. These ignition systems have an extremely fast voltage rise (a few kV/λ s) and, hence, effectively resist shunts from deposits on the spark plugs. A disadvantage is that the very short combustion time of approximately 100 μ s can lead to misfiring when there are inhomogeneous mixtures, and the strong spark current can increase spark plug erosion. An additional is “AC voltage ignition” as is utilized in the Mercedes 12-cylinder (V12) [13-9]. In this case, a capacitor is connected as an energy accumulator with a weakly coupled ignition transformer, which is connected to a resonant circuit with a resonance frequency of approximately 20 kHz. After flashover, energy is delivered in the spark from the secondary side of the coil, while the capacitor is recharged (reverse converter principle). In contrast to CDI, the spark does not cease, since enough energy remains in the system to maintain oscillation. The danger of misfiring from inhomogeneous mixtures is much less than CDI. With AC ignition, a type of ignition is obtained in which the combustion time is freely settable, independent of the

572 | Internal Combustion Engine Handbook

6606_Book.indb 572

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.1 Gasoline Engines

provided ignition voltage in contrast to transistorized coil ignition. With a combustion time tailored to the demand (energy controlled ignition), spark plug wear is less and e.g. the ionic current can be measured at the spark plugs to detect misfiring after the controlled end of the spark [13-9]. All of the discussed ignitions considered alongside the TSZ ignitions require in additional components in addition to the coil, such as capacitors and power supplies (100–800 V) for generating the required charge voltages, which thereby increase costs and hinders the acceptance and dissemination of such ignition systems.

EGR to ensure repeatable and sufficient arc development [13-10]. One can assume that this value will fall with improvements in mixture control.

13.1.8 Spark Plug 13.1.8.1 Requirements for Spark Plugs The spark plug represents the electrode necessary for ignition in the combustion chamber and, hence, has to satisfy quickly changing engine requirements. Electrically, the spark plug must ensure high-voltage transmission and isolate the required ignition voltages of over 30 kV, prevent arcing, and consistently resist dielectric loads from high field strengths and quickly changing fields during its service life. Mechanically, the spark plug should seal the combustion chamber against pressure and gas, and absorb the mechanical forces that arise when screwing in the plug. Thermally, good heat conduction protects the spark plug against loads from small thermal shocks in each combustion cycle and keeps down the temperature of the spark plug. Electrochemically, the spark plug must resist attacks from spark erosion, combustion gases, and residue such as hot gas corrosion, oxidation, and poisoning from sulfur in the fuel, and it must resist the formation of deposits on the insulator.

13.1.7 Summary/Outlook

To increase operational reliability, ignition systems should have low source impedance and/or a fast voltage rise (shunt resistance). Furthermore, ignition systems must provide sufficiently high voltage. In future ignition systems, we can anticipate a further rise in the demands on available voltage (lean operation, high EGR rates, turbocharging, gasoline engines—DI). In particular, the required ignition voltage in a lean-running engine with direct injection under a partial load in stratifiedcharge operation is higher than for a comparable engine in stoichiometric operation, since the charge dilution from excess air and/or exhaust recycling increases the gas density in the cylinder and, hence, raises the breakdown voltage at the time of ignition. However, the demand for maximum ignition voltage that is typically attained under a full load in homogeneous operation is comparable in both instances; the demands on an engine with direct injection therefore remain unchanged in regard to maximum electrical insulation resistance of the ignition coil, wire and spark plug, in comparison to an engine with multipoint fuel injection [13-10]. Only when the ignition system has a large enough capacity to store energy, can a sufficiently large plasma channel be generated. The energy requirements are also higher for an engine with direct injection under partial load in stratifiedcharge operation, in contrast to an engine with intake manifold injection, since more energy must be supplied to the mixture (70–100 mJ) because of the charge dilution from excess air or Nickel-plated plug body Quintuple leakage current barrier Connection for with grooved profile the plug cap

13.1.8.2 Design Given the above requirements, the basic design of the spark plug has scarcely changed over the course of the development of the engine (Figure 13.9). Nevertheless, primarily over the last 20 years, changes have been made in the form of constructive details and improved materials because of the increased need to adapt the spark plug to the specific conditions of each engine; this has led to a substantial increase in the change interval. Different surface ignition approaches have become possible with the use of unleaded fuels. The insulator of the spark plug consists of an aluminum oxide ceramic, that provides strong electrical arcing resistance, and it is usually provided with a ribbed insulator flashover barrier on the insulator neck. Embedded in the insulator, the center electrode and igniter are connected gas sealed by a

Captive outer sealing ring Inner sealing ring Introduction application Breathing space Insulator base Mass electrode Middle electrode

Firing pin

Insulator made of aluminum oxide

Electrically conductive glass melt

Figure 13.9 Design of a spark plug.

Internal Combustion Engine Handbook | 573

6606_Book.indb 573

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

1 Air spark route 2 Air/sliding air spark route 3 Sliding air spark route 2

1

2

1

1

3

3

Figure 13.10 Different spark paths.

special, electrically conductive glass seal. With corresponding additives, this conductive glass seal can be provided with a specific resistance to improve erosion resistance and interference suppression. The gas sealed connection between the insulator and metallic body is created with an internal sealing ring, and the initial mechanical force on the sealing ring arises from the spark plug body that is first beaded onto the insulator and then electro-coated in a special heating procedure. Welded onto the spark plug body are one or more ground electrodes, that form the gas discharge path together with the center electrode. The various spark plug types are distinguished according to their electrode or spark path (Figure 13.10). Air gap: In spark plugs with hook electrodes (J-type), goodto-optimum mixture accessibility is provided by the open spark path through the gas chamber (air). Surface gap: If the spark glides across the insulator when arcing, deposits and combustion residue can be burned up. Electrical shunts are avoided, but the ignition spark must be energy rich to compensate for the cooling that arises while the spark glides over the insulator. Simultaneously, the lower required voltage sometimes permits a longer spark path and, hence, greater mixture accessibility. Semi-surface gap: By arranging the electrode, spark paths can be set that partially traverse the air and partially run across the insulator. By combining mutually independent air and surface gap paths, the rise in the required ignition voltage from electrode erosion can be reduced, which greatly extends the life of the spark plugs.

13.1.8.3 Heat Range The heat range is a measure of the thermal resistance of a spark plug, and describes the maximum operating temperature that arises in the spark plug from the equilibrium between the absorption and release of heat. After starting the engine, the spark plug should reach the “self-cleaning temperature” of > 400°C as quickly as possible to oxidize (burn off) deposits on the insulator to prevent electrical shunts. At the same time, the thermal conductivity must be sufficient so that the stationary end temperature does not exceed 900°C at any point on the spark plug which could produce uncontrolled autoignition. In terms of the design, the heat range of the spark plug is controlled by the geometric shape of the insulator nose and the breathing space, as well as the electrode’s position, geometry, and thermal conductivity (Figure 13.12). Spark plugs with long insulator paths up to the internal seal and open breathing space form large heat absorbing surfaces with poor heat conduction. These spark plugs are termed “hot”; spark plugs, those with short insulator noses are correspondingly called “cold.” „Hot plug“

„Cold plug“

The electrode position determines the spark position in the combustion chamber (Figure 13.11).

Figure 13.12 Warm or cold spark plugs.

Figure 13.11 Normal and advanced spark position.

By using compound electrodes such as nickel electrodes with a copper core—copper is unsuitable to be used directly in the combustion chamber, but its thermal conductivity is very good—the heat removal from the electrode is substantially improved.

574 | Internal Combustion Engine Handbook

6606_Book.indb 574

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.1 Gasoline Engines

13.1.8.4 Required Ignition Voltage The difference between the high voltage offered by the ignition coil and the required ignition voltage (Figure 13.13) defines the voltage reserve. The arising electrode erosion increases the electrode spacing and, hence, the required voltage (Figure 13.14) and, together with the voltage reserve, determines the maximum possible service life (length of use) of the spark plug. A one-sided increase in the available voltage of the ignition coil, to allow the spark plug to be operated longer, is counterproductive: It produces problems with the high-voltage capacity of the feed lines and increases electrode erosion because of the high ignition energy. The required ignition voltage in Figure 13.13 and Figure 13.14, as displayed according to the amount and frequency of occurrence, is calculated in a mixture of overland travel and a

When the spark position is extremely advanced in the combustion chamber, specialized adaptation of the cross section and the heat-absorbing surface of the insulator nose tip enables a rapid achievement of the free burning temperature and a quasi, self-regulation of the upper on the insulator below 900°C. Thus, this type of spark plug is suitable for use in combustion chambers with relatively low and high temperatures [13-16]. Stratified combustion usually require spark plugs that extend further into the combustion chamber [13-17]. This can lead to an increased mechanical and thermal stress on the electrodes. To prevent vibration fractures, the thread insert is therefore lengthened. This allows shorter and, hence, colder ground electrodes. All the electrodes are also equipped with a copper core.

35000 Uz [V]

Spark Plug 60.000 km

30000

Spark Plug, new

25000

Voltage range

20000 15000 10000 5000 0

0

1000

2000

3000 4000 n [1/min]

5000

6000

7000

Figure 13.13 Required voltage (min, max) and available voltage.

5000 Spark Plug 60,000 km

4500

Spark Plug new

4000

Frequency

3500 3000 2500 2000 1500 1000 500

Ignition voltage classification

–29 kV

–27 kV

–25 kV

–23 kV

–21 kV

–19 kV

–17 kV

–15 kV

–13 kV

–11 kV

–9 kV

–7 kV

–5 kV

–3 kV

0

Figure 13.14 Frequency distribution of the required ignition voltage.

Internal Combustion Engine Handbook | 575

6606_Book.indb 575

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

circular track test with a high acceleration component. A clear rise in the required voltage over the operating time can be discerned at the spark plugs with two lateral Cr-Ni electrodes. One of the tasks of the spark plug is to keep down the ignition voltage itself, and the additional rise of the ignition voltage over the time of operation. The reduction of the electrode spacing to lessen the essential ignition voltage is subject to narrow restrictions because of the required mixture accessibility, in particular, with lean mixtures, and the occurrence of quenching, etc. If the spacing is too small, misfiring occurs from the combustion of a volume that is too small for initial ignition, or is because of poor mixture accessibility. If the spacing is too small, then there are misfires occurring from the combustion of a volume which is too small for the initial ignition or by the poor mixture accessibility. The reduction of the electrode cross-sections causes an increase in the electric field strength from a peak effect with reduced ignition voltage. This necessitates the utilization of high-grade metal electrodes that reduce electrode erosion because of increased electron discharge work and higher material melting and boiling points (Figure 13.15). At the same time, the heat-absorbing surface is reduced. Because of the temperature of the electrode, it is preferable for the polarity of the ignition voltage to be negative since the hotter center electrode enhances electron discharge and, hence, lowers the required voltage. 13.1.8.5 Ignition Characteristic (and Mixture Ignition) In addition to the cited features, the spark plug is also evaluated for its ability to reduce cyclic combustion fluctuations, and to shift the lean limit to influence the smooth running of the engine as well as the exhaust gas and fuel consumption. Spark plugs with small electrodes are optimally suitable for reducing the required ignition voltage and the contact surface of the flame with the electrodes to prevent heat loss. Large ignition gaps with favorable mixture accessibility can be advantageously implemented in spark plugs with surface gaps that limit the rise in voltage resulting from electrode erosion, and therefore offer a favorable electrode orientation at average flows of 2–5 m/s; the flame is moved away from the

electrode but not extinguished [13-18]. The influence of these measures on electrode wear is not taken into consideration. Surface gap spark plugs (including those with several electrodes) are, according to a set of investigations, not suitable for igniting leaner mixtures since the insulator heat loss and the mixture accessibility are inferior [13-19], [13-20] however, the sparking distance is greatly increased at the cost of a greater required ignition voltage. This improves mixture accessibility, and the shunt resistance of this type of spark plug is more effective. In modern engines, flow speeds of more than 10 m/s are normal, and they can achieve 30 m/s in SI engines with direct fuel injection. This causes the orientation of the electrodes to become irrelevant because of turbulence in the combustion chamber and has no influence on the inflammation, however the influence of the spark position is clearly recognizable [13-21]. According to these investigations, the spark plug with a spark position (spark air gap), which is advanced extremely far into the combustion chamber of engines with intake manifold injection, engines with stratified charge operation (FSI from VW) [13-21], function better, however the overall behavior of surface gap spark plugs is improved. The self-cleaning behavior on the insulator is probably the decisive factor. Optimum arc formation increases the combustion speed, but the greater combustion speed but, however, the increased combustion chamber temperature enhances the formation of NOx. Cold starts are particularly demanding on spark plugs where the spark plug must ensure a faultless start without shunts and especially flawless engine acceleration (load assumption). More voltage is required for acceleration, and this can cause electrical shunts and, hence, misfiring when there are deposits on the spark plugs. Similar problems occur with repeated starts, continuous short-distance travel, or slow driving in which the spark plug does not become sufficiently hot. Technical assistance is provided by equipping the spark plugs with surface gaps on the insulator (cleaning from the surface spark) or providing corona edges facing the insulator on the high-voltage center electrode (cleaning from additional ionization). Sharp edged or pointed electrodes in the spark-over path reduce the required voltage and, hence, the tendency to shunt.

6000

Temperature [°C]

5000 4000 Melting point

3000

Boiling point

2000 1000 0

W Re

Ir

Ru Pt

Rh Pd Au Co Elements

Ni

Cr

Ag

Figure 13.15 Melting and boiling points of different metals.

576 | Internal Combustion Engine Handbook

6606_Book.indb 576

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.1 Gasoline Engines

In summary, it must be noted that spark plugs need to be specially re-adapted to each engine and engine variation (supercharging, EGR rate, etc.). No general conclusion can be made regarding which spark plug type is best suited for each application. The best possible adaptation to the parameters of thermal behavior, spark geometry, and the required ignition voltage is required. At the same time, the spark plug should be located at the site of the most favorable flow conditions (advanced spark position—SI engines with direct fuel injection), which poses additional demands on the selected electrode material, and the design of the insulator nose. 13.1.8.6 Wear The spark plug electrodes are subject to several wear mechanisms. 1. The thermal stress from internal engine processes arising from compression and ignition wears the material of the electrode extending into the combustion chamber from hot gas corrosion and scaling. 2. Another cause of wear is chemical reactions such as oxidation of the electrodes triggered by fuel, additives, and combustion gases. Notable wear of the electrode material occurs at high temperatures from aggressive gases. 3. The spark erosive attack on the electrode causes the materials to partially melt and evaporate from the high temperatures in the plasma channel. This creates a demand for materials with high melting and boiling points. Nickel is the primary electrode material that is utilized, and alloyed with aluminum and chrome as oxide formers and manganese and silicon against sulfur in the oil and fuel to improve chemical resistance (Figure 13.16). With a

Test temperature: 1000 °C

+500

4

±0

Weight displacement [g/m2]

1 2 3 4 5 6 7 8

NiCr2 Mn NiMn3 Si NiMn2 NiCr5 Mn NiCr2 MnSI NiCr2 MnSiCe NiCr 7615 NiAl 11 (15398) 8

7

–500

3 2

–1000 –1500

6

1

–2000 –2500

5 24

48

72

96

120 144

Annealing [h]

168 192

Figure 13.16 Hot gas resistance of various Ni alloys [13-24].

216

melting point of only approximately 1,450°C, the material is not resistant against attack from hot gas and spark erosion (Figure 13.17). Nevertheless, operating service lives of 60,000 km and more are possible with optimized alloys and suitable geometric designs. Funk phase Increase

duration

Energy

EDM

60 ∝s

Breakthrough

2 ns

0,5 mJ

12 · 10 –12 g/mJ

Bend

1 ∝s

1 mJ

210 · 10 –12 g/mJ

Glow

2 ms

60 mJ

3,5 · 10 –12 g/mJ

Figure 13.17 Wear by different spark phases [13-22].

Platinum fulfills the demands for a high temperature and oxidation stability. Chemical attacks on the grain boundaries sulfur and silicon, which are poisonous for platinum, increase wear. The arc of the spark partially melts the electrode surface, which can then more easily react with combustion gases. Iridium has even higher melting and boiling points but it is unsuitable as an electrode material when it is pure. To exploit its high temperature resistance; platinum, palladium or rodium are alloyed and form oxides that therefore protect the surface of the iridium [13-23]. In high-performance spark plugs, precious metal electrodes are particularly suitable. However, because of the high cost of the precious metals, the amount of material is restricted, and chrome-nickel electrodes are utilized where only the areas that form the arcing path are reinforced with precious metal. With a suitable design, the demands can be combined for high performance (service life) with nearly unchanged required ignition voltage, favorable mixture accessibility and idling stability, and reduced shunt sensitivity and superior cold start behavior. Figure 13.18 represents the principle of flow guidance with two-material electrodes. Small anchoring sites (1) made of materials with high discharge and low vaporization rates (such as Pt) are on both electrodes with inverse properties (such as Cr-Ni), and they determine the required ignition voltage and arcing site. This geometry and the selected materials force the first spark to arc via the anchoring sites, but the discharge immediately transitions to the areas of the support electrodes formed as sacrificial zones (3). The erosion of the anchoring sites is minimal, and the electrode spacing (2) and required ignition voltage remain constant. The erosion (Figure 13.19a, b) is shifted to specified areas of the base electrode; the effective spark length increases over time, and thereby even favors the ignition capability [13-22]. Since the ignition voltage remains nearly constant due for the service life of the spark plug due to the reduced electrode erosion, the electrode spacing can be larger and a more favorable electrode geometry can be chosen, which improves the ignition capability and idling stability. A restriction of the life from deposits (4) on the insulator is prevented by an auxiliary spark path (5), which eliminates these deposits with occasional creeping discharges. At the same time, these additional creepage spark paths improve

Internal Combustion Engine Handbook | 577

6606_Book.indb 577

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

cold-start behavior and prevent misfiring in operating conditions with a very high required voltage.

13.2 Diesel Engine 13.2.1 Autoignition and Combustion

3

1

2

3

5

Autoignition characterizes diesel engine combustion. Combustible fuel is injected toward the end of the compression cycle into the hot, compressed cylinder charge, mixed with it, and ignited. During the ignition lag (between injection and the start of autoignition), a series of complex physical and chemical subprocesses occur such as spray formation, vaporization, mixing, and chain branching (initial chemical reactions) without any notable conversion of energy. The ignition is therefore dependent on the starting conditions of mixture formation: •• The pressure and temperature of the charge

4

Figure 13.18 Principle of flow guidance [13-22].

•• The temperature, viscosity, vaporization characteristics, and ignitability of the fuel •• The pressure, time, and characteristic of injection, as well as the nozzle geometry that determines the spray formation (size, distribution, and pulse of the droplets) •• Charge movement •• Charge composition, i.e., the oxygen component and the changes in the specific thermal capacity from the EGR, etc. •• The combustion chamber geometry.

Figure 13.19a Erosion behavior of a Cr–Ni electrode for a standard spark plug after 28,000 km, alteration in the electrode spacing from 0.7 to 1.1 mm; b Erosion behavior of a platinum-reinforced electrode, long service life spark plug after 105,000 km, alteration of the electrode spacing from 1.00 to 1.05 mm.

13.1.8.7 Application In principle, spark plugs must be redesigned for each engine since the requirements are very different due to the type of mixture guidance, the EGR rate, the position of the spark plug, etc. The thermal suitability of a spark plug is ideally evaluated in the original aggregate. The heat range is adapted by measuring ionic current, wherein the changes in combustion are observed, and the preignition and postignition can be observed by blanking individual ignitions (self-ignition). The postignition is uncritical for the engine. In addition, investigations are carried out using thermocouples on the spark plugs in the engine under different speed and load conditions to determine the hottest cylinder, the maximum electrode temperature, and other component temperatures. The spark plug needs to be dimensioned in such a ways so that preignition cannot occur under a full load. From measurements of special “heat range measuring engines,” the temperature of the spark plug can be clearly increased on the test bench by advancing the ignition angle, and the temperature of the individual spark plug components can be determined pyrometrically with an optical access to the cylinder, and the suitability for preignition can be inspected. The heat range reserve can be indicated in degrees for the crankshaft angle to shift the ignition in an early direction without causing preignition.

Autoignition starts locally in the areas with completely evaporated fuel mixed with sufficient atmospheric oxygen. During this phase, injection typically continues and combustion and mixture preparation occur simultaneously. The ignition process is strongly inhomogeneous, since liquid and gaseous phases simultaneously exist with a complex dynamic interaction. The local temperature is the decisive factor in determining the ignition lag and related processes. The fuel-air mixture prepared during the ignition lag burns quickly upon the onset of ignition. The combustion of the fuel prepared subsequently occurs with a slower diffusion combustion. The fuel preparation is further accelerated from the increasing release of energy. High conversion rates in autoignition generate high-pressure gradients and, therefore, usually high noise emissions. To avoid this, the combustion of premixed components is limited as much as possible e.g. by introducing preinjection. The start of combustion, or the moment of ignition timing, must be optimized in relation to exhaust gas emissions, fuel consumption, performance, and noise. Compromises are required since the measures taken within the engine are mutually influential. In passenger cars, engines with direct injection have been predominant in recent years compared to concepts with divided combustion chambers [13-26]. The injection engineering, as well as the equipment for cold-start assistance, have been considerably developed further. We now can have several injections per work cycle at a high maximum injection pressure, a largely free start of injection, and injected fuel quantity to improve fuel consumption and smooth running. Components to support cold starts such as glow plugs were improved to

578 | Internal Combustion Engine Handbook

6606_Book.indb 578

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.2 Diesel Engine

offer reliable support at extremely low start temperatures, faster heating speed, less required energy, and longer life. Passenger car diesel engines are fitted with electric motor starter systems, which have a design which is oriented to the cold start limit temperature which the engine requires to start reliably [13-27]. The ignition is strongly dependent on the initial conditions. In particular, during a cold start, these starting conditions are so poor that satisfactory ignition cannot take place without additional measures.

13.2.2 Cold Start Diesel Engine

The ignition is dependent on the initial conditions to a great extent. So a cold start assist is performed by changing the injection times and quantities already at temperatures below +60°C. As the engine warms up, the smooth running, throttle response or load assumption are enhanced, and the pollutant emissions are reduced. More extensive measures are necessary at temperatures below freezing since the starting quality worsens disproportionately until the temperature decreases so much that the engine cannot be started.

13.2.2.1 Important Influence Parameters

Diesel engine combustion is optimized for the operation of the engine in a hot engine operation status. The selection thereby for the following external parameters substantially influences cold start quality: •• The engine construction type (DI/IDI) •• The number of cylinders

•• The charge volume or respective surface/volume ratio •• The compression ratio •• The starter features (starter output, battery) •• The injection system •• The air guidance and charge •• The internal losses (oil viscosity, gearbox, auxiliary systems …). In contrast to combustion in a hot-running engine, the conditions during a cold start and the following warm-up of the engine are much poorer for autoignition and the subsequent complete combustion of the fuel. The most important influential parameters on the start behavior, and the relationships of the parameters to each other, are shown in the diagram in Figure 13.20. Attention has, therefore, been given to the development of cold-start components and injection systems with more degrees of freedom. For the sake of clarity, the representation of additional relationships such as the direct influence of the temperature on the charge loss (gap dimensions/oil film), or the final compression temperature, are not shown. Low temperatures reduce the battery performance and increase the drive train friction so that the attainable starter speed is lower from the increase in required torque. This increases charge and heat losses because of the longer end phase of compression. The revolution speed of the engine increases at very low ambient temperatures in the range of compression or ignition dead center so that the long dwell time of the hot compressed charge in the combustion chamber significantly decreases the temperature and the charge pressure [13.2-3]. The conditions for the mixture formation and ignition deteriorate dramatically, the temperature versus pressure has a much greater impact on starting quality [13-28], [13-28], [13-30].

Outside temperature/air pressure

Oil viscosity, power unit friction

Battery status

Fuel

Cold start components

Starter rpm

Charging losses from the cylinder

Heat losses of the cylinder charge Injection application

Pressure/temperature of the load

Mixture formation/ignition

Start

Figure 13.20 Important influencing parameters during cold start [13.2-33].

Internal Combustion Engine Handbook | 579

6606_Book.indb 579

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

min–1

Rotary piston, gasoline engines

200 Direct. . Diesel engine without starting aid 160

Rpm n

120

Diesel engine with starting aid

80

40

0

Injection gasoline engines

°C

–25

–20

–15

Carburetor petrol engines –5

–10

0

Temperature u

greatly improved mixture formation. At greater starter speeds, however, the chemical ignition predominates so much that the presented results are still applicable. 100 °KW

Ignition delay [° CA]

As temperatures decrease, higher starter speeds are required to ensure safe cold start. The required minimum start speed and, hence, the cold-start threshold temperature can be greatly lowered by means of start aids that enable starting at temperatures of around −20°C and below (Figure 13.21). The output of the starter and battery is designed for the required coldstart threshold temperature, where a fully charged battery is assumed. If the battery is only half charged, then the threshold temperature rises from, for example, −24°C to −20°C [13-27]. An important factor in this context is the ignition lag,which describes the time from the beginning of injection to ignition. The beginning of injection is determined by the needle lift signal, the solenoid valve, and injector flow or, in the case of optically accessible aggregates, by the exit of the fuels from the injection orifice. The start of combustion can be obtained from the cylinder pressure signal, from an ionic current signal, or optically from light signals. The ignition lag increases exponentially as the charge temperature decreases [13-31] and it has a minimum at an average start speed of approximately 200 min-1 accordingly [13-28], Figure 13.22. This is explained by the overlap of the physical ignition lag with the chemical ignition lag. While the physical ignition delay decreases with increasing starting speed because of improved mixture preparation, the chemical ignition lag thereby decreases [13-32]. This is due to the kinetics of the initial reaction whose time duration is nearly constant. In the illustrated example, the chemical ignition lag at ϑ 0 = −20°C above 200 min–1 is approximately constant at 6 ms. The chemical ignition lag in degrees crankshaft angle thereby increases proportionally to rpm. In contrast to the in-line fuel injection pump used in the cited investigations, modern injection techniques have further lowered the physical ignition lag and

Figure 13.21 Minimum start speed [13.2-2].

u0 = –20 °C

10 u0 = +20 °C 1

Chemical ignition delay

0,1

0,01

0

100

Physical ignition delay

200 300 Rpm [min–1]

400

500

Figure 13.22 Minimum start speed [13-28].

The longer ignition lag under cold-start conditions cannot be compensated at will by advancing injection. Fuel injected too early mixes with the slightly compressed charge until it falls below the ignition threshold or deposits on the combustion chamber walls. It is no longer available for combustion when the pressure and temperature necessary for autoignition are attained because of increasing compression. 13.2.2.2 Start Evaluation Criteria In passenger cars, a reliable, independent start and subsequent stable and smooth engine running are required. There is still no regulation of the exhaust emissions in cold starts below the

580 | Internal Combustion Engine Handbook

6606_Book.indb 580

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.2 Diesel Engine

freezing point in passenger car diesel engines. In evaluating the starting quality, the impairment of driver comfort is the primary focus. This is based on the perception of noise or odor, visible exhaust gas clouds (soot, blue and white smoke), vibrations, the waiting time until the start, the starting time itself, and a poor reaction of the engine to acceleration. The quality of cold starts can be evaluated by measuring the noise level, smoke density, and other exhaust emissions—in particular, HCs—and the evaluation of the speed fluctuations during idling and increases in speed as a response to the quantity of injected fuel Figure 13.23. Despite the technical measuring possibilities which are available for assessing the cold start, the ultimately subjective decision is the driver’s impression, which is much more complex and the absolutely measured quantities are weighted completely differently.

13.2.3 Components for Supporting Cold Starts

As the temperature decreases, the conditions for quick ignition and complete combustion worsen even under otherwise favorable conditions. Without a cold-start aid, the start quality decreases until the start becomes too long for the driver at temperatures below −10°C or even becomes impossible. Aids for supporting cold starts have the task of improving ignition conditions until the combustion in the cylinder is highly effective within the available time limits. The limits are set by the engine processes in the power cycle and are set, on the one hand, by the optimum beginning of injection so that the injected fuel can ignite before it deposits on the combustion chamber wall, or mixes so thoroughly that it falls below the ignition threshold; on the other hand, the limits are set by the maximum available time for complete combustion.

Furthermore, a sufficient amount of fuel must be converted to continue accelerating the engine by a release of energy that exceeds internal losses. To fulfill these requirements, the start of combustion or the maximum rise in pressure must be at top dead center. Typically, the combustion in the cold start is subject to strong cyclical fluctuations so that considerable instabilities arise, including misfiring [13-33]. The task of the cold start aids is to compensate for the deterioration of the starting conditions, particularly in the delayed mixture preparation and to initiate the punctual and smooth ignition for stable combustion. This is done with glow plugs by electrically generated heat that is directly introduced into the combustion chamber and locally promotes mixture formation and ignition. Another approach especially targeted for engines with large displacement is to heat the intake air with flame glow plugs, or electrical heating flanges, that heats the entire air charge in the intake tract to a substantially higher level, so that the injected fuel conditions correspond to those in a hot-running engine. 13.2.3.1 Glow System A heating system with glow plugs as the active heating elements in the combustion chamber, and an electronic control, interprets the commands of the engine management, prepares information on the status of the system, and returns it to the control unit. In modern passenger car diesel engines, the glow plug has become a standard component. For engines with a divided combustion chamber, it is indispensable as a cold start aid to also ensure the start in frequent temperature range of 10 to 30°C. Because of the drastic deterioration in start quality below the freezing point, the glow plug is used as a cold start aid for the direct injection diesel engine [13-29].

2400

2,4

Rotational speed

2000 1800 Temperature, rpm [°C, rpm-1]

2,2

∅n = 980 min–1

2 1,8

∅n = 37 min–1

1600

12,6 1,4

1400 ∅t = 1 s

1200

1,2 1

1000 800 600

0,8

Glow plug temperature

s ∫ Edt = 17 m

0,6

s ∫ Edt = 0,4 m

400

0,4 0,2

200 0

Extinction [m–1]

2200

0

0

2

4

6

8

10

12

14

16

18

20

22

Time [s]

Figure 13.23 Start assessment criteria.

Internal Combustion Engine Handbook | 581

6606_Book.indb 581

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

13.2.3.1.1 Principle The glow plug is typically close to the injection nozzle, but is not directly positioned in the injection jet and extends approximately 3 to 8 mm into the combustion chamber. It offers a comparatively lower heat performance in the form of a hot surface directly in the combustion chamber. The power input is equal to 30 to 150 W in a state of equilibrium, depending on the construction model. The glow plug surface temperature thereby achieves temperatures of around 800 to 1,100°C. The physical and the chemical ignition lag is reduced in the vicinity of the hot heating pin through the accelerated evaporation of fuel droplets and by the rapidly occurring initial reactions at higher temperatures [13-34]. Subsequently, local combustion must be supplied with sufficient energy to independently maintain the flame and ignite fuel injected by the injection jets far enough away from the glow plug, so that all the introduced fuel is completely combusted in the remaining time. The glow plug, therefore, acts as an indirect, local ignition aid; the energy for igniting the majority of the fuel originates from the fuel itself. The glow plug continues to be supplied with current after starting for up to three minutes, depending on the engine temperature (post glowing), to ensure favorable and constant ignition conditions during the engine warm-up phase. The energy introduced into the combustion chamber by preheating the charge or the combustion chamber walls is not decisive, although it still must be taken into consideration. Good start qualities can also be obtained by using quickly heating glow plugs without additional preheating. Moreover, the thermal mass of the metal combustion chamber walls is high enough so that a significant increase in temperature cannot occur within the cited performance range within 3 bis 15 s. The heat supplied to the air charge during the pre-glow process is lost with the first gas exchange. The experience that long preheating phases improve the start quality is based on the fact that a self-regulating glow plug heats a greater area as the heating time increases and, hence, saves more heat energy. The glow plug cools less while the starter operates, as is the case with shorter preheating phases. It is frequently assumed that a “hot spot,” a comparatively minuscule hot point, is sufficient for ignition. Since the locations with favorable ignition conditions fluctuate strongly from cycle to cycle, in particular, during a cold start, and a large thermal mass reduces temperature fluctuations at the glow element, a “hot area” or “hot volume” is necessary in practice.

Plug-in connection

Internal pole

13.2.3.1.2 Requirements The glow plug should provide a sufficiently high temperature over the shortest possible time to support ignition, and maintain this temperature independent of the momentary conditions or even adapt it to the conditions. The available installation space for the glow plug is particularly limited in modern engines in four-valve technology with pump-nozzle elements or injectors, so that the glow plugs must be designed as slim as possible, but on the other hand, must have a certain robustness [13-35]. Linked to this is often an installation situation that results in high costs for changing the glow plugs, so that the glow plugs should have a “life of the engine” service life. Since the load on the vehicle electrical system is particularly critical during cold starts, the glow plugs require a minimal power input. Legal regulations require a permanent monitoring system (OBD) for emissions-relevant components. This is realized in glow plug systems by monitoring each individual glow plug and providing feedback to the engine control unit. There are other possibilities of influencing emissions within the engine, with an electronic glow plug system. By intermediate heating, i.e., turning on the glow plugs again when the aggregates cooled during overrunning, controlled combustion is ensured with minimum emissions. 13.2.3.1.3 Design Glow plugs consist of a metal resistance heating element wound into a coil that is protected from combustion chamber gases by a metal sheath resistant to hot gas corrosion. In this glow tube, the coil is embedded in compressed magnesium oxide powder that provides electrical isolation, good heat transmission, and mechanical stability. This component forms the heating element together with the supply of current to the helical heating wire. This is pressed into a body with a sealing seat, a thread, and a hexagonal head that is used to screw the glow plug into the cylinder head and thereby creates the ground contact. The current is transferred to the heating element with a threaded, or plug-in, connection. A standard dimension for a glow plug is an M10 thread and a 5 mm diameter heating element. The length and the head shape vary depending on the requirements (Figure 13.24). (a) Self-Regulating Glow Plug In self-regulating glow plugs, the coil consists of a combination of a helical heating wire and a control filament. The

Body

Sealing seat Glow tube

Wendel combination

Figure 13.24 Glow plug design.

582 | Internal Combustion Engine Handbook

6606_Book.indb 582

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.2 Diesel Engine

clocked and therefore reduces the voltage present at the glow plug to effectively 5V. The desired temperature on the glow plug can therefore be retained, as soon as a vehicle electrical system voltage of more than 5 V is available. The glow plug temperature is, hence, independent from the vehicle system voltage that is frequently only 7 bis 9 V, particularly during starter operation.

helical heating wire consists of a high-temperature resistant material whose electrical resistance is largely independent of temperature, whereas the resistance of the control filament has a large positive temperature coefficient. With cold glow plugs, first a high current arises that quickly heats the helical heating wire. Through heat conduction and self-heating, the control filament subsequently becomes increasingly hot so that overall resistance increases and the current decreases. We, therefore, have a combination of fast heating with independent regulation at a top constant temperature. By selecting the control material and the resistance division between the heating and control filament, various characteristics in the temperature curve can be represented.

In running engines, the glow plug is cooled by the charge cycle and air movement in the compression phase. The temperature of the glow plug decreases as the rpm increases at a constant glow plug voltage and injected fuel quantity, whereas the temperature increases with increasing injected fuel quantity and constant glow plug voltage and rpm. These effects can be compensated for with the aid of the electronic control unit, by always supplying the optimum effective voltage to the glow plugs which is required for the respective operating point. Other influencing variables are compensated in an analogous manner. The glow plug temperature can, hence, be applied depending on the operating status.

The glow plug is controlled via a relay or an electronic switch, and its nominal voltage corresponds to the voltage provided by the vehicle electrical system when the engine idles. The glow plug is cooled by the movement of air during starting, and while the engine is running. This is, however, compensated by the higher available vehicle electrical system voltage, so that the desired temperature is maintained during post glowing (Figure 13.25).

Furthermore, the combination of a low-voltage glow plug with an electronic control unit is used to heat the glow plug extremely quickly, by applying the full vehicle system voltage to the glow plug for a predefined time, and only subsequently cyclically applying the required effective voltage. The hitherto conventional preheating time is reduced to very low temperatures to a maximum of 2 s. This enables start times similar to gasoline engine.

(b) ISS Glow Plug System (Instant Start System) A quick-start glow system comprises an electronic control unit and a quick-start glow plug [13-36]. The design is somewhat similar to the self-regulating glow plug, whereby the coil combination is, however, considerably shortened and the glowing area is reduced to about one third. In diesel engines with direct injection, this corresponds to the part of the heating element extending into the combustion chamber. On the downside, the power demand is two to three times less, which is particularly important in engines with 8 or more cylinders.

A start without preheating is also possible due to the high dynamics of the glow plug system. At low temperatures, it is nevertheless logical to set short preheating times which can coincide with the necessary initializations, checks etc. In principle, improved ignition conditions are in place from the very beginning with existing hot glow plugs.

The glow plug is designed to operate with a reduced nominal voltage of e.g. 5 V compared to vehicle system voltage, with which this glow plug reaches a steady-state temperature of approximately 1,000°C. By utilizing the electronic control unit, the on-board network voltage is

The electronics also assume protective functions for the glow plug and communicate with the engine control unit (for OBD). With its expanded degrees of freedom, the glow plug system will be used in the future in the application

1200

5000 Glow plug temperature

5000

800

4000

600

3000 Rotational speed

400

2000

200 0 –15

Rpm [min-1]

Glow plug temperature [°C]

1000

1000

–10

–5

0

5

10

Time [s]

15

20

25

30

35

0

Figure 13.25 Start with a selfregulating glow plug.

Internal Combustion Engine Handbook | 583

6606_Book.indb 583

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

phase to optimize the internal engine combustion processes and the service life of the glow plugs. 13.2.3.2 Heating Flange Today, electrical heating flanges are primarily used for commercial vehicle engines with a piston displacement greater than 0.8 liter per cylinder. They allow a reliable start at low temperatures and the reduction of smoke emissions [13-37]. As demands increase for the reduction of emissions during cold starts and the improvement of driving comfort, correspondingly adapted electrical heating flanges are becoming interesting for passenger car applications. 13.2.3.2.1 Principle The heating flange with a 0.5 to 2 kW, connecting cable is installed in front of, or inside, the intake manifold. The electrical output is converted into heat in the heating flange and released to the intake air. Normally, metal heater elements do not have temperaturedependent resistance. Moreover, there are the heating flanges with PTC characteristics, which are realized with metallic or ceramic elements [13-37]. The optimum characteristic for good starting performance characteristics can be supported with appropriate, controllable power electronics. The heating flange should heat up the intake air temperature by at least 30 K. Figure 13.26 illustrates the relationship of the polytropic compression T2 = T1 ⋅ ε n–1 calculated theoretical increase of compression temperature T2 on the intake air temperature T1 for various compression ratios and a polytropic exponent of n = 1.37.

300 e = 18,5

n = 1,37 250 e = 21

e = 16

T2 [K]

200 150

T2 theor. = 457°C. This value would not be achieved due to the heat and charge losses and the lower intake air density. For the relative estimation of the “reinforcement” of the air temperature increase in the polytropic compression, this relationship still has validity despite the absolute lower final compression temperature. In the literature, polytropic exponents of n = 1.2 to 1.3 for the calculation of the thermodynamic state are commonly utilized in the cold start. These exponents result from pressure measurements, taking into consideration the heat and charge losses. However, in an integrating combustion chamber temperature measurement, Rau [13.2-3] has shown that the exponent for calculating the temperature is close to the theoretical exponent, which is in the relevant range (250 K < T1 < 830 K) n = 1.38. Because of the globally higher charge temperature, the heating flange improves mixture preparation and substantially reduces ignition lag. 13.2.3.2.2 Requirements For the electrical heating flange, a short heating time and good heat transfer from the heater element to the air are required with minimal flow resistance in the intake air duct. The electrical connecting cable must heat the intake air as much as possible without overloading the vehicle’s electrical system. The available installation space is determined by the intake air cross section. The dynamic changes in the intake air flow speed require that the heating flange have sufficient thermal mass to prevent fast cooling and overheating of the element. For the OBD, the function of the heating flange can be monitored with the aid of an electronic control unit and sent to the engine control unit. 13.2.3.2.3 Design The electrical heating flange consists of an approximately 20 mm wide frame or flange in the intake air guide. It assumes a sealing function, connects to and guides power, and takes on the control electronics and the heater element, including insulation. The heating element consists of one or more metal strips that typically meander-shaped with approximately five windings in a ceramic insulator and are connected at one side to the frame for a ground connection (Figure 13.27).

100 Framework

50 0

0

20

40

60

80

Mass connection

Heating cable

100

Support for ceramic insulation

T1 [K]

Figure 13.26 Compression temperature increase with intake air.

As a result, for example, this leads to a compression ratio of 18.5 ε = with an increase in the intake air temperature to ∆T1 = 50 K and an increase in the final compression temperature to ∆T2 = 147 K. The aforementioned relationship yields much higher values for the compression temperature in cold start. With a given T1 = −25°C, there would be, e.g., with the above data

Supply connection

Figure 13.27 Heating flange.

584 | Internal Combustion Engine Handbook

6606_Book.indb 584

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.2 Diesel Engine

13.2.3.2.4 Function When current is applied to the heating flange, the heater element achieves 900 to 1,100°C and heats the stationary, surrounding air. Preheated air is inducted and compressed once the starter is activated. The higher global charge temperatures improve ignition conditions. The heating flange heats the now flowing air in the intake tract by approximately 50°C and is thereby cooled to 500 to 600°C (Figure 13.28). The thermal mass of the heater element buffers fast changes in the air flow, and slow changes are compensated by the self-regulating behavior of the heating tape or an electronic control. Because of the heat output of the heating flanges, which is substantially greater than that of a glow plug, ignition conditions are achieved quickly in the entire combustion chamber that, together with the adaptation of fuel injection, substantially reduce smoke emissions in the warm-up phase [13-37] (Figure13.29).

13.2.4 Outlook 13.2.4.1 Combined Systems The glow plug system is the suitable cold start aidfor diesel engines in passenger cars to ensure the fastest start with a minimum drain on the vehicle electrical system. In contrast, electrical heating flanges have the potential to further reduce warm-up emissions, improve smooth engine running, and improve load assumption. An electrical heating flange therefore makes sense in view of increasingly stringent exhaust regulations to combine both systems to attain fast starts with minimum emissions and maximum smooth running. This solution is also particularly recommendable when there are many cylinders and a large charge volume. 13.2.4.2 Ionic Current Measurement The ionic current measurement is already being used to gain information about the combustion directly from the combustion

50 Heating tape temperature

800

40

Intake air temperature [°C]

Heating tape temperature [° C]

1000

30

600

400

Intake air temperature

Start

20

10

200

0

0

10

20

30

Time [s]

40

50

60

0

Figure 13.28 Heating behavior of a heating flange.

100

80

Opacity [%]

Without jump-start 60

40 2,2 kW heating flange 20

0 –25

–20

–15

–10

–5

Temperature [°C]

0

5

10

15

Figure 13.29 Opacity 30 s after the start [13-37].

Internal Combustion Engine Handbook | 585

6606_Book.indb 585

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 13 Ignition

chamber in the field of gasoline engines [13-38]. To prevent the requirement for additional probes being introduced into the combustion chamber, glow plugs have a favorable position for diesel engines [13-39] and the possibility of soot oxidation on the electrodes on. If the heating element is isolated from the glow plugs and a voltage is applied, then an electrical field forms in the combustion chamber around the glow plug tip. The charges of the particles in the field flow through the electrode. A current of a few microamperes to milliamperes can be measured by a suitable circuit, then amplified and possibly prepared and sent to the engine control unit (Figure 13.30). Insulation

I

Auxiliary power

Cylinder head

RM

attain the isolation of the heating element from the cylinder head required for measuring ionic current, a separate ground connection is necessary between the heating element to the cylinder head, which can then be interrupted for the ionic current measurement. A switching circuit for this implementation will be integrated in the glow plug, so that the external design of the glow plug is not altered. 13.2.4.3 Regulated Glow Plug System The self-regulating glow plugs that are frequently used today will be increasingly replaced in the future by electronically controlled systems. The next objective is to develop regulated systems that do not require complex calculations of the control output depending on the engine parameters. Instead, a higher-level engine control unit should only transmit the required amount of heating in the form of a set point to the glow plug control device, which then interprets the set point and correspondingly regulates the required voltage sent to the glow plug. To attain this objective, glow plugs must be developed that can return an easily evaluated and stable temperature signal to the glow plug control device.

Bibliography

13-1. Heywood, J. B. 1989. Internal Combustion Engine fundamentals. McGraw: New York.

Ionic current Measured voltage

13-2. Autoelektrik. 1987. Autoelektronik am Ottomotor. Bosch: VDI-Verlag. 13-3. Albrecht, H., Maly, R., Saggau, B., and Wagner, E., “Neue Erkenntnisse über elektrische Zündfunken und ihre Eignung zur Entflammung brennbarer Gemische,” Automobil-Industrie (4): 45–50, 1977.

Figure 13.30 Ionic current measurement principle.

The diesel engine combustion is especially subjected to local stochastic fluctuations [13-40]. This means that the ionic current measured at the glow plug, in contrast to the integrating cylinder pressure signal, can sometimes only be partially taken for an indirect determination for thermodynamic information such as combustion function, the location of the combustion center of gravity, etc. with increased mathematical complexity. Measuring ionic current using glow plugs is cheaper in comparison to indexing cylinder pressure and represents a robust internal engine sensing mechanism that can be continuously evaluated. Potential applications of ionic current measurement are, for example •• Detection of misfiring •• Cylinder equalization at the start of combustion, balancing tolerances in the injection and intake system, etc. •• Fulfillment of OBD requirements by direct feedback from the combustion chamber •• Compensation for differing fuel quality. To realize “ionic-current-regulated” diesel engines, substantial efforts are being made at present to develop corresponding evaluation algorithms and regulating structures to represent the correlation of the measured signals with the processes in the combustion chamber. Furthermore, the position of the sensor and its design for long-term use must be optimized. To

13-4. Maly, R., and Vogel, M. “Ignition and Propagation of Flame Fronts in Lean CH4-Air Mixtures by the Three Modes of the Ignition Spark,” Proceedings of 17th International Symposium on Combustion, pp. 821–831, The Combustion Institute, 1976. 13-5. Schäfer, M. “Der Zündfunke,” Dissertation at the University of Stuttgart, 1997. 13-6. Hohner, P. “Adaptives Zündsystem mit integrierter Motorsensorik,” Dissertation at the University of Stuttgart, 1999. 13-7. Maly, R. “Die Zukunft der Funkenzündung,” Motortechnische Zeitschrift 59(7/8), 1998. 13-8. Herweg, R. “Die Entflammung brennbarer turbulenter Gemische,” Dissertation at the University of Stuttgart, 1992. 13-9. Schommers, J., Kleinecke, U., Miroll, J., and Wirth, A. “Der neue Mercedes-Benz Zwölfzylindermotor mit Zylinderabschaltung,” Teil 2, Motortechnische Zeitschrift 61(6), 2000. 13-10. Stocker, H., Archer, M., Houston, R., Alsobrooks, D., and Kilgore, D. “Die Anwendung der luftunterstützten Direkteinspritzung für 4-Takt Ottomotoren—der Gesamtsystemansatz‘,” 7. Aachener Kolloquium Fahrzeug- und Motorentechnik, pp. 711ff, 1998. 13-11. van Basshuysen, R. Ottomotor mit Direkteinspritzung—Verfahrene Systeme × Entwicklung × Potenzial. Vieweg+Teubner:Wiesbaden, 2008. 13-12. Tschöke, H., Schultalbers, M., Gottschalk, W., Huthöfer, E.-M., and Jordan, A. “Thermodynamische Optimierungskriterien für die Zündzeitpunktabstimmung moderner Ottomotoren,” Motortechnische Zeitschrift, Wiesbaden, 2011. 13-13. Babic, G., and Bargende, M. “Betriebsstrategien für die Benzinselbstzündung,” Motortechnische Zeitschrift, Wiesbaden, 2010. 13-14. Groß. V., Kubach, H., Spicher, U., Schiessl, R., and Maas, U. “Laserzündung und Verbrennung im Ottomotor mit Direkteinspritzung,” Motortechnische Zeitschrift, Wiesbaden, 2010.

586 | Internal Combustion Engine Handbook

6606_Book.indb 586

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

13.2 Diesel Engine

13-15. Heuermann, H., Theobald, J., and Olivetti, G. M. “Hochfrequenzzündung—Lohnt sich die Technik?—Forum der Meinungen,” Motortechnische Zeitschrift, Wiesbaden, 2010.

13-29. Petersen, R. “Kaltstart- und Warmlaufverhalten von Dieselmotoren unter besonderer Berücksichtigung der Kraftstoffrauchemission,” VDI-Fortschrittsbericht, 6(77), 1980.

13-16. Meyer, J.; Niessner, W.: Neue Zündkerzentechnik für höhere Anforderungen. In: ATZ/MTZ Sonderausgabe (Special Edition), System Partners 97

13-30. Reuter, U. “Kammerversuche zur Strahlausbreitung und Zündung bei dieselmotorischer Einspritzung.” Dissertation RWTH Aachen, 1990.

13-17. Eichlseder, H.; Müller, P.; Neugebauer, S.; Preuß, F.: Innermotorische Maßnahmen zur Emissionsabsenkung bei direkteinspritzenden Ottomotoren, TAE Esslingen, Symposium: Entwicklungstendenzen Ottomotor, 7./8. 12. 2000

13-31. Pischinger, F. “Verbrennungsmotoren. Vorlesungsumdruck.” RWTH Aachen, 1995.

13-18. Pischinger, S.; Heywood, J. B.: Einfluss der Zündkerze auf zyklische Verbrennungsschwankungen im Ottomotor. In: MTZ 52 (1991) 2 13-19. Lee, Y. G.; Grimes, D. A.; Boehler, J. T.; Sparrow J.; Flavin, C.: A Study of the Effects of Spark Plug Electrode Design on 4-Cycle Spark-Ignition, Engine Performance, SAE, 2000-01-1210 13-20. Geiger, J.; Pischinger, S.; Böwing, R.; Koß, H.-J.; Thiemann, J.: Ignition Systems for Highly Diluted Mixtures in SI-Engines, SAE, 1999-01-0799 13-21. Kaiser, Th.; Hoffmann, A.: Einfluss der Zündkerzen auf das Entflammungsverhalten in modernen Motoren. In: MTZ 61 (2000) 10 13-22. Maly, R.: Die Zukunft der Funkenzündung. In: MTZ 59 (1998) 7/8 13-23. Osamura, H.; Abe, N.: Development of New Iridium Alloy for Spark Plug Electrodes, SAE, 1999-01-0796 13-24. Brill, U.: Krupp-VDM, private communications, 1994 13-25. van Basshuysen, R.: Ottomotor mit Direkteinspritzung—Verfahrene Systeme × Entwicklung × Potenzial. Wiesbaden. Vieweg+Teubner, 2008 13-26. Bauder, R. “Die Zukunft der Dieselmotoren-Technologie,” Motortechnische Zeitschrift 59 (7/8), 1989. 13-27. Henneberger, G. 1990. Elektrische Motorausrüstung. Vieweg: Wiesbaden. 13-28. Rau, B. “Versuche zur Thermodynamik und Gemischbildung beim Kaltstart eines direkteinspritzenden Viertakt-Dieselmotors.” Dissertation Technische Universität Hannover, 1975.

13-32. Sitkei, G. Kraftstoffaufbereitung und Verbrennung bei Dieselmotoren. Berlin: Springer, 1994 13-33. Zadeh, A., Henein, N., and Bryzik, W. “Diesel Cold Starting: Actual Cycle Analysis under BorderLine Conditions.” SAE 909441, 1990. 13-34. Warnatz 1996. Technische Verbrennung. Springer Verlag:Berlin 13-35. Endler, M. “Schlanke Glühkerzen für Dieselmotoren mit Direkteinspritzung,” Motortechnische Zeitschrift 59 (2), 1998. 13-36. Houben, H., Uhl, G., Schmitz, H.-G., and Endler, M. “Das elektronisch gesteuerte Glühsystem ISS für Dieselmotoren,” Motortechnische Zeitschrift 61 (10), 2000. 13-37. Merz, R. “Elektrische Ansaugluft-Vorwärmung bei kleineren und mittleren Dieselmotoren,” Motortechnische Zeitschrift 58 (4), 1997. 13-38. Schommers, J., Kleinecke, U., Miroll, J., and Wirth, A. “Der neue Mercedes-Benz Zwölfzylindermotor mit Zylinderabschaltung; Teil 2,” Motortechnische Zeitschrift 61 (6), 2000. 13-39. Glavmo, M., Spadafora, P., and Bosch, R. “Closed Loop Start of Combustion Control Utilizing Ionization Sensing in a Diesel Engine.” SAE paper 1999-01-0549. 13-40. Ernst, H. “Zündverzug und Bewertung des Kraftstoffs,” Deutsche Kraftfahrtforschung, (63), VDI-Verlag: Düsseldorf,1941. 13-41. Mollenhauer, K. and Tschöke, H. “Handbuch Dieselmotoren,” 3. Auflage. Springer: Berlin; Heidelberg, 2007.

Internal Combustion Engine Handbook | 587

6606_Book.indb 587

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 588

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14 Combustion 14.1 Fuels and Fuel Chemistry Fuels for gasoline spark ignition (SI) engines and compression ignition (CI) diesel engines are normally refined from mineral oil in a distillation process and consist of hundreds of individual components. This composition has a very great influence on the physical and chemical properties and, thus, on the combustion characteristics. During the production of fuels from coal, a synthesis gas is initially produced by hydrogenation or gasification, which is subsequently converted using an appropriate synthesis method to produce an alternative fuel, for example, methanol (CH3OH) or methane (CH4). Fuels produced from coal, as well as nonfossilized fuels such as rape oil or rape methyl ester that are obtained from biomass, are only of subordinate importance for motor vehicle applications and are utilized at best in niche markets today. We focus in this chapter on the classification and chemical structure of simplified hydrocarbons, the so-called CxHyOz compounds, insofar as it is necessary for an understanding of the oxidation of hydrocarbons. Hydrocarbon compounds are generally divided into alkanes (previously paraffins), alkenes (olefins), alkines (acetylenes), cycloalkanes (naphthenes), and aromatics. The alkanes (paraffins) are chain-like or “aliphatic” hydrocarbons with purely single bonds (monovalent), while a distinction is made between the normal alkanes with a straight-chain structure and isoalkanes with a branched-chain structure. The alkenes (olefins) are chain-like hydrocarbons with one or two double bonds, while alkenes (monoolefins) have one and alkadienes (diolefins) two double bonds. The alkines (acetylenes) also have a chain-like structure and a triple bond. Figure 14.1 illustrates the structural formulas for these aliphatic hydrocarbons.

Alkanes CnH2n+2 (previously paraffins) Aliphatic built-up hydrocarbons with only mit Kettenförmig aufgebaute Kohlenwasserstoffe single bonds nur Einfach-Bindungen Iso-Paraffine Normal-Paraffine Normal-Paraffine Iso-Paraffine Straight-aliphatic Branched-aliphatic gerad-kettenförmig verzweigt-kettenförmig H CH3 H

H H H C

C H

H H

H

C

C

C

H

H CH3 H propane 2,2 Dimethyl Dimethylpropan

Ethan

Alkenes CnH2n (previously Olefin) Kettenförmig aufgebaute Kohlenwasserstoffe mit Aliphatic built-up hydrocarbons having double bond (DB)(DB) Doppel-Bindung Alkenes (mono-olefins) Alkadienes (Diolefines) Alkene (Monoolefine) Alkadiene (Diolefine) Aliphatic, oneeine DB DB Aliphatic, two DBDB kettenförmig, kettenförmig, zwei H H C

CnH2n-2 H

C

H

C C C

H H Ethen

H H Propadiene(Allen)

Alkanes CnH2n-2 (previously acetylene) Aliphatic built-up hydrocarbons with only mit Kettenförmig aufgebaute Kohlenwasserstoffe triple bonds einer Dreifach-Bindung H C

C H

Ethyne Ethin

Figure 14.1 Aliphatic hydrocarbon compounds.

Internal Combustion Engine Handbook | 589

6606_Book.indb 589

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

The structural formulas of the cycloalkanes (naphthenes), with their circular structure and purely single bonds, and of the circular aromatics with their double bonds whose basis is the benzene ring are illustrated in Figure 14.2 accordingly.

Cyclo-Alkanes CnH2n (previously naphtene) n Ring-formed built-up hydrocarbons with single t bonds

H

H H H H

C C

C

H H

Cyclopropane

H H C H C H

H

C

C

H C H C H H

H H 1,2 Cyclohexanol

Aromatics Ring-formed built-up hydrocarbons with conjuncted double binding basic modules is the benzene ring

H

CH3

C

C

Alcohole , R - OH Containing a Hydroxyl Group –OH Methanol Ethanol (Methyl alcohol) (Ethyl alcohol) H H H C2 H5 OH CH3 OH H C

OH

H C

H H H

C

C

H H O C

H H

C

H

H H

Diethyl ether C2 H5 O C2 H5

Ketone, R1 - CO - R2 Are bonded together with hydrocarbon residues via a CO-Carbonyl Group H C

H C

C

Acetone CH3 C

H

H O H

C

H

H C

C

H

H C

C

H

H C

C

CH3

C

C

H

H

Benzene

1,3 Dimethyl Benzene

Figure 14.2 Alicyclic and aromatic hydrocarbon compounds.

Oxygenated hydrocarbons are chain-like compounds for which a distinction is made between alcohols, ethers, ketones, and aldehydes. Alcohols contain a hydroxyl group (R–OH). The simplest alcohols are methyl alcohol (methanol: C3H–OH) and ethyl alcohol (ethanol: C2H5–OH). Ethers are hydrocarbon residues linked together by an oxygen bridge (R1–O–R2) and ketones residues linked by a carbonyl group (R1–CO–R2). Aldehydes contain a CHO group, for example, formaldehyde (HCHO). The structural formulas of the oxygenated hydrocarbons are illustrated in Figure 14.3, where the CHO groups should not be confused with the OH group (–COH) attached to the carbon.

OH

H H H Ether, R1 - O - R2 Are bonded together with hydrocarbon residues (R1,R2) via an O-Bridge

H

H C

C

CH3

O

Aldehyde, R - CHO Containing a CHO-Group Formaldehyde H H C O Figure 14.3 Aliphatic hydrocarbon compounds.

Two-component substitute fuels are used to determine the ignition quality of SI and CI engine fuels, namely the substitute fuel consisting of: •• n-heptane (C7H16) with the octane number OZ = 0 and •• iso-octane (C8H18) with the octane number OZ = 100 consisting of substitute fuel for benzene and those from: •• α -methyl naphtalene (C11H10) with the cetane number CZ = 0 and •• n-hexadecane (cetane: C16H34) with the cetane number CZ = 100 consisting of substitute fuel for the diesel fuel, whereby the octane number is defined as the iso-octane content and the cetane number as the cetane content of the two component substitute fuel. The structural formulas of the components of the two substitute fuels are illustrated in Figure 14.4.

590 | Internal Combustion Engine Handbook

6606_Book.indb 590

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.2 Oxidation of Hydrocarbons

equation, but has a very complex reaction pattern based on elementary reactions that only today has started to be roughly understood and is illustrated schematically in Figure 14.5.

Components for the comparative combustion for cetane number determination CH H 3

H

H

H H

H

C

C ... C

C

H

H H

H

C H

C

H

C

C

C

H

H

C

C

C

H

C

Alkane

ROOH

C

H H α-Methyl Naphthalene C11 H10 CZ = 0

Cetane C16 H34 CZ = 100

Dehydrating

H, O, OH

Components for the comparative combustion for octane number determination H CH3 H CH3 H H

H

C

C

ISO-Octane C8 H18 OZ = 100

C

C

H CH3 H H

H

H

H H H

C

H H H C

H

C

C

C

H H H

C

H

C

No alkanes

C

H H H

H

Thigh

n-Heptan C7 H16 OZ = 0

Aldehyde HCHO, CH3CHO

While a lower ignition quality and, hence, a higher knock resistance is required for SI gasoline engines, the opposite is the case for CI (diesel) engine fuels. The octane number drops with increasing numbers of hydrocarbon atoms for n-alkanes and alkenes, while increasing the number of branches for isoalkanes and the number of components with double bonds. The lower net calorific value for the combustion of hydrocarbon compounds lies in the range of: 40.2 MJ/kg (benzene) < Hu < 45.4 MJ/kg (n-pentane) The maximum laminar flame speed of the liquid fuel components in air at 1 bar is only approximately 2 m/s, on the other hand turbulent flame speeds of up to 25 m/s occur during combustion of these components in the engine.

14.2 Oxidation of Hydrocarbons With complete combustion, hydrocarbon compounds (CxH) are converted into carbon dioxide (CO2) and water vapor (H2O). This reaction can be generally described by the gross reaction equation y y C x H y + ⎛⎜ x + ⎞⎟ O 2 ! x ⋅ CO 2 + H 2 O + ∆ R H , ⎝ 4⎠ 2

Slight alkenes, alkadines C2H4, C3H6, (C2H2)

Pollutants

Figure 14.4 Structural formulas for the components of the substitute fuels for gasoline and diesel engines.



Chain design Important for all subsequent reactions

low RQ

(14.1)

where the reaction enthalpy ∆R H represents the heat released by the combustion process. In reality, however, the combustion does not take place as described by this gross reaction

CO, H2, H2O medium OH

medium RQ high RQ

CO2, H2O

Figure 14.5 Hydrocarbon oxidation process.

In a first reaction phase, hydrocarbon peroxides (ROOH) are formed that are broken down into small alkanes by dehydrogenization. These reactions are the decisive factors in the ignition process in motorized applications and will be dealt with in more detail in Section 14.3. The subsequent reactions with the radicals H•, O•, and OH• from the fuel create a β -disintegration [14-1]. Chain propagators initially create light alkenes and alkadienes, and ultimately aldehydes, such as acetaldehyde (CH3CHO) and formaldehyde (HCHO). At higher temperatures the formation of hydrogen peroxides is circumvented. Alkenes are directly formed instead. The formation of the aldehydes, during which only about 10% of the released heat is produced, is accompanied by the occurrence of a cold flame. Further, CO, H2, and already H2O are formed in the following subsequent blue flame and, in the last stage, CO2 and H2O are ultimately formed in the hot flame. During the oxidation of the hydrocarbons to CO, approximately 40% is released and, finally, during the oxidation of the CO to CO2, the remaining 45% of the thermal energy stored in the fuel is released. The main release of heat, therefore, only takes place initially at the end of the reaction process during the oxidation of CO to CO2.

Internal Combustion Engine Handbook | 591

6606_Book.indb 591

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

16

CxHy, O2

CO2, H2O [%]

ci Temperature H, O, OH

CO2

C2H4, C3H6, C2H2 HCHO, CH3CHO, H2

H2O

Figure 14.6 Curve for the temperature and concentration process over time during hydrocarbon combustion.

To calculate the temperature and concentrations in the flame front, it can be assumed that the eight components H, H2, O, O2, OH, CO, CO2, and H2O in the flame front are in partial equilibrium because of the high temperature prevailing there. This so-called OHC-system is therefore described by the five reaction equations

H2 = 2 H

(14.2)



O2 = 2 O

(14.3)



H2O =

1 H 2 + OH 2

(14.4)

H2O =

1 O2 + H2 2

(14.5)

1 CO 2 = CO + O 2 , 2

(14.6)

whereby the following are valid for the five equilibrium constants

K1 = [H] [H2 ]

(14.7)



K 2 = [O] [O2 ]

(14.8)

−1

2

−1

2



K 3 = [H2 ]



K 4 = [O2 ]



K 5 = [ CO ][ O 2 ]

1/2

[ OH ][ H 2 O ]



(14.9)

[ H ][ H O ]

−1



(14.10)

−1

.

(14.11)

1/2

−1

2

2

1/2

[ CO ] 2

CO2

12

H2O

3

10

CO

8 6

Together with the atom balances for the atoms O, H, and C (better CO) and the condition that the sum of the partial pressures of all the components has to be equal to the total pressure, we ultimately obtain a nonlinear equation system that can be clearly solved using conventional numeric integration methods, for example, the Newton–Kantorovitch method. Figure 14.7 illustrates an example for the concentration distribution of the OHC components as a function of the temperature for the total pressure of 1 bar.

2

O2

OH

2 0 1500

t



14

4

CO



4

2000 2500 Temperature [K]

1

CO, O2, OH, H2, O, H [%]

Figure 14.6 illustrates qualitatively the temporal concentration and temperature curve during the hydrocarbon combustion.

H2 H O 0 3000

Figure 14.7 Partial equilibrium of OHC components as a function of the temperature for the total pressure 1 bar.

If, for a subsequent calculation of the thermal nitrogen formation, only the oxygen atom concentration is required, this can also be approximately calculated from the equation

[ O ] = 130[ O 2 ]

1/2



exp ⎛⎜ − ⎝

29468 ⎞ ⎟ T ⎠

(14.12)

accordingly. Refer to [14-2] for further details.

14.3 Autoignition Ignition is the transition of an nonreactive fuel–air mix to combustion. Ignition processes can be subdivided in the categories of thermal explosion and chain explosion. According to Semenov’s analysis [14-3], a thermal explosion occurs when the chemical thermal production exceeds the thermal loss on the combustion chamber walls. This type of ignition has a direct temperature increase without any delay. The opposite is the case with chain explosions whereby there is normally an ignition-delay time at constant temperature in the process. Radicals which serve as the chain carriers are initially formed during this time. Once there are a certain amount of this radicals present in the system, there is enough heat release available for a temperature increase and a subsequent explosion. The reactions of a chain explosion are subdivided into start-up reaction, reproduction reaction, branching reaction, and abort reaction. Important radicals are, for example, the atoms O – and H–, as well as the hydroxyl radical (OH–), the hydroperoxyl radical (HO•2) and the methyl radical (CH•3). Start reactions form radicals from stable species such as for example in the reaction between methane and molecular oxygen:

CH 4 + O 2 → CH i3 + HO i2

(14.13)

Reproduction reactions receive the number of radical species

CH 4 + OH i → CH i3 + H 2 O

(14.14)

Excess, superfluous radicals are formed in branching reactions

CH 4 + O i → CH i3 + OH i .

(14.15)

592 | Internal Combustion Engine Handbook

6606_Book.indb 592

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.3 Autoignition

The number of radical species will be reduced in the abort reactions, for example, in the recombination reaction for methyl radicals: CH i3 + CH i3 → C2 H 6 .



(14.16)

Chain aborts can also result when radicals collide with combustion chamber walls, a mechanism that is particularly important at lower pressures.

14.3.1 The H2–O2 System

The H2–O2 system contains a comparatively simple ignition mechanism and plays an important role both in examining hydrogen combustion, as well as also as a subset in reaction mechanisms for more complex fuels. Despite the simple characteristics of fuels, there are already approximately 25 reactions between eight different species, H2, O2, OH–, H2O, H–, O•, HO•2, and H2O2, which have to be considered for hydrogen ombustion. The most important reactions with regard to ignition are [14-1]

H 2 + O 2 ! HO i2 + H i

(14.17)



i

H 2 + OH ! H 2 O + H

(14.18)



H i +O 2 ! O i + OH i

(14.19)



O + H 2 ! H + OH

(14.20)



H → 0.5 H 2

(14.21)



H + O 2 + M ! HO + M .

(14.22)

i

i

i

i

i

i 2

Reaction (14.21) hereby illustrates a wall abort, the trimolecular reaction (14.22) is in principle a reproduction reaction, but it can still be considered to be a chain abort as the resulting HO•2is a relatively inert radical. The influences from the various reactions on the ignition can be clarified on the basis of an explosion diagram, that is, Figure 14.8. An ignition will not occur at constant temperature

Ex

Explosion

plo

sio

nL

The ignition of hydrocarbons, like the hydrogen ignition, is considered a chain process. However, hydrocarbons have significantly more complex ignition mechanisms with many more participating species and reactions. As with the hydrogen combustion, there are three explosion limits in explosion diagram in hydrocarbons. At high temperatures and pressures above about 1100 K, the reaction (14.19) is the dominant chain branching in hydrocarbons. In this area, the oxidation of the fuel takes place according to the scheme discussed in Section 14.2. In motor applications, the temperature is usually below 1000 K after compression. In this area occur with hydrocarbons, particularly in alkanes, additional, more complex firing mechanisms. The branching reaction (14.19) is strongly temperature dependent and disappear at T < 1100 K rapidly. The ignition in the lower- and medium-temperature range is characterized by the appearance of the so-called two-stage ignition. Here, the first heat release increases by a first ignition phase and then goes above about 900 K. After exceeding about 1000 K, a second firing phase is introduced which results in the complete oxidation of the fuel. The exact temperatures are thereby very dependent on the pressure. The two-staged ignition clarifies the occurrence of the schematically illustrated, negative temperature coefficients (NTC) in Figure 14.9, which describes the fact that the ignition delay time will be greater with increasing output temperature within the NTC system.

im

it

Slower reaction

n Lim

losio

xp 2. E

14.3.2 Ignition of Hydrocarbons

Ignition delay

Pressure p

3.

with low pressure, as the formed radicals will rapidly diffuse onto the combustion chamber walls and recombine in a diffusing reaction (14.21). When increasing the pressure, the diffusion is slow, so did the chain branching (14.19) predominates and a first explosion limit is reached. Upon further increase in the pressure, the second reaction limit is reached. In this area, the highly pressure-dependent response (14.22) is gaining in importance and the H2–O2 mixture is again stable. At the third explosion limit is another chain branching on the previously inert HO•2radical important, along with a rising by the higher pressure heat release per unit volume, it comes back to the ignition.

it

Explosion

1. Explo

sion Lim

it

Temperature T Figure 14.8 H2–O2 explosion diagram.

Temperature [K]

Figure 14.9 Schematic representation of the NTC.

Internal Combustion Engine Handbook | 593

6606_Book.indb 593

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

There is a more complex chain branch mechanism in the lower temperature range at T < 900 K [14-4]. In the initial stage, a hydrocarbon atom will be divided from a fuel molecule RH. The now formed alkyl radical R• is then subjected to an O2-addition

RH + OH i ! R i + H 2 O

(14.23)



RH + O 2 ! R i + HO i2

(14.24)



R + O 2 ! RO .

(14.25)

i 2

The equilibrium constant from reaction (14.25) is very dependent on the temperature. The equilibrium is on the right-hand side with lower temperatures, but displaces to the left with increasing temperature. The RO•2-radicals which are formed then pass through an internal hydrocarbon abstraction:

RO i2 ! QOOH i



R i + O 2 ! Alkene + HO i2

(14.27)



HO i2 + HO i2 ! H 2 O 2 + O 2

(14.28)

The temperature slowly increases as a result until hydrogen peroxide is extremely quickly decomposed subsequently at temperatures above approximately 1000 K and the second ignition phase is introduced: H 2 O 2 + M → 2OH i + M

10

(14.26)

Further reactions with a renewed addition of oxygen and an internal hydrocarbon abstraction form three radicals, including two OH•-radicals, so that a strong chain branch results which then leads to an initial important release of heat. The low temperature oxidation proceeds until the equilibrium from reaction (14.25) displaces at a temperature of approximately 900 K. This displacement will lead to an interruption of the chain branching via the isomerization reaction (14.26) and there will be an increased formation of alkenes and HO•2 radicals in the mid-temperature range. The HO•2 radicals react further to hydrogen peroxide, H2O2, which is initially relatively inert:



The two-staged ignition is easily identified in tests with rapid compression machines. A homogeneous fuel–air-mix is compressed by a single compression stroke in a compression machine and the piston is retained at the upper dead point. Figure 14.10 illustrates the pressure process in such a device during the test time. An initial ignition delay time exists after the completion of the compression at approximately 9.3 ms. The pressure and temperature increase rapidly in the first ignition stage until the mid-temperature range is reached with the reactions (14.27) and (14.28). The second ignition stage with the subsequent combustion begins after a second, longer, ignition time.

(14.29)

This process is designated as a degenerated chain branch and is the cause of the NTC. After the displacement of the equilibrium from reaction (14.25), there are not enough radicals formed to continue the ignition process. Only after the decomposition of the hydrogen peroxide is completed can large volumes of OH• radicals be produced, so that the ignition is accelerated and leads to a second release of heat which then introduces a high temperature mechanism. The negative temperature coefficient is most affected by long-chain alkanes. Compared to these alkanes, the alkenes and aromatics display a weak and/or no NTC behavior [14-5]. The predetermined reaction process is the significant mechanism for autoignition in diesel and HCCI-engines, rapid compression machines as well as for the engine knocking, leading autoignition in gasoline engines.

8

Pressure [ MPa]

i

14.3.3 Rapid Compression Machines

6

2. Ignition phase 1. Ignition phase

4 2 0 0.000

0.010

0.020

0.030

Time [s]

Figure 14.10 Pressure process in a rapid compression machine.

14.3.4 Diesel Engine

Diesel engine combustion comprises many partial processes, among others including injection, drop disintegration, drop evaporation, autoignition, combustion, and pollutant formation. The individual partial processes operate mainly simultaneously and are combined interactively. The first ignition process results in diesel engines at a local air ratio of 0.5 < λ < 0.7. As the injection in diesel engines, with direct injection, commences close to the upper dead point, the ignition delay time is relatively short.

14.3.5 HCCI Engine

A lean, homogeneous fuel–air mix is compacted in the HCCI process. Most of the mixture is ignited homogeneously close to the upper dead point. Figure 14.11 illustrates a typical pressure process for HCCI combustion (fuel: diesel). The two contrasting ignition stages are clearly identifiable. There are very long ignition delay times in this process because of the early mixing of fuel and air. As the heat release therefore mainly results during the ignition, the HCCI combustion is mainly dominated by the aforementioned kinetic procedures.

594 | Internal Combustion Engine Handbook

6606_Book.indb 594

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.3 Autoignition

Refer to Chapter 15.2 for a detailed description of knocking combustion.

80

2. Ignition phase

Cylinder pressure [bar]

70

14.3.7 Modeling the Autoignition

fired

60

1. Ignition phase

50 40 30 20

towed

10 0 -40

-20

0

20

40

Crankshaft angle [° CA]

As stated in the previous sections, the kinetic procedure for autoignition in combustion engines is very complex. Figure 14.13 Illustrates typical dimensions for some reaction mechanisms for hydrocarbons. These mechanisms include all the known processes and are therefore designated as comprehensive mechanisms. It must be recognized that the complexity of the mechanisms increases significantly with the fuel complexity, thus approximately 1200 species are utilized for cetane. It is obvious that such types of complex models can only be utilized in simple applications. This is especially the case when considering real, existing fuels formed from different hydrocarbons. For this reason, there have been many simplified models created for describing autoignition.

Figure 14.11 Pressure process in HCCI combustion.

14.3.6 Engine Knock

Engine knock is an undesired phenomenon in gasoline engines. After introduction of the combustion by the ignition spark, the unburned mixture is compacted again by the flame front and therefore additionally heated up. If temperature and pressure are thereby high enough, and there is sufficient time available, then an autoignition commences according to the described mechanisms in reactions (14.23)–(14.29). This practically isochoric residual gas combustion results in steep pressure gradients, which spread in the form of pressure waves in the combustion chamber, causing the familiar knocking or ringing noise. Figure 14.12 sketches the qualitative pressure process of a knocking combustion, whereby the knock commencement is indicated. The pressure waves occurring during both processes can result in mechanical material damage, and the increased thermal load can also lead to fusion of elements on the piston and cylinder head.

Ignition timing

Pressure [bar]

60

40 30 20 10 0

Knock beginning 330

345

Species

Reactions

H2

8

25

CH4

30

200

C3H8

100

400

C6H14

450

1500

C16H34

1200

7000

Figure 14.13 Typical dimensions of comprehensive reaction mechanisms.

14.3.7.1 Intervention Mechanism The simplest modeling for autoignition is provided by utilizing a global intervention reaction, whereby the reaction ratio is described via an Arrhenius equation. The ignition delay time is often calculated directly from the largest pressure, temperature, and air ratio:



t ZV = A

l −E ⎞ exp ⎛⎜ ⎝ RT ⎟⎠ p2

(14.30)

It is necessary to therefore subsequently calibrate the model constants A and E when, for example, the temperature range which leads to autoignition significantly alters. The simplest intervention model can often be practically utilized with good results depending on the task requirement, the previous complex kinetic procedure based on NTC range cannot, thus, be reproduced.

70

50

Fuel

390 ZOT 375 Crankshaft angle [°CA]

Figure 14.12 Pressure process in knocking combustion.

405

420

14.3.7.2 Shell Model To achieve a realistic illustration for ignition, a series of semiempirical multiple step models will be developed. The most probable of the mostly utilized models of these types is the Shell model [14-6], which was originally developed for predicting the engine knocking in gasoline engines and was subsequently expanded for modeling diesel ignition [14-7]. The model comprises eight reactions between five generic species and is able to represent the negative temperature coefficients. Moreover, the ignition delay times under various conditions can be predicted effectively.

Internal Combustion Engine Handbook | 595

6606_Book.indb 595

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

While the Shell model displays good results for diesel engine combustion and for engine knocking in gasoline engines, there is normally still a requirement for detailed mechanisms for calculations for HCCI processes because of the longer ignition delays involved. Examples for various complex mechanisms can be found in [14-8] and [14-9].

14.4 Flame Propagation 14.4.1 Turbulent Scales

urms =

( u′(t))2



(14.31)

urmslI n

Ret ≡



The flow in a combustion engine is generally turbulent and significantly influences the combustion process. To classify the various flame types, it initially makes more sense to define some of the typical characteristic curves, maps, and scales for the flows. A more detailed, comprehensive outline for the fluid mechanics and turbulence can be found, for example, in [14-10] and [14-11]. Turbulent flows form when there is instability in the flow cannot be sufficiently dampened by viscosity. Turbulence is characterized by chaotically varying flow dimensions and three dimensional (3D), vortical flow fields. A detailed description of the separate vortices is, even allowing for today’s modern computation systems, not feasible. A statistical computation of the turbulent fluid mechanics is normally customary. One respectively divides the Reynolds number for the torque values u, v, and w for the turbulent acceleration components in the mean value u, v, w and the fluctuation values u′, v′, and w′. This decomposition is clarified graphically in Figure 14.14 for the measured speed at a fixed position vi. a dimension for the turbulence intensity is provided by the “root mean square” value of the fluctuation size:

Two axial canals are especially important when considering turbulent vortices. The integral axial scale lI represents the average expansion of the largest resulting vortice in the flow field. The Kolmogorov axis lK is associated with the smallest existing vortice. The molecular viscosity is important for this magnitude, and a dissipation of the turbulent kinetic energy in the internal energy of the fluid takes place. A Reynolds’ number for the turbulence can be defined with the integral axis and the turbulence intensity: (14.32)

Time scales can be derived from the axial scales. The integral time and the Kolmogorov time describe the revolution time of the respective vortice and are defined as tl =



lI urms

(14.33)

lK uK

(14.34)

and tK =



Figure 14.15 shows the typical turbulent characteristic values for the flow field in a gasoline engine. Parameter

Dimension

Turbulence intensity urms

2 m/s

Turbulent Reynold number, Ret

300

Integral axial size

2 mm

Kolmogorov axis

0.03 mm

Integral time measurement

1 ms

Kolmogorov time

0.06 ms

Figure 14.15 Typical turbulent characteristic values in a gasoline engine, λ = 1.0; n = 1500 min–1 [14-12].

14.4.2 Flame Types

vi vi = vi' + vi vi' vi

t Figure 14.14 Reynolds’ mean value illustration of a turbulent flow which is transient in the middle.

Combustion processes can be introduced in processes where there is heat release as a result of a flame or flame-free. Flames can be subdivided into premixed and nonpremixed flames. Premixed flames are a homogeneous mixture of fuel and oxidizing agents which are mixed before combustion, the mixing and combustion process for nonpremixed flames occurs simultaneously. An additional important differential characteristic when considering combustion ranges results according to the type of the existing flows in laminar and turbulent flames. Turbulence accelerates the fuel implementation for both the premixed combustion via increasing the reaction zone, as well as for the nonpremixed combustion via improving the mixture. Figure 14.16 illustrates a simplified motorized application, in which the stated combustion process occurs. High turbulence intensities normally exist in the combustion chambers of combustion engines, which means that the turbulent flames are of particular importance. An example for a turbulent premix flame is illustrated by the gasoline engine with intake manifold injection. The heat release in a gasoline engine with

596 | Internal Combustion Engine Handbook

6606_Book.indb 596

1/19/16 8:45 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.4 Flame Propagation

direct injection lays etween both extremes for premixed and nonpremixed combustion and is therefore designated as partial premixed combustion. Turbulent premixed flame Spark-Ignition Engine

DE-spark-ignition engine (stratified) Gas burners Turbulent diffusion flame

Da =



HCCI - engine Homogeneous combustion

14.4.2.1 Premixed Flames A laminar premixed flame can be subdivided into three areas of preheating zone, reaction, and postoxidation zone [14-14]. The preheating zone, which takes up most of the flame, creates the thermal conduction and material dissipation as well as the first initial reactions. The actual reaction zone, in which most of the radical chain reactions occur, is very thin. The slower reactions dominate the postoxidation zone, for example, the oxidation from CO to CO2. Normally the burning speed sL is characterized by utilizing a laminar flame, a defined theoretical flame thickness δ L assisted by the flame speed, as well as the thickness of the internal reaction zone, δ i. The characteristic time scale for the laminar flame can then be defined as

(14.35)

accordingly. Turbulent premixed flames take on various forms according to the framework conditions. A division into the various regimes can be executed by utilizing the Borghi diagram [14-14], [14-15], refer to Figure 14.17. Various areas can be identified in this diagram by utilizing the 3D key performance indicators of the turbulent Reynolds’ number, Karlovitz number, and Damköhler number.

Regime 1 Ret < 1 planare laminar flame front

Regime 2 Ret > 1, Ka < 1, Da > 1, u'/sl < 1 corrugated laminar flame front

t F ⎛ dL ⎞ = tl ⎜⎝ lK ⎟⎠

Regime 3 Ret > 1, Ka < 1, Da > 1 folded flame "islanding"

(14.37)

Da < 1

106

A flame-free combustion exists in an ideal HCCI process.

dL sL

(14.36)

2

Ka =

Figure 14.16Flame types in motorized applications [14-13].

tF =

tl tF

The Karlovitz number is the ratio of the laminar time scale and the Kolmogorov time

Knocking spark-ignition engine DI - Diesel

The dam charcoal number describes the behavior of the turbulent integral time scale, which can be utilized as a measurement for the mixing time and the time scale for the laminar flame:

u' sl

Da = 1 10

Da > 1

4

102

Ret = 1 1

Ka > 1

5

Kai = 1

Ka = 1 4

1 Ret < 1 1

Ka < 1 3

2 102

104

106

lI δl

Figure 14.17 Illustration of the premixed combustion in the Borghi diagram.

Figure 14.18 schematically illustrates the various occurring combustion areas in a Borghi diagram. The line with Re = 1 separates the laminar flame areas (Re < 1) in the left bottom corner from the turbulent area (Re > 1). The turbulence intensity is lower than the laminar burning speed in area 2 of the turbulent area, so that there is only a slight wavy flame formed. This flame still indicates the characteristic for laminar flames and the burning speed will be primarily set from the laminar speed. Flames with laminar-typical characteristics are described in English linguistic usage as flamelets. More creased flames can be found above the area of the wavy flames

Regime 4a Ret > 1, Ka > 1, Da > 1, distributed flame reaction zone

Regime 4b Ret > 1, Ka > 1, Kai < 1 thin flame reaction zone

Regime 5 Ret > 1, Ka > 1, Da < 1 distributed flame stirred reactor

Figure 14.18 Schematic illustration of the various types of flames for premixed combustion.

Internal Combustion Engine Handbook | 597

6606_Book.indb 597

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

and they can individually form into island formations (area 3). The flame itself still remains locally laminar. On the limiting line Ka = 1, which is defined as the Klimov–Williams criteria, the smallest turbulence structures still have the same size arrangements as the laminar flame. According to the classical conceptions, small vortices occur in the flame which, on the one hand, can lead to localized thickening and, on the other, also to localized distorted flames and localized extinguishing. This regime will therefore also be defined as a divided reaction area. Above the line Da = 1 in area 5, the reactions will be very slow compared to the turbulence (Da ≪ 1), so that a comprehensive turbulent mixture of the reactants can be assumed before the reaction. The Borghi diagram from Peters [14-16] was recently increased with a second Karlovitz number, Kai. This describes the ratio between the dimensions of the internal reaction zones for the laminar flame and the Kolmogorov vortice. The smaller vortice structures still intrude into the preheating zone in the area between Ka = 1 and Kai = 1, but the actual reaction zone still remains unaffected. It can therefore still be assumed that a laminar-type behavior still exists in this area. New studies from Dinkelacker [14-17] also still indicate that a thickening of the preheating zone in this area only occurs with relatively large turbulent Reynolds’ numbers. This, and other points, would lead the author back to the conclusion that the Kolmogorov vortice, which intrudes into the preheating zone, is dissipated by the higher temperature. 14.4.2.2 Nonpremixed Flames Nonpremixed flames, which for example occur in diesel engines as direct injection, are normally designated as diffusion flames. This designation creates the impression that the combustion is dominated by the mixing process, and therefore via molecular and turbulent diffusion. It should, however, also be observed that the diffusion process is also of core importance in premixed flames. As the time scales for the mixing for diffusion flames are generally significantly larger than the time scales for the reaction, an assumption of a nonending rapid chemistry is often taken. However, in actual processes there are always local areas in which this assumption is also not fulfilled. The speed of the chemical reaction is particularly important for pollutant formation. The burn-off of soot, which is required for engine designs, initially takes place at sufficiently high temperatures which, however, simultaneously lead to the creation of increased amounts of nitrogen oxides. 14.4.2.3 Partial Premixed Flames The area for the partial premixed flame lies between both extremes for the complete mixing and the complete separation of fuel and air before the reaction. An example for this type of combustion is the gasoline engine with direct injection with layered charging mode. Because of the lean global air ratio (λ > 1), there must be a layering of the injected cylinder charge in the combustion chamber, so that an injection-capable mixture exists on the spark plug at the ignition timing [14-2]. The characteristic flame form for a partial premixed flame is schematically illustrated in Figure 14.19. As the burning

speed, along with the temperature and pressure, is especially dependent on the air/fuel ratio,λ , the flame generates itself the quickest along a line with λ = 1. A lean premix flame creates itself on the lean side, a thicker premix flame is created on the thicker side.

y

Premixed flame front λ=1

λ>1 Lean premixed flame

Diffusionsflame

λ<1 Full pre-mixed flame

Triple point

x

Figure 14.19 Schematic illustration of a partial premixed flame [14-18].

Between these flames there is a diffusive exchange between unburnt fuel on the thick side and superfluous oxygen on the lean side, so that a diffusion flame is formed in this position. The occurrence of three various flames creates the terminology triple flame, the point at which the three flames collide is called the triple point.

14.5 Model Development and Simulation A prerequisite for a numerical simulation is the creation of a technical process for the described model. One can understand model development to be a target-oriented simplification of the reality by utilizing abstraction. A prerequisite hereby is that the real process is divided into individual process sections and therefore subdivided into partial problems. A series of demands must be set for the resulting model, also refer to [14-2]: •• The model must be formally correct, that is, it should be consistent. •• The model must be described as close as possible to the reality. A model is still only a proximity and can therefore never be fully comprehensive with reality. •• The required effort for the numerical solution must be feasible within the framework of the task requirement. •• The model should be as simple as possible, and as complex as necessary. Once we have utilized the model conception, we are then able to actually understand the physical and chemical process flow.

598 | Internal Combustion Engine Handbook

6606_Book.indb 598

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.5 Model Development and Simulation

An essential prerequisite for creating the mathematical model for simulating the time and space altering flow, temperature and concentration fields with chemical reactions is the knowledge of the basic principles for thermodynamics, the fluid dynamics and the technical combustion principles. It must be considered with a simulation of flow fields with chemical reactions that physical and chemical processes can flow with very differing time and axial scales. The description for these process flows is mainly simpler when the time scales are very different, because simplified assumptions can then be made for the physical process and/or chemical process, and it is normally very complex when the time scales are of the same size arrangement. The numerical simulation has developed into an essential tool for developing engines, drive units, and vehicles during recent years. Its importance with regard to the increasing complexities with additional reductions for development time for engines and vehicles will still continue to increase dramatically.

are solved by utilizing turbulence models and additional physical and chemical submodels. The essential difference of this model is illustrated in Figure 14.20. 0D models form the basis for the complete process analysis, that is, the simulation of stationary and transient behavior for engines, drive units, and complete vehicles. Phenomenological models are implemented for the simulation of processes in the combustion chamber, that is, the mixture formation, ignition, combustion, and pollutant build up. Further, 3D-CFD models are particularly suitable for the required computation time only for specific and very detailed tasks.

14.5.2 0D Models

•• Phenomenological models, in which the heat release is modeled by the combustion by utilizing physical and chemical concepts.

14.5.2.1 Substitute Combustion Process The combustion process describes the time of the release of heat by the combustion. One can designate the integral of the combustion process as the total cumulative combustion curve and/or as throughput combustion function. One utilizes various concepts for modeling the combustion and/ or mathematical modeling which all have the goal of real liberation of heat, by utilizing the combustion by way of the so-called substitute burning processes to precisely describe the process as accurately as possible. The best known are singleand double-Vibe functions, polygon hyperbolic function, and neural networks. The following briefly clarifies the simple Vibe substitution combustion process, for more information also refer to [14-2] and [14-19]. Starting from “Triangular combustion processes,” [14-20] has on the basis of reaction-kinetic considerations, the relationship

•• 3D-computational fluid dynamics (CFD) model, in which the contents’ comparison for mass, energy, and impulse

EB = 1 − exp ( −a ⋅ y m+1 ) EB,ges

14.5.1 Classifications for Combustion Models

Combustion models are normally divided into three different categories, as listed hereunder: •• 0D or thermodynamic models, whereby not only the heat release from the combustion as well as the heat transfer between the cylinder charge and the combustion chamber limiting walls are described by utilizing half-empirical models, for example, vibe-substitution combustion processes and/or Woschni’s heat transference model.

Thermodynamic (0-dimensional)

· ·

·

Empirical Burning Process No pollutant build up

Customary DGL’s (t)



CFD (Multi-dimensional)

Phenomenological (Quasi-dimensional)

· ·

Quasi-dimensional local dissipation Physical and chemical sub-model

·

No turbulence Flow

·

Customary DGL’s

(14.38)

· ·

··

Turbulence Flow Field (Navier-Stokes-Gl.) detailed physical and chemical sub-model

partial DGL's (t,x y ,z , ) Computing Time: Hour-Day

Figure 14.20 Classification arrangements in reciprocating piston engines [14-19].

Internal Combustion Engine Handbook | 599

6606_Book.indb 599

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

where

c = −1/lRB ,

EB, ges = mB ⋅ HU



d = −0, 0375 − ( lRB − 1.17 ) / 15,

(14.39)

a = (0.05 − d)/[ lRB ⋅exp(−1) − exp(c)] ,

is created for the maximum releasable amount of heat and y=

j − jBB ∆jBD

(14.40)

for the dimensionless crankshaft angle with the burning time ∆jBD = jBE − jBB



(14.41)

with φ BE = burning and φ BB = combustion start. Figure 14.21 plots the combustion process dEB = f (j, m) dj



(14.42)

as a function of dimensionless crankshaft angle for various shape parameter, m. For the cumulative combustion curve or the burnout function there results: EB = ∫ f (j, m)⋅ dj = F(j, m)



(14.43)

At the end of the burning time, that is, at φ = φ BE and/or at y = 1, there should be μ u,ges-percent of the total of the fuelsupplied energy released. This, thereby follows the relationship EB = hU , ges = 1 − exp(−a) EB, ges



η u,ges

0.999

0.990

0.980

0.950

a

6.908

4.605

3.912

2.995

For the degree of conversion [14-21], the empirical relationship



⎧ 1 l > lRB ⎪ = ⎨ a ⋅ l ⋅exp(c ⋅ l) − b 1 ≤ l ≤ lRB ⎪ 0.95 ⋅ l + d l≤1 ⎪⎩

b = a ⋅exp(c) − 0, 95 − d



∆jBD ⎛ l0 ⎞ ⎛ n ⎞ =⎜ ⎟⋅ ∆jBD,0 ⎝ l ⎠ ⎜⎝ n0 ⎟⎠



m ⎛ ∆jZV ,0 ⎞ = m0 ⎜⎝ ∆jZV ⎟⎠

0.5

⋅

0.5

⋅ hU0.6,ges

pT0 ⎛ n0 ⎞ ⎜ ⎟ p0T ⎝ n ⎠

(14.47)

0.3



⎡ 7800 ⎞ ⎛ 0.135 4.8 ⎞ ⎤ ⋅ ∆jZV = 6 ⋅ n ⋅10−3 ⋅ ⎢ 0.5 + exp ⎛⎜ + 1.8 ⎟ ⎥ ⎝ 2 ⋅T ⎟⎠ ⎜⎝ p 0.7 p ⎠⎦ ⎣ n jBB = jFB + ∆jEV ,0 + ∆jZV , n0

(14.48)

(14.49)

(14.50)

with the start of delivery φ FB and the injection delay ∆φ EV,0. Refer to [14-2] for further details. 14.5.2.2 Heat Transfer Models The heat transfer from the hot flue gas in the combustion chamber is performed by convective heat transfer and by thermal radiation glowing soot particles. Because of the

Relative Crankshaft Angle

has been indicated with

(14.45)

(14.46)

whereby λ RB is the air/fuel ratio, in which an exhaust gas blackening with soot digit according to Bosch RB = 3.5 is achieved. The interval 1.17< λ RB 2.05 is given as the scope of validity. The Vibe substitution combustion process is determined by utilizing the three Vibe parameters: combustion start φ BB, combustion time ∆φ BD, and shape parameter m are set. Only three parameters can be set for a given operating point. The adjustment is therefore executed so that the start of combustion φ BB, the ignition pressure pz and the mean pressure pm,i correspond with those of the real engine processes. The conversion of Vibe parameters on any operation is carried out using semiempirical points and functions depending on the main factors: air/fuel ratio I, N speed, power, ignition delay ∆φ ZV, and start of combustion φ BB accordingly:

(14.44)

and from this numerical values

hU , ges





Relative Crankshaft Angle

Figure 14.21 Combustion process and throughput combustion function for Vibe [14-20].

600 | Internal Combustion Engine Handbook

6606_Book.indb 600

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.5 Model Development and Simulation

formation of layers of soot at low load, and those which burn at full load, the description of heat transfer is more difficult. An overview is given in [14-2]. The heat transfer model shown in the following goes back to Woschni [14-22] and is still the state-of-the-art today. From a dimensional analysis follows, for the dimensionless heat transfer coefficient, the Nußelt number for a stationary and fully turbulent pipe flow

Nu = C Re0.8 Pr 0.4



aD l

(14.52)

rwD h

(14.53)

v a

(14.54)

Nu =

the Reynolds’ number Re =

whereby, for the intake swirl cu/cm, the validity range 0 < cu/cm < 3 will be specified. The speed corrected with the “combustion link” provided for towed engines and in the lower load range leads to lower values. Therefore, the characteristic speed for the relationship

(14.51)

with the Nußelt number







If one considers the mixture in the combustion chamber as an ideal gas with the thermal status equation r=



p RT

(14.55)

and takes the expansion for the temperature dependence of the correlation x



y

l ⎛T⎞ h ⎛T⎞ Pr = 0.74; = ; = l0 ⎜⎝ T0 ⎠⎟ h0 ⎜⎝ T0 ⎠⎟

(14.56)



with r = 0.8(1 + y) − x and the assumption that the characteristic speed w in the engine is equal to the average piston speed Cm. By comparison with measured values, the exponent for the temperature dependence r = 0.53 and the constant C* = 127.93 will be determined. For fired engines, a modification of the characteristic speed must be introduced accordingly, because the combustion increases the turbulence and therefore the heat transfer dramatically.

V ⋅T w = C1 ⋅ cmC2 h 1 ( p − p0 ) . p1 ⋅V1

(14.58)

The second term in eq. (14.58) is the so-called “combustion link,” with the pressure curve p(φ ) in the fired mode and p0(φ ) in lag mode. V1, p1 and T1 are the values at “input closes.” For the constants C1 and C2 are obtained by adjustment to measured values



⎧ 3.24 ⋅10−3 m / (s ⋅ K) Tw < 550 K ⎪⎪ −3 −3 (14.62) C2 = ⎨ 5.0 ⋅10 m / (s ⋅ K) + 2.3⋅10 ⎪ TW > 550 K T − 550 K m / (s ⋅ K) ( ) w ⎪⎩



∂2 T ∂T = a 2 , ∂t ∂x

⎧6.18 + 0.417 ⋅ cu /cm : Charge cycle C1 = ⎨ ⎩2.28 + 0.308 ⋅ cu / cm : Compression/Expansion



(14.59)

(14.63)

with the temperature conductivity a=

(14.57)

(14.61)

Because of the chronological fluctuations of the gas temperature in the combustion chamber, there are resulting temperature fluctuations in the combustion chamber bounding walls according to Figure 14.23. The energy transport by thermal conduction in solid bodies is described by the Fourier’s differential equation

then, one obtains finally a = C∗D−0.2 p 0.8 cm0.8T − r

2 ⎡ V −0.2 ⎤ w = C1 ⋅ cm ⎢1 + 2 ⎛⎜ c ⎞⎟ ⋅ pm!i ⎥ ⎝ ⎠ V ⎣ ⎦

will have a proposal and recommendation that the respective largest numerical value will be utilized. For diesel engines with direct injection, the constant C2 at higher wall temperatures must be corrected accordingly. For more details, please refer to the literature cited.

and the Prandtl number Pr =

⎧⎪6.22 ⋅10−3 m / (s ⋅ K) : Antechamber engine , (14.60) C2 = ⎨ −3 ⎩⎪3.24 ⋅10 m / (s ⋅ K) : DI-Engine



l rc p

(14.64)

accordingly. Figure 14.22 illustrates the process for gas temperature, thermal density, and heat transfer coefficient for a gasoline engine at full load. For more details, please refer to the literature cited.

14.5.3 Phenomenological Models

Phenomenological combustion models are available with regard to the complexity and detailing between 0D burning functions and combustion models, which are utilized in 3D-CFD calculations. Unlike 0D models, the phenomenological models normally provide a quasi-dimensional resolution and chemically and physically well-founded submodels so that the combustion process can be calculated in advance, and not only shown empirically. Compared with models for 3D-CFD simulation, the submodels utilized are generally much more simplified. Ordinary differential equations are usually solved by phenomenological models. Altogether, phenomenological models thereby exhibit significantly shorter computing times as 3D-CFD models. However, because of the lack of 3D resolution, the turbulent flow structures in the combustion chamber cannot be illustrated.

Internal Combustion Engine Handbook | 601

6606_Book.indb 601

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

3000

3000 4-Stroke-Gasoline Engine, Full load n = 2200 min–1

T[K] 2000

a i[

TG

1000

. q α 0

180

360

. q = a (T-T W)

540

14.5.3.1 Gasoline Engine Combustion Models for turbulent premixed combustion in gasoline engines are primarily based on the so-called Entrainment model from Blizard and Keck [14-23]. It is hereby assumed that the flame front generated from the spark plug will propagate spherically, so that the area and location of the flame front from the turbulent burning speed, and a geometric view of the combustion chamber can be calculated. To calculate the turbulent burning speed, first the laminar burning speed sl must be evaluated, for example, via the empirical relationships of Metghalchi and Keck [14-24] and Rhodes and Keck [14-25]. Diverse correlations were developed to calculate the turbulent burning speed, sr. A simple formulation, based on a correlation from Damköhler with two empirical constants C and n, reads as follows: n



W ] m2K

2000

1000

0

. ] q [ kW m2

⎛ u′ ⎞ st = 1 + C ⎜ ⎟ . sl ⎝ sl ⎠

(14.65)

In addition to the uncertainty of the correlation in itself, the determination for the correlation for the necessary turbulent characteristics such as the fluctuation speed u′, are particularly problematic within the framework of phenomenological models. Another difficulty is the description of the ignition and ignition phase. One approach for describing the ignition phase has been introduced by Herweg and Maly [14-26]. Here, the effective burning speed in the ignition phase has been described as the sum of turbulent burning rate sr and a plasma speed spe. The plasma speed as hereby been calculated under the assumption that the air–fuel mixture in the flame core is heated by the ignition energy to the adiabatic flame temperature. If, instead of premixed combustion in conventional gasoline engine applications, the partially premixed combustion in gasoline engines with direct injection and heterogeneous operation are considered, then an even more complex description is necessary. One approach for such combustion can be found for example in Koch [14-27]. 14.5.3.2 Diesel Engine Combustion In conventional diesel engine combustion, the injection has the dominant influence on the heat release. As the injection

0 720

Figure 14.22 Process for gas temperature, heat flowdensity, and heat transfer coefficient for a gasoline engine at full load.

f [°KW]

can be described quite well phenomenologically, a series of phenomenological models for the diesel engine combustion have been developed and used successfully in the past [14-19]. Two models should be briefly described here as examples. For alternative approaches and further details, refer to [14-2] and the literature cited. The package model was originally proposed by Hiroyasu [14-28] and later developed by other authors, for example, Stiesch [14-29]. Prior to injection, the combustion chamber will only be dissipated with one zone as in the 0D models. During the injection, and thereafter, the injection jet is then divided into several zones or packages, refer to the schematic representation of the package model in Figure 14.22 (center). It can thereby be assumed that the injection jet is rotationally symmetrical, so that a description in axial and radial direction takes place and that the packages have a ring shape. An interaction between individual packages is usually neglected. After the packages are introduced into the combustion chamber, they initially contain only liquid fuel. The package speed decreases with increasing distance from the nozzle and/or with increasing service life and is described empirically as follows:

(



)

⎛ pinj − pcyl DD2 ⎞ m vax = 1.48 ⎜ ⎟ ⋅t ⎜⎝ ⎟⎠ rL

(14.66)

In Eq. (14.66) pinj and pcyl are the injection and cylinder pressure, respectively, Dn is the nozzle diameter, and ρ L the fuel density. It is assumed that the drawn-in air mass into the package can be calculated from the impulse reception in the package. This so-called entrainment has a significant influence on the calculated combustion rate for the package model. In the individual packages, submodels for calculating the beam and drops decay, evaporation, ignition, and combustion are used. By quasi-dimensional resolution, mixture composition and temperature can be displayed in the injection beam, which also allows the calculation of pollutant emissions such as soot and nitrogen oxides. Another widely used model for the diesel engine combustion is the model from Barba [14-30]. In describing the heat release, a differential is made between the premixed combustion of

602 | Internal Combustion Engine Handbook

6606_Book.indb 602

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.5 Model Development and Simulation

a pilot injection and the premixed mixture and controlled combustion of the main injection. Here, the introduced fuel in the preinjection pilot fuel is described as a single zone that mixes with air. After a calculated ignition delay time via a decisive reaction [compare with Eq. (14.30)], turbulent flames initially propagate from individual, in the subsequent process from multiple, ignition sites whereby the turbulent burning speed is described in a similar manner as in models for gasoline engine combustion via an empirical relationship for the laminar burning speed and a Damköhler relationship for turbulent burning speed [Eq. (14.65)]. In contrast to the package model, the model according to Barba has the injection jet of the main injection only in discrete axial segments, whereby the penetration depth is also described in an empirical manner. The implementation of the fuel, which is introduced until the ignition delay time is achieved, is described via the model for premix combustion as in the pilot injection. The implementation of the subsequently introduced fuel is represented by a diffusion model. In this case, the reaction rate is inversely proportional to the turbulent time scale, which is derived via further relationships. Phenomenological models offer the possibility for really good predictions for the firing process of the diesel engine combustion although, in this case, the utilization of models with different engines must usually have an adaptation of empirical parameters. The quality of the introduced injection process is of particular importance, especially for the quality of the calculation. Although it is not strictly speaking relevant for phenomenological models, the so-called stochastic reactor models (SRM) should also be mentioned at this point, refer, for example, to [14-31]. These are primarily used to describe the homogenized diesel engine combustion (HCCI). In these models, ignition and combustion are described via probability density functions, whereby especially detailed kinetic particular reaction mechanisms can be utilized.

14.5.4 3D-CFD Models

Multidimensional flow problems are described via conservation equations for mass, momentum, and energy in the form of partial differential equations, the Navier–Stokes equations. Except model problems, for which certain simplifications are possible, these equations are too complex to be solved analytically, so that a numerical solution is necessary. In this case the flow-through component, hereby the combustion chamber, is discretized with a computational grid. In principle, the Navier– Stokes equations are valid for both laminar and turbulent flows. The smallest, occurring structures in a turbulent flow structures can be described by the Kolmogorov axial (refer to 14.4.1). A resolution of this structure, the so-called direct numerical simulation (DNS), results in industrial question formulations in a too high number of cells of the computational grid, and hence longer resulting computation times. As a rule, therefore, the conservation equations for the Reynolds evaluated (and/or Favre evaluated) variables (compare with Section 14.4.1) are dissolved, whereby the information about

the fluctuating quantities will be lost. Therefore turbulence models are required to finalize the conservation equations but these are however not universal, and in principle must be adapted to the respective flow conditions. In recent years, the so-called large-Eddy simulation (LES) has therefore become more important. The large turbulent eddies are resolved in LES and a modeling is only undertaken for the smaller flow structures. This approach has the advantage that a refinement of the computational grid automatically leads to a gain of information, which is not the case with Reynolds evaluated concepts. Figure 14.23 shows exemplary simulation results of a shear layer calculated with LNS and DNS.

Figure 14.23 2D section through a shear layer calculated with LES and DNS [14-33]. See color section page 1086.

Because of the great influence of the turbulence on the processes during combustion, turbulence models have a significant impact on the overall quality of a combustion calculation. Therefore, future developments in this area are of particular importance. For engines with direct injection, the interaction between the processes for jet development, drop decay and drop coalescence, droplet evaporation, ignition, combustion, and pollutant formation plays an important role. A good description of the mixture formation processes is therefore a prerequisite for modeling the combustion. Refer at this point to [14-2] and [14-32] for models for describing the mixture formation. For describing the combustion within the 3D fluid mechanics in the past, a wide variety of different models were developed. Most of the models for the description of the premixed combustion, such as that which occurs in conventional gasoline engines, are based on the so-called Flamelet adoption [14-16], which states that the fanned by turbulence flame front can be treated locally as a laminar flame. The fundamental difficulty in the description of turbulent, premixed flames is in the detection of the flame front. Here, in particular, two different model approaches were particularly in

Internal Combustion Engine Handbook | 603

6606_Book.indb 603

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 14 Combustion

practical applications in recent years. In flames surface models (coherent flamelet models) [14-34] there is first a release for a transport equation for a progress variable, which assumes the values between 0 (no mass transfer) and 1 (mass transfer fully completed). Second, a transport equation for the flame surface density (flame surface per unit volume) is solved, the flame surface density occurs as a source term in the equation for the progress variable. In the second model approach, the flame front is described by the so-called G-equation [14-14], which is based on the level set concept, a method for describing the moving surfaces. In the main phase of the diesel engine combustion, the heat release is determined by the turbulent mixing between air and fuel. An important role in most models is hereby taken by the turbulent timescale

t turb = k/e

(14.67)

whereby k describes the turbulent kinetic energy and ε its dissipation. A simple, but often successfully used in the past model for diesel engine combustion, is the characteristic time-scale model, wherein the reaction rate of a species i based on the difference of the actual species density ρ i and the local, instantaneous equilibrium density ρ *iis calculated:



dri ri − ri∗ = dt tc

(14.68)

Thereby, τ c is the characteristic time to reach equilibrium. It is assumed that the characteristic time τ c for all considered species is equal. Further, τ in this case formed from a laminar portion τ lam, which describes the influence of the reaction kinetics depending on the temperature and the air ratio and the turbulent proportion τ turb together:

t c = t lam + f ⋅ t turb

(14.69)

The turbulent proportion takes into consideration that the starting materials have to be mixed prior to the reaction at the microscopic level by turbulence. The delay factor f strives with sequential combustion from zero to one, and thereby marks the transition from premixed mixture to a controlled combustion. The direct involvement of the turbulent time scale comprises a very strong dependence on the results of the combustion model for utilized turbulence models, as an incorrect calculation of turbulent parameters directly affects the heat release rate. Another widely used model for calculating the diesel engine combustion is the flamelet model [14-16]. It is hereby assumed that the turbulent flame from an ensemble of laminar diffusion flames (flamelets) is jointly composed. This assumption allows for a transformation of the 3D structure of combustion to a dimension in the direction of the mixture fraction, a standardized air/fuel ratio between 0 and 1. By converting the original 3D conservation equations to a 1D problem, very complex reaction mechanisms can be solved with a reasonable computational effort. As it is no longer possible to clearly separate premixed and nonpremixed combustion forms for modern combustion

processes, for example, for gasoline engines with direct injection and spray-guided processes, or with the homogenized diesel engine combustion, there is an increasing need for combustion models which can cover several different regimes. An example of such a model is the ECFM-3Z model [14-36] which has been increasingly utilized in recent times. For a detailed treatment of the here presented, and additional combustion models, please refer to the literature cited and [14-2], [14-16], [14-19], and [14-34].

Bibliography

14-1. Glassmann, I. 1996. Combustion. Academic Press: USA 14-2. Merker, G. P., Schwarz, C. (Hrsg.). 2009. “Grundlagen Verbrennungsmotoren.” in Simulation der Gemischbildung, Verbrennung, Schadstoffbildung und Aufladung, 4th Ed. Vieweg + Teubner. 14-3. Semenov, N. 1935. Chemical Kinetics and Chain Reactions. Oxford Univ. Press: London. 14-4. Curran, H. J., Gaffuri, P., Pitz, W. J., and Westbrook, C. K. 1998. “A Comprehensive Modeling Study of n-Heptane Oxidation,” Comb. Flame. 114(1–2):pp. 149–177. 14-5. Leppard, W. R. 1990. “The Chemical Origin of Fuel Octane Sensitivity.” SAE Technical Paper 902137. 14-6. Halstead, M., Kirsch, L., and Quinn, C. 1977. “The Autoignition of Hydrocarbon Fuels at High Temperatures and Pressures—Fitting of a Mathematical Model,” Comb. Flame. 30. 14-7. Kong, S.-C., Han, Z., and Reitz, R. D. 1995. “The Developement and Application of a Diesel Ignition and Combustion Model for Multidimensional Engine Simulations.” SAE Technical Paper 950278. 14-8. Tanaka, S., Ayala, F., and Keck, J. C. 2003. “A Reduced Chemical Kinetic Model for HCCI Combustion of Primary Reference Fuels in Arapid Compression Machine,” Comb. Flame. 133. 14-9. Ogink, R., Golovitchev, R. 2001. “Gasoline HCCI Modelling: Computer Program Combining Detailed Chemistry and Gas Exchange Processes. SAE Technical Paper 2001-01-3614. 14-10. Merker, G. P., Baumgarten, C. 2000. Fluid- und Wärmetransport: Strömungslehre. Vieweg+Teubner Verlag—Germany. 14-11. Pope, S. B. 2000. Turbulent Flows. Cambridge Univ. Press: England 14-12. Heywood, J. B. 1988. “Combustion and its Modeling in Spark-Ignition Engines.” 3rd International Symposium COMODIA. 94. 14-13. Otto, F. Strömungsmechanische Simulation zur Berechnung motorischer Prozesse. In Lecture notes, Univ. Hannover. 14-14. Peters, N. 1986. “Laminar Flamelet Concepts in Turbulent Combustion,” Proceeding of the Combustion Institute, 21. 14-15. Borghi, R. 1985. “On the Structure and Morphology of Turbulent Premixed Flames.” in Recent Advances in Aeronautical Science, edited by C. Casci. Plenum Press. 14-16. Peters, N. 2000. Turbulent Combustion. Cambridge Univ. Press. 14-17. Dinkelacker, F. 2001. Struktur turbulenter Vormischflammen. Leipertz (Hrsg.): Berichte zur Verfahrenstechnik, Heft 1.4. 14-18. Kech, J. M., Reissing, J., Gindel, J., and Spicher, U. “Analyses of the Combustion Process in a Direct Injection Gasoline.” 4th International Symposium COMODIA. 98. 14-19. Stiesch, G. 2003. Modeling Engine Spray and Combustion Processes. Springer Verlag: Berlin, Heidelberg. 14-20. Vibe, R. R. 1970. Brennverlauf und Kreisprozess von Verbrennungsmotoren. VEB-Verlag Technik: Berlin. 14-21. Betz, A. 1985. Rechnerische Untersuchung des Stationären und Transienten Betriebsverhaltens Ein- Und Zweistufig Aufgeladener ViertaktDieselmotoren. Druckladen.

604 | Internal Combustion Engine Handbook

6606_Book.indb 604

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

14.5 Model Development and Simulation

14-22. Woschni, G. 1970. Die Berechnung der Wandwärmeverluste und der Thermischen Belastung der Bauteile von Dieselmotoren. MTZ. 31:pp. 491–499. 14-23. Blizard, N. C., Keck, J. C. “Experimental and Theoretical Investigation of Turbulent Burning Model for Internal Combustion Engines.” SAE Technical Paper 740191. 14-24. Metghalchi, M., Keck, J. C. 1982. “Burning Velocities of Mixtures if Air with Methanol, Iso-octane and Indolene at High Pressure and Temperature,” Comb. Flame, 38:pp. 143–154. 14-25. Rhodes, D. B., Keck, J. C. 1985. “Laminar Burning Speed Measurements of Indolene-Air Diluent Mixtures at High Pressures and Temperatures.” SAE Technical Paper 850047.

14-30. Barba, C., Burkhardt, C., Boulochos, K., and Bargende, M. 2000. “A Phenomenological Combustion Model for Heat Release Rate Prediction in High Speed DI Diesel Engines with Common Rail Injection.” SAE Technical Paper 2000-01-2933. 14-31. Su, H., Mosbach, S., Kraft, M., Bhave, A., Kook, S., and Bae, C. 2007. “Two Stage Fuel Direct Injection in a Diesel Fuelled HCCI Engine.” SAE Technical Paper 2007-01-1880. 14-32. Baumgarten, C. 2006. Heat and Mass Transfer in Sprays. Mixture Formation in Internal Combustion Engines (Heat and Mass Transfer), Springer, USA. 14-33. Chumakov, S. 2001. “Large-Eddy Simulation for Subgrid Scalar Transport,” M.Sc. Thesis, Univ. Wisconsin.

14-26. Herweg, R., Maly, R. R. 1992. “A Fundamental Model for Flame Kernel Formation in SI Engines.” SAE Technical Paper 922243.

14-34. Poinsot, T., Veynante, D. 2005. Theoretical and Numerical Combustion, 2nd ed. RT Edwards, Inc.

14-27. Koch, T. 2002. “Numerischer Beitrag zur Charakterisierung und Vorausberechnung der Gemischbildung und Verbrennung in einem direkteingespritzten, strahlgeführten Ottomotor,” Proceeding of the ETH Zürich.

14-35. Kong, S. C., Han, Z., and Reitz, R. D. 1995. “The Developement and Application of a Diesel Ignition and Combustion Model for Multidimensional Engine Simulations.” SAE Technical paper 950278.

14-28. Hiroyasu, H., Kadota, T., and Arai, M. 1992. “Development and Use of a Spray Combustion Model to predict Diesel Engine Efficiency and Pollutant Emissions, Part 1: Combustion Modeling,” Bull JSME, 26(214):pp. 569–575.

14-36. Colin, O., Benkenida, A. 2004. “The 3-Zones Extended Coherent Flame Model (Ecfm3z) for Computing Premixed/Diffusion Combustion,” Oil Gas Sci. Technol. Rev. IFP, 59(6):pp. 593–609.

14-29. Stiesch, G. 1999. Phänomenologisches Multi-Zonen-Modell der Verbrennung und Schadtstoffbildung im Dieselmotor, Fortschritt-Berichte VDI, Reihe 12, Nr. 399, VDI-Verlag: Düsseldorf.

Internal Combustion Engine Handbook | 605

6606_Book.indb 605

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 606

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15 Combustion Systems 15.1 Diesel Engines 15.1.1 Diesel Combustion 15.1.1.1 General Overview Combustion is understood to be chemical reactions in which a substance releases heat (exothermic reaction) while bonding to molecular oxygen (oxidation). Combustion starts with ignition. This can take place only under certain conditions that may be described in a simplified manner as follows: •• The reaction partners must possess a minimum energy level, the so-called activation energy. Only molecules that have reached this energy level can react with each other. The proportion of molecules in a mixture of reaction partners with a sufficiently high energy level increases exponentially as the mixture temperature rises. The reaction mixture must have a specific composition. When there is a large share of one or another reaction partner, the potential molecular collisions are insufficient to trigger a stable, self-supporting reaction. Accordingly, ignition reliably occurs only within the ignition limits (an air-fuel ratio of approximately 0.7–1.3). These ignition limits expand as the mixture temperature rises. Inert gas components in the reaction mixture (such as exhaust) reduce the reaction speed similar to a shift in the mixture composition toward a “lean” ignition limit. The mode of operation of diesel engines is based on autoignition of the fuel introduced into the combustion chamber. The fuel is supplied to the combustion chamber by injection with a suitable system, the injection system (see Chapter 12.4.1). To achieve reliable autoignition of the fuel, the air must be sufficiently hot in the combustion chamber. This is essentially attained by a correspondingly high engine compression ratio. Mixing the fuel with the available air and, hence, creating optimum conditions for the ignition of the forming air-fuel mixture is a necessary prerequisite for the subsequent combustion.

In addition to the charge movement in the combustion chamber, the combustion chamber geometry, the thermal state of the cylinder charge, the walls neighboring the combustion chamber, and the type of fuel injection influence mixture formation. In conventional mixture formation and combustion processes in diesel engines, this is done in the combustion chamber itself. For this reason, the mixture formation in diesel engines is also termed internal mixture formation in contrast to classic spark-injection (SI) engines (mixture formation in the intake manifold). The degree of homogeneity of the concentrated field of oxygen and fuel (liquid and vapor) that forms in the combustion chamber during fuel injection and changes according to time and place is a measure of the mixture formation quality. The quality substantially depends on the local and temporal processes and the completeness (pollutant formation) of combustion in the diesel engine. The measurable pollutant emissions in the engine exhaust arise from the interaction between the pollutant formation and the pollutant decomposition in the combustion chamber and the exhaust system. This is especially true for the emissions of soot, hydrocarbons (HCs), and carbon monoxide (CO). The amount of heat released by combustion after ignition determines the gas pressure and temperature characteristics in the combustion chamber in conjunction with the exchange of heat between the fuel, the walls neighboring, the combustion chamber, and the liquid fuel. This also determines the success of energy conversion (mean pressure and fuel consumption) and the mechanical and thermal loads on the engine components. In addition, the change over time of the gas pressure substantially influences the noise emitted by a combustion engine (combustion noise). The injection of the fuel, the decay of the fuel jet into a spray, the evaporation of the fuel, the mixture of the fuel with the air, the heat transfer between the working substance, combustion chamber walls, and fuel, the air movement created or intentionally generated (swirl ducts) by the piston movement as physical processes, and the combustion (oxidation) of the

Internal Combustion Engine Handbook | 607

6606_Book.indb 607

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

fuel as a chemical process occur simultaneously to a degree under continually changing conditions and are mutually influential. These processes, therefore, must be considered in terms of their interrelationships. Figure 15.1 schematically illustrates the relationships and interactions of the processes occurring in the combustion chamber of a diesel engine. Because of the complexity of mixture formation and combustion in diesel engines, their theoretical and experimental study is extremely problematic. An additional difficulty is that conventional engine fuels are not pure substances but are mixtures of different HCs that cannot be exactly defined. It is, therefore, difficult to determine physical and chemical properties and chemical reactions under engine combustion conditions, and these properties, sometimes, can only be approximated. This is also the reason why a high level of understanding has been gained for a series of subprocesses, but all the details of the overall processes involved in diesel engine combustion have not yet been explained. 15.1.1.2 Fuel Injection The design and construction of the injection system, including the injection nozzle, determine how the fuel is fed to the combustion chamber. The injection itself can be essentially characterized by the following: given a single injection per power cycle •• the time of the start of injection and the length of injection given divided injection

•• the point in time and the length of the individual injections •• the time characteristic of the injection rate •• the geometry, the number, and the spatial alignment of the nozzle openings in relation to the combustion chamber (see also Chapter 12.4.1). 15.1.1.3 Mixture Formation The goal of mixture formation is to generate the optimum local mixture of fuel and air (micromixture formation) and the optimum distribution of the air-fuel mixture to the combustion chamber volume (macromixture formation). The goal of optimization is to attain the maximum engine work with minimum fuel consumption and simultaneous minimum exhaust emissions. The limits to the engine noise and the mechanical and thermal loads on components also need to be maintained. Because some of the measures to attain these goals counteract each other, optimization can be only a compromise between the individual demands. A notable example is the contradictory behavior of nitrogen oxide (NOx) emissions and the specific fuel consumption. If we examine the different parameters (such as exhaust regulations) that exist for the individual types of engines, such as large diesel ship engines as opposed to passenger car diesel engines, it becomes clear that there can be no general quantitative formulation of optimum mixture formation conditions for all diesel engines. Nonetheless, there are a few basic considerations that must be considered when designing and optimizing all diesel engines.

Engine behavior Noise, consumption, Emissions, dynamics

Pollutant formation Nox , soot, HC

Heat release

Pollutant decomposition

Ignition system

Mixture formation vaporization characteristics, Duration of liquid phase Jet spread Penetration depth, spray dispersion angle, droplet size distribution

Jet/wall interaction

Flow, turbulence

Primary jet decomposition internal nozzle flow, cavitation

Injection system Injection process, stability

Injection nozzle Bore geometry, HE degree of rounding

Piston movement

Inlet duct geometry Swirl, tumble, squish

Recess geometry K factor

Figure 15.1 Processes involved in mixture formation and combustion in diesel engines [15-1].

608 | Internal Combustion Engine Handbook

6606_Book.indb 608

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

Cavitation bubbles Surface and turbulence angles waves

Solid core

Disturbances from cavitation bubbles and turbulences

Low-cavity core

Gas bubbles

Bands

Droplets

Accumulation

Jet breakdown caused by aerodynamic interactions

Figure 15.2 Two-phase spray model [15-2], [15-3].

The forming fuel jet consists of many individual fuel droplets of different sizes (2–30 × 10–3 mm) and shapes [15-4]. Depending on the parameters of the injection system and the gas state in the combustion chamber, each fuel jet has its own characteristic statistical distribution of droplet size. The droplet size essentially depends on the following influences. The arising droplets are smaller: •• the smaller the nozzle bore diameter •• the higher the exit speed from the nozzle •• the greater the air density in the combustion chamber •• the lower the fuel viscosity •• the smaller the surface tension of the fuel. Additional air movement in the combustion chamber increases the relative movement between fuel and air and, hence, the atomization quality, and the micromixture and macromixture formations. A typical distribution of the droplet sizes in the fuel jet is shown in Figure 15.3.

25

Quantity distribution [%]

For normally used hole-type nozzles [in direct injection (DI) engines], the mixture formation process presently consists of the following steps. Mixture formation starts directly with fuel injection. Depending on the injection system, the fuel jet leaves the injection nozzle at different speeds (>100 m/s). Given the high relative speed of the exiting fuel in comparison to the surrounding air and the implosion of cavitation bubbles from the injection orifices immediately after they leave the nozzle, and the fuel jet disperses almost immediately directly after the nozzle exit. Figure 15.2 shows a measurement-based model of this process.

CFD simulation

20

Hohmann measurement z = 30 mm

15 10 5 0

0

5

10 15 Diameter [μm]

20

25

Figure 15.3 Droplet size distribution in a fuel jet 30 mm from the nozzle [15-5].

Because the fuel is injected at the end of the compression stroke, the fuel droplets in the injection jet are immediately exposed to the high gas temperature in the combustion chamber. This produces an intensive exchange of heat from the heated combustion chamber air to the relatively cold fuel droplets. As the temperature exchange continues between the air and the fuel, the evaporation at the droplet surface increases. The forming fuel vapor then mixes with the surrounding air. Differences in concentration and temperature then form in the environment around the droplets (see Figure 15.4) and, therefore, also in the entire fuel jet (heterogeneous mixture) that subsequently trigger diffusion processes near the individual fuel droplets and in the jet [15-6]. In the middle of Figure 15.4, we see the change over time of the air conditions at the edge of a fuel jet of approximately 26.5 mm from the nozzle exit for three different injection pressures [15-2]. Figure 15.4 (bottom) shows a snapshot of the distribution of the air/fuel conditions in a fuel jet [15-7]. It becomes clear that the ignition conditions in a diesel fuel injection jet (ignition lag) can always attain •• a mixture composition within the ignition limits •• a sufficiently high mixture temperature after a certain amount of time. The primary influential parameters on the development of a fuel jet in diesel engines are schematically illustrated in Figure 15.5. 15.1.1.4 Ignition Lag, Ignition and Combustion [15-4] The physical and chemical processes that start in the combustion chamber at the beginning of fuel injection need time before ignition conditions are attained. This span of time, the ignition lag extending from the start of injection to the first ignition, is very important for the subsequent combustion process. Depending on the conditions existing in the engine upon the injection of the fuel, the ignition lag is up to 2 ms.

Internal Combustion Engine Handbook | 609

6606_Book.indb 609

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

pK

pK

p02

p02 T Droplets

xz

Mini blind-hole nozzle kalte

0.9

0.6

Combustion

me

0.7

Flam

60 MPa 90 MPa 120 MPa

0.8

Air ratio

x

Combustion zone

Soot

1.0

Soot

0.5 0.4 0.3 100

300

500

700

900

1100

1300

Time after start of injection [μs]

With a short ignition lag, relatively little fuel is injected and physically and chemically optimized up to the start of combustion. After ignition, this produces a moderate rise in the pressure and the temperature in the combustion chamber. Because a rise in pressure and maximum pressure in the combustion chamber are major causes of combustion noise, and they are set at a relatively low level. With a low maximum pressure, the mechanical component load is also lower. The maximum gas temperature, the formation of NOxs related to the high gas temperature, and the thermal component load are relatively low. On the other hand, combustion occurs at comparatively low pressures and temperatures, yielding a higher specific fuel consumption and increased soot formation. The latter is because of the relatively large amount of fuel that is injected into the developing hot flame after ignition, and the excessively slow mixture of the forming fuel vapor with the air. Local zones of insufficient air and high temperatures promote soot forming crack reactions. A comparatively long ignition lag produces correspondingly opposite effects. The cited physical and chemical reasons for the ignition lag also provide indications of how to effectively influence it. A short ignition lag is produced by (a) Physical influences •• high gas temperature and high gas pressure at the beginning of injection •• strong atomization of the fuel •• high relative speeds of the fuel and the air.

0

Distance to nozzle [mm]

20

These influences cause the fuel to quickly evaporate, which permits the rapid distribution and mixture of the fuel with the air in the combustion chamber. The gas pressure and the temperature at the beginning of injection can be increased by the following constructive measures:

l 0,6 0,8

•• high compression ratio

1,0 40

•• late injection time

1,6

•• supercharging •• high coolant temperature and suitable cooling channel

3,2

•• combustion chamber design (influence of the wall temperature) •• use of ignition aids (glow elements, earlier glow plugs, and intake air preheating).

60

The atomization of the fuel is mainly determined by the selected injection system as well as the injection time (gas state) and the temperature-related fuel properties (see above). 80

–10

0 10 Radius [mm]

Figure 15.4 Top: oxygen, fuel, soot concentration, and temperature in the environment of a burning individual droplet. Middle: change over time of the air-fuel ratio at a site in the injection jet up to the beginning of cold flame reactions and when ignition conditions are attained for different injection pressures. Bottom: momentary local distribution of the excess air factor in a fuel jet.

The relative speeds of the fuel and the air can be influenced by the constructive design and harmonization of the injection system, combustion chamber shape, and intake duct. The different approaches to the solution of this task include the essential differences between the diesel engine combustion processes. (b) Chemical influences •• high ignitability of the fuel (high cetane number)

610 | Internal Combustion Engine Handbook

6606_Book.indb 610

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

Jet development

Fuel charge – – – –

Combustion chamber wall – Geometry – Temperature

Pressure Temperature Viscosity Speed (swirl, turbulence) according to amount and direction

Jet characteristic

Injection nozzle geometry – – – – – – – – – – –

Needle stroke Needle seat geometry Distribution of inlet areas Inlet conditions into the injection hole (blind-hole or seat-hole nozzle) Rounding radii of the injection hole Injection hole elevation angle //d-Injection hole ratio Bore pattern of the injection hole wall (swirl ribs) Injection hole wall roughness Rounding radii of the injection hole outlet K factor

Fuel flow dynamics – Delivery pressure wave with reflection phenomena – Cross-section change with flow narrowing and cavitation – Needle dynamics – Flow velocity – Transient phenomena – Turbulence – Temperature – Combustion chamber pressure

•• high fuel temperature, high gas pressure, and temperature at the start of injection ensure fast chemical preparation of the fuel. The actual combustion in the diesel engine can be presently described as follows: The slowest processes control the combustion process. Directly after ignition, the fuel that was physically and chemically well prepared during the ignition lag burns quickly and with a high conversion of energy. This first phase is also termed premixed combustion, and is mainly controlled by the still relatively slow chemical processes (low temperature). After this, combustion transits into a second phase that is characterized by the continued injection of fuel into the existing flame, and, hence, by a strongly inhomogeneous charge composition and temperature. In this phase, combustion is again controlled by the slowest process, mixture formation (diffusion). The speed of the chemical reaction is much faster because of the quick rise in temperature in the first phase. As combustion progresses into the third phase, the reaction rate decreases, because the zones of decreasing oxygen and gradually sinking gas temperature from expansion reduce the reaction speed. In addition, the charge movement initiated by the intake process decreases in this phase. The phase is then controlled by the slower mixture formation and the decreasing reaction speed. This produces a thermodynamically problematic delay of combustion far into the expansion stroke. Figure 15.6 shows

Fuel properties – – – – –

Composition Viscosity Compressability Surface tension Boiling characteristics

Figure 15.5 Influences on the development of the injection jet [15-8].

the typical qualitative characteristic of the injection and combustion rate for a diesel engine with direct fuel injection. There is significantly less time available for internal mixture formation in diesel engines in comparison to the classic SI engine. In addition, the boiling curve of the diesel fuel is much higher. This is a disadvantage for the diesel engine in comparison to the SI engine when used in vehicles. With increasing speed, the time problem becomes more pronounced. The unavoidable inhomogeneity of the cylinder charge yields lower average pressures because of the less efficient exploitation of air (smoke limit). Both lower the power output per liter for diesel engines. Hence, today, diesel engines are usually supercharged. Compensating for this disadvantage remains an essential goal in the development of combustion processes for diesel engines. 15.1.1.5 Pollutant Formation In developing combustion processes for diesel engines, major considerations include particle and NOx emissions. The exhaust particles consist of a varying number of HC and/or sulfur compounds that are deposited on soot. HC and CO emissions play less of a role in diesel engine combustion. The formation of pollutants is directly related to the local conditions of ignition, mixture formation, and combustion in the combustion chamber. According to [15-9], soot and NOx formation is strongly influenced by reaction kinetics. These processes are not

Internal Combustion Engine Handbook | 611

6606_Book.indb 611

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

Injection start

Injection end

1. Phase ZV

mK q

2. Phase

3. Phase

Injection duration

q

mK

–20

fEB

20

ZOT fEE

40

completely understood, however. From many investigations of flames and shock wave tubes, we have gleaned Figure 15.7.

Soot yield

50 40 30 20 10 0 0,4

Air

rat

io

0,5

l

0,6

00

00 16

18

00

20

00

e

tur

era

p em

T

K

14

Figure 15.7 Soot yield as a function of the temperature and the excess air factor [15-9].

The soot formation is limited by the temperature and the excess air factor. This restriction is independent from the predominant pressure. The so-called soot yield (soot mass/overall HC mass) increases with pressure, however. At the temperatures of approximately 1600–1800 K and the excess air factors of <0.6, the soot yield approaches a maximum. For many HCs, the soot formation thresholds are very similar, allowing these considerations to also be applied to diesel engines. The previously described heterogeneous mixture formation in diesel engines means that despite an overall excess air/fuel factor >1, local excess air/fuel factors of <0.6 can arise. As long as the

60 f° Crankshaft angle

Figure 15.6 Qualitative characteristic of fuel injection and heat release [15-6].

mixture temperature remains below approximately 1450 K, soot formation is largely excluded. When a burning, relatively “rich” mixture cools (close to the wall, for example) or fuel is heated that is insufficiently mixed with air, intense soot formation occurs. Because of the influence of the wall (quenching) and the fuel composition, the start of soot formation in diesel engines must be assumed when there is an excess air/ fuel factor of <0.8. Figure 15.8 shows a temperature/excess air factor diagram in which the states of the mixture and combusted material close to ignition TDC are plotted in addition to the soot formation range. In addition, the range of intense NOx formation (components formed within 0.5 ms) is also shown. As we know, the highest NOx formation rates occur with an excess air factor of 1.1. In this range, a part of the formed soot can be recombusted, as the reaction time of soot particles (d = 40 nm) in this range shows. As the excess air/fuel factor further increases toward the average combustion chamber value, there is a decrease in the combustion temperature and, hence, NOx formation. The graph also provides an explanation of the typical contradictory behavior of soot and NOx emissions in diesel engines. Relatively low temperatures and a deficiency of air promote soot formation and lessen NOx formation. High temperatures and excess air have opposite effects. A clear reduction of both emissions is basically attainable only when the fuel-air mixture is leaned off or homogenized as quickly as possible at a very low temperature before ignition while avoiding “rich” areas, or the mixture is combusted at excess-air factors of between 0.6 and 0.9.

15.1.2 Diesel Four-Stroke Combustion Systems

The combustion strategy that arose over the course of developing diesel engines can be explained and understood based on the above-described processes in the combustion chamber. Rudolf Diesel did not have an opportunity to use industrially manufactured, highly developed injection techniques during his lifetime. His attempt to use the high-pressure fuel injection system with which we are familiar today failed because of the technical restrictions of his time. As an emergency solution, he

612 | Internal Combustion Engine Handbook

6606_Book.indb 612

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

Soot oxidation time 0,4 ms τ = 1,0

Temperature [K] 3000

Combusted Soot formation

ppm NO 5000 3000 1000 500

2500

2000

1500 1,0 1,5 2,0

4,0

λ Mean Air-fuel-ratio

1000

Mixture 500 Target range

0 0

0,5

1,0

Local air-fuel ratio λ ö

developed a method in which the liquid fuel was blown into the combustion chamber of the engine using compressed air. According to [15-10], this combustion process is distinguished by very even and “gentle” engine running. The exhaust was soot-free and odorless. Given today’s knowledge, this result can be explained as follows: •• Blowing in the air produces a very fine fuel atomization (fast evaporation and mixing). •• Before the actual combustion phase, there is an intense mixture of fuel and air in the nozzle. •• Because of the cooling effect of the blown air and further cooling as the fuel flows into the cylinder (expansion), soot formation is largely prevented in the arising mixture. The decisive disadvantage of the later developments of this procedure was the large amount of work to drive the required air compressor. This resulted in correspondingly high fuel consumption. Because of the direct introduction of the fuel into the combustion chamber, this method can be termed a DI method, although it is fundamentally different from the high-pressure fuel-injection procedure used today. The DI of fuel is, therefore, the historically oldest combustion system. In the described form, it was limited only to relatively low-speed engines, that is, engines with large combustion chambers. To make the diesel process useful for high-speed engines and vehicle use, additional developmental steps were necessary. An important prerequisite was the introduction of

1,5

2,0

Figure 15.8 States of the mixture and combusted material in diesel engine combustion [15-7].

high pressure injection that became technically feasible at the beginning of the 1920s, which was cheaper and metered fuel better. The disadvantage of high-pressure injection in contrast to blowing in air was that, without additional measures, the mixture was formed only by the injection nozzle (without air support). To a certain degree, the charge turbulence that rises in the combustion chamber as an rpm increase is sufficient to compensate for the shorter time because of the increased mixing speed and, hence, increasing combustion speed. Given accelerations within ranges that are covered by today’s mediumhigh-speed and high-speed engines, solutions had to be found that permitted a corresponding acceleration of the mixture formation and combustion processes. The recognized important variables were the relative speeds of the fuel and the air, and the influence of the combustion chamber wall on the speed of the mixture formation processes. The relative speeds of the fuel and the air are most effectively influenced by •• the fuel speed in the combustion chamber (injection pressure) •• the air speed in the combustion chamber (combustion chamber and intake duct design). The best results in engines are attained by the optimum harmonization of fuel injection and the air movement in the combustion chamber. The diesel engine combustion system arose against this background.

Internal Combustion Engine Handbook | 613

6606_Book.indb 613

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

15.1.2.1 Methods Using Indirect Fuel Injection In engines that use this method, the combustion chamber is divided. It consists of a main combustion chamber and a secondary combustion chamber. The secondary combustion chamber is in the cylinder head. The main combustion chamber is formed by the cylinder and a recess in the piston head that is adapted to the position of the chamber mouth. These engines are, therefore, also termed indirect-injection or divided chamber engines. The main and secondary combustion chambers are connected by one or more channels with a cross section that is narrower than that of the main and secondary combustion chambers. The secondary combustion chamber is designed as a whirl chamber or prechamber. Both methods have the following in common. The fuel is injected under a moderate pressure (<400 bar) by means of an inline or distributor injection pump into the secondary chamber. Throttling pintle nozzles (small injected quantity during the ignition lag) are used as the injection nozzles. After the first amount of fuel is quickly mixed with the air followed by a relatively short ignition lag (high chamber wall temperatures), ignition occurs in the secondary chamber. The air overflowing from the main combustion chamber at a high speed into the secondary chamber because of the displacement effect of the piston during the compression phase greatly supports mixture formation in the secondary chamber. Directly after ignition, the pressure and the temperature quickly rise in the secondary chamber above the values in the main combustion chamber. The higher chamber pressure causes the air-fuel mixture forming and partially burning in the secondary chamber to intensely flow into the main combustion chamber. This causes the outflowing mixture to thoroughly mix with the sufficient amount of air in the main combustion chamber. Combustion processes that use indirect injection tend to form a greater amount of soot. The reason is the lack of air at relatively high temperatures in the secondary combustion chamber after ignition. Under high engine loads, a part of the soot formed in this phase can still burn in the main combustion chamber. Under a partial load, however, the temperatures for effective afterburning are too low. The formation of NOxs is largely suppressed in the indirect fuel injection (IDI) method. The air deficiency in the chamber is then an advantage in this regard. When the mixture is expelled from the secondary chamber, it is quickly diluted so that high local temperatures are largely avoided along with excess air factors that support NOx formation. The intense mixture formation in the IDI method also produces favorable HC and CO emissions in these engines. Another advantage of the intense mixture formation is the resulting relatively low rise in cylinder pressure that produces correspondingly low noise. This combustion process also allows high air utilization (close to the stoichiometric mixture composition) under a full load and, simultaneously, at high engine speeds. The described properties of the IDI method secured indirectinjection engines a predominant position among high-speed engines for a long time, especially passenger car diesel engines.

Even among medium fast-speed engines, indirect-injection engines were represented in the upper speed range. 15.1.2.1.1 Prechamber system [15-11] This method arose in the 1920s, and its development is largely concluded by now. Figure 15.9 shows a prechamber according to [15-12].

Figure 15.9 Combustion chamber arrangement of a four-valve prechamber engine (DaimlerChrysler AG).

In contrast to a two-valve design, the design shown here in a four-valve engine has greater potential for savings in fuel consumption and lower exhaust emissions because of the symmetrical and central arrangement of the prechamber in relation to the main combustion chamber. All prechambers are rotationally symmetrical. The actual chamber area in which the fuel is injected can be spherical to egg-shaped. The chamber is connected to the main combustion chamber via a duct that ends in several combustion holes. The number, direction, and diameter of combustion holes must be optimized with the piston recess. According to [15-4], the chamber volume should be approximately 40%–50% of the compression volume. This percentage strongly influences soot and NOx formation and should be correspondingly optimized. The optimum cross section of all the combustion holes is 0.5% of the piston cross section. A greater number of holes reduce soot emissions. The direction of the combustion holes is set in relation to the thermal load of the piston heads. The injection nozzle is on the top of the chamber opposite the duct. The compression ratio of these engines is between 21:1 and 22:1. In the above example, the ratio is 21.7:1. The prechamber system is not particularly suitable for very small cylinder-stroke volumes. The mixture

614 | Internal Combustion Engine Handbook

6606_Book.indb 614

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

formation in the chamber can be optimized with a spherical pin whose geometry and position is adapted to the chamber (see Figure 15.9). This pin is perpendicular to the direction of the fuel injection jet, and supports the preparation of the contacting fuel jets and the fuel distribution in the secondary chamber. Because of the special shape of the bottom of the spherical pin, a slight swirl is generated in the secondary chamber during the inflow of the air coming from the main combustion chamber in the compression phase that provides a more intense mixture formation in the secondary chamber. It is best for the injection nozzle to be at a slight angle in relation to the lengthwise chamber axis. Despite the relatively high compression ratio, the method does not work without ignition assistance. It is helpful to place a glow element or glow plug downwind from the air stream in the secondary chamber following the injection nozzle. To support the warm-up phase, the glow element can be operated for up to 1 min. 15.1.2.1.2 Whirl chamber system [15-13] In this type of combustion system, the secondary chamber is in the cylinder head as is the case with the prechamber system (Figure 15.10).

Injection nozzle Glow element

Figure 15.10 Whirl chamber with an injection nozzle and glow element (Opel Omega 2,3 D) [15-13].

The chamber can be disk-shaped, spherical, or oval. The main combustion chamber and the secondary chamber are connected via a passage with a relatively large flow cross section. The transfer passage ends tangentially in the actual combustion chamber so that, as the piston moves upwards, the air flowing into the secondary chamber is subject to a forceful swirling movement. The ratio of the whirling speed to the engine speed is related to the operating state of the engine, especially the rpm, and is between 20 and 50. The whirling in the chamber at the beginning of compression initially corresponds to a solid swirling that becomes more similar to a potential swirling in the last phase before TDC. The whirling in the chamber at the beginning of compression initially corresponds to a solid swirling that becomes more similar to a potential swirling in the last phase before TDC. The

maximum circumferential speed increases with the rpm and always occurs in the range of 10–20° crank angle before TDC. The piston recess suppresses the burning flame at the edge of the recess and, hence, reduces the danger of incompletely combusted fuel being transported to colder regions of the piston head and causing increased soot formation there. At the present stage of development [15-4], the optimum chamber volume is approximately 50% of the compression volume. The optimum overflow cross section is 1%–2% of the piston cross section. The injection nozzle is in the top part of the secondary chamber, so that the fuel jet enters the secondary chamber tangentially against the incoming air directed at the hot, opposing chamber wall in order that the swirling air in the secondary chamber penetrates the jet at a right angle. As the fuel jet passes through the whirl chamber, a part of the fuel is quickly evaporated and becomes ignitable. Most of the injected fuel reaches the 900-K chamber wall where it vaporizes relatively slowly. Ignition greatly accelerates this process. The forming fuel vapor is mixed quickly and intensely by the whirling motion of the air in the chamber. The remaining combustion process is the same as in a prechamber engine. The compression ratio of these engines is between 22:1 and 23:1. The stroke/bore ratio for these types of engines is between 0.95 and 1.05. The whirl chamber system can be used at up to the speeds of approximately 5000 rpm (somewhat greater than the prechamber system), and is, hence, especially suitable for use in passenger cars. The combustion properties and attainable average pressure at the soot limit are comparable with those of prechamber engines. The whirl chamber system also does not work without ignition assistance. The position of the glow element or the glow plug in the secondary chamber strongly influences engine performance, so that it requires a specially optimized design. The disturbing influence of the glow element on chamber flow can be compensated to a certain degree by reducing the cross section of the transfer passage (increase in speed). 15.1.2.2 Direct Fuel Injection Method In the direct fuel injection method, the combustion chamber is not divided [15-4], [15-14], [15-15], [15-16] (Figure 15.11). The actual combustion chamber is formed by a recess in the piston head. Up to 80% of the compression volume can be incorporated in this piston recess. Diesel engines with a cylinder diameter larger than approximately 300 mm usually work without additional air movement in the combustion chamber. The mixture is formed exclusively by the injection system, especially the injection nozzle. Multihole nozzles are used as injection nozzles with up to twelve nozzle holes depending on the engine size. An engine with four valves allows the combustion chamber to be symmetrical (which has a positive effect on mixture formation and the thermal load) because of the centrally located injection nozzle aligned with the cylinder axis. Figure 15.11 shows an asymmetrical design with two valves.

Internal Combustion Engine Handbook | 615

6606_Book.indb 615

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

Figure 15.11 Combustion chamber arrangement of a two-valve DI engine with a pump-nozzle injection system (Audi AG).

The injection pressure (1300 to more than 2000 bar) and the nozzle bore diameter determine the size of the fuel droplets and the relative speeds of the fuel and the air in the injection jet. The combustion chamber is largely open and adapted to the shape and position of the injection jets. In smaller, faster running engines, the air movement generated by the intake process, fuel injection, and piston movement is no longer sufficient for good mixture formation. Special measures must be taken to increase the relative speeds of the fuel and the air in the combustion chamber. By designing the intake duct as a swirling and/or tangential duct, the air swirls intensely around the cylinder axis as it flows into the combustion chamber.

This swirling overlaps the turbulence that already exists in the combustion chamber and causes the rapid distribution and mixture of the fuel vapor arising upon injection directly next to the injection jet with the air in the combustion chamber (macromixture formation). Another possibility of increasing the relative speeds of the fuel and the air in the combustion chamber is to narrow the piston recess near the piston head. During the compression stroke, the air above the piston head is, thereby, directed toward the piston recess. When the air flows into the recess, an intense whirling movement arises, the so-called squish (Figure 15.12). The advantage of the squish in contrast to swirling is that it increases in intensity as the piston approaches TDC (fuel injection phase) while the swirling generated during induction decreases. As the speed of the engine increases, both methods are combined. To attain the best values for fuel consumption and exhaust emissions, the intake ducts, the combustion chamber geometry, and the fuel injection must be optimized and harmonized with each other (Figure 15.13). A reduction in the number of nozzle holes requires an increase in the swirl and vice versa. If the swirl is too high, the fuel in the individual fuel jets becomes mixed. This produces local “over-enrichment” of the mixture and, hence, worse air utilization and high exhaust emissions. In vehicle engines, it is particularly difficult to optimally harmonize the mixture formation over the entire operating range. Simulation methods (3D) and improved experimental possibilities (such as the transparent engine) help solve these tasks. Swirling must be adapted to the load and the rpm for an optimum engine operation. High-speed engines need compression ratios between 15:1 and 19:1 and, like engines with IDI, must have glow elements as a starting aid for reliable cold starts and warm-up. Today, these engines attain the maximum speeds of up to 4500 rpm and an efficiency of 43% with exhaust turbocharging at their best point. In large engines, compression ratios between 11:1 and 16:1 are necessary depending on the amount of turbocharging. In addition, efficiencies of more than 50% are attained. The described relationship between engine speed (engine size), air movement, and combustion chamber shape is revealed in the comparison of typical combustion chamber shapes in engines that use DI with increasing speeds (Figure 15.14).

Cylinder head

Squish flow

4-hole injection nozzle

Swirling

Jet spread without air swirl Jet spread with air swirl

Figure 15.12 Flow processes in the combustion chamber of a diesel engine that uses DI and mainly air-distributed fuel [15-4].

616 | Internal Combustion Engine Handbook

6606_Book.indb 616

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

increased with simultaneously smaller diameters causing a reduction in the swirl.

2000

8

4,5 1500

4

1000

500

2

3,5

Integral swirl number [-]

Injection pressure

Swirl

4,0

Injection pressure bar

Swirl number

6

3,0 Flow factor spread in various series engines

2,5 2,0 1,5 1,0

0

0

2

4 6 Nozzle hole number

8

0

Figure 15.13 Typical relationship between injection pressure, swirl number, and nozzle hole number [15-16].

Figure 15.14 shows typical combustion chamber designs of engines with DI at increasing speed. We can clearly see the increasing narrowing and deeper piston recess as the speed increases. This increases the squish, and the swirl is sustained into the expansion stroke. The necessary swirl increases in the same manner (Figure 15.15). There is likewise a simultaneous reduction in the number of nozzle holes. It becomes difficult to optimally harmonize the combustion process as the engine speed rises, because the system becomes more sensitive to the combustion chamber geometry. In passenger car combustion chambers, particular attention needs to be paid to the specific shape of the recess edge (turbulence ring) [15-17]. The design with four valves and a central injection nozzle is becoming increasingly popular with smaller cylinder sizes. Because of the injection pressures that can be attained today, the number of nozzle holes can be

Cylinder diameter decreases, speed increases required swirl number increases

0,5 0 50

60

70

80 90 100 110 120 130 140 150 Cylinder diameter [mm]

Figure 15.15 Typical relationship of the required swirl number to cylinder diameter [15-17].

15.1.2.3 Comparison of Combustion Systems The previously discussed combustion systems are primarily compared regarding their specific fuel consumption, exhaust emissions, and combustion noise [15-4], [15-18]. They basically differ according to how they generate the relative speeds of the fuel and the air that are required for mixture formation. The IDI methods work at low injection pressures and relatively low fuel speeds, and, therefore, require high air speeds. In DI methods, high fuel speeds are attained by high injection pressures. They can, therefore, get by lower air speeds. The swirl ducts that are necessary in engines that use DI to generate air movement restrict the cylinder charge at high speeds and increase charge cycle losses. The necessary flow speeds near TDC tend to be lower with DI, a bit faster in the whirl chamber system, and highest in the prechamber system.

Figure 15.14 Influence of engine size (speed) on the combustion chamber recess shape and required air movement in diesel engines that use DI as described by [15-4].

Internal Combustion Engine Handbook | 617

6606_Book.indb 617

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

of combustion systems with DI (Figure 15.18) is supported by additional advantages.

2

1

0

5 g kWh 4

25

50

75 % 100

3

n = 2500 min–1 Full-load Gasoline engine DI diesel engine IDI diesel engine

180 160 140 120 100

1000 Frequency f

10000

Hz

Million engines

6

Indirect injection Direct injection

4

2 40 % 0

1996

63 %

1999

87 %

2002

95 %

2005

Figure 15.18 Production development of a passenger car with a diesel engine in Western Europe (Bosch GmbH).

Engines that use DI experience a lower thermal load. This makes them highly suitable for exhaust turbocharging that can be used to reduce exhaust emissions. The progress over recent years in turbocharger development (such as variable

10

2

0

dB 200

15 g kWh

1 0

220

Figure 15.17 Gas pressure excitation spectra of different diesel combustion systems in comparison to the SI engine [15-4].

NOx-emission

Commercial vehicle DI engines Car DI engines Indirect injection engines

HC-emission

Smoke number

Bosch 3

Cylinder pressure excitation level Lpz

As the flow speed in the combustion chamber rises, the flow losses increase. In addition, faster flow speeds cause a greater transfer of heat and, hence, a greater loss of wall heat and higher thermal load; this is further pronounced in indirect-injection engines in contrast to engines that use DI because of the larger combustion chamber surfaces. Because of the greater flow and heat transfer losses, combustion systems using IDI have an approximately 15% higher fuel consumption than engines using DI. Because of the less favorable surface/ volume ratio (30%–40% higher than DI) of the combustion chambers, engines using IDI exhibit worse cold start behavior that cannot be fully compensated by a higher compression ratio. The higher injection pressures required in engines that use DI lead to more expensive injection systems subject to greater loads. The higher charging speeds with the IDI method allow better air utilization. This allows low air-fuel ratios at the smoke limit, which, in turn, compensates for worse delivery and fuel consumption in comparison to DI engines. More or less equally high average full load pressures are, therefore, attainable. Black smoke emissions are more pronounced for the IDI method, particularly in the lower load range in comparison to DI. As the engine load increases, NOx emissions are better for the IDI method than for direct fuel injection. In terms of HCs, IDI has advantages over DI (Figure 15.16). The enormous advances in the development of injection technology, especially increasing injection pressure, have reduced the emission advantages of IDI. Only in the case of NOxs, DI needs substantial improvement because of the higher constant volume component in the supply of heat, which is also the reason for the noisier combustion (Figure 15.17). Because DI engines have a substantially higher EGR rate than indirect-injection engines, the disadvantage of the emissions can be compensated. Common-rail injection systems allow the realization of nearly any temporal injection progresses and any partition of fuel injection, and therefore, we can assume further improvements in the future. Now, the future appears to lie with combustion systems with DI, especially because of the clear fuel consumption advantage. This positive development in favor

0

25

50 Load

75 % 100

5

0

0

25

50

75 % 100

Figure 15.16 Comparison of exhaust emissions in different combustion processes [15-9].

618 | Internal Combustion Engine Handbook

6606_Book.indb 618

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

turbine geometry) makes high-speed turbodiesel engines with DI serious competitors of SI engines for use in passenger cars. In summary, it can be stated that combustion systems using DI have largely displaced combustion systems with divided combustion chambers independent of engine speed and size. 15.1.2.4 Development Trends 15.1.2.4.1 Homogeneous compression ignition [15-19], [15-20] Continuously stricter exhaust regulations are motivating the search for improved diesel engine combustion systems. The current activities are directed, in particular, toward a solution or mitigation of the NOx/particle issue (see also Chapter 15.1.1.5). At the same time, fuel consumption will only increase marginally or not at all. One option under consideration is the autoignition of heavily homogenized diesel fuel/air mixtures. 15.1.2.4.2 Principle The principle is the extensive mixture homogenization before combustion start. By compressing the homogeneous mixture, ignition occurs simultaneously at many locations in the combustion chamber (spatial combustion). The combustion start and combustion process, here, are controlled mostly by the chemical kinetics. The reaction process of the diesel fuel has two stages. It starts with the so-called cold flame reaction (below approximately 900 K) and continues, after a short phase of a reaction with negative temperature coefficient, as a hot flame reaction (above approximately 1000 K) with significantly increased intensity. High combustion speeds occur (short combustion period), causing steep pressure rises in the cylinder. If an optimal combustion center can be maintained, the thermal efficiency will also increase. To attenuate the combustion speed to normal levels, the charge must be diluted, which can be achieved by leaning down the mixture (high λ values) or, more expediently, EGR. Because of the mixture homogenization and the fact that the burning fuel must also heat the entire cylinder charge, high local temperature peaks can be avoided. However, diesel fuel has, because of its high boiling range (moderate vaporization) and its high cetane number (early start of combustion), less favorable preconditions for the generation, that is, combustion of a homogeneous mixture, when compared with gasoline. 15.1.2.4.3 Options of mixture homogenization As a rule, a homogeneous air/fuel mixture can be obtained by transporting the fuel (in liquid or gas or vapor form) into the intake manifold or, directly, into the combustion chamber. The vaporization of fuel requires an additional energy and system expense, but has the advantage that vaporized fuels can be mixed with air significantly faster and more evenly than a liquid fuel. By feeding the fuel into the intake manifold, an additional and/or very different mixture formation system is required, compared with a conventional diesel engine. There is also the risk of an increased fuel deposit at the intake manifold’s walls. For this reason, processes with internal mixture formation (DI) are preferred in the development of homogeneous diesel combustion:

•• Homogenization by early injection during the intake or compression stroke (long ignition lag, more time for mixture formation but the risk of wall wetting and lubrication dilution). •• Homogenization by late injection during the expansion phase (long ignition lag, more time for mixture formation but less thermodynamically favorable). •• Multiple injections (improved local distribution of the fuel in the combustion chamber, avoidance of wall wetting). •• Use of injection nozzles with, for example, up to 40 holes (they may be laser-drilled and have a diameter of <0.1 mm); injection nozzles with different spray hole diameters (formation of smaller fuel droplets with faster vaporization); load-dependent adjustment of the spray hole diameters (varionozzle). •• Blowing fuel vapor into the intake manifold or the cylinder (vapor can be quickly mixed with air, and is more expensive). •• Homogenization can be supported by an optimized adjustment of the load movement and the prewarming of the combustion air (promotes fuel vaporization that may cause a thermodynamically unfavorable early ignition timing). 15.1.2.4.4 Problems of homogeneous diesel combustion The problems of homogeneous diesel combustion are the following: •• Combustion start and combustion progress (combustion noise, combustion duration, and so on) cannot be controlled via the start of injection but by the state of the cylinder charge at inlet closing and during compression, as well as the charge composition (for example, variable intake air temperature, variable compression ratio, and variable EGR rate). •• Tendency toward premature ignition because of the high cetane value of diesel fuel (lowering the compression ratio and the charge dilution required). •• Generating an optimal ignition window (λ -T range). •• Achieving high mean pressures is restricted by knocking phenomena and/or λ ⪢ 1 or high EGR rates.

•• The motor speed is limited toward the top range as the mixture formation (homogenization) is time-controlled.

•• The HC and CO emissions increase because of intensified wall influences (wall wetting and flame extinguishing) and an incomplete conversion of both the components because of the lower temperatures (countermeasure: oxidation catalyst; but higher expense and light-off problem because of lower exhaust gas temperature caused by the high dilution of the charge). •• Mixture homogenization and complex process control significantly increase the necessary effort. •• Wall wetting by the injected fuel (in particular, when the fuel is fed into the intake manifold or at an early injection into the combustion chamber during the compression stroke) must be avoided as much as possible. •• The cold start becomes more difficult because of the reduced compression ratio to avoid premature autoignition. •• The control of the process is very complex, in particular, during transient operational states.

Internal Combustion Engine Handbook | 619

6606_Book.indb 619

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

15.1.2.4.5 Conclusion The combustion of a mixture that is fully homogenized before the compression ignition occurs would be ideal because of its low emission of pollutants. This ideal process is called the homogeneous charge compression ignition (HCCI) process. From the current point of view, the operation of a diesel engine with homogeneous combustion is only possible in the lower load and speed ranges. Figure 15.19 demonstrates the potential of this combustion process. The so-called “knock limit” limits the engine operation to high mean pressures caused, as a rule, by the necessary charge dilution by a leaning-down process (λ min). This limit can be expanded to the corresponding ignition limit by an increasing EGR rate. The maximum mean pressure is then achieved with λ ≈ 1 and high EGR rates. Toward the lower end, the mean pressure is defined by the smallest possible fuel amount. The maximum speed is determined by the time required for forming a homogeneous mixture. Turbocharging can still expand the upper limit for the mean pressure. Because of the given, relatively tight, limits for mean pressure and speed and the problems associated with the aforementioned properties of diesel fuel, the engine must be operated with a heterogeneous mixture in the upper load and speed ranges. To realize both the mixture formation processes in an engine, very high demands are made on the optimal combustion chamber design, the injection system, and the EGR regulation because of the extremely variable injection times. A major challenge is the regulation of transient engine operation and the seamless transition from the homogeneous compression ignition to the conventional diesel operation (using combustion chamber sensors, for example). Furthermore, for practical reasons, only those processes can be used that are capable of optimizing the degree of homogenization of the cylinder charge in respect to NOx and particle emissions and the fuel consumption in the entire engine map by the process control system (fuel injection, exhaust gas return— internally and externally—and charging, that is charge air pressure and temperature, in particular). The main point is to

achieve a far-reaching mixture homogenization in the largest possible partial load and speed range, while maintaining the aforementioned limits, and to optimally adjust the degree of homogenization (partial homogenization) of the charge to the operational conditions in the engine. 15.1.2.4.6 Proven processes (selection) The operation of dilution-controlled combustion system by Toyota is the same as a conventional diesel engine, except for the necessary very high EGR rates of approximately 75%. Because of this high charge dilution, the working substance temperatures are lowered under the threshold values for intensive NOx and soot formation. Mean pressures of the high-pressure process of approximately 10 bar are reached at very low NOx and soot emission values. The HC emissions are considerably increased. The indexed efficiency is very much lowered. The modulated kinetic process presented by Nissan uses a partial homogenization affected by late injection above the upper TDC. By extending the ignition lag because of lower working substance temperature, an EGR rate of approximately 40% is supported. To prevent soot formation, a rigid temporal separation of the injection and the combustion phase is important. The achievable indexed mean pressure of the high-pressure process is approximately 8 bar. It is limited by the injection time, which increases with the engine load and the shortening of the ignition lag that occurs because of the simultaneously increasing process temperatures. Efficiency, HC, and CO emissions are approximately the same as in the conventional diesel engines. This process is also known as highly premixed late injection (HPLI). If the partial homogenization and high EGR rate characteristics are combined, one can obtain the so-called homogeneous charge late injection (HCLI) process. This process is characterized by a comparably early injection. The temporal distance between the injection end and the combustion start is significantly higher than in the HPLI process. No soot is generated at very low NOx levels and efficiencies that are

Effective mean pressure [bar]

20 16 12

DE Diesel

8

0

1500

Untreated emissions 487 348

300

Compression 200 ignition (homogeneous) 100

4 0 500

500 % 400

7,4 NOx

5,6

Soot HC (From FSN)

2500 Engine speed [1/min]

3500

CO 4500

Figure 15.19 Potential of homogeneous compression ignition in a DI diesel engine [15-19].

620 | Internal Combustion Engine Handbook

6606_Book.indb 620

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.1 Diesel Engines

close to a conventional diesel engine. The HC and CO emissions are similar to the ones of modern SI engines with DI. The process is mostly impervious in respect to the setting parameters. Indexed mean pressure of the high pressure up to approximately 6 bar can be attained. Figure 15.20 displays the working ranges of the aforementioned combustion processes in the λ -T diagram. In addition, Figure 15.21 represents each position of the fuel injection and the energy conversion in the crankshaft arc. 10

DCCS

NOx

HPLI Local air ratio [-]

HCCI HCLI

1

DI-Diesel

Alternative combustion methods

0,1 1000

Ruß

1500 2000 2500 Local flame temperature in K

3000

Figure 15.20 Working ranges for homogeneous or partially homogeneous diesel combustion processes [15-19].

15.1.2.4.7 Applications For practical application, a combination of several complementary processes is useful:

Needle stroke

•• up to an effective mean pressure of approximately 4 bar, use of the HCLI process

100%

HCCI

HCLI

•• beyond this value, up to an effective mean pressure of approximately 6–8 bar, use of the HPLI process •• beyond this value up to full load, conventional diesel combustion. A retrofit and stationary-operated passenger car diesel engine with 2.2-L displacement showed these results. Calculated for the New European Driving Cycle from the EURO 4 Emission Directive, a vehicle test weight of 1590 kg, reduction potential of approximately 60% for the NOx and 70% for particle emissions, was obtained. 15.1.2.4.8 Outlook The future developments are oriented toward an increase in the load limits for the HCLI process to eliminate the HPLI process. This requires further research for improving the charge dilution by splitting the necessary EGR volume between internal and external EGR. Multiple injections during compression stroke to generate a homogenized partial mixture and its defined ignition by main injection are one option. The interaction of the modified combustion processes with the requirements of exhaust gas turbocharging and operating point-dependent control and regulation of the processes must be further research and optimized. Although a fully homogeneous diesel combustion (HCCI) in the entire engine map must be considered to be unrealistic, and the homogeneous or partially homogeneous diesel combustion will increasingly affect the continued development of the combustion processes for the diesel engines of the future. Further improvements can be anticipated because of an optimal adaptation of the fuel to the combustion process. The development of synthetic fuels will possibly show further potential. Because the homogeneous autoignition is also interesting for SI engines, the combustion process for diesel and SI engines will probably become more similar. The exhaust gas aftertreatment system must be adjusted for the modified circumstances.

DI-Diesel DCCS HPLI

0%

Combustion rate dQB [J/°CA]

300 HCLI

250

HCCI

DI-Diesel

DCCS

HPLI

200

150 100

50 0 –50 –120

–100

–80

–60

–40

–20

0

20

40

Crankshaft angle [°CA]

60

80

100

120

Figure 15.21 Energy conversions for homogeneous or partially homogeneous diesel combustion processes [15-20].

Internal Combustion Engine Handbook | 621

6606_Book.indb 621

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

15.1.2.5 Special Methods and Features 15.1.2.5.1 MAN-M process This method completely broke new ground. Whereas the rule of thumb was to keep the fuel away from the combustion chamber wall, in this case, the fuel was directly applied to the wall. The spherical combustion chamber is centrally located in the piston head. This arrangement yielded the name middle sphere method (Figure 15.22). The fuel is injected with a one- or two-hole nozzle at a relatively low pressure tangential to the combustion chamber wall where it first spreads as a film. Only a small part of the fuel is distributed by the air to start ignition. Combustion systems with DI can, hence, be divided into wall-distributing and air-distributing systems. Neither method can be fully represented, however. Hence, the designed combustion processes, especially for small cylinders, are better described as primarily wall-distributing or air-distributing. In the MAN-M method, applying the fuel to the wall reduces the fuel speed to nearly zero, and the liquid fuel is not exposed to a high combustion chamber temperature (a wall temperature of approximately 340°C at full load). To attain a high relative speed of the fuel and air, we need a high air speed in the combustion chamber that is attained by means of swirl ducts. During the ignition lag, a slight amount of fuel evaporates from the combustion chamber wall. A correspondingly slight amount of fuel is prepared for combustion, which yields a very low pressure rise and combustion noise. After ignition, the fuel film rapidly evaporates off the wall because of the high gas temperature. The intensely swirling air quickly mixes the air and the fuel. In the whirl itself, there is a separation of the hot combustion gases that travel along spiral paths to the center and the relatively cold air that is forced outward toward the fuel. Because the fuel is initially separated from the high gas temperature, the soot emissions are relatively low. This combustion system is, therefore, distinguished by good air utilization, and it attains high average pressures at the smoke limit. Disadvantages are the high flow and heat transfer losses

that increase fuel consumption thermal load, especially on the piston and cylinder head. For this reason, this method is not particularly suitable for supercharging. In the partial load range, mixture formation worsens because of the falling temperature which, in particular, leads to increased HC emissions. The disadvantages of this method in comparison to the primarily air-distributing combustion system with DI are the reasons that it is no longer used today. The M method was chiefly used in commercial vehicles. 15.1.2.5.2 FM method In developing the M method, it was found that it was also suitable for burning low-boiling fuels (multifuel suitability). This led to the development of the FM method. The inner mixture formation, combustion chamber shape, and load regulation are taken from the M method. Ignition occurs with the aid of a spark plug [F = Fremdzündung (externally supplied ignition)] as in an SI engine. The process closely corresponds to the constant pressure process. The behavior of the exhaust emissions is somewhat better than with the M method. Because of the combination of features of the classic diesel and SI methods, the FM method is counted among the hybrid combustion systems. 15.1.2.5.3 Ignition jet method In this method, a small amount of diesel fuel (up to 10% of the full load heat consumption) is injected to ignite a fuel-air mixture that is homogeneously premixed outside of the engine cylinder and introduced into the combustion chamber. The homogeneously premixed fuels can be lean mixtures of diesel, alcoholic, or gaseous fuels. In practice, this method is mainly used in ignition jet gas–diesel engines. These engines can also be operated as fuel-changing engines; that is, the quantity of diesel fuel can be increased from the ignition quantity to the full load quantity while simultaneously reducing the quantity of gas. The engine then runs completely on diesel. The advantage is that such engines can also be operated when a continuous gas supply cannot be ensured and/or when

Fuel jet

Fuel film

Air swirl

Movement of hot reaction zones

Figure 15.22 Combustion chamber—fuel and air movement in the MAN-M method [15-6], [15-21].

622 | Internal Combustion Engine Handbook

6606_Book.indb 622

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

full diesel engine performance is needed when using weak gases, for example. 15.1.2.5.4 Gas diesel engine In this method, air is compressed as in the conventional diesel engines. Highly compressed combustion gas is blown into the compressed air (internal mixture formation). While the combustion gas is being blown in, local regions with a combustible gas–air mixture arise in the combustion chamber. The autoignition temperature of the combustion gases is usually higher than that of diesel fuel. During the resulting longer ignition lag, a relatively large amount of ignitable gas–air mixture forms. This burns almost immediately upon autoignition. Impermissibly high-pressure gradients and maximum pressures arise in the combustion chamber. For this reason, a slight amount of diesel fuel is injected (pilot fuel). The gas is blown in only after the pilot fuel ignites. This allows the gas to burn almost immediately with an acceptable rise in cylinder pressure. 15.1.2.5.5 Particularities of heavy oil operation Heavy oil operation that is required today for medium-speed four-stroke diesel engines with piston diameters of approximately 200–600 mm and speeds between approximately 1000–400 rpm has a few particularities regarding combustion. The vanadium and sodium content of the heavy oil can produce deposits in the combustion chamber during combustion. The result is high-temperature corrosion. This seriously restricts the continuous operation of an engine or makes it even impossible. With the water rising during combustion, the sulfur in the heavy oil causes the formation of sulfuric acid and sulfurous acid when the oil falls below the dew point and, hence, causes low temperature corrosion. This makes it necessary to design the cooling system for the components forming the combustion chamber, so that critical temperatures are not reached over the entire range of operation.

15.2 Spark-Injection Engines Combustion in SI engines uses spark plugs for externally supplied ignition. The air-fuel mixture can be prepared in different ways: •• homogeneous mixture preparation by external mixture formation [port fuel injection (PFI)] •• homogeneous mixture preparation using fuel injected directly into the combustion chamber during the intake phase [DI spark ignition (DISI) (homogeneous)] •• stratified mixture preparation using fuel directly injected into the combustion chamber toward the end of compression [DISI (stratified)]. In homogeneous mixture preparation, the output is set by adjusting the charge (quantity control). In stratified mixture formation, the output is set by varying the excess air/fuel ratio (quality control), which enables throttle-free load control. In the following, we first discuss combustion by homogeneous

mixture formation in PFI engines. The particular features of DISI engines are treated in Chapter 15.2.1.

15.2.1 Combustion Processes in Port Fuel Injection Engines 15.2.1.1 Combustion of HCs Normally, fuels for SI engines consist of a mixture of approximately 200 different HCs (alkanes, alkenes, alcohols, and aromates). In PFI engines, a largely homogeneous air-fuel ratio occurs at the end of compression that is ignited shortly before TDC by an electrical ignition spark. There must be an ignitable mixture near the ignition spark. An air-fuel ratio of approximately 0.8 ≤ λ ≤ 1.2 must exist near the ignition spark. To enable chemical reactions and, hence, combustion in the air-fuel mixture, the reaction partners must possess activation energy that is produced by the ignition spark. This required ignition energy is 30–150 mJ per combustion. The ignition spark generates the local temperatures of 3000–6000 K. To ensure reliable ignition, an ignition voltage at the spark plug of 15–25 kV is required with a spark duration of 0.3–1 ms (depending on the environmental state and charge movement). For flame reliable propagation, the energy released from combustion must be greater than the heat transported to the evaporating fuel and the walls delimiting the combustion chamber. The heat is released by the combustion of HCs with oxygen according to the following overall reaction equation: CxHy + (x + y/4) O2 → x CO2 + y/2 H2O Because the probability is slight of all the required reaction partners occurring simultaneously, the HCs are oxidized by many elementary reactions [15-28], in which alkanes rise in a first reaction phase via the dehydration of HC peroxide, and the alkanes, in turn, form aldehydes by reacting with H, O, or OH radicals. The formation of the aldehydes requires approximately 10% of the entire released energy, and is accompanied by cold flames. In the subsequent blue flame, CO, H2, and H2O (30% of the saved energy) are formed. In the following hot flame, CO2 and H2O rise, and 60% of the energy saved in the fuel is released. 15.2.1.2 Cylinder Pressure Characteristic, Internal Efficiency, and Flame Propagation The energy released during combustion leads to a rise in the temperature and pressure of the cylinder charge in the combustion chamber that is detected only in the cylinder pressure characteristic analysis after a delay following the start of ignition (Figure 15.23). This is influenced by the local heating to ignition temperature of the mixture directly next to the spark plug, and is approximately 1 ms independent of the rpm. The combustion time can be calculated with the aid of physical models for the conversion of energy and heat dissipation from the combustion chamber [15-29]. These produce the combustion function that indicates the ratio of combusted to utilized fuel mass as a function of the crankshaft angle. This allows us to evaluate the state and length of combustion, as well as its thermodynamic effect. For homogeneous mixtures,

Internal Combustion Engine Handbook | 623

6606_Book.indb 623

1/19/16 8:46 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

the center of gravity of combustion is most efficient at 8° crank angle after TDC, and the effective combustion period is 30–50° crankshaft angle depending on the working point and combustion process. We are limited in our ability to portray the detailed effects of changes in the engine parameters (such as the variation of the control time and the changes in charge movement) using the previously mentioned simulation models. To accomplish this, the differential pressure characteristic analysis (DPA) method [15-30] is suitable for evaluating the indicated work based on the measured pressure characteristic at each degree crank angle [Figure 15.23 (bottom)]. With the aid of the used fuel mass and a few simple calculations, the advantages and disadvantages of different engine configurations can be compared and optimized. The flame front proceeding from the spark plug is thin and widens during normal combustion at a rate of approximately 20–25 m/s. Shortening the combustion period increases efficiency by approaching isochoric energy conversion and can be attained by the following: •• fast flame front speed from greater charge movement (swirl, tumble, or squish)

•• shorter flame paths from a compact combustion chamber design with a central spark plug or several ignition sites •• higher charge density from a higher compression ratio. In dual ignition, the cylinder charge burns faster because of the shorter combustion paths, and the flame tends to reach the combustion chamber walls. This reduces the tendency of the flame to become extinguished before the cylinder wall (flame quenching) and substantially reduces the amount of uncombusted HCs in the exhaust. The fast conversion of energy also reduces cyclic fluctuations in SI engine combustion. 15.2.1.3 Cyclic Fluctuations and the Influence of the Ignition Angle The fluctuations in the cylinder pressure characteristic from power cycle to power cycle (Figure 15.24) are typical for SI engine combustion and arise from the fluctuations in the turbulent velocity field and local charge composition that influence the propagation of the flame front and, hence, energy conversion. Figure 15.25 illustrates the substantial influence of the moment of ignition on the maximum cylinder pressure and efficiency-influencing position in reference to TDC.

Cylinder pressure [bar]

35

n = 2000 1/min pme = 4 bar ZZP = 26 °CA before TDC

30 25 20

Ignition system

15 10 5

Combustion function [%]

0 120 100 80 60 40

Ignition system

50 %

20 0

0.010

0.50

0.008

0.40

0.006

0.30

0.004

0.20

0.002

0.10

0.000

0.00

–0.002

–0.10

–0.004

0

90

180

270

360

450

Crankshaft angle [°CA]

540

630

–0.20 720

hi cumulative [-]

hi incremental [-]

–20

Figure 15.23 Pressure characteristic and pressure characteristic analyses.

624 | Internal Combustion Engine Handbook

6606_Book.indb 624

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

40

Cyclical fluctuations in peak pressure in 10 consecutive cycles

Cylinder pressure [bar]

35 30

n = 2000 min–1 pme = 2 bar

25 20 15 10 5 0

270

300

330

390

OT

420

Crank position [°KW]

450

15.2.1.4 Influence of the Compression Ratio By increasing the compression ratio, we can partially compensate the combustion-inhibiting influence of low cylinder pressure under a partial load that arises when setting output by throttling the intake air. Figure 15.26 shows the increase and the loss in fuel consumption that arises by changing the compression ratio starting from ∑ = 10. 15.2.1.5 Knocking Combustion As the load increases, limits are set on increasing compression and advancing ignition because of the tendency of uncombusted residual mixture from the cylinder charge to self-ignite. Important properties beyond compression and the

480

510

540

Figure 15.24 Cyclical cylinder pressure fluctuations.

moment of ignition are the fuel properties, the temperature of the combustion air, the combustion chamber shape, the component temperatures, and the charge state (composition and flow field). The theory preferred in [15-33] concerning the origin of engine knocking is based on the secondary ignition of the uncombusted mixture. The additional progress of engine knocking is determined by the propagation of these secondary reaction fronts triggered by these autoignition sites. Given extremely fast energy conversion, local changes in pressure can arise that generate pressure oscillations of the cylinder charge around 5–20 kHz, and that can be detected in the cylinder pressure signal detected (Figure 15.27).

Cylinder pressure [bar]

40 35

Variation of ignition timing 20 ° CA before TDC 24 ° CA before TDC 27 ° CA before TDC 29 ° CA before TDC

30 25

n = 2000 min–1 pme = 2 bar

20

Cylinder pressure [bar]

15 20 15 10 5 –30

–30 –20 –10 –20

–10

Crank position [°CA]

0

0

10 20 30 40 50 60 70 80 90

Crank position [°CA]

Figure 15.25 Influence of the ignition angle on the cylinder pressure curve.

Internal Combustion Engine Handbook | 625

6606_Book.indb 625

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

pme = 2 bar

+7 %

be [g/kWH]

l = 1,0

–6 %

6

8

10

12

14

16

Compression ratio e

Figure 15.26 Impact of compression ratio on partial load consumption [15-32].

2

bar

60

1

40

0

20

–1

p

80

bar

0 –40

–20

ZOT

20

a

40

60

°CA

80

p filtered (5 kHz high-pass)

n = 2000 min–1

The high-frequency oscillations die asymptotically and last up to 60° crankshaft angle. The pressure waves occurring in knocking combustion excite the cylinder charge to form characteristic resonance oscillations that can be calculated using the general wave equation applied to a hollow cylinder and using the Bessel function [15-34]. Figure 15.28 shows a typical calculated resonance oscillation pattern in a hollow cylinder. The resonance oscillations are a function of the cylinder diameter. Figure 15.29 illustrates their influence on the frequency of the most important oscillation modes. The spontaneous propagation of the reaction fronts is frequently inhomogeneous because of the sequential, apparently random ignition of neighboring mixture components at shock wave propagation speeds of up to 600 m/s that are, hence, close to the velocity of sound of the final gas, and can trigger thermal explosions that can damage the engine. If the final gas is burned by heat conduction and diffusion processes, many isolated, autoignition sites arise distributed over the end gas, and pressure waves are not generated

–2

Figure 15.27 Cylinder pressure characteristic and filtered cylinder pressure in knocking combustion.

f = 0°

B

m = 2, s = 1 p

0°

f = 360°

A

B

A

p Figure 15.28 Calculate pressure distribution of a resonance oscillation in the cylinder.

626 | Internal Combustion Engine Handbook

6606_Book.indb 626

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

[15-33]. Figure 15.30 shows typical flame propagation during knocking combustion. 25

Mode 5 Mode 4

15

Mode 3 Mode 2

10

Mode 1

5 0

20

a = 900 m/s Mean flame speed wF [m/s]

Frequency [kHz]

20

25

65

70

75 80 85 90 Cylinder diameter [mm]

95

100

192,6°

192,7°

192,7° 192,6° 192,5° 192,4°

192,5° 192,4° 191,9° 192,5°

192,6°

p1

191,8°

p3

A

E

p2

pme = 8,5 bar n = 2400 min–1 aZ = 162 °CA after BDC

Figure 15.30 Flame propagation in knocking combustion (fiber-optic measuring) [15-35].

15.2.1.6 Flame Speed The flame speed of normal combustion is the product of the combustion speed and transport speed of the local fresh gas. The combustion speed is strongly influenced by the local charge composition and increases with the charge turbulence in the combustion chamber. The transport speed is influenced by the piston movement, the squish flows, and the charge movement triggered by the intake process (swirl and tumble). Figure 15.31 shows the average flame speed in relation to the air-fuel ratio. The probability that the reaction partners will contact each other is greatest at λ = 0.8–0.9. The maximum work is accomplished at λ = 0.8–0.9 because of the fast combustion. The flame speed slows greatly with richer or leaner mixtures, and must be corrected by advancing the ignition angle. Lean mixtures lower the energy remaining in the exhaust because of their low thermal capacity and the lower final combustion temperature resulting from the dilution of the charge [15-36].

10

n = 32 s–1 5

105

Figure 15.29 Cylinder pressure resonance frequencies in dependence on the cylinder diameteπr.

192,8° 192,8°

15

0

0,4

0,6

0,8 Air ratio l

1,0

1,2

Figure 15.31 Average flame speed in relation to the air-fuel ratio.

The efficiency of the combustion engine, therefore, increases when the charge is diluted. Because the flame speed and the charge dilution have a contradictory influence on the efficiency of real combustion, the optimum efficiency arises at ë = 1.1–1.3 in the conventional PFI SI engines with a homogeneous mixture distribution. 15.2.1.7 Cylinder Charge Dilution The charge can be diluted with environmental air or recycled exhaust. By increasing the air-fuel ratio or the share of residual exhaust gas, the share of components not participating in the chemical reaction is increased. To more easily compare the influence of EGR and over-stoichiometric charging, these inert components can be summarized as the parameter “inert gas component” or IG [15-37]:

mIG = mN2 + mRG + mO2 ,( l>1) + mH2O,L

(15.1) (15.2)



mIG IG = mB ⋅ LSt

Figure 15.32 shows the influence of charge dilution on the combustion speed and the indicated thermal efficiency in the high-pressure phase, and on the overall process under a partial load [15-38]. The axes were scaled for the share of residual exhaust and excess air factor assuming an “equivalent inert gas component.” Diluting the charge lengthens the ignition phase. The combustion period remains initially constant and efficiency increases. As charge dilution increases, the cylinder pressure necessary for ignition limits the advance of the ignition angle, which makes the reaction phase longer, cyclic fluctuations increase, and efficiency drops.

Internal Combustion Engine Handbook | 627

6606_Book.indb 627

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

240 Converted fuel energy

I-variation xRG = 20 %) EGR variation lV = 1,0)

220

Crank angle [°CA after BDC]

90 % 200 50 % 180 5% 160

140

ZZP

n = 2000 1/min pme = 2 bar Without swirl 120 40 hIHD

hi,HD [%]

35

hI

30

25

0

2

4

6

external 1.00

1.05

8

10

12

1.10

1.15

14

16

18

20

%

EGR 1.20

lV

Given an equivalent IG component, the ignition phase is longer when exhaust is added than when air is added; when exhaust is added, the advance of the ignition angle is restricted, and the efficiency tends to worsen. When exhaust gas is recirculated, the oxygen partial pressure drops, which slows flame propagation. With the recirculation of external exhaust gas, the indicated thermal efficiency η i does not fall as much as the indicated high-pressure efficiency ρ i, HD while dilution increases. The reason for this is the thermal unthrottling resulting from the

1.25

1.30

Figure 15.32 Combustion progression and efficiency in relation to the charge dilution [15-38].

high intake air temperature in the intake manifold, which reduces charge cycle loss. Despite the reduction in the charge cycle loss, combustion mixtures diluted with exhaust gas do not attain the internal efficiency of air-diluted mixtures, because they enable greater dilution of the charge. EGR is used to lower fuel consumption in λ = 1 approaches, because this allows three-way catalytic converters to be kept for the treatment of exhaust gas. In addition to external EGR, the EGR rate can be internally controlled by means of variable valve timing. Figure 15.33 illustrates how, by varying the

628 | Internal Combustion Engine Handbook

6606_Book.indb 628

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

exhaust camshaft position by a 40° crankshaft angle, the valve overlap increases when ignition is retarded, and the efficiency changes. Differential pressure characteristic analysis shows that, given an equivalent working point, compression work increases as the valve overlap increases. The reason for this is that the cylinder charge increases as the EGR increases. During expansion, a higher share of residual exhaust gas produces a longer combustion period and lower cylinder peak pressures. However, because of the superior charge characteristics and later EO, there is an improvement in efficiency. The unthrottling in the intake phase that occurs as the EGR

Valve stroke [mm]

14 12

Aö: 230 °CA before LCDC

10

Aö: 190 °CA before LCDC

8

increases reduces the charge cycle losses and, accordingly, further increases the efficiency of variants with a greater valve overlap by an overall 4%. In addition to efficiency, the level of untreated emissions is an important factor in evaluating combustion processes. Figure 15.34 shows the fuel consumption/emission trade-off of 4V engines with intake valve shutoff when the charge dilution is varied at a stationary working point. In contrast to the starting point (λ = 1, no EGR), increasing the EGR to 17.5% lowers fuel consumption by 4% and HC + NOx emissions by 50%. The alternative type of charge dilution using a

40°

Inlet

Exhaust

6 4 2 0 35

Cylinder pressure [bar]

30 25 20 15 10 5 0 0.50 0.40

ηi cumulative [-]

0.30 0.20 0.10 0.00

–0.10

Difference ηi cumulative [%]

–0.20 8 4

Aö 190° – Aö 230°

+4 %

0 –4 –8 –12 –16

0

90

180

270

360

450

Crankshaft angle [°CA]

540

630

720

Figure 15.33 Internal EGR by varying the discharge control time. ZZP = consumption-optimized, pressure progression analysis.

Internal Combustion Engine Handbook | 629

6606_Book.indb 629

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

lean mixture allows a maximum excess air factor of λ = 1.4. In contrast to the basic variant, this lowers fuel consumption by 9% and HC-NOx emissions by 40%.

a rotating vortex whose axis is parallel to the cylinder axis. Swirling flows remain during intake and compression and dissipate only during expansion. Turbulence ducts generate whirling in the cylinder whose axis is perpendicular to the cylinder axis. Whirling arises from a one-sided inflow through the intake valve as a result of the burbling in the flow in the intake duct. Tumble flows generally last up to the time of compression and decay into microturbulences close to ignition top dead center.

15.2.1.8 Charge Movement Lean adjustment is primarily improved by increasing the charge movement. This can be achieved by giving the intake duct a special shape for the inflowing fresh cylinder charge. Swirl ducts or an intake duct shutoff in 4V engines generate

220 90 %

Crankshaft angle [degree afterH]

200

50 %

180

10 % 160

Converted fuel energy

140 ZZP 120

0

4

8 12 External EGR [%]

16

20

16

20

16

20

0.4

spmi [bar]

0.3

-

1 EV, valve shut-down (hv,sek = 0.66 mm)

Tumble -

1 EV, two-valve operation (hv = 8.40 mm)

Swirl

0.2 0.1 0.0

0

420

8

EGR [%]

12

n = 2000 1/min pme = 2 bar, l = 1,0 ZZP = be optimal

410

be [g/kWh]

4

400 390 380 370

0

4

8 12 External EGR [%]

Figure 15.34 Consumption/ emission trade-off in charge dilution variation [15-38].

630 | Internal Combustion Engine Handbook

6606_Book.indb 630

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

is also faster in the swirl approach than the tumble approach. The faster energy conversion during swirling means that less preignition is required, thereby resulting in more favorable ignition conditions at the moment of ignition. Cyclic fluctuations (⌠ pmi) are much lower as the EGR rate increases in the swirl variant. The improved combustion stability and the short combustion period give the swirl variant its advantage in fuel consumption.

Figure 15.35 shows an example of engine behavior when there is external EGR of swirl and tumble flows in a 4V engine. The charge swirl was created by shutting off an intake valve. In contrast to the tumble approach, the swirl variation has a much shorter ignition lag in this type of engine. Because of the large charge movement, the flame core can more quickly reach a larger mixture area after the start of ignition and induce a detectable energy conversion. The combustion phase 220

90 %

Crankshaft angle [degree afterH]

200

50 %

180

10 % 160

Converted fuel energy

140 ZZP 120

0

4

8 12 External EGR [%]

16

20

16

20

16

20

0.4

spmi [bar]

0.3

-

1 EV, valve shut-down (hv,sek = 0.66 mm)

Tumble -

1 EV, two-valve operation (hv = 8.40 mm)

Swirl

0.2 0.1 0.0

0

420

8

EGR [%]

12

n = 2000 1/min pme = 2 bar, l = 1,0 ZZP = be optimal

410

be [g/kWh]

4

400 390 380 370

0

4

8 12 External EGR [%]

Figure 15.35 Combustion progression of swirl and tumble flows in a 4V engine [15-38].

Internal Combustion Engine Handbook | 631

6606_Book.indb 631

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

15.2.1.9 Combustion Chamber Shape The combustion chamber shape influences the following properties in SI engines:

These requirements are met by roof-shaped combustion chambers with valves arranged in a V. Because of charging advantages, most current engines are four-valve engines with two intake and two exhaust valves. Figure 15.36 shows an example of a 4V standard combustion chamber. Because of cost advantages, two-valve engines with parallel suspended valves and a camshaft are also used.

•• inflow of the cylinder charge •• charge movement in the cylinder •• speed of energy conversion •• untreated emission level

15.2.1.10 Influence of Load and Loss Analysis The performance of PFI engines is set by throttling the intake air. The reduced density of the inducted fresh charge reduces the cylinder pressure that, in turn, reduces the flame speed. As shown in Figure 15.37, this substantially lowers efficiency in the high-pressure phase under low loads. In addition to losses in the high-pressure phase, the efficiency of the engine under a partial load is worsened by the greater charge cycle losses arising from the throttling of the intake air and loadindependent engine friction.

•• knocking. The combustion chamber design must, therefore, meet the following requirements: •• unhindered inflow to the valve seat •• high flow speeds for the cylinder charge at the ignition TDC •• short flame paths arising from a centralized spark plug position and compact combustion chamber geometry •• minimal dead space (fire land height and valve pockets) •• avoidance of hot components. 50

hv

Efficiency h [%]

∆hi,HD hi,HD ηe 30

∆hi,LW + ∆ hr

Without swirl, EGR = 0 n = 2000 1/min

10

0

A

1

2

3

4

7

8

9

10

Figure 15.37 Efficiency characteristics of PFI engines relative to the load [15-38].

A1–A1

E

A1

5 6 pme [bar]

A1

A

E Figure 15.36 4V combustion chamber of a series SI engine.

632 | Internal Combustion Engine Handbook

6606_Book.indb 632

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

Full-load (pme = 9,5 bar)

No ideal combustion Friction

55,3 %

Exhaust gas energy in ideal process

52,7 %

Partial load (pme = 1,5 bar)

Exhaust gas energy in ideal process No ideal combustion Leak

Leak

5,8 % Charge cycle

Cooling system Useful work

15,2 %

Loss energy

29,5 %

8,9 %

0,9 %

33,0 %

12,5 %

0,9

14,3 %

4,9 %

Loss energy

Useful work

Figure 15.38 shows the energy loss distribution for a PFI engine under partial and full loads. Whereas approximately 30% of the used energy is available as useful work under a full load, and only approximately 15% can be used under a partial load. At more than 50%, most of the used energy is lost as exhaust energy because of the selected process parameters (compression ratio, homogeneous cylinder charge with λ = 1, and combustion chamber shape). Under a partial load, loss components further rise to approximately twice that of full load operation because of engine friction, throttling, and too-slow combustion.

15.2.2 Combustion Process of Direct Injection Spark Ignition Engines

In comparison to a similar SI engine with PFI, we can expect a potential fuel consumption reduction of 5%–20% in partial load by the DISI SI engine, depending on the working point and the operating mode (Figure 15.39).

Intake manifold injection Direct injection Fuel Consumption

Air ratio homogeneous three-ways catalytic converter

lean homogeneous De-NOx-cat

lean stratified De-NOx-cat

Figure 15.39 Fuel savings, partial load operation [15-39].

1,8 %

%

7,1 %

Friction 3,6 % 1,8 %

Charge cycle

Cooling system

Figure 15.38 Loss analysis of a four-cylinder PFI engine (1.6 l displacement) [15-38].

Considering the efficiency chain

hi = hV − ∆hVB − ∆hWW − ∆hLW − ∆hLeck



hV = 1 −



1

e keff −1

the reasons are •• an increase in efficiency of the ideal engine η V because of a higher effective and geometric compression ratio enabled by internal mixture cooling and corresponding knocking behavior and an increase in the isentropic exponent k by designing a DISI combustion process with over-stoichiometric combustion air ratio •• the reduction in the throttle losses ∆η LW •• the minimization of the wall heat loss ∆η WW in stratified combustion processes •• η VB = combustion efficiency and ∆η = leakage losses for the theoretical efficiency of an SI engine with DI. A comparison of PFI and DISI engines is shown in Figure 15.40, where the loss distribution with speed is n = 2000 min–1 and the effective mean pressure is pme = 2 bar. Because of quality regulation that theoretically do not require a throttling of the air flow into the combustion chamber, charge change work is reduced to a third in a DISI engine with stratified combustion, compared with a PFI engine. The comparatively low temperature difference between the combustion chamber wall and the process media close to the wall significantly lowers the thermal losses in a stratified DISI combustion system—up to 60% less heat is injected into the coolant. The higher exhaust gas mass flow with a slightly higher portion of chemical energy and the increase in the friction work (piston group and high-pressure pump) diminish the previously shown advantages of the stratified DISI concept. Overall, a fuel consumption advantage of approximately 13% is achieved.

Internal Combustion Engine Handbook | 633

6606_Book.indb 633

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

Power distribution / kW

Radiation + convection Exhaust gas enthalpy flow Chemical losses (HC + CO) Cooling water Charge cycle Friction Effective output

engine with DI with one featuring PFI. It must be noted that this is the same engine that was operated once as an engine with PFI and once as a DI configuration. The slightly increased level of all emission values shown for the SI engine with DI compared with the PFI is quite obvious. The standard emission values such as HCs, CO, and NOxs, but the indicators for soot and/or particulate emissions that are expressed by the filter smoke number and the particle number concentration, emphasize the difficulty of achieving a mixture preparation that is as good as in PFI concepts. Because the slightly higher potential of standard emission values does not represent an insoluble problem because of the use of three-way catalytic converters, the fuel savings potential is the focus of attention. Only the problem of increased soot and/or particulate emissions must be still resolved. As described above, the potential of reducing fuel consumption is the essential motivation in the development of a DISI engine. This potential is frequently reduced in the overall vehicle by requirements, such as •• exhaust gas treatment under current emission standards

Figure 15.40 Loss distribution. n = 2000 min . pme = 2 bar [15-40]. –1

The advantageous fuel consumption characteristics of SI engines with DI are achieved, however, by a somewhat poorer emission characteristic. In stratified but homogeneous operation, there is less time for mixture preparation which, as a rule, favors the generation of pollutants. Furthermore, the design of the fuel spray and the combustion chamber wall interaction also represents a major challenge for the developers, because any wall wetting also causes an increase in pollutant levels. Figure 15.41 compares the emission characteristics of an SI

•• thermal behavior •• diagnosis •• system costs •• long-term stability. In evaluating a combustion process, however, these parameters should be considered, because they may directly influence the combustion process. As an example, the NOx emission behavior in lean operation in combination with NOx storage catalytic converters is shown. Because of the accumulation of untreated NOx emissions in the catalytic converter storage

0.90

1400

3400

0.80

1300

0.70 0.60

3.60·1013

0.4

3.24·1013

3300

0.4

2.88·1013

1200

3200

0.3

2.52·1013

1100

3100

0.3

3000

0.3

2900

0.2

1000

Homogeneous operation n = 3000 min-1 pmi = 6 bar

Smoke number in FSN

0.5

NOx in ppm

0.50

3500

2.16·1013 1.80·1013

0.40

900

0.30

800

2800

0.2

1.08·1013

0.20

700

2700

0.1

7.20·1012

0.10

600

2600

0.1

3.60·1012

0.00

500

2500

0.0

2.00·109

CO

HC

NOx

Smoke number

Particle number

1.44·1013

Particle number in 1/cm3

1500

HC in ppm

CO in Vol %

Left bar: Intake manifold injection - Right bar: DISI engine 1.00

Figure 15.41 Comparison of the raw emission profile of an SI engine with jet-directed DI (homogeneous operation) and PFI.

634 | Internal Combustion Engine Handbook

6606_Book.indb 634

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

system, a regeneration process is necessary. This occurs in under-stoichiometric engine operation. The fuel consumption behavior is influenced by the frequency of this process. A reduction of the NOx emissions can, hence, indirectly influence the fuel consumption. As a means of lowering untreated NOx emissions, EGR is frequently used. This can occur both externally and internally [15-41]. Another task must be accomplished in a DI approach to ensure the long-term stability of the vehicle propulsion system that the customer takes as “a given.” Because no fuel film is introduced into the intake manifold, deposits that can arise in exhaust recirculation systems or from crankcase ventilation cannot be decreased from corresponding fuel additives. The goals of combustion process development are guided by several parameters, some of which have been discussed above. Some examples of essential developmental tasks are [15-32] •• generating charge movement in air-directed and wall-directed systems and for stratified operation •• ensuring a mixture preparation that is comparable to concepts with intake manifold injection •• ensuring robust injection and spray formation, especially in jet-directed systems, for homogeneous and stratified combustion processes

well as the throttling loss, are comparable to the values of SI engines with PFI. This type of operation is used in the fullload range and for lean operation in the partial load range. The advantages are the knocking behavior under a full load from higher internal cooling of the directly introduced fuel and the exhaust gas treatment at a stoichiometric air-fuel ratio from the use of a conventional catalytic converter system [15-43]. The advantage of reduced knock tendency becomes apparent in DISI concepts with supercharging, in particular, and is, thus, the current focus for many developers, especially for small displacement SI engines. 15.2.2.1.1 Operating mode strategy between homogeneous and lean operation The time required for mixture formation in stratified operation is not available at higher speeds; in this instance, homogeneous operation is used with its early injection timing. Another limitation at higher loads is mixture formation with expanded zones of over-rich ignition. The operation strategy in Figure 15.42 also includes lean operation with a homogeneous mixture, in which the advantages of lean operation can be exploited, without having to accept a reduced mixture formation time.

•• meeting the demands of exhaust gas treatment in stratified and lean operation •• representation of a suitable exhaust gas return, possibly as a stratified EGR •• ensuring the stability of the combustion process •• providing the necessary control unit function •• ensuring sufficient cold-start capacity •• avoiding throughput-reducing and spray image-distorting deposits at the injection valves

Mean pressure / bar

•• limiting the wetting of components with fuel homogeneous, lambda=1 and < 1 homogeneous lean or lambda=1+ EGR Layered with EGR

Engine speed / 1/min

•• using a robust ignition system. 15.2.2.1 Operating Modes of an SI Engine With Direct Injection In contrast to systems using ported fuel injection, several types of engine operation are possible with DI that we identify and qualitatively evaluate in the following. •• Stratified operation An essential parameter of mixture preparation is injection timing. In particular, controlling the injection during the compression stroke prevents the mixture from being completely mixed in the combustion chamber before the moment of ignition; stratification of the fresh charge results. This makes the mixture leaner than the introduced mixture mass. In certain methods, high air-fuel ratios (λ = 6) can be measured on the test bench [15-42]. •• Homogeneous operation Injection during the intake stroke is a characteristic of homogeneous operation. The fuel consumption and leaning, as

Figure 15.42 Operating mode strategy of a wall-directed combustion process.

To lower the greater untreated NOx emissions in comparison to the PFI approach and maintain emission thresholds with the exhaust treatment system during over-stoichiometric engine operation, most DISI engines must operate with high exhaust recirculation rates that additionally promote fuel evaporation. 15.2.2.1.2 Comparison of homogeneous and stratified operation In addition to measuring the fuel consumption and pollutant emissions, analyzing the cylinder pressure characteristic is an important tool for comparing homogeneous and chargestratified operation. In the following, we, therefore, consider cylinder pressure characteristics at a speed of n = 2000 rpm and an effective mean pressure of pme = 2.0 bar, and the development of the internal efficiency under the same load. Figure 15.43 shows the cylinder pressure characteristics of both types of operation that are typical for partial load operation.

Internal Combustion Engine Handbook | 635

6606_Book.indb 635

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

High-pressure loop

Charge change loop Homogeneous operation Stratified operation

Cylinder volume

Cylinder pressure

Homogeneous operation Stratified operation

TDC

Cylinder volume

BDC

TDC

In stratified charge operation, the compression and the combustion peak pressure are higher, because substantially more air mass must be compressed because of the unthrottled operation. In addition, the charge cycle work is much lower in stratified operation. Figure 15.45 shows the progress of the indicated thermal efficiency η i. Starting at the charge cycle TDC upon induction, more work has to be expended in homogeneous operation up to the beginning of the compression stroke than in stratified

Cylinder volume

BDC

Figure 15.43 Cylinder pressure characteristics of the homogeneous and the stratified charge operation of the DISI engine. n = 2000 min–1. pme = 2 bar.

operation (negative internal efficiency) because of throttling. Only upon the beginning of the compression stroke does the piston experience force in the direction of TDC in homogeneous operation because of the greater difference in pressure between the cylinder and the crankcase. The gradient in the indicated thermal efficiency curve is, hence, initially positive. Only after approximately one-half of the piston stroke is the cylinder pressure greater than the environmental pressure as can be seen in Figure 15.44. We now have force acting

positive

approx. 11%

negative

Internal efficiency progression

Homogeneous operation Stratified operation Difference stratified - homogeneous

Intake

Compression

Work

Crank angle / °CA

Exhaust Figure 15.44 Thermal efficiency characteristics of the homogeneous and the stratified charge operation of the DISI engine. n = 2000 min–1. pme = 2 bar.

636 | Internal Combustion Engine Handbook

6606_Book.indb 636

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

against the piston stroke. The gradient of the indicated thermal efficiency curve again becomes negative. Toward the end of the compression stroke, the greater air mass is compressed in stratified operation, and more work is, therefore, expended. During the power cycle, the curve for stratified operation has the greater positive gradient and attains greater indicated efficiency than homogeneous operation toward the end of the power cycle. The portrayed advantage is partially explainable by the lower wall heat loss and the superior charge properties in stratified operation. However, this positive difference in efficiency cannot be completely maintained up to the end of the emission cycle, because in contrast to homogeneous operation, a greater exhaust mass must be expelled. The indicated efficiency curve of stratified operation, therefore, has the greater negative gradient. 15.2.2.1.3 Expansion of the stratification-capable operating range Because of its significant fuel savings potential in over-stoichiometric engine operation, DI is considered to be the concept with the largest individual savings potential. This potential is limited by the unrestricted map range but the still-too-high level of untreated NOx emissions. Another enormous fuel reduction potential can be developed without significantly increasing costs, if, in future, it is possible to implement suitable measures to further lower the level of untreated NOx emissions by, for example, an improved EGR compatibility of the combustion process or an efficient aftertreatment system, and to expand the map range of the over-stoichiometric operating range with an optimal stratification of the fuel-air mixture and ambient air. Clearly, the injection system assumes a key part here, for which reason, it will be further discussed in this chapter. As mentioned, modern stratification-operated DISI concepts are limited upward in speed and in load. On the one hand, there is not sufficient time for mixture preparation; on the other hand, not enough fresh air reaches the cylinder to stratifiedly prepare the large volume of fuel at higher loads. To push the limitation in speed upward, it may be helpful to significantly raise the injection pressure. The limitation in load can be expanded only in the context of increasing the fresh air volume driven into the cylinder. For this purpose, such a stratification concept may be combined with, for example, the supercharging process. However, it must be said that the supercharging system assumes here a key part as it is supposed to work to a high efficiency but represent a very high spread of air flow to be able to deliver large air volumes in partial load

jet-directed

wall-directed

without limiting the rated power by, for example, blocking the supercharging system. The requirements listed will rather exclude standard exhaust gas turbocharger concepts that are available today. Overall, it is a vision for the future that the SI engine can be operated unthrottled over its entire operating map; this would allow significant fuel savings. Technology modules such as supercharging, injection, and ignition assume key parts in this respect, and will be the focus of future developments. 15.2.2.2 Characteristics and Specifics of the SI Engine Combustion Process With Direct Injection and Their Technology Modules and Combinations 15.2.2.2.1 Description of the different combustion process concepts Most of the first-generation DI systems now used in series production, which usually feature a lateral injector position are no longer designed for stratified operation. This fact can be traced to the limited fuel savings potential for the customer and to the presently available second-generation combustion system concepts, which feature a jet-directed mixture preparation. However, it applies to both generations that stratified operation, that is, the DI of fuel into the compression stroke imposes high demands on the combustion process in respect to mixture formation, mixture transportation, ignition, and conversion: •• The mixture must be formed relatively quickly; liquid fuel or zones with over-rich mixture need to be attenuated by the time of ignition. •• The mixture must be transported to the spark plug in a controlled manner that can be reproduced cycle for cycle. •• A clear stratification of the ignitable mixture should be discernible with the goal of lower wall thermal losses using exhaust or air. •• Zones with nonflammable lean mixture and rich zones must be avoided, in particular, in the vicinity of the spark plug. These requirements not only must be met for a map area as large as possible, but also must be complied with relative to the operating point to create a sufficient stability window. The stratified combustion methods presently being developed and used in series can be classified into three combustion approaches, and are characterized as follows (Figure 15.45): •• Air-directed combustion method The fuel is transported by a charge movement generated from the site of introduction to the spark plug. The combustion chamber walls are not wet when the method is properly

air-directed

Figure 15.45 Schematic combustion process concepts [15-29].

Internal Combustion Engine Handbook | 637

6606_Book.indb 637

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

applied. The precise timing of injection and a stable charge movement are decisive for the quality of the method. The quality of the mixture formation that is supported by the charge movement is high in the corresponding designs. However, the stability of this combustion process depends mostly on the reproducibility of the charge movement from cycle to cycle and from keeping the tolerance of the airguiding components. •• Wall-directed method The fuel is guided by a correspondingly shaped combustion chamber wall (the piston in this instance) to the site of ignition. This method is associated with greater fuel deposits on the combustion chamber walls; evaporation before ignition cannot usually eliminate the entire fuel film. Because this method is based on uniform conditions, the process is stable. The higher untreated emissions and the comparatively low potential fuel consumption prevent this method from being widely used. Literature, however, indicates some inadequacies of both the combustion process types and their hybrid forms in respect to the representation of the theoretically expected fuel consumption benefits and further disadvantages shown below [15-44]: •• An insufficient stability in stratified operation at low loads limits the maximum possible leaning and may require throttling. •• At higher loads, the stratification-capable map area is limited because of increasing disadvantages in mixture formation with the subsequent increased CO and soot emissions. •• Combustion chamber and piston wetting and flame extinguishing will lead to increased HC emissions in stratified operation. •• Because of the necessary use of a NOx catalytic converter storage system in lean operation, the permissible NOx mass flows are limited, relative to the exhaust gas cycle in stratified and in homogeneous lean operation. Therefore, during normal driving operation, the theoretical fuel consumption potential is reduced, compared with the PFI engine, because of the restricted stratification-capable map area, the regeneration of the NOx catalytic converter storage, and the enriching required for component protection. •• Jet-directed method This combustion method is differentiated into a near and remote position, depending on the position of the injector and ignition system. Although the remote position has the advantage of a longer mixture preparation time, the spray is more exposed to any charge movements. When comparing the processes, the transport of the fuel into the direct vicinity of the ignition site theoretically has the highest potential for stratification; values of the air-fuel ratio between one and eight are anticipated. However, the corresponding advantage in fuel consumption characteristics is contradicted by major challenges described below with possible approaches for solutions:

•• The formation of an ignitable mixture is performed only in a small partial area of the fuel jet/spray, and little time is provided for mixture formation. One approach would be the use of a high-pressure injection pump with an injection pressure of ≥200 bar. •• The use of a charge movement to support the mixture preparation is a problem as there is the risk of the mixture blowing away from the ignition site. The previously held opinion was that the combination of a jet-directed DI with strong charge movement would create problems. More recent research, however, has shown that jet-directed DI returns very good results in stratified operation, even in combination with high tumble levels. Special attention must be paid, however, to the design of the combustion chamber geometry, including the position of injector and spark plug. Overall, it can be determined, in such a combination, that the stable working area becomes smaller in stratified operation, although a very high efficiency and stability of the process can be determined in the drivable map area. Furthermore, this combination allows a higher residual gas and EGR tolerance, resulting in a further reduction of the NOx emissions. •• The spark plug must be enclosed by a flammable air-fuel mixture for a sufficient ignition window, depending on the operating point and cyclical fluctuations. A stable position of the spray is achieved using injectors opening to the exterior, which will not narrow when the backpressure rises. •• Because of wetting with fuel, the spark plug is exposed to a significant alternating thermal load. In addition to material optimization, the use of modern ignition systems, such as laser ignition, for example, may be a solution [15-45]. In respect to the exhaust gas aftertreatment, four issues are important in a jet-directed system when compared with a wall-directed or air-directed process: •• In stratified operation with high excess air in the low-load range, the occurring exhaust gas temperatures may be so low that the NOx catalytic converter storage and the threeway catalytic converter will lose most of their capabilities. •• Expanding the stratified operation to higher loads will usually cause an increase in NOx mass flow and corresponding higher consumption because of more frequent nitrogen depletion in the catalytic converter storage. •• Today, the use of selective catalytic reduction (SCR) systems is being discussed more for, in particular, establishing the promising stratified operation in jet-directed combustion processes. But even this system for NOx exhaust gas treatment requires a specific temperature window to ensure full functioning. This window, however, is placed much more favorably as the one for a lean NOx trap catalytic converter. Depending on the used reduction agent, the SCR catalytic converter achieves a 90% conversion of NOx in a temperature window between approximately 150°C to approximately 500°C (see Figure 15.46).

638 | Internal Combustion Engine Handbook

6606_Book.indb 638

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

Catalytic converter temp

•• By eliminating a global charge movement, the residual gas tolerance is significantly reduced, causing an increase in the NOx mass flow. The stratification of residual gas is considered as a solution, in addition to the combination of stratification-operated combustion process with charge movement [15-46]. The quantities of injection timing, moment of ignition, and combustion period are different in stratified operation without exhaust recirculation in comparison to the process shown in Figure 15.47.

Operating points 2000 1/min, p mi = 2,8 bar jet-directed wall-directed air-directed Injection

Ignition system

Combustion

Crank angle [°CA after BDC]

Figure 15.47 Process curve of various DISI combustion methods in the operating point. n = 2000 L/min. pmi = 2.8 bar [15-47].

The differing times between the end of injection and the moment of ignition are noteworthy. The time is the shortest for the jet-directed method, because local ignition starts after the end of injection. The time of mixture formation is contrastingly longest for the wall-directed method, because the mixture is guided a relatively long way across the piston surface. In the subsequent ignition phase (time between the actual ignition and a 5% reaction rate), a relatively poor mixture treatment of the jet-directed process with long lag time can be seen, which indicates a required improvement of the

Figure 15.46 Temperature windows of different NOx exhaust gas treatment systems. See color section page 1087.

mixture formation and, therefore, the injection system. The good mixture formation of the two other methods yields a similar combustion lag despite different modes of operation (length of mixture formation in the wall-directed method and intense charge movement in the air-directed method). During the subsequent conversion, the time up to an 85% reaction rate is observed. This observation considers the incomplete conversion of the wall-directed process. The unprepared fuel on the piston surface restricts the maximum conversion to 88%. The air-directed method has the shortest combustion period because of the favorable mixture quality and the intensity of the charge movement. This relatively fast conversion continues until the end of combustion. The two other methods have a longer combustion time because of the poor local mixture quality and the lesser charge movement. An advantage of the wall-directed method is the relatively late and efficient location of combustion. 15.2.2.2.2 Advantages and disadvantages of the different concepts (concept benchmark) Each of the combustion methods shown has advantages and disadvantages, and a detailed evaluation of the potentials is shown in Figure 15.48 (bottom). The quantitative data of this evaluation do not meet quantitative requirements but should represent quite realistic trends. Overall, the jet-directed combustion process has the most potential for consumption improvements, but only when it is combined with stratified and/or lean operation. If DI is realized as a homogeneous concept and in combination with supercharging in particular, the lateral injector position will be able to offer sufficient potential. Which combustion process design will ultimately be implemented in a particular development must be decided from case to case and depends on other factors, in addition to the purely technical assessment. These factors include market penetration (sufficient fuel availability), combination with other technological modules, costs, and so on.

Internal Combustion Engine Handbook | 639

6606_Book.indb 639

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

Performance Combustion process

Quantitative evaluation

Fuel Emissionen Fuel conCosts consumption sumption HC NOx Particles

Homogeneous, stoichiometric and lean, lateral installation, possibly with charging Stratification, wall-directed, lateral installation, swirl valve Stratification, jet-directed, central installation, multi-hole valve Stratification, jet-directed, central installation, multi-hole valve

Application effort

Particles

+

+

O

-

+

-2%

+50%

O

++

-

-

---

++

-8%

+400%

O/-

+++

-

-

--

O

-13%

+200%

-

+++

+

--

-

--

-15%

+50%

--/---

+++ very good (low effort) / O neutral (neutral effort) / --- Poor (high effort) Comparison basis is a homogeneously working SI engine with intake manifold injection

same compression. This means an increase in the efficiency of approximately 22% compared with an engine with intake manifold injection. As shown, homogeneous operation also has a potential to increase efficiency by some percentage when compared with an otherwise identical SI engine with intake manifold injection. This advantage is based, as described, on the higher geometrical compression, which can be converted at higher loads because of the effects of internal cooling in intake-synchronous injection and, therefore, less knocking (see Figure 15.50). Figure 15.50 describes the advantage achieved from DI to increase compression. If, starting with an engine with intake manifold injection and a compression of 10.5, an engine is operated with identical compression and DI, the degree of delivery increases and knocking decreases. Both phenomena can be traced to the charge cooling because of direct transport of the fuel during the induction stroke. In the course of the subsequent polytropic compression, the charge cooling

Thermal efficiency

15.2.2.2.3 Characteristics of direct injection—combination with other technology modules As explained above, DI carries a non-insignificant potential to improve fuel consumption in SI engines. The most potential is provided by stratified or lean operation, as can be seen from the efficiency of the ideal engine in Figure 15.49. When looking at the theoretical efficiency of air intake and a compression of 11, a stoichiometric mixture of λ = 1 will result in a thermal efficiency of 47%, representatively for an SI engine with intake manifold injection. By considering the fact that the compression of an SI engine with DI leads to a reduced knock tendency because of the internal cooling of the fuel vaporizing in the cylinder, and the compression can be raised by approximately 1–2 to 1.5–12.5 units in our example—resulting in a thermal efficiency of approximately 50%. If such a concept is now combined with lean operation, a thermal efficiency of approximately 58% can be attained at realistically assumed air ratios of, for example, λ = 4 at the

Figure 15.48 Evaluation of different combustion process according to specific advantages and disadvantages [15-48].

Compression ratio

Figure 15.49 Influence of efficiency of the ideal engine with constant volume combustion [15-48].

640 | Internal Combustion Engine Handbook

6606_Book.indb 640

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

Output / torque

Rising the compression ratio

Increase if degree of delivery and charge cooling

Direct injection ε = 10,5

Improvement 2-5%

Direct injection ε = 12

ε = 10,5 Intake manifold injection

Knocking limit

Figure 15.50 Influence of DI on performance and knocking limit [15-48].

Preignition in °CA before TDC

downsizing and down-speeding engine concepts could not be realized to this extent.

Base: PFI 2.0lt aspirator 1470 kg: 70Km/h = 2000 rpm

% ptio

sum

con n

Ring gap nozzle (A nozzle)

1350 1/min (de-rate factor for aspirators: 0.68)

-15

disproportionally affects the temperature progression in the cylinder, significantly reducing knocking; this can be seen in the illustration at the larger preignition. The reduction of knocking can now be used to increase efficiency by raising compression so much that a similar knocking behavior is attained at the same center of gravity of combustion, as can be seen in Figure 15.49. Preignition is again reduced. However, the efficiency benefit of the DISI combustion process is retained over the entire map area because of the raised compression. In addition, the efficiency benefit is retained in the entire operating range, which means an additional performance yield of approximately 2–5. Based on the advantages presented, there are two additional characteristics of DISI combustion processes that have been commercially successful: the very fuel-efficient overstoichiometric stratification operation and the stoichiometric homogeneous operation in combination with supercharging. Because of the sensitive and still being developed NOx aftertreatment systems that are mandatory to reduce the potentially higher untreated NOx emissions, in particular, the NOx mass flow in lean operation, and the trend to smaller displacement engine concepts using the downsizing approach, we now see an obvious trend to supercharged DISI engines with stoichiometric operation. A combination of supercharging and DI with their specific benefits, even in stoichiometric homogeneous operation, leads to a further significant rise in fuel efficiency compared with naturally aspirated engines with DISI homogeneous operation. The frequently realized so-called downsizing, that is, a reduction in the displacement at equal performance of an engine with larger displacement, enables, depending on the degree of downsizing and downspeeding, a reduction in the fuel consumption of approximately 15%–20%, compared with the engine with larger displacement (Figure 15.51). Without a combination with DI, the above-mentioned reductions in fuel consumption could not be achieved, because, without the benefits of the increased degree of delivery, the potential raise of the geometric compression and the use of variabilities in the injection and charge change system, the

Engine speed [1/min] Figure 15.51 Downsizing and downspeeding (reducing displacement and speed) [15-49].

As a fundamental cornerstone of future technology combinations, DI will further establish its importance in the marketplace. After all, it permits the achievement of significantly higher fuel efficiencies for many other technologies, such as homogeneous autoignition, as a module in hybrid concepts because of direct start options. 15.2.2.2.4 Particle emissions—a challenge for direct injection DI leads in diesel and SI engines to a reduction in the mixture formation time, so that mixture formation and combustions will, in part, progress at the same time. As a result, the emissions of unburnt HCs and CO will increase. In addition, the particle output (soot) must not yet be considered as a relevant pollutant in SI engine combustion [15-50]. For the soon-to-be implemented EURO 5 and EURO 6 exhaust regulations, limit values for particle mass and particle numbers have been introduced or are under discussion. For EURO 5, a limit value for particle mass of 4.5 mg/km has been

Internal Combustion Engine Handbook | 641

6606_Book.indb 641

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

decided. Furthermore, beginning with EURO 6, a limitation of the particle number is under discussion; the limit value considered is 6 × 1011/km. According to current development, the particle mass can be contained while the limitation of particle numbers presents major problems to all SI engines with DI. Mastering the discussed limit value will only be possible with very high development efforts. In particular, measures to reduce untreated emissions must be implemented. These are optimizations at the injection system and the combustion chamber geometry, as well as application-related measures such as ignition point, injection timing, ignition timing, fuel temperatures, and intake temperature. Increasing the injection pressure can be a measure to reduce the particle number. The fundamental influence of this measure is discussed below as this parameter has a decisive effect on all DISI concepts. 15.2.2.2.5 Effects of the injection system on direct injection—layout and pressure level 15.2.2.2.5.1 Layout Up to now, three fundamentally different injectors or nozzle variants were generally used. Figure 15.52 shows the difference in the spray pattern of these nozzle layouts. Because of its simplicity, the swirl nozzle was used in the early concepts with DI and stratified operation. The swirl generator in the nozzle provides atomization of fuel. The fuel axially enters the nozzle, and is tangentially diverted in the swirl generation. This decreases the kinetic energy of the fuel droplets in axial direction and reduces the penetration depth. A disadvantage of this nozzle is the narrowing of the fuel spray relative to the back-pressure. In the case of a compression stroke injection, different spray cone angles occur at any changed injection timing, including the pressure level in the cylinder. For this reason, this nozzle is not suited for jet-directed DI with stratification operation, because the reproducibility of the spray directly affects the process stability. In homogeneous operation, that is the intake manifold injection, this behavior is not decisive; for this reason, this concept is still used in homogeneous operation, because this nozzle represents an interesting solution from a cost point of view. A nozzle layout variant that is, by now, significantly more often discussed, is the multihole nozzle. Because of its spray characteristics, it can be used for DI concepts with either a lateral or a central injector position. As the name indicates, the multihole nozzle features not just one hole in the valve

Swirl nozzle

Multi-hole nozzle

seat, as does the swirl nozzle, but several holes. The most common number of holes is between five and eight. More or fewer holes are also possible. The multihole nozzle allows very flexible jet patterns, because the arrangement and number of the individual injection jets can be adapted for the combustion chamber and the requirements of the combustion process. For example, asymmetric spray contours can be realized to avoid a wetting of the inlet valve or the spark plugs. Because of its design, the multihole valve has, however, a disadvantage that will repeatedly cause problems over the service life of such a nozzle layout. The dead volume present in this nozzle causes vaporization of the residual fuel after any injection process and encourages the coking of the nozzle tip. With a corresponding design of the nozzle hole geometry and application of the nozzle in the cylinder head, this adverse behavior can be mostly controlled. The third nozzle shown in Figure 15.52 opens to the outside by releasing a ring-shaped gap, hence the name ring gap nozzle. A thin fuel film leaves this gap at high velocity. Because of the specific layout, that is the geometry of needle and valve seat, a uniform hollow cone spray is generated without prespray and a very high atomization quality, represented by Sauter diameters (SMDs) of less than 15 μ m. The spray cone angle remains highly stable regardless of the combustion chamber counter-pressure, making this nozzle type suitable for use in jet-directed combustion processes, in particular. Because of the very large released surface in this nozzle type, large quantities of fuel exit rapidly with the result that this nozzle attains the by far the best vaporization rate, that is, fuel volume per time unit expressed over the degree of the crankshaft angle (see Figure 15.53). Compared with swirl and multihole nozzles, a two- to three-times higher vaporization rate can be achieved. This type of nozzle may be electromagnetically driven or actuated by a piezoactuator. A piezocontrolled nozzle needle has the major advantage of a high valve stroke reproducibility which, on the one hand, results in a high reproducibility of the injected fuel quantity (<2% quantity deviation). On the other hand, extremely short switching times (<0.2 ms) can be realized, which is very suitable for the realization of multiple injections in particular. Because this nozzle tip is subject to significantly higher temperature fluctuations because of its design when compared with the nozzle tip of swirl and multihole injectors, the A nozzle has proven to be much less sensitive to coking. All nozzles mentioned are now operated at injection pressures of maximum 200 bar. However, various continuing

Ring gap nozzle (A nozzle)

Figure 15.52 Different nozzle designs of injectors, used in SI engines with DI.

642 | Internal Combustion Engine Handbook

6606_Book.indb 642

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

Swirl nozzle

Mean vaporization rate in mg/°KW

Vaporization rate in mg/°KW

5

3

Multi-hole nozzle A nozzle

2

1

0

4 3 2

Dralldüse Mehrlochdüse

1

A-Düse

0

0

5

10

15

20

25

0

Figure 15.53 Vaporization rates for different injection nozzle concepts.

research allows the assumption that raising the injection pressure beyond the currently common values of 200 bar will have very positive effects in respect to mixture formation and, ultimately, engine performance. 15.2.2.2.5.2 Influence of a raised injection pressure If raising the injection pressure is discussed, the question arises as to whether this measure will be equally profitable for every injector type. As a rule, the vaporization rate of an A nozzle already attains very high values (Figure 15.53), while an increase in the vaporization rate seems to be quite desirable for swirl and multihole injectors. This can be achieved when the injection pressure is significantly increased. Figure 15.54 demonstrates precisely this issue, that is, the characteristic of the mean vaporization rate relative to the injection pressure for the previously discussed nozzle types.

200

400

600

800

1000

Injection pressure in bar

Angle in °CA after ESB

Figure 15.54 Extrapolated mean vaporization rate of a swirl nozzle and a multihole nozzle relative to injection pressure between 0 and 1000 bar, compared with an A nozzle up to 200 bar.

If a vaporization behavior of an A nozzle with 200 bar is to be achieved in a swirl or multihole nozzle, the injection pressure must be increased to approximately 800 bar. Because this is a linear relationship, any further increase of the injection pressure will also increase the vaporization rate. Although the vaporization rate of nozzles opening to the exterior is already very high, a pressure increase will also increase their vaporization rate. However, a significant increase in the injection pressure in an A nozzle would cause problems for closing and keeping closed the nozzle as these actions would be executed against the injection pressure applied. In addition to the rate increase, it would have to be determined how the spray of a multihole nozzle develops in different injection pressures between 200 and 1000 bar. This behavior is shown in Figure 15.55 (bottom). Optical images of the fuel

3 separate areas of vaporized fuel in

Striation technique

1 large area at 1000 bar

Figure 15.55 Pressure chamber results of a multihole nozzle at different injection pressures under constant chamber conditions—constant injection mass for all pressures. See color section page 1087.

Internal Combustion Engine Handbook | 643

6606_Book.indb 643

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

spray during injection into a pressure chamber at constant marginal conditions are shown for all the injection pressures. When injecting the same fuel mass, the injection at 1000 bar is completed first. This moment also represents the time of the captured images for all the nozzle types. The surface colors of these striation images make it possible to differentiate between the liquid (black) and vaporized (gray) fuel portion at every injection. The larger portion of fuel in vapor state at 1000 bar injection pressure compared with 200 bar can be clearly seen. Furthermore, a significantly larger surface (orange background) of the vapor fuel portion is captured, which means a significantly increased air acquisition and, therefore, improved mixture preparation. The images also confirm that the penetration depth of the liquid fuel portion is not increased when injection pressures are raised but reduced, a fact that is interesting in respect to a reduction of wall wetting in particular. The increase in the injection pressure can, therefore, represent a measure to combine SI engines with DI at central injector positions and very small bores, which is, as a rule, considered to be critical because of the risk of wall wetting. A further indication of the positive effects of an increase in pressure on the mixture formation characteristics is the analysis of the mean droplet size in a fuel spray at different injection pressures relative to the counter-pressure (see Figure 15.56) [15-51].

Average droplet size in -meter

14

Injection pressure 200 bar Injection pressure 300 bar Injection pressure 500 bar

13 12 11 10 9 8 7 6

2

4

6

8 10 12 14 Chamber pressure in bar

16

18

Figure 15.56 Characteristics of the fuel droplet size versus the counter-pressure in the injection chamber and for different injection pressures [15-51].

Because of the decreasing droplet size at increasing injection pressure, it can be assumed that a positive effect on the general mixture formation behavior will be achieved. The smaller droplets overall cause an enlargement of the surface of the injected fuel and, thus, allow faster absorption of heat, which will subsequently result in a faster vaporization of the fuel. Figure 15.57 shows the effects of raising the injection pressure on the engine operation at the example of an operating point in stratification operation. The illustration shows different engine-relevant values at different injection pressures for the operating point pmi = 3 bar and n = 2000 min–1. The injection pressure variation

starts at 200 bar and ends at 1000 bar; in between, all the relevant parameters are shown for pressures of 500 and 800 bar, equivalent to the pressure chamber results shown in Figure 15.55. The above discussed improved mixture preparation or homogenization at an increased injection pressure can be clearly seen at a 6% reduction in fuel consumption. As well, the obvious drop of all the emissions shown emphasizes the improved mixture preparation and the resulting more efficient combustion. The CO emissions that again increase between 500 and 800 bar are indicative of locally overly enriched areas that can have different causes, such as wall wetting, because of a nozzle layout poorly adjusted for the combustion chamber. Because these values are below the initial values at 200 bar, while the advantage in fuel consumption and/or the very major advantage of a nearly 40% reduction in NOx emissions is fully maintained, the significant benefit of an increase of injection pressure represents interesting technology for the future, even for the SI engine. Because of the obvious NOx advantage that can be traced back to improved homogenization and fewer locally leaned and overly-rich areas, the increase in the injection pressure can also be a central approach for future stratification-operated lean concepts. The benefit because of improved mixture preparation for stratification operation can be, as a rule, transferred to homogeneous operation. The more homogeneous the basic mixture, the more efficient the fuel conversion. For highly charged engine concepts, in particular, this presents a major advantage because the fewer cyclical fluctuations occur because of insufficient homogenization of the fuel-air mixture, and the less the cylinder charge tends to incur knocking at high loads. Consequently, the ignition timing must be adjusted less toward delayed to prevent knocking, relative to the efficiency-optimized point, which likewise brings a significant efficiency advantage. In this respect, the increase in the injection pressures represents a very interesting approach even for homogeneous and, in particular, highly charged engine concepts. In addition to the obvious benefits in respect to mixture preparation, the injection pressure increase has, because of the considerably-reduced injection times, much potential relative to multi-injection in homogeneous but in stratified lean operation. 15.2.2.2.6 Operating modes for SI engines with direct injection Multi-injection is understood as the distribution of the injection event to several points in time. This process can be used for the optimization of several engine operating states in an SI engine with DI in homogeneous and in stratified operation. In any case, it is intended to contribute to the improvement of mixture formation and/or engine running characteristics, because, as a rule, the considerably shortened time for mixture formation is a major challenge in DISI concepts, compared with SI engines with intake manifold injection. 15.2.2.2.6.1 Multi-injection in stratified engine operation To represent in the entire engine map, the optimal settings for fuel supply relative to emission behavior and the stability

644 | Internal Combustion Engine Handbook

6606_Book.indb 644

1/19/16 8:47 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

250 240 230 220 210

0.30 0.25 0.20 0.15 0.10 0.05 0.00

H Ci [g/k W h]

20

σpmi [bar ]

b i [g/k W h]

260

18 16 14 12 10

0.25 0.15 0.10

FSN [-]

0.20

0.05 0.00

26 22 18 14 10

15 13 11 9 7 200

400

600

800

Injection pressure [bar]

of the combustion process, multi-injection is indispensable for jet-directed combustion processes in particular. It enables the representation of very compact mixture clouds and wider transition areas from a rich injection jet up to the ambient air within the cylinder. In addition, it is possible to keep the area of stoichiometric air-fuel ratio more stable at the spark plug and to enlarge it at the same, which is very important in stratified combustion processes in particular. The demand on the combustion process and, therefore, the injection strategy becomes even more stringent in a combination of stratification and supercharging, that is, an enlargement of the map relative to the motor load and the spread capacity of fuel dosing. Figure 15.58 below emphasizes the advantage of multi-injection relative to mixture formation and combustion in stratified operation. Figure 15.58 (left) shows the fuel droplets and the lambda distribution at the spark plug at different injection strategies.

1000

5

NOxi [g/k W h]

C O [g/k W h]

30

Figure 15.57 Characteristics of engine-specific parameters when varying the injection pressure between 200 and 1000 bar at a stratification operation point pmi = 3 bar and n = 2000 min–1 [15-49].

Directly next to it, one can see the flame propagation, represented by the temperature development near the spark plug. The analysis of both the image sequences recognizes a clear advantage of multi-injection compared with single-injection. The result of these images is confirmed by an analysis of the inflammation shown in the bottom right of the illustration. Multi-injection not only ensures a faster ignition initiation but also enlarges the ignition window. 15.2.2.2.6.2 Multi-injection in homogeneous engine operation In homogeneous operation, multi-injection is successfully used mostly in combination with supercharging. Thus, doubleinjection can be used, for example, so that the fuel mass is distributed in a homogeneous basic mixture and a stratified portion to reduce the knocking trend under full load (see Figure 15.59).

Internal Combustion Engine Handbook | 645

6606_Book.indb 645

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

Injection process Injection rate [-]

Combustion

Mixture formation

Injection

single

Ignition stable Double injection Triple injection

double

Inflammation characteristics

Variation ZZP

triple

12 °CA

Analysis Fuel conversion

Fuel conversion 8 °CA after ZZP

Air ratio [-] rich

1.0

fast inflammation

6 °CA

slow inflammation

lean

Acceptance time ZZP

Acceptance time 8 °CA after ZZP

Ignition timing [°CA after ZOT]

Figure 15.58 Multi-injection in stratified operation at 2000 min , pmi = 3.0 bar (simulation results) [15-52]. See color section page 1088. –1

1. Injection

2. Injection

AnteilPortion der Zweiteinspritzung of secondary injection an derofGesamtmenge total volume in in %%

-360

TDCFiring

35

360

Einfacheinspritzung: Single injection 33 °CA 33 °KW 26.99 26.36

28.37

30

26.97

28.38

28.04

28.59

27.00

25 30.07

20

29.24

27.81 28.00 28.40

15

28.90

31.98

10

30.09

30.36 33.08

30.30

50 % conversion point °CA after TDCHD 5

80

70

60

50

40

30

20

Time of secondary / °CA /before TDC Zeitpunkt der zweiteninjection Einspritzung °KW vor OTHD HD

10

0

Figure 15.59 Development of the 50% fuel conversion point at a variation of the fuel mass distribution in a homogeneous and stratified component and variation of the injection timing of the stratified injection at constant full load (n = 1500 min–1, pme = 19.5 bar, and ignition permanently at the knocking limit). See color section page 1088.

646 | Internal Combustion Engine Handbook

6606_Book.indb 646

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.2 Spark-Injection Engines

This illustration shows the characteristics of the 50% fuel conversion time in a variation of the injection timing of the secondary injection and the distribution volume of primary and secondary injection when using double-injection in fullload operation. The load was kept constant over the variation, and the ignition timing was tracked to the knocking limit. The positive effect of the double injection on the knocking behavior is clearly visible. While the 50 fuel conversion point is at approximately 33° crankshaft angle after TDC at single injection, it was possible to adjust this point by approximately 6° crankshaft angle to early when using double injection at constant load. The basic mixture that is very lean after intake-synchronous primary injection tends less to knocking events and, after termination of the stratified secondary injection, the time window between injection end and ignition initialization is so shortened that even the full mixture will tend to knocking only under considerably more stringent conditions in the combustion chamber. If the portion of the stratified injection is increased beyond the 30% shown, the load can no longer be kept constant, and the HC and CO emissions will rise severely complete mixture preparation or combustion, which represents the limit of a useful utilization of double-injection in homogeneous full-load operation. Another option is the additional injection of a homogeneous fuel portion in the area of charge change, to improve homogenization [15-53]. This will also cause a reduction in knocking because of the reduction in the compression discharge temperature.

and inflammation, an increase in the exhaust gas temperature, when compared with single injection (Figure 15.60). This will ultimately result in a faster response of the catalytic converter after a cold start. 15.2.2.2.6.4 Stratified start In most current series engines, a DISI engine is started at a relatively low pressure in the fuel rail that is provided by the electric fuel pump. To raise this pressure level in the fuel supply and attain the corresponding mixture formation quality, high-pressure storage chambers in the fuel supply system are considered. It is expected to result in a significantly improved starting and emission behavior. Another option to optimize the starting behavior is the so-called direct start. Almost without the support of a starting aid, such as a starter for example, an engine start can be realized, with a precisely adjusted injection and ignition, with the first rotation of the crankshaft drive [15-54]. These functionalities and the correspondingly improved starting behavior require, however, the use of additional components such as a pressure accumulator and/or crank-angle index mark generators of a higher quality.

Valve stroke [mm]

15.2.2.2.6.3 Catalytic converter heating While in SI engines with intake manifold injection very often secondary air pumps are used for catalytic converter heating, the SI engines with DI provide a new freedom in this discipline, using multi-injections in particular. Upon termination of an initial intake-synchronous injection, a relatively lean basic mixture is present, which is maintained end-to-end and provides, in combination with one or more injections near the time of ignition ensuring reliable ignition

Cylinder pressure [bar]

15.2.2.2.7 Effects of the ignition system on direct injection Direct-injection SI engines pose high demands on the ignition system, in particular, in stratified operation. On the one hand, the spark plug may be, in jet-directed DI for example, subject to very high temperature fluctuations or thermal shock stresses because of wetting with fuel, and tends to the formation of deposits. On the other hand, it must compensate for the presence of inhomogeneities near the spark plug that are typical for DI and to still ensure reliable inflammation. As a rule, a locally very limited area with ignitable air-fuel mixture, which is a globally lean mixture because of stratification, must be around the spark plug in stratified operation. This is one of the focal points in the development of combustion processes with DI; that is, the reliable provision of ignitable mixture in the area of the spark plug’s electrodes, regardless

Cylinder pressure Inlet

Exhaust

BDC

TDC TDCangle Crank

BDC strategy Triple injection coupled to ignition timing Double injection coupled to ignition timing Double injection

Figure 15.60 Injection strategies for heating exhaust gas catalytic converters [15-52].

Internal Combustion Engine Handbook | 647

6606_Book.indb 647

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

of the operation point, that is, the load, speed, injection timing, injection point, and so on. The ignitable mixture must be in the range between approximately λ = 0.7 and 1.3. Within this range, the inflammation and formation of a flame front by the ignition can be assured. If the mixture quality is subject to stronger cyclical fluctuations, inflammation cannot be assured for every cycle, resulting in poor engine running behavior with an increase in emissions and higher fuel consumption. A first step to ensure inflammation despite inhomogeneities and/or cyclical fluctuations of the mixture quality at the spark plug is the increase in the ignition energy of the conventional ignition system, from approximately 40 mJ to values in the 80–120 mJ range. A quiescent homogeneous stoichiometrically composed fuel-air mixture requires approximately 1 mJ ignition energy, while an quiescent but lean or rich mixture requires approximately 3 mJ [15-55]. Because of inhomogeneities, dilution caused by returned exhaust gas and transmission and heat losses in feed lines and electrodes, this requirement will significantly increase. Nonquiescent mixtures also have a higher risk of sparks blowing off. Some of the aforementioned items are clearly more prominent in DI, for example, the diversion of the spark because of the jet impulse in jet-directed combustion processes and the subsequent compression stroke injection, resulting in a considerable increase in the flow velocity at the spark plug and between the plug’s electrodes in particular. In total, a significant increase in the ignition voltage is the result. To ensure reliable inflammation even under more difficult conditions of DI, new ignition systems have been tested in the past for their fitness, such as plasma ignition systems, laser ignition systems, and so on, but none of them was so satisfactory that it would have attained the benefit-cost ratio of a conventional ignition. However, new systems are constantly under development as they may open a further important advantage in respect to emission and consumption. It would be a major benefit for example, if the ignition limit could be expanded in the lean range with the result, for instance, that, in stratification, it would no longer be required to generate the fuel-air core around the ignition site to be as stoichiometric as possible, but that it could be significantly over-stoichiometric. In a clearly over-stoichiometric mixture cloud, the untreated NOx emission level could be significantly lowered, because the NOx formation potential within the stoichiometric mixture cloud could be heavily reduced. If a truly major leap would be possible in this respect, further reductions in fuel consumption could be achieved, in addition to less emissions and increased residual gas tolerances, because it would be possible to expand lean operation and/or stratification beyond their current limits. All in all, a further developed ignition system would mean a clear thrust for DI. As a matter of course, such an ignition system development would be beneficial not only to stratified operation but also to supercharged homogeneously operated DISI combustion processes, because these also require high ignition voltages because of the charging of highly compressed cylinder charges, to ensure fast and reliable inflammation for the mixture not to have too much time for early reactions that may cause knocking.

15.2.2.2.8 Development process and tools 15.2.2.2.8.1 Development process To answer the complex questions arising in the development of concepts with DI, all the necessary tools for each development step are used—from the analysis of familiar combustion process and initial simulations to the presentation of a concept vehicle. There exists a broad range of variations and possibilities available for the short time normally available for the development of SI engines, especially in optimizing the combustion method. To restrict the amount of experimentation to reasonable limits, statistical models or computational fluid dynamics (CFD) calculations must be used (Figure 15.61). Idea

Construction: - Piston for entire engine - Piston for optical engine - CFD calculation model Series engine

Manufacturing test components - Piston for entire engine - Piston for optical engine

Prototype engines

Examinations: - Optical engine, mixture formation - Entire engine: Thermodynamics - CFD: Mixture formation

Figure 15.61 Development tool [15-51].

Optical investigation methods are used to represent and assess processes in the engine interior. For example, Doppler global velocimetry is capable of visualizing the complex 3-D actions of the charge movement stationary and a dragged engine and to use suitable parameters for the assessment in a relatively short time [15-56]. An exemplary overview of a few optical investigative methods is listed in Figure 15.62. To accomplish these investigations quickly, developers must deal with organizational problems in addition to technical demands. Specific tools are, therefore, used in the individual CFD and optical investigations that limit the normally timeconsuming research to essential conclusions and to substantially reduce the required time. The results of these investigations largely serve as a comparison with CFD calculations. The tools used in the subsequent developmental process for harmonizing engine control unit (ECU) functions are also model-supported. To work efficiently, the torque-based

648 | Internal Combustion Engine Handbook

6606_Book.indb 648

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.3 Combustion Process (Two-Stroke Engine)

Dim.

Measuring technology

used for

Video stroboscopy

Injection (liquid phase)

+ Statistics easily possible; + can be combined with endoscopes; - max 1 image per cycle

High-speed film

Injection (liquid phase)

+ recording of total cycle; - image processing, statistics

LIF

Injection (liquid and vapor phase)

+ visible possibly in both cycles - Quantification difficult - max 1 to 2 data per cycle

Fuel concentration

+ quantitative - low signals, prone to faults - max 1 to 2 data per cycle

Flow velocity

+ accurate, statistics easily possible + (recording of total cycle) + established and mature - line information requires much effort

Flow velocity (possibly simultaneous injection)

+ 2D flow information + Statistics easily possible + established and mature - max 1 data per cycle

functional structure of the ECU requires the use of statistical experimental design [15-58]. Most models are not precise enough regarding stochastic behavior such as cyclic fluctuations in the combustion characteristic, for example. The entire engine must, therefore, be represented in the testing of thermodynamic variables when developing combustion processes. The actual fuel consumption potential of SI engines that use DI, therefore, can be realized only by closely linking the development of the combustion process, exhaust gas treatment, and operation strategy with comprehensive models [15-42], [15-43].

15.3 Combustion Process (Two-Stroke Engine) Despite a certain popularity of two-stroke diesel engines in the 1950s and 1960s as small stationary engines, tractor engines (Lanz, Hanomag, F&S, ILO, Stihl, O & K, and Hirth), as well as in commercial vehicles (Krupp and Ford) (see also [15-67] and [15-68]), two-stroke diesel engines are presently not used to power passenger cars or commercial vehicles. The reasons for the irrelevance of the two-stroke engine in this market segment are the increased requirements concerning engine life, lubricating oil consumption, and emissions that cannot be sufficiently met with conventional, simple engines (crankcase scavenging pump, symmetrical timing diagram, and limitation to three moving parts). Other reasons are the limited developmental status of scavenging blowers that are an alternative to the crankcase scavenging pump, as well as problems with cooling, lubrication, and materials. The increasing use of exhaust turbocharging in four-stroke diesel engines also lessens the performance advantage of two-stroke diesel engines. The advantages of the two-stroke diesel engine for lower excitation of drive train oscillations in engines with few cylinders, the torque characteristic, weight-to-power ratio, cold-start behavior, engine heating after a cold start, untreated

Remarks

Figure 15.62 Summary of optical investigative methods [15-57].

NOx emissions, and exhaust gas treatment conditions make the two-stroke diesel engine especially interesting for oneto three-cylinder engines in low-consumption passenger cars. In the 1990s, this led to related development projects by companies including Toyota [15-69], AVL [15-70], Yamaha [15-71], and Daihatsu [15-72]. In two-stroke diesel engines, the choice of the combustion process is strongly influenced by the scavenging approach. In two-stroke diesel engines with uniflow scavenging using intake ports and exhaust valves, it is fairly easy to generate a swirling flow in the cylinders by the design of the scavenging ports and intake ports. The swirl can be influenced as a function of the engine load and speed by placing valves in front of the scavenging ports. For this reason, similar mixture formation conditions can be generated, and comparable combustion processes can be used in comparison with the DI four-stroke diesel engines that are primarily used today. Chapter 7.25 shows a uniflow scavenging three-cylinder two-stroke engine by AVL for use in passenger cars (see also [15-70]). When camactuated injection pumps are used (distributor injection pump and pump nozzle), the injection frequency that is twice that of four-stroke engines must be considered when designing the pump and the cam. In particular, the use of common-rail injection allows the engine to be operated within certain ranges of the program map (such as low load at high speed) with four strokes. Independent of the selected injection system, effective cooling of the nozzle holder or injection nozzle must be provided when designing the cylinder head. In two-stroke engines with loop scavenging (head loop scavenging or piston-controlled loop scavenging, for example, according to Schnürle), a swirling flow does not form in the combustion chamber around TDC but a more- or lesspronounced tumble flow. For this reason, loop-scavenged two-stroke diesel engines in older vehicles used the IDI method almost without exception (prechamber and whirl chamber). In small stationary engines with loop scavenging (F & S and ILO [15-67]), DI was contrastingly used, sometimes with a

Internal Combustion Engine Handbook | 649

6606_Book.indb 649

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

radial arrangement of the nozzle holder assembly. Modern direct fuel injection systems with injection pressures up to 2000 bar and a greater number of injection orifices need only comparatively low-swirling combustion air to provide satisfactory mixture formation, so that a DI combustion method remains a possibility for loop-scavenged two-stroke diesel engines, in certain circumstances with a slightly radial squish flow at TDC. Figure 15.63 shows an example of a diesel engine with loop scavenging with a lengthwise section and cross section of a two-cylinder two-stroke diesel engine with 1-L displacement by Yamaha [15-71] that was intended for use in small cars. The stroke is 93 mm, and the bore is 82 mm. At 33 kW, the rated output is 4000 rpm. The maximum torque of 80 Nm is attained at 2500 rpm. The engine with an overall weight of 95 kg is designed for use in 3-L vehicles, and should meet the future Euro 4 thresholds. The cylinder crankcase made of an aluminum alloy with a Ni–P–SiC-coated cylinder barrel has four transfer passages per cylinder through which the fresh gas from the crankcase reaches the cylinder. The cylinder barrel and the crankshafts with roller bearings or connectingrod bearings are intentionally provided with lubricating oil via map-controlled total loss lubrication to enable minimum consumption of lubricating oil. The outlet of the loop-scavenged cylinder is provided with two superimposed exhaust ports. To improve the torque characteristic, the top exhaust port can be closed by means of a throttle valve, while the ratio in engine operation varies between 13:1 and 18:1. Apparently, given the difficulties in generating a characteristic combustion chamber swirl for DI combustion, a chamber combustion method was used. In this type of combustion process (Figure 15.64), strong swirling is generated via four tangential blow ports after the start of combustion in the chamber with low overthrust losses in the cylinder. According to Yamaha [15-71], this allows complete combustion to occur with low fuel consumption and emissions.

Figure 15.64 Representation of the whirl chamber of the Yamaha 1-L two-stroke diesel engine [15-71].

15.4 Two-Stroke SI Engines In contrast to two-stroke diesel engines, the two-stroke SI engine for passenger cars has enjoyed a long tradition. In particular, the positive experiences in the development and production of two-stroke motorcycle engines formed the basis in the 1920s for the introduction into the market of passenger cars with two-stroke SI engines by the companies DKW, Aero, Jawa, and Ceskoslovensko Zbrojovka. The large demand for reasonably priced vehicles after the mass acquisition of cars following World War II provided the foundation, especially in Germany, for the development and production of many passenger cars with two-stroke SI engines. In addition to DKW, two-stroke passenger cars were produced by Lloyd, Goliath, Gutbrod, Glas, and others. The market share of two-stroke passenger cars by the end of the 1950s in West Germany was approximately 20%. In East Germany up to the cessation of production at the beginning of the 1990s, the brands Wartburg and Trabant achieved a maximum market share for two-stroke passenger cars of over 60%. Against the background of increasing customer and public sensitization to directly perceptible

Figure 15.63 Longitudinal and cross section of the Yamaha 1-L two-stroke diesel engine [15-71].

650 | Internal Combustion Engine Handbook

6606_Book.indb 650

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.4 Two-Stroke SI Engines

HC emissions (blue smoke), rough idling, service life problems, and comparatively high fuel consumption under a full load, DKW in Ingolstadt, Germany, stopped producing passenger car two-stroke engines in 1966 (Saab of Sweden stopped in 1968), practically simultaneous to the cessation of production of passenger cars with two-stroke engines in Wartburg and Sachsenring at the beginning of the 1990s, publications and presentations by Orbital [15-73], [15-74], AVL [15-75], Subaru [15-76], Toyota, GM, and Ficht, and so on (see also [15-77] and [15-78]), reawakened interest in two-stroke SI engines. According to these publications, improving mixture formation (DI) and using alternative scavenging methods could overcome the specific disadvantages of traditional passenger car twostroke engines and produce low emissions and consumption in engines especially for small passenger cars. The essential characteristic of the two-stroke method is that the engine undergoes one complete power cycle per rotation in contrast to the four-stroke method. The combusted charge is removed, and the fresh gas is introduced (scavenging process) into the cylinder at the same time within a crankshaft angle range around BDC. Because the gas volume communicates with the atmosphere via the open exhaust organs, compression always begins after the intake and exhaust organs close (apart from gas-dynamic influences and supercharging and boosting effects—even when the intake air is throttled under a partial load) at a cylinder pressure that approximately corresponds to the atmospheric pressure. In contrast to throttle-controlled four-stroke SI engines, comparatively high final compressions also occur under a partial load. As shown in Section 10.3, there are various scavenging methods available for the charge cycle of two-stroke engines that are associated with the respective advantages and disadvantages. Because of the simple and compact design and the demand for comparatively high nominal speeds, two-stroke SI engines were developed for passenger car use almost exclusively with loop scavenging and a crankcase scavenging pump. In contrast to throttlecontrolled four-stroke SI engines, the charge cycle work falls in the conventional two-stroke engines with a crankcase scavenging pump when approaching a partial load (see also [15-79]). This principle of load control yields a high exhaust component in the cylinder under a partial load because of the “open” gas exchange; during a charge cycle, only the amount of exhaust is expelled from the cylinder that corresponds to the fresh gas entering the cylinder as determined by the degree of intake port throttling. A high exhaust component in the cylinder lowers NOx emissions and improves the physical preparation of the fuel because of the increased temperature under a partial load. On the other hand, the high inert gas component under a partial load and especially during idling drastically worsens ignition conditions. A large amount of residual exhaust gas in connection with a high final compression pressure under partial load gives rise to the demand for an ignition system with high ignition energy. If, under these conditions, the scavenging process does not place an ignitable mixture near the spark plug, misfiring occurs. In the following scavenging process, more of the air-fuel mixture is scavenged in the cylinder, which improves ignition conditions. If ignition

occurs after one or more scavenging processes, the subsequent combustion, as a result of the prior reactions in the mixture during preceding compression cycles, is characterized by high-energy conversion rates, pressure gradients, and peak pressures. This behavior of mixture-purged two-stroke SI engines leads to inferior tractability under a partial load and especially during idling. Furthermore, the expulsion of uncombusted mixture components increases fuel consumption and HC emissions. Because of the influence of gas fluctuations in the intake system, especially in two-stroke engines with a crankcase scavenging pump, the cylinder charge changes with the rpm, and so does the mixture composition, particularly when mixture formation is external (in a carburetor). In addition to the residual exhaust gas, there are other influences on ignition, tractability, and emissions. As the load rises, the increasing fresh gas component in the cylinder produces smoother engine running. Experience shows that the fuel consumption is comparatively good in mixture-scavenged two-stroke engines under an average partial load at average speeds. As a full load is approached, the increasing amount of mixture scavenged in the cylinder, depending on the gasdynamic design of the intake and exhaust systems causes a more-or-less marked increase in the loss of fresh gas and, hence, an increase in fuel consumption and HC emissions. According to the present state of technology, a prerequisite for maintaining strict current and future exhaust pollutant thresholds in passenger car four-stroke SI engines (at least in some areas of the program map) is the oxidation of uncombusted HCs and CO, as well as the simultaneous reduction of NOxs in three-way catalytic converters with a stoichiometric air-fuel ratio (λ = 1 control). The basic condition for the three-way catalytic converter to function in a two-stroke SI engine is for the uncombusted mixture leaving the cylinder during the charge cycle to have the same (stoichiometric) air-fuel ratio as the fresh charge. This state is obtainable in theory in two-stroke SI engines with external mixture formation. However, the described misfiring under a partial load and the related serious time-related changes in the exhaust composition produce regulatory difficulties in maintaining a narrow λ “window.” In the internal mixture formation process (DI), the cylinders are scavenged with air. Depending on the quality of the scavenging method, the mean pressure (load), and the amount of scavenging air that may be used to cool the cylinder, the oxygen directly introduced into the exhaust must be compensated by enriching the mixture remaining in the cylinder (increase in the injected quantity) for operation at λ = 1. A rise in the oxidizable exhaust components (HC and CO) from the enrichment of the mixture in the cylinder is undesirable from the perspective of fuel consumption, the thermal load on the catalytic converter, and the limited reduction of pollutants from the conversion rates of the catalytic converter. In contrast to the use in passenger cars, the low weight, the small required area, the mechanical robustness, and the lowmaintenance operation of two-stroke SI engines have, at least, partially secured their dominant position as outboard engines, jet ski and snowmobile engines, small two-wheeler engines,

Internal Combustion Engine Handbook | 651

6606_Book.indb 651

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

and power tool engines. The technical and environmental demands that are increasing in this market segment have also led to the development and introduction to the market of many technical improvements, some of which have drastically reduced fuel consumption and/or pollutant emissions. This includes, among others, the scavenging supply (fresh gas or air supply) in small stationary engines, the introduction of oxidation catalytic converters in connection with the optimization of loop scavenging, and a lean mixture adjustment in mopeds and scooters, the use of secondary air systems in small two-wheelers, and the mass production of electronic DI for outboard engines and two-wheelers [15-82]. Above all, the strict pollution thresholds in the most important markets for passenger cars and the high demands on comfort and service life as presented in [15-83] that have not been sufficiently met in at least some new approaches represent a serious hurdle for the use of two-stroke SI engines in passenger cars. To successfully market two-stroke SI engines in automobiles, the following central design features must be considered along with the related developmental tasks (see also [15-84]): •• Use or optimization of a scavenging method that offers a reliably combustible mixture to the spark plug with minimal fresh gas losses, even under a partial load. •• Transition from external mixture formation (carburetor/PFI) to a direct fuel injection system that ensures good mixture preparation and the presentation of a reliably combustible mixture to the spark plug even at low, partial-load operating points within the short time available for mixture formation. Air-supported direct fuel injection systems offer good mixture formation within the short time available, but they need to be optimized for the high system cost and high power input.

•• Based on the technological knowledge of exhaust treatment in DI four-stroke SI engines, exhaust treatment systems need to be adapted to the requirements for two-stroke SI engines to allow fulfillment of strict future pollutant thresholds, even when not operating the engine at a stoichiometric air-fuel ratio. •• An interesting approach is a combined two-cycle and fourcycle operation in vehicle SI engines, which is, principally, applicable for diesel engines as well. Corresponding to the results of the fundamental research by Ricardo [15-85], [15-86], the two-cycle operation (scavenging concept reversed-head scavenging) in the high load and low engine speeds map points would enable major downsizing combined with considerable fuel consumption savings. The development focus for these concepts is mostly directed toward the realization of series-capable charging concepts, scavenging concepts, working mode change strategies, and switching concepts for variable valve actuation times. Figure 15.65 provides an exemplary view of a loop-scavenged three-cylinder two-stroke engine by Orbital. The engine has a stroke of 72 mm and a bore of 84 mm. At 58 kW, the rated output is 4500 rpm. The maximum torque of 130 Nm is attained at 3500 rpm. The overall weight of the engine is 85 kg. According to statements in [15-87], Euro 3 thresholds are maintained after running continuously for 80,000 km with a sufficient safety margin. The engine was, and is, intended for use in Indonesian passenger cars for the brands Maleo and Texmako.

•• Use and optimization of ignition devices and ignition methods that reliably, stably, and consistently ignite mixtures that are difficult to ignite under a partial load under real vehicle operating conditions. •• Develop and use scavenging and supercharging blowers that permit freely selectable cylinder scavenging and supercharging over the entire program map of the engine with minimal power input. Electrically supported turbochargers, perhaps with variable turbine geometry, offer the option of using a part of the otherwise unused exhaust energy and simultaneously compensating for the disadvantages of symmetrical timing diagrams (loop scavenging) because of the collection of exhaust in front of the turbine. •• Not using the crankcase as a scavenging pump makes it possible to use crankshafts mounted on a plain bearing with better service life, cost, and acoustics, and effectively cools the strongly heated piston, perhaps with a cooling channel, by means of a force-fed lubrication system and oil injection nozzles. The cylinder/piston stroke combination and the piston ring assembly are essential factors in optimizing the minimum required lubricating oil and maximum oil scraping effect of the piston rings and sufficient mechanical and thermal strength of these valve train components.

Figure 15.65 Sectional view of the 1.2-L three-cylinder two-stroke engine by Orbital [15-84]. See color section page 1089.

The water-cooled cylinder crankcase made of an aluminum alloy has several overflow passages per cylinder, and is divided in the crankshaft midplane. The forged crankshaft consists of a single piece and has divided roller bearings at the pins

652 | Internal Combustion Engine Handbook

6606_Book.indb 652

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.4 Two-Stroke SI Engines

for the main and connecting rod bearings. When the piston moves upward, intake air is sucked into the respective crankcase through the intake manifold. Reed valves in front of the crankcase prevent a return flow of gas into the induction tract during compression in the crankcase. The crankshaft bearing and the cylinders are supplied with fresh oil by an electronically controlled lubricating oil pump. The fuel-oil mixture ratio is normally between 1:50 and 1:200. To attain a high torque over the entire speed range, there is a barrel controller in the exhaust ducts near the exhaust ports that can change the exhaust timing. The barrel controller is adjusted via a direct current motor. A particular feature of the engine is air-supported gasoline DI (see also [15-87] and [15-88]). The main element of this injection system is an electromagnetically controlled valve in the cylinder head for injecting an air-fuel emulsion into the combustion chamber. The liquid fuel is precisely metered by an injection valve of a conventional ported fuel injection system and injected into a mixture chamber. By injecting air compressed in a reciprocating piston supercharger into this chamber, an air-fuel emulsion forms that is finely atomized and blown into the combustion chamber. According to [15-87], an average SMD of less than 8 μ m is attained. A good mixture quality is attained (for example, at 3000 rpm at the end of injection around 25–30° crank angle before TDC in the stratified air-fuel mixture) that permits a lean adjustment of the air-fuel ratio up to 100:1, given stable combustion under a partial load.

Bibliography

15-1. Renner, G. and Maly, R. R. 1998. “Moderne Verbrennungsdiagnostik für die dieselmotorische Verbrennung.” in Dieselmotorentechnik 98, edited by Essers, U. (Hrsg.). Expert-Verlag: Renningen-Malmsheim.

15-11. Armbruster, F.-J. 1991. “Einfluss der Kammergeometrie auf den Energiehaushalt und die Prozesssimulation bei Kammerdieselmotoren.” Fortschrittberichte VDI, Reihe 12: Verkehrstechnik/Fahrzeugtechnik, Nr. 149. VDI Verlag: Düsseldorf. 15-12. Fortnagel, M., Moser, P., and Pütz, W. 1993. “Die neuen Vierventilmotoren von Mercedes-Benz.” MTZ 54, 9: pp. 392–405. 15-13. Sun, D. 1993. “Untersuchung der Strömungsverhältnisse in einer Dieselmotor-Wirbelkammer mit Hilfe der Laser-Doppler-Anemometrie.” Disseration. Universität Stuttgart. 15-14. List, H. and Cartellieri, W. P. 1999. “Dieseltechnik—Grundlagen, Stand der Technik und Ausblick.” MTZ Sonderausgabe “10 Jahre TDI-Motor von Audi. 15-15. Spindler, S. 1992. “Beitrag zur Realisierung schadstoffoptimierter Brennverfahren an schnelllaufenden Hochleistungsdieselmotoren.” Fortschritt-Berichte VDI, Reihe 6: Energieerzeugung, Nr. 274. VDI Verlag: Düsseldorf. 15-16. Dietrich, W. R. 1999. “Die Gemischbildung bei Gas- und Dieselmotoren sowie ihr Einfluss auf die Schadstoffemissionen—Rückblick und Ausblick–Teil 1.” MTZ 60, 1: pp. 28–38; “Teil 2.” MTZ 60, 2: pp. 126–134. 15-17. Thiemann, W., Dietz, M., and Finkbeiner, H. 2000. “Schwerpunkte bei der Entwicklung des Smartdieselmotors.” in Dieselmotorentechnik 2000, edited by Bargende, M. and Essers, U. (Hrsg.). Expert-Verlag: Renningen-Malmsheim. 15-18. Kirsten, K. 1986. “Vergleich unterschiedlicher Brennverfahren für kleine schnelllaufende Dieselmotoren.” Dissertation. RWTH Aachen. 15-19. Willand, J., Vent, G., and Wirbeleit, F. 2000. “Können innermotorische Maßnahmen die aufwendige Abgasnachbehandlung ersetzen?” 21. Internationales Wiener Motorensymposium, Wien Mai; Fortschrittberichte VDI, Reihe 12: Verkehrstechnik/ Fahrzeugtechnik Nr. 420; Band 1: erster Tag S. 33; VDI-Verlag: Düsseldorf. 15-20. Chmela, F., Piock, W. F., and Sams, Th. 2003. “Potenzial alternativer Verbrennungsverfahren für Otto- und Dieselmotoren.” in VKM-THD Mitteilungen, H. 83; 9. Tagung “Der Arbeitsprozess des Verbrennungsmotors” Graz, September 2003. pp. 45–59. 15-21. Grote, K.-H. and Feldhusen, J. (Hrsg.) 2007. Dubbel—Taschenbuch für den Maschinenbau. 22. Aufl. Springer: Heidelberg.

15-2. Schünemann, E. et al. 1999. “Analyse der dieselmotorischen Gemischbildung und Verbrennung mittels mehrdimensionaler Lasermesstechniken.” in IV. Tagung Motorische Verbrennung (Essen, März 1999). Haus der Technik: Essen.

15-22. Bürgler, L. et al. 2004. Brennverfahren,Abgasnachbehandlung, Regelung—Kernelemente der motorischen HSDI Diesel Emissionsentwicklung. 13. Aachener Kolloquium Fahrzeug- und Motorentechnik 2004, Band 2. VKA, ika, RWTH Aachen, VDI (Hrsg.).

15-3. Fath, A., Fettes, C., and Leipertz, A. 1999. “Modellierung des Strahlzerfalls bei der Hochdruckeinspritzung.” in IV. Tagung Motorische Verbrennung (Essen, März 1999). Haus der Technik: Essen.

15-23. Coma, G. et al. 2004. “HCCI Verbrennung: Traum oder Realität?” 13. Aachener Kolloquium Fahrzeug- und Motorentechnik, Band 1. VKA, ika, RWTH Aachen, VDI (Hrsg.).

15-4. Mollenhauer, K. and Tschöke, H. (Hrsg.). 2007. Handbuch Dieselmotoren, 3rd Edition. Springer: Berlin.

15-24. Kahrstedt, J. et al. 2004. “Homogenes Dieselbrennverfahren für EURO 5 und TIER2/LEV2—Realisierung der modifizierten Prozessführung durch innovative Hardware und Steuerungskonzepte.” 25. Int. Wiener Motorensymposium, Band 2. Fortschritt-Berichte VDI, Reihe 12, Nr. 566.

15-5. Adomeit, Ph. and Lang, O. 2000. “CFD Simulation of Diesel Injection and Combustion.” SIA Congress “What challenges for the Diesel engine of the year 2000 and beyond” (Lyon, May 2000). SIA: Suresnes, France. 15-6. Urlaub, A. 1995. Verbrennungsmotoren: Grundlagen, Verfahrenstheorie, Konstruktion, 2nd Edition. Springer: Berlin. 15-7. Dietrich, W. R. and Grundmann, W. 1993. “Das Dieselkonzept von DEUTZ MWM, ein schadstoffminimiertes, dieselmotorisches Verbrennungsverfahren.” Fortschritt-Berichte VDI, Reihe 6: Energieerzeugung, Nr. 282. VDI Verlag: Düsseldorf. 15-8. Heinrichs, H.-J. 1986. “Untersuchungen zur Strahlausbreitung und Gemischbildung bei kleinen direkteinspritzenden Dieselmotoren.” Dissertation. RWTH Aachen. 15-9. Pischinger, F., Schulte, H., and Jansen, J. 1988. “Grundlagen und Entwicklungslinien der dieselmotorischen Brennverfahren.” in VDI Berichte Nr. 714 (Tagung: Die Zukunft des Dieselmotors, Wolfsburg Nov. 1988). VDI Verlag: Düsseldorf. 15-10. Meurer, J. S. 1988. “Das erstaunliche Entwicklungspotenzial des Dieselmotors.” in VDI Berichte Nr. 714; (Tagung: Die Zukunft des Dieselmotors, Wolfsburg Nov. 1988). VDI Verlag: Düsseldorf.

15-25. Kahrstedt, J., Manns, J., and Sommer. A. 2010. “IAV Berlin: Brennverfahrensseitige Ansatzpunkte für Pkw-Dieselmotoren zur Erfüllung künftiger EU- und US-Abgasstandarts.” 10. Internationales Stuttgarter Symposium. 15-26. Tomoda, T. et al. 2010. “Verbesserung der Dieselverbrennung bei ultra-niedriger Verdichtung.” 19. Aachener Kolloquium. pp. 4–6. 15-27. Wenzel, W. et al. 2010. “Ersatz von Sensoren im Luft- und Abgaspfad von Verbrennungsmotoren unter Verwendung des Zylinderdrucksignals einer Druckmessglühkerze.” 9. Internationales Symposium für Verbrennungsdiagnostik, AVL. pp. 8 and 9. 15-28. Merker, G. P. and Stiesch, G. 1999. Technische Verbrennung, Motorische Verbrennung. Teubner Verlag Stuttgart: Leipzig. 15-29. Pischinger, S. 1998. “Verbrennungsmotoren, Vorlesungsumdruck.” RWTH Aachen, 19. Auflage. 15-30. Wurms, R. 1998. “Differenzierte Druckverlaufs-Analyse—eine einfache, aber höchst wirkungsvolle Methode zur Interpretation von Zylinderdruckverläufen.” 3. Internationales Indiziersymposium 22./23.04.1998.

Internal Combustion Engine Handbook | 653

6606_Book.indb 653

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 15 Combustion Systems

15-31. Niefer, H. et al. “Verbrennung, Ladungswechsel und Abgasreinigung der neuen Mercedes-Benz V-Motoren mit Dreiventiltechnik und Doppelzündung,” MTZ 58: pp. 392–399. 15-32. Pischinger, F. and Wolters, P. 2000. Ottomotoren—Teil 2, in Vieweg Handbuch Kraftfahrzeugtechnik, edited by Braess H.-H. and Seifert, U.

15-54. Kulzer, A. et al. 1808. “Einige Aspekte bezüglich Gemischbildung und Verbrennung im Rahmen des Direktstarts von Ottomotoren mit Benzin-Direkteinspritzung, Kraftstoffe und Antriebe der Zukunft.” VDI-Bericht. 15-55. 1987. Autoelektrik, Autoelektronik am Ottomotor. VDI Verlag.

15-33. Stiebels, B. 1997. “Flammenausbreitung bei klopfender Verbrennung, Forschr.-Ber.” VDI-Reihe 12 Nr. 311. VDI Verlag: Düsseldorf.

15-56. Dingel, O. et al. 2003. “Dreidimensionale Messung der Ladungsbewegung mit Doppler Global Velocimetry.” Motortechnische Zeitschrift.

15-34. Adolph, N. 1983. “Messung des Klopfens an Ottomotoren.” Dissertation RWTH Aachen.

15-57. Stiebels, B., Krebs, R., and Zillmer, M. 2001. “Werkzeuge für die Entwicklung des FSI-Motors von Volkswagen.” Leipertz, A.: Motorische Verbrennung, Tagung HdTMärz.

15-35. Kollmeier, H.-P. 1987. “Untersuchungen über die Flammenausbreitung bei klopfender Verbrennung.” Dissertation RWTH Aachen. 15-36. Pischinger, R. et al. 1989. Thermodynamik der Verbrennungsmaschine. Die Verbrennungskraftmaschine. Neue Folge, Band 4. Springer Verlag: Wien, NY. p. 99. 15-37. Südhaus, N. 1988. “Möglichkeiten und Grenzen der Inertgassteuerung für Ottomotoren mit variablen Ventilsteuerzeiten.” Dissertation RWTH Aachen. 15-38. Fischer, M. 1988. “Die Zukunft des Ottomotors als PkW-Antrieb— Entwicklungschancen unter Verbrauchsaspekten.” Dissertation TU Berlin, Schriftenreihe B—Fahzeugtechnik—des Institut für Straßen- und Schienenverkehr. 15-39. Eichlseder, H. et al. 1988. Beurteilungskriterien für ottomotorische DI-Verbrennungskonzepte. Spicher, U. u. A.: Direkteinspritzung im Ottomotor. Expert Verlag. 15-40. Krebs, R. and Theobald, J. 2001. “Die Thermodynamik der FSI-Motoren von Volkswagen.” 22. Int. Wiener Motorensymposium, 26./27.04.2001. 15-41. Dahle, U., Brandt, S., and Velji, A. 1988. Abgasnachbehandlungskonzepte für magerbetriebene Ottomotoren. Spicher, U. u. A.: Direkteinspritzung im Ottomotor. Expert Verlag. 15-42. Karl et al. 1996. “Thermodynamische Analyse eines DI-Ottomotors.” 17. Internationales Wiener Motorensymposium. 15-43. Zhang, H. et al. “Doppeleinspritzung am Otto-DI-Motor: Anwendungsmöglichkeiten und deren Potenzial.” 22. Int. Wiener Motorensymposium, 26./27.04.2001. 15-44. Fröhlich, K., Borgmann, K., and Liebl, J. 2003. “Potenziale zukünftiger Verbrauchstechnologien.” 24. Internationales Wiener Motorensymposium. 15-45. Geringer, B. et al. 2004. “Laserzündung. Ein neuer Weg für den Ottomotor.” MTZ. 15-46. Peters, H. 2004. “Experimentelle und numerische Untersuchung zur Abgasrückführung beim Ottomotor mit Direkteinspritzung und strahlgeführtem Brennverfahren.” Dissertation Universität Karlsruhe. 15-47. Grigo, M. 1999. “Gemischbildungsstrategien und Potenzial direkteinspritzender Ottomotoren im Schichtbetrieb.” Dissertation RWTH Aachen. 15-48. van Basshuysen, R. (Hrsg.). 2008. Ottomotor mit Direkteinspritzung, 2. Auflage. Vieweg + Teubner Verlag. 15-49. Prevedel, K. and Piock, W. F. 2004. “AufladungbeimDirekteinspritz-Otto-motor.” VDI-Tagung Innovative Fahrzeugantriebe, Dresden. 15-50. Thiemann, J. 2000. “Laserdiagnostische Untersuchungen der Rußbildung in einem direkteinspritzenden Ottomotor.” Dissertation RWTH Aachen. 15-51. Nauwerck, A. 2006. “Untersuchung der Gemischbildung in Ottomotoren mit Direkteinspritzung bei strahlgeführtem Brennverfahren.” Dissertation Universität Karlsruhe (TH), LOGOS Verlag.

15-58. Fischer, M. and Röpke, K. 2000. “Effiziente Applikation von Motorsteuerungsfunktionen für Ottomotoren.” MTZ: p. 562. 15-59. Altenschmidt, F. et al. 2010. “Das strahlgeführte Mercedes-Benz Brennverfahren—Der Weg zum effizienten Ottomotor.” TAE, 9. Symposium Ottomotorentechnik, 2. und 3. 15-60. Buri, S. et al. 2010. “Reduzierung von Rußemissionen durch Steigerung des Einspritzdruckes bis 1000 bar in einem Ottomotor mit strahlgeführtem Brennverfahren.” 9. Internationales Symposium für Verbrennungsdiagnostik, AVL. pp. 8 and 9. 15-61. Hammer, J. et al. 2010. “Künftige Anforderungen und Systemlösungen für die Kraftstoffzumessung bei modernen Ottomotoren.” TAE, 9. Symposium Ottomotorentechnik. pp. 2 and 3. 15-62. Pritze, S. et al. 2010. “GM’s HCCI—Erfahrungen mit einem zukünftigen Verbrennungssystem im Fahrzeugeinsatz.” 31. Internationales Wiener Motorensymposium. pp. 29 and 30. 15-.63. Schaupp, U. et al.: “Benzin-Direkteinspritzung der 2. Generation: Kombination von Schicht- und homogenem Brennverfahren.” 10. Internationales Stuttgarter Symposium. pp. 16 and 17. 15-64. Dobes T. et al. 2011. “Maßnahmen zum Erreichen künftiger Grenzwerte für Partikelanzahl beim direkteinspritzenden Ottomotor.” 32. Internationales Wiener Motorensymposium. Pp. 5 and 6. 15-65. Kratzsch M. 2011. “Der qualitätsgeregelte Ottomotor—Ein konsequenter Weg mit Zukunftspotenzialen.” 32. Internationales Wiener Motorensymposium. pp. 5 and 6. 15-66. Lückert P. et al. 2011. “Potenziale strahlgeführter Brennverfahren in Verbindung mit Downsizingkonzepten.” 32. Internationales Wiener Motorensymposium. pp. 5 and 6. 15-67. Frese, F. and sowie Fuchs, A. 1965. Bussien, Automobiltechnisches Handbuch, Bd. 1, 18. Aufl. Technischer Verlag Herbert Cram: Berlin. pp. 757–788 and 789–791. 15-68. Scheiterlein, A. 1964. Der Aufbau der raschlaufenden Verbrennungskraftmaschine, 2. Edition. Springer Verlag: Vienna. 15-69. Nomura, K. and Nakamura, N. 1993. “Development of a new Two-Stroke Engine with Poppet-Valves: Toyota S-2 Engine.” in A New Generation of Two-Stroke Engines for the Future?, edited by Paris: P. Duret. Editions Technip. pp. 53–62. 15-70. Knoll, R., Prenninger, P., and Feichtinger, G. 1996. “2-Takt-Prof. List Dieselmotor, der Komfortmotor für zukünftige kleine Pkw-Antriebe.” 17. Internationales Wiener Motorensymposium. In: VDI Fortschritt-Berichte Reihe 12 Nr. 267. VDI Verlag und AVL Infounterlagen: Düsseldorf. pp. 25 and 26. 15-71. vom März. 1999. http://www.yamaha-motor.co.jp. And information from: N. N.: Diesel Progress International Edition (ISSN1091 3696) XVII, 4, Skokie, IL, USA: pp. 42 and 43. 15-72. N. N.: IAA. 1999. “Motoren und Komponenten.” MTZ 60, 11: p. 719.

15-52. Waltner, A. et al. 2006. “Die Zukunftstechnologie des Ottomotors: strahlgeführte Direkteinspritzung mit Piezo Injektor.” 27. Internationales Wiener Motorensymposium.

15-73. Schunke, K. 1989. “Der Orbital Verbrennungsprozess des Zweitaktmotors, Vortrag beim.” 10. Internationalen Wiener Motorensymposium. Fortschritt-Berichte VDI Reihe 12 Nr. 122. VDI-Verlag: Düsseldorf. pp. 27 and 28.

15-53. Stiebels, B. et al. 2003. Die FSITechnologie von Volkswagen—nicht nur ein Verbrauchskonzept. Spicher, U. u. A.: Direkteinspritzung im Ottomotor IV. Expert Verlag.

15-74. Cumming, B. S. 1991. “Opportunities and challenges for 2-stroke engines, Beitrag zum.” 3. Aachener Kolloquium Fahrzeug- und Motorentechnik. Aachen: 15.–17. 10. 1991.

654 | Internal Combustion Engine Handbook

6606_Book.indb 654

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

15.4 Two-Stroke SI Engines

15-75. Plohberger, D. and Miculic, L. A. 1989. “Der Zweitaktmotor als Pkw-Antriebskonzept-Anforderungen und Lösungsansätze, Vortrag beim.” 10. Internationalen Wiener Motorensymposium. Fortschritt-Berichte VDI Reihe 12 Nr. 122. VDI-Verlag: Düsseldorf. pp. 27 and 28. 15-76. N. N. 1991. “Neuer Subaru-Zweitaktmotor im Versuch.” MTZ 52, 1: p. 15. 15-77. Appel, H. (Hrsg.). 1990. “Der Zweitaktmotor im Kraftfahrzeug, Abgasemission, Kraftstoffverbrauch, neue Konzepte.” Tagungsband gemeinschaftliches Kolloquium am 28. Uni Berlin, ISBN 37893 13695. 15-78. N. N. 1993. “Fahrzeugmotoren im Vergleich.” Tagung Dresden 3–4. VDI Gesellschaft Fahrzeugtechnik, VDI-Berichte 1066. VDI-Verlag: Düsseldorf. 15-79. Groth, K. and Haasler, J. 1962. “Gaswechselarbeit und Ladungsendzustand eines Zweitakt- und eines Viertaktottomotors bei Teillast.” ATZ 62, 2: pp. 51–53. 15-80. Mugele, M. 2002. Numerische Analyse eines Spülvorlagenkonzeptes zur Emissionsreduzierung bei kleinvolumigen Zweitaktmotoren. Logos Verlag: Berlin. 15-81. Jäger, A. 2005. Untersuchungen zur Entwicklung eines Zweitaktottomotors mit hoher Leistungsdichte und niedrigen Kohlenwasserstoffemissionen. Logos Verlag: Berlin. 15-82. Gegg, T. 2007. Analyse und Optimierung der Gemischbildung und der Abgasemissionen kleinvolumiger Zweitaktottomotoren. Logos Verlag: Berlin.

15-84. Meinig, U. 2001. “Standortbestimmung des Zweitaktmotors als Pkw-Antrieb: Parts 1–4.” MTZ 62, 7–11. 15-85. Rebhan, M. and Stokes, J. 2009. “Kombinierter Zweitakt- und Viertakt-Ottomotor für weitreichendes Downsizing.” MTZ 70, 4: pp. 316–322. 15-.86. Osborne, R. et al. 2008. “The 2/4 Sight Project-Development of a Multi-Cylinder Two-Stroke/Four Switching Gasoline Engine.” Warren dale, PA: SAE-paper 20085400 Presentation 384. 15-87. Shawcross, D. and Wiryoatmojo, S. 1997. “Indonesia’s Maleo Car, Spreareads Produktion of a Clean, Efficient and Low Cost, Direct Injected Two-Stroke Engine.” IPC9 Conference. Nusa Dua, Bali, Indonesia. 15-88. Stan, C. (Hrsg.). 1999. Direkteinspritzsysteme für Otto- und Dieselmotoren. Springer Verlag: Berlin, Heidelberg. 15-89. Blair, G. P. 1996. Design and Simulation of Two-Stroke Engines. SAE-Verlag: Warrendale, PA. ISBN 1.56091-685-0. 15-90. Heywood, J. B. and Sher, E. 1999. The Two-Stroke Cycle Engine. Its Development, Operation, and Design. SAE, Verlag Taylor, and Francis: Warrendale, PA. ISBN 0-7680-0323-7. 15-91. Dixon, J. C. 2005. The High-Performance Two-Stroke Engine. Haynes Publishing: Sparkford, UK. ISBN 1844250458. 15-92. Kirchberger, R. et al. 2010. “Können umkehrgespülte Zweitaktmotoren für Freizeitanwendungen die zukünftigen Emissionsgrenzwerte erfüllen?” 31. Internationales Wiener Motorensymposium. pp. 29 and 30.

15-83. Braess, H. H. and Seiffert, U. (Hrsg.). 2000. Vieweg Handbuch Kraftfahrzeugtechnik. Friedrich Vieweg & Sohn Verlagsgesellschaft mbH: Braunschweig/Wiesbaden.

Internal Combustion Engine Handbook | 655

6606_Book.indb 655

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 656

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16 Electronics and Mechanics for Engine and Shift Control Transmission Management 16.1 Environmental Requirements

16.1.1 Classes of Installation

The environmental requirements on engine management and shift control transmission management are determined mostly by these parameters: •• temperature

The environmental conditions are chiefly a function of the installation site (Figure 16.1) in the passenger car, and are categorized according to four main classes of installation: •• passenger compartment or electronic box (E box) •• engine compartment (surface mounted on the chassis)

•• vibration •• protection against certain media (pressurized and unpressurized liquids, solids, and so on). Due to increased functionality and resulting power loss, self-heating must be taken into increased consideration for the thermal design of controllers. Ultimately, the suitability of every design must be proved for the defined environmental requirements. For this purpose, the finite-element method is used in simulations during the early stages of development and, finally, testing for environment-related suitability.

•• surface mounted on the aggregate •• integration in aggregate. The definition of the different classes (Figure 16.2) allows the development of specific housing approaches and, for example, overall integration approaches for transmission shift controls. This standardization reduces project-related expenditures, that is, chiefly the costs of materials and tools. Standardization also allows the manufactured units to be simplified and supports a global manufacturing strategy.

Integration in aggregate (e.g., installation in transmission) Aggregate add-on (e.g., at the engine, transmission, air filter) Engine compartment Electronic box Passenger compartment

Figure 16.1 Installation spaces.

Internal Combustion Engine Handbook | 657

6606_Book.indb 657

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

Passenger compartment, E box

Engine compartment

… . 5g

… 16 g

… 28 g

… 40 g

… 80°C

… 105°C

… 125 °C

… 140 °C

Dust-tight

Dust-tight

Dust-tight, vapor-jet-tight

Gear oil-tight

Vibration values (depending on frequency) Thermal class (ambient temperature) Seal

Aggregate add-on

Integration in aggregate

(e.g., at the en- (e.g., installation gine, transmission, in transmission) air filter)

Figure 16.2 Installation classes.

There are different reasons for selecting an installation site, such as •• cost savings in the wiring harness (engine space, engine, and transmission mounting) •• EMC reduction with a shortened wiring harness (engine compartment, surface-mounted on engine, and transmission) •• installation in the passenger compartment and concentration of the electronic control units (ECUs) (E box) •• engine tests including the ECU before installation (surface mounted on the engine) •• integration options (system approach); for example, intake module (near the engine and transmission management). For some time, we have seen a tendency toward installation near the engine or transmission instead of the passenger compartment. Figure 16.2 shows the predominant environmental conditions at the installation sites.

The environmental conditions for the housing become increasingly harsher closer to the engine or transmission, which is reflected in the design of the device (selected materials, manufacturing principles, functioning, and so on). Control devices are differentiated as stand-alone products and integrated products. Stand-alone products are engine management systems and transmission shift controls that are installed as independent units in passenger cars in contrast to integrated products that are combined with other functional units (for example, the transmission). These two approaches are described in detail in Chapter 16.2 and Chapter 16.4. Figure 16.3 shows the different approaches using the example of a transmission shift control.

16.1.2 Thermal Management

In the past, the housings had the principal task of protecting the electronics from environmental conditions such as water or dust and mechanical effects. Due to increased functionality

Modern drivetrain architecture Comparison: Stand-alone vs. mechatronic solution Stand-alone control unit Sensors/actuators

Mechatronic transmission module (MTM) Sensors/actuators

Internal wire harness

Plug Transmission

TCU

Wire harness transmission-TCU Plug Plug

Transmission

Plug

Vehicle wire harness

Vehicle wire harness Vehicle signals

Vehicle signals

Figure 16.3 Power train: Representation of stand-alone and integrated products.

658 | Internal Combustion Engine Handbook

6606_Book.indb 658

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.1 Environmental Requirements

and electrical output but increased ambient temperatures, thermal management is gaining in importance. Hence, the thermal parameters have become a decisive factor in the selected approach. Figure 16.4 provides an overview of maximum temperatures occurring at various installation sites in a passenger vehicle. While, in the past, electronics were frequently installed in the passenger compartment or thermally protected by an E box, and they are now increasingly installed in the engine compartment. This causes the ambient temperatures to frequently exceed the previously common 85°C. Even within the engine compartment, moderate zones are increasingly occupied by other electronic components, resulting in the increased installation of engine and transmission management systems on or in the aggregate. In the transmission interior, the temperatures of up to 140°C can occur. For engine management systems, installation sites in the immediate vicinity of the exhaust system are avoided to limit the maximum temperatures to 125°C. These temperatures mostly occur during the post-heating phase, the so-called “hot soak.” In this event, the vehicle with a maximum heated engine is parked at a wind-protected site. Thus, there is no cooling by airflow, and the entire heat stored in the engine and the cooling system heats up the engine compartment.

The maximum temperatures define the range in which the functioning of the engine management system must be assured. Regarding the reliability over the service life, an analysis of the maximum temperature would result in undesirable, that is, cost-driving reserves. It is more expedient to use the distribution of ambient temperatures over service life, that is, the so-called thermal load profile. Figure 16.5 shows a typical temperature profile. The main focus of the distribution is typically approximately 30–40°C below the maximum temperatures. Using common service life formulas (Arrhenius equation) and acceleration factors, the service life on the basis of the temperature profile is approximately 10 times the service life when assuming a permanent operation at maximum temperature. In addition to the ambient temperature, the self-heating of the device must be considered because of the continuous increase in the power loss. Power loss is driven by measures for fuel consumption savings and emission prevention such as direct injection or variable valve train. If the power loss in the past was approximately 15 W, modern engine management systems discharge 40 W and more into the ambient air. Figure 16.6 shows a basic design for heat removal in modern control units and a simplified thermal resistance model.

Passenger compartment - Moderate zones 85 °C - Roof 120 °C

Trunk 85 °C

Engine compartment

- Moderate zones 85 °C - at engine 125 °C - in the transmission 140 °C - at the exhaust system 205 °C

Chassis

- Insulated 85 °C - at heat sources 175 °C

Figure 16.4 Maximum temperatures at various installation sites in a passenger vehicle.

25%

Frequency (%)

20%

15%

10%

5%

0% 0

10

20

30

40

50

60

70

Temperature at installation site (°C)

80

90

100

Figure 16.5 Temperature distribution in operation across service life.

Internal Combustion Engine Handbook | 659

6606_Book.indb 659

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

TBE PBE

Thermal pad Thermal interface

Rth int Theat sink PSG Rth ext Heat sink

Tambient

Cooling fin

Figure 16.6 Thermal management—simplified resistance model.

While the printed circuit board (PCB) is always populated on both sides to optimize the package density, components with high power loss are populated only at one side and placed on a so-called thermal bank. In the area of the thermal bank, the PCB is connected to the heat sink by a thermal interface (heat-conducting foil or thermal compound). The thermal interface has the task to electrically isolate and, at the same time, to ensure proper thermal contact. To improve the heat throughput via the PCB, the so-called thermal vias (electrically and/or thermally conductive connection from the upper conductor plane to the one below) are used—through-plating without electrical function, which increases the effective vertical thermal conductivity value of the PCB because of the copper portion. The combination of the PCB soldering pad and an optimized arrangement of thermal vias is called a thermal pad. The heat sink is usually manufactured from aluminum die-cast, and discharges heat mostly by convection into the ambient air. To improve the thermal transition, it features cooling fins that are placed mostly in the area of the thermal banks.

Load case 1 Rotational speed: 700 U/min

The thermal path can be simply displayed so; the component discharges its power loss P BE via the internal thermal resistance Rth int and the heat sink. The heat sink conducts the total power loss of the device PSG to the ambient air via the external thermal resistance Rth ext. The following applies:

TBE = TAmbient + PSG ⋅ Rth ext + PBE ⋅ Rth int



Depending on the component size, Rth int is usually between 3 and 15 K/W. Rth ext strongly depends on the housing size and the air approach flow. For quiescent ambient air (natural convection) and typical housing sizes, this value is approximately 1 K/W and steeply decreases with increasing air velocity. In the course of the development of a control device, the thermal situation is verified in regular intervals to optimize the design of the cooling fins and the circuit carriers. For this purpose, thermal simulations are used in the beginning. Figure 16.7 shows the result of such a virtual validation; at a constant ambient temperature and an air approach flow (here: 90°C and 0.5 m/s), the speed of 4000/min represents the most critical load event. The temperatures are approximately 20–30°C above the level at idle speed. However, even in the most critical load event, the components stay below 150°C, and their electrical functioning is assured. The results from the other load events are weighted and flow into service life analyses. Using simulations, it is possible to verify the thermal condition of a control device before the first hardware has been produced. In addition, designs can be optimized in very short loops—without the familiar waiting times caused by prototyping. Furthermore, it is possible to analyze effects that can be accessed only with difficulty or not all by the conventional measuring instruments, for example •• temperature distribution in the enclosed device •• local heating in the ICs interior •• transient states (short power loss peaks).

Load case 2 Rotational speed: 1700 U/min

Load case 3 Rotational speed: 4000 U/min

Ambient air: 90°C / approach flow: 0.5m/s Figure 16.7 Virtual validation using simulation—analysis of different load states. See color section page 1089.

660 | Internal Combustion Engine Handbook

6606_Book.indb 660

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.2 Stand-Alone Products

By now, thermal simulations have an excellent consistency with measurable results. However, their quality is very dependent on the quality of the input variables, such as

the metallic housing parts. To keep the path of transported heat as short as possible, the power semiconductors are usually placed on the edge of the PCB.

•• power loss of the components •• assumption regarding ambient temperature and approach velocity •• deviations of real components and control device designs from the nominal designs. For this reason, measurements complement and, in part, replace the simulations in the course of the development of the control device. The final validation is made exclusively using real arranged hardware.

16.2 Stand-Alone Products The core functions of the housing for vehicle electronics are •• protecting the electronics from ambient influences (dust, water, and aggressive liquids) •• protecting the electronics from mechanical stresses (vibration and mechanical shocks) and stabilization of the arrangement •• electrical interface to the wiring harness •• thermal interface to the environment •• mechanical interface to the vehicle. In many applications, the housing also serves as the interface to the air pressure to allow the ambient pressure to be measured via an internally installed sensor. In this case, a so-called pressure compensation element is used that basically consists of a semi-permeable membrane. The housing must also have a variety of fastening options (insert options, screwing, and clamping) and resist local vibrations. The development team is also supported by stress analyses with the goal of optimizing weight (dimensioning parts to distribute stress under different load conditions). This is reflected in the trend toward light construction in automotive design. Figure 16.8 represents and explains various housing types. The housing in Figure 16.8 is a classic representation of the installation under moderate conditions. The thermal design consists of a PCB with a special layered construction that conducts the released heat from the electrical components to

Figure 16.8 Example of housing for use in the passenger compartment.

Figure 16.9 shows the housing type for installation in the engine compartment. This design fulfills the conditions that are required by most customers. Three representatives are shown, covering the band width from a basic functionality to standard size to a high-end version. The number of connector pins varies by a factor of 2 and the used PCB surface and/or number of components by a factor of 5. The modular design permits adaptation to different types of fixation, including an integrated module built onto the intake system (Figure 16.10). The variability of the concept is mostly achieved by adapting the heat sink, which is made from aluminum die-casting. In the course of efforts to reduce vehicle weight and fuel consumption, weight optimization of this component gains importance. The following options are available: •• minimizing the material use in cooling fins •• reducing the wall thickness •• use of alternative materials such as magnesium die-casting. Despite many customer-specific design options, the technical design and, hence, the related manufacturing process remain mostly unchanged. Various plug connectors reflect the appropriate functionality (corresponding to the number and type of connector pins) and wiring harness approach (number of plug-in modules) (Figure 16.11).

Figure 16.9 Housing assembly kit for the engine compartment. Left: basic functionality. Center: mainstream with scalable length. Right: high-end with two PCBs and water cooling.

Internal Combustion Engine Handbook | 661

6606_Book.indb 661

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

clamped in E-Box Water cooling

Air filter Fins “across” Fins “longitudinal” Assembly with bolts Figure 16.10 Adaptation of the installation space by variation of the die-cast component.

Due to tightened exhaust gas standards and the efforts to reduce fuel consumption, more and more sensors and actuators are connected to the engine management system. This causes these trends: •• increased number of pins •• additional high-current-capable pins •• further differentiation of pin numbers in the product palette.

Number of pins

Figure 16.11 Modularity of the plug connector.

Figure 16.12 shows the constantly rising requirement of connector pins. While ten years ago, 121 connector pins were sufficient for engine control, and the current VDA plug with 154 pins was introduced in 2003. At this point, plug connectors with up to 196 pins are under development. Requirements exceeding this number are realized using modular designs with multiple plug connectors for each engine management system. At the same time, there is a rising

First use in series Figure 16.12 The platform approach meets the increasing demand for connector pins.

662 | Internal Combustion Engine Handbook

6606_Book.indb 662

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.3 Connection Systems

demand for engine control systems with basic functions with approximately 100 connector pins. The housing shown in Figure 16.13 is an adaptation of the standard design for location at the air filter or intake module near the engine. The resistance against vibration as a PCB device should be emphasized. It is achieved by stabilizing the PCB with additional screwing points. The die-cast part is directly exposed to the air flow in the air filter. This ensures an effective form of thermal management.

This very expensive solution has been mostly replaced by devices based on PCB [Figure 16.14 (right)] in recent years. The fundamental disadvantage of the standard technology is mostly compensated by double-sided population and the improved adaptability to the outside contour specified by the installation. Sufficient ruggedness for this installation location was achieved using high-temperature-resistant circuit boards, the press-fit technology, and a low-resonance design.

16.3 Connection Systems

Figure 16.13 Example of housing for air filter installation.

Mounting devices on the engine (Figure 16.14) pose the greatest demands on materials, design, and manufacturing. Until now, the solutions [Figure 16.14 (left)] fundamentally differed from the aforementioned principles. Ceramic materials are required as the substrate, and the ICs used as the electrical components are bare-die packages. The investment in the manufacture of these devices is substantial. The advantages to the customer of this approach are the potential miniaturization and the possibility of integration in the engine or drivetrain (integrated transmission shift controls and smart actuators).

The plug connector or “plug” is the result of extensive coordination with the customer and the suppliers. The design of this part includes such considerations as engine management, customer system approach (distribution in the engine and chassis: wiring harness architecture), contact system (cross sections and surfaces), seal approaches (single core or collective seal), plug-in force, direction of installation, locking strategy, anti-theft strategy, resistance to vibration, flexural strength, installation procedure (flow soldering, reflow soldering, and bonding), and selection and combination of materials to cite the most important. The result is a part that is chiefly defined by the following criteria: •• seal •• number of pins •• number of modules (chambers) •• plug-in direction of the plug (perpendicular or parallel to the PCB). Because the developmental effort is substantial, automobile manufacturers are striving toward uniformity in design and classes of requirements in collaboration with the suppliers of ECUs and connectors. The electrical connection between the connector plug and the PCB in the interior of the engine management system

Figure 16.14 Engine management systems for installation on the engine. Left: ceramics technology with unhoused ICs. Right: PCB technology with SMT components.

Internal Combustion Engine Handbook | 663

6606_Book.indb 663

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

has been traditionally realized by soldering. The “press-fit” technology offers a very promising innovation in this respect. The male connector pins are here given flexible press-in zones, which are pressed in bores with tight tolerances on the PCB, creating an extraordinarily rugged and wear-free connection. This technology is currently used mostly for engine and transmission management systems with high ambient demands (temperature and vibrations). Because the “press-fit” technology additionally meets the lead-free requirement of the new European legislation, it will become more popular in future.

16.4 Integrated Products (MTM) In addition to the actual controlling of switching components, a transmission shift control includes the detection of the peripheral variables that influence path sensors, angle sensors, speed sensors, temperature sensors, and pressure sensors. A mechatronic transmission module (MTM) integrated in the transmission covers the functions of electronic control and peripheral sensors, as well as the subfunctions of potential and signal distribution, contacting the shifting components and other hydraulic and electronic interfaces. It is ideally a complete, independent module. In contrast to “stand-alone” transmission shift control units, ECUs integrated in the transmission have maximum mechatronic potential, because all important input and output components are directly in the transmission. If the transmission and the integrated control can be designed together in the concept phase, the maximum amount of integration can be achieved, because the arrangement of all the required components can be optimized in reference to their location, orientation, and technology (Figure 16.15).

Figure 16.15 Modern MTM for a Volkswagen direct manual transmission system.

A controller integrated into the transmission is ideally attached to the hydraulic control plate. In longitudinally installed automatic multistage transmissions, the control plates and controllers are usually installed in the lowest point of the transmission in the oil sump, because the pressure control valves work here and can be directly contacted and

actuated. Crosswise-installed dual clutch transmissions, on the other hand, have usually “standing” arranged control plates and controllers, because the hydraulic interfaces and position sensors orient on the arrangement of the gear shift shafts. The MTM is subject to extremely harsh environmental conditions such as high, continuous vibration, continuously high temperatures, and, occasionally, very aggressive oils. An MTM must offer a hermetically sealed holder for the electronics, long-lasting internal and external interface contacts, high mechanical strength and stability to survive the vibration and pressure, a high level of reproducible precision to provide an exact reference for attached sensors, and 100% compatibility with the used transmission fluid. In addition, the MTM must be capable of manufacturing by means of automation and, of course, be economical. We now discuss an example of the design and technology of a modern MTM. The example shown in Figure 16.16 is a series-produced dual clutch transmission management for Volkswagen (direct manual transmission). An aluminum base plate serves as a base and interface element of the MTM to the hydraulic control plate of the transmission, and it also functions as a heat sink and heat transfer body for the ceramic substrate [low temperature co-fired ceramic (LTCC)] adhered to the control plate with thermally conductive adhesive. Laminated to the base plate is a flexible PCB that is the signal and potential distribution element of the transmission shift control electronics on the LTCC substrate to all peripheral sensing and signal detection components. Pressed into the aluminum base are pressure sensors with their own seals. The electrical contact of the substrate to the pressure sensors and flexible PCB is made with 200-μ m aluminum thick-wire bonds. The overall electronics and pressure sensors are hermetically sealed against the environment with a plastic cover and insertable seal to the laminated, flexible printed circuit and then riveted. The seal configuration and the seal material are designed to provide a sufficient and permanent seal restoring force under all tolerances and extreme temperatures. To protect the electronic components and bonds against vibration, the electronics compartment must be filled with a silicone gel. Contacts of the flexible PCB are created to the peripheral components such as speed and position sensors, control valve terminals, selector switch, vehicle plugs, and internal plugs via an approved laser-welding process. When mounting the MTMs on the control plate, all the pressure control valves are simultaneously contacted, the hydraulic interface to the pressure sensors is established, and all the sensor reference positions are set. The design and flexibility of the flexible PCB can compensate for a 3-D malpositioning of the valves in the hydraulic system. Although a great deal of value is placed on reusable technology, customer requirements and special demands will always need adaptable approaches. To achieve another and higher level of integration, it is possible in integrated transmission management systems to integrate the pressure regulation as a component of the module and even the control electronics as an integral component of

664 | Internal Combustion Engine Handbook

6606_Book.indb 664

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.5 Electronic Design, Structures, and Components

Speed sensors Position sensors

Position sensors

Cover

Internal plug Insertable seal and silicone gel

Flexible printed circuit board

LTCC substrate

Alu floor plate

Pressure regulator, contact bar Vehicle plug

Integrated pressure sensors

PCB cover

Figure 16.16 Main components of a modern MTM. See color section page 1090.

the hydraulic control plate. This would create an electrohydraulic module. A further design example is the 7G-Tronic, an MTM that has been series-produced since December 2002. This device integrates the sealed control electronics, the sensors for the selectable range, the turbine, the intermediate and drive speed, and the corresponding signal and potential distribution.

Eight coded pressure regulating valves with spring contacts are flatly pressured onto open-flexible PCB areas and, thus, connected to the control electronics. The electronic interface MTM/vehicle is formed by an oil-tight five-pole plug that projects through the transmission housing. This plug is sealed against the housing, movable within the defined limits, and, thus, allows the offset of all the assembly tolerances (Figure 16.17).

16.5 Electronic Design, Structures, and Components 16.5.1 Basic Structure

Figure 16.17 7G-Tronic: Integrated mechatronic transmission management for Daimler vehicles.

Figure 16.18 illustrates the basic signal flow and the essential function blocks in a block diagram. The signals detected by the sensors are sent via an input filter structure to the computer. These signals are then converted and are sent via the final stages to the actuators. By means of digital interfaces, contact can be made with other ECUs or shop diagnosis devices. A voltage regulator ensures the required supply of voltage and the current to the components. In addition, complex reset logic is required to ensure proper functioning.

Internal Combustion Engine Handbook | 665

6606_Book.indb 665

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

Inputs

Signal processing

Outputs

Micro-controller (8/16/32bit) Storage units

Sensors

Input signal filter

Driver stages

Actuators

possibly monitoring unit Diagnosis

Diagnosis

Communication interfaces

Power supply, reset logic

Figure 16.18 General signal flow.

16.5.2 Electronic Components

We now look at a few typical examples of electronic components in engine and transmission shift controls.

This is done via a serial interface. The serial interface is also used to program the variables that can be set in the component (Figure 16.19).

16.5.2.1 IC Knocking Input Filter Component Up to two knocking sensors can be connected to this building block. Their signals are processed in the component’s filter and sent to the microcontroller for evaluation.

16.5.2.2 Driver Stage Component Multiple building blocks are frequently used. Figure 16.20 and Figure 16.21(a) and (b) provide examples for a four-output and a sixteen-output driver stage. These multiple driver stages

FEATURES

- Supply voltage range

4.0 V .. 5.5 V

- Temperature range

40 - to 125 °C

- CMOS, TTL compatible inputs - SPI Interface to the failure register - microprozessor programmable 1 gain 2 filter frequencies 3 integrator time constant 4 clock prescaler

BLOCK DIAGRAM differential amplifier 1

channel select

antialiasing filter

+1 –1

gain

bandpass filter

- switched capacitor technology - various extrenal clock capability (prescaler) - Package SO20 full wave rectifier

output buffer/ converter

integrator

Out +2 –2 differential amplifier 2

logic block power supply

midrail generator

control

interface

OSCin osc

OSCout clk /CS Data /Data Int/Hold /Test

VCC

GND

Vmid

Figure 16.19 Filter for IC knocking.

666 | Internal Combustion Engine Handbook

6606_Book.indb 666

1/19/16 8:48 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.5 Electronic Design, Structures, and Components

FEATURES - Supply voltage range

4.0 V .. 5.5 V

- Short current protection with current limit

3A

- Average current (for each output)

2.5 A

PACKAGING ≤ 0.5 Ω

- On resistance (@ Tj = 150 °C ) - Output clamping voltage

50 V typ.

- Temperature range

–40 to 125 °C

- Slewrate-control 1 pos slewrate (rise-time) 2 neg. slewrate (fall-time)

10 .. 55 V/µs 5 .. 20 V/µs

20

- Individual thermal shutdown - Undervoltage Reset and controlled power-up and down

1

- Controlled output voltage-slewrate

PSO 20 Power-Package

- SPI Interface to the failure register - Destingtion between 3 kinds of failure for each powerstage 1 overcurrent (SCB) or overtemperature 2 short circuit to GND (SCG) at the off-state 3 open load (hot OL detection) at the on-state

BLOCK DIAGRAM Vcc

S

NON2

OUT1

URES

R

RES

NON3

dV/dt control

driver

RES URES

NON4

charge pump

Trigger

NON1

VCC

FR RESET

ON1

OUT2

over-temp. detection

OUT3

I_SCB filter t_ISCB

IRES

OUT4

ON1

SDI

CLK

VCC

shift register VCC

failure register

I_OL fitler t_IOL NON1

(FR)

SCG filter t_SCG V_REF

NCS IRES

SDO

URES

under voltage RESET

GND

RES

Reset

OSC

Oscillator

NRES

Figure 16.20 Four-output driver stage.

Internal Combustion Engine Handbook | 667

6606_Book.indb 667

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

FEATURES MULTIPLE LOWSIDE DRIVER

PACKAGING

- Supply voltage range 4.5V .. 5.5V - Short current protection: for channel 1 and 2 8A >3A for channel 3 – 8 >1A for channel 9 – 16 - Average current for channel 1 – 8 2.5A for channel 9 – 16 0.7A On resistance (@ Tj = 150°C ) for channel 1 – 8 0.4 for channel 9 – 16 1.5 - Output clamping voltage 50V typ. - Temperature range 40 to 125C programmable - Slewrate-SPI voltage slewrate 0.6 .. 7.2 V/s current slewrate 0.15 .. 2.3 A/s - Individual thermal shutdown - SPI Interface to the failure register - Destingtion between 3 kinds of failure for each powerstage : -: overcurrent (OC) or overtemperature short circuit to GND (LVT) at the off-state open load (hot OL detection) at the on-state

QuadPowerFlatpack MO188

BLOCK DIAGRAM Vbat

CP Vcc

Over Temperature Latch

SDI SDO

Failure Latch

CLK NCS

SPI Block

4x8 4x6 4

l o g i c

out 9...16 out 3...8 out 1,2

Charge pump control

Slew Rate Setting (from SPI)

slew rate

mode

Temp high/low

Vcc

VSR_th

dV/dt dl/dt

ref

VLVT filter

OUT1

ref

Vcc OUT16

NON1 Control (from SPI)

Temp. Sense

Gate Drive

Drive Signal

Temp Threshold Setting (from SPI)

S Over Load R Latch

Ioc filter

NON16

Iref

filter NDIS1

Iuc

ref ref

CFB1

NDIS4 Vcc CFB2

reset

OSC

NRES NRES1

undervolt. reset

internal reset

oscillator GND

Figure 16.21 (a) Example of a sixteen-output driver stage (part 1). (b) Example of a sixteen-output driver stage (part 2).

668 | Internal Combustion Engine Handbook

6606_Book.indb 668

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.5 Electronic Design, Structures, and Components

IPeak

IHold

tPeak

t Hold Figure 16.22 Peak-and-hold current profile.

are directly controlled by the microcontroller and are able to control the actuators. A complex diagnostic system monitors the outputs for fault states such as overcurrent, short circuit, excess temperature, or open-load (wire break). The error bits are separately stored for each output, and can be read, evaluated, and stored by the microcontroller via a serial interface. H-bridge circuits are used to actuate DC motors, either as integrated circuits (ICs) or, for high performance, with individual transistors. These components enable a change in the direction of rotation reverse/forward and of the speed by varying the impulse-pause ratio (see Figure 16.31 to Figure 16.33). Peak-and-hold circuits are used to achieve very fast valve opening times. For this purpose, the valve is first actuated with a high current to quickly open. The system then regulates to lower a current value holding the valve open. Such a current profile is shown in Figure 16.22 (peak-and-hold circuit). 16.5.2.3 Microcontroller The microcontrollers used have been specifically designed for applications in automotive engineering. They combine high computational performance with a high degree of integration with peripheral components that are necessary for evaluating the input signals and controlling the driver stages. Figure 16.23 and Figure 16.24 each show an example for a microcontroller with the bandwidths of 16 bit or 32 bit and their essential function blocks.

16.5.2.4 Voltage Regulator This component comprises three regulators—the main regulator and two downstream regulators with substantially less power. The main regulator is responsible for the components in the ECU, while the two downstream regulators can be used, for example, to control sensors outside of the ECU. Furthermore, a monitoring unit and release logic are integrated in the component (Figure 16.25). 16.5.2.5 DC/DC Converter For the provision of the necessary electrical energies, shorter cut-in times of electromagnetic injection valves or piezovalves require voltages many times above the battery voltages available in the vehicle. The transformation of the 14-V vehicle system voltage to the operation voltages of these high volt loads is realized by DC/DC converter circuits (Figure 16.26). They can be electrically described as current-limited voltage sources in parallel to a storage capacity at the output. The electric charge stored in an inductance flows cycled from a power switch via a diode into a storage capacitor and charges the same in steps to a defined voltage value Vout, which may be multiple of the input voltage Vin and, thus, delivers the prerequisites for the fast opening of electromagnetic valves. The DC/ DC circuits used in control units also contain comprehensive circuitry for protection from excess current and polarity reversal and output-end short circuits to ground or supply voltage.

Internal Combustion Engine Handbook | 669

6606_Book.indb 669

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

FEATURES Microcontroller C167-Family of Infineon – CPU: 16 bit „von Neumann“ register oriented architecture. – CPU: 4 stage pipeline with 16bit ALU – 4kRAM build of 2k Dual Port RAM and 2k XRAM

PACKAGING

– PEC for fast data transfer from peripheral to RAM – ADC unit with 10bit, 16 channel and channel injection – 2 CAPCOM unit with 16 CAPCOM channels each – PWM unit with 4 PWM channels – General purpose timer unit with 5 Timers – 2 serial interfaces (UART and SPI) – watchdogtimer – up to 61 digital I/O channels when external bus enabled – fast interrupt inputs with min. 300ns response time at 20 MHz – operation temperature range –40 °C ... 125 °C – full CAN – optional: 32 kByte ROM

PQFP 144

BLOCK DIAGRAM

ROM 32kByte (optional)

Instruction Bus 32 bit Data

Dual Port RAM

C167 - CPU Core

2kByte

PLL Oscillator

Watchdog Timer PEC

CAN module Interrupt Controller

XRAM 2kByte

Interrupt Bus

Port 6 8 Bit Port 0 16 Bit Port 4 8 Bit

External ADC Bus 10bit Controller 16 channel

Port 1 16 Bit

Port 5 16 Bit

USART module

SPI module

GPT 1 module

Port 3 16 Bit

GPT 2 module

CAPCOM 1 + 2 PWM module modules 4 channels 32 channels

Port 2 16 Bit

Port 7 8 Bit

Port 8 8 Bit

Figure 16.23 Microcontroller.

670 | Internal Combustion Engine Handbook

6606_Book.indb 670

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Figure 16.24 Block diagram of the function blocks of a 32-bit microcontroller.

Figure 16.25 Example for a voltage controller.

Internal Combustion Engine Handbook | 671

6606_Book.indb 671

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

L

Q C out1

Vin (14V)

16.6.3 Signal Evaluation

CR

Vout

Rs

PWM CONTROL

Figure 16.26 Principles of a DC/DC converter circuit.

16.6 Electronics in the Electronic Control Unit 16.6.1 General Description

The engine electronics assumes the tasks of a central control unit for the combustion process in the engine. Depending on the engine and vehicle-side input signals, the integrated computer unit calculates the required variables of the actuators relevant to the engine functions. The ECU comprises the following main functional groups.

16.6.2 Signal Processing

Analog and digital signals pass from sensors, switches, and other control units throughout the engine compartment and vehicle interior via the wiring harness into the ECU. Here, the different types and amplitudes of signals are converted into digital voltages and frequencies that represent information readable by the microcontroller. The adaptation of the input signals from the knock sensor, lambda sensor, and induction-type pulse generator on the crankshaft is particularly complex. In the stochastic signal from the knock sensor, the higher signal of the knocking engine is filtered out of the permanent engine noise level, amplified, rectified, and integrated. This is done with the aid of an IC that allows the midfrequency and amplitude to be preset to any desired level via a programmable register. Finally, a normalized signal is transmitted during a definable time slot to the A/D converter of the computer. During operation, the lambda sensor yields a current proportional to the oxygen component in the exhaust gas. This current flow at the internal resistance of the sensor generates at the output a voltage that passes via an adapter circuit to the A/D converter. The special feature of the crankshaft signal is the dependence of the signal amplitude from the speed. It ranges from a few hundred millivolts at low speeds to several hundred volts. The signal is converted into the digital rectangular shape of the same frequency by zero transition point detection, where interference is suppressed by a variable negative feedback.

The computer unit itself comprises the CPU, the ROM for program code and parameters, the variable data memory, and the monitoring unit for the safety tests in E-gas systems. The digitally prepared input signals serve as the variable actual values of the engine functions represented in binary code. Program maps and characteristics form the variable manipulated variables for the programmed arithmetic operations. The results of many individual calculations are transferred in the form of level/time information to the output ports of the microcontroller. A predefined safety approach is used in devices with throttle valves set by an electric motor. Algorithms are calculated simultaneously in the CPU and the monitoring unit, and the results are exchanged via the serial interface and compared. In the case of deviations, the safety function is activated, which then turns off the redundant throttle valve, fuel injectors, and ignition to stop the vehicle.

16.6.4 Signal Output

The logical level of the output port driver of the controller is used directly as a control signal of the respective driver stage that, in turn, operates the actuators installed in the vehicle. The driver stages can be classified in three categories. Low-side drivers control inductive and ohmic loads that are connected against the battery voltage, such as valves, relays, and ignition coils, as well as heating resistors and the logical interface of other electronic controls (Figure 16.27).

Low-side driver

Ubat

Ubat

High-side driver

Inductive, ohmic load Current flow Current flow

Inductive, ohmic load

Figure 16.27 Principles of a low-end or high-end driver circuit.

High-side drivers connect the current flow for consumers that are connected to ground at one side. In the case of bridge driver stages, the consumer is connected with both terminals of the ECU. This type of connection is employed particularly for operating DC motors that require a continuous adjustment of forward and reverse movements. All driver stages have a self-protection function that prevents the component from being destroyed in the case of electrical short circuits at the output to the battery or ground or a load

672 | Internal Combustion Engine Handbook

6606_Book.indb 672

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.6 Electronics in the Electronic Control Unit

short circuit. In addition, these interruptions are detected by circuitry and buffered in a fault register. The arithmetic-logic processor then can retrieve the error code from the driver stage via the available serial interface and implement reactions such as limp-home functions, the triggering of an error light, and entry into the internal fault storage.

16.6.5 Power Supply

This circuit element draws the current for the ECU from the vehicle system voltage. Depending on the level and load on the battery (for example, the starter), 6–18-V variables are converted into a stable DC voltage of 5 V (additional voltages are used in modern systems, such as 3.3 and 1.8 V) to operate the electronics. Very often, linear voltage regulators, the so-called linear regulators, are used. The unit compares the output voltage with an internally generated reference voltage and, in the event of deviations, actuates the transistor to regulate to the target value. The voltage difference between input and output is converted to heat in the transistor. Contrary to this, the switching regulator principle causes, during the cut-in phase, the energy to be stored in the magnetic field of an inductance, which is then discharged to the output during the cut-out phase. Switching frequency, duty cycle, and circuitry decisively influence the properties and the efficiency, which is usually between 70% and 90% and, thus, considerably higher than the values of linear regulators. The suppression of interference from the vehicle electrical system (up to ±150 V and 100 ms pulse width) by protective measures with semiconductors and capacitors contributes to the flawless operation of the electronics. In addition, this circuit block offers up to three stable 5-V voltages for supplying external potentiometers or sensors with current.

16.6.6 Interfaces 16.6.6.1 CAN Bus Interface The controller area network (CAN) is a serial bus system. It was created especially for linking intelligent sensors, actuators, and electric motor/transmission controls (ECU/TCU) in the vehicle. CAN is a serial bus system with multimaster

properties. The CAN bus protocol was developed especially for applications affecting safety in the automobile industry. All CAN elements can transmit data; several nodes can simultaneously query the bus. The serial bus system has real-time properties. It was declared an international standard in ISO 11898. The object-oriented messages contain information, such as speed and temperature, and are available for all the receivers. Each receiver independently decides, based on the transmitted identifier, if the message is to be processed or not. The arbitration between the bus elements is prioritycontrolled by the identifier. The maximum data transmission rate is 1 Mbit/s. 16.6.6.2 LIN Bus Interface Local interconnect network (LIN) describes an inexpensive communication standard for comfort electronics, intelligent sensors and actuators, and engine control components with safety relevance. The communication is based on a bit-serial single-wire conductor following the SCI (UART) data format. The maximum transmission rate is 20 Kbit/s. The individual nodes are synchronized without a stabilized time base. The specification follows ISO 9141. By differentiation in the master and one or more slaves, a collision of messages on the data line is avoided as only the master can initiate communication. 16.6.6.3 FlexRay Bus Interface The very fast FlexRay field bus is a time-controlled and errortolerant communication interface meeting the requirements for safety-critical systems in the vehicle. This bus defines a standard, across manufacturers, with a data rate of up to twice 10 Mbit/s, fixedly defined latency times and transmission cycles. FlexRay uses the time division multiple access principle. Fixed time windows are assigned to the bus elements or messages, during which they have exclusive access to the bus. These so-called “time slots” repeat in the defined intervals, that is, the periods during which the information is on the bus can be precisely predetermined. The disadvantageous influence of this fixed assignment on the band width is offset by the division in static and dynamic segments, where the very short dynamic “minislots” are used priority-controlled only when communication is needed. Both the transmission channels allow an error-tolerant transmission of the data. Figure 16.28 depicts the communication cycle of the data transfer.

communication cycle dynamic segment

channel 1

static segment

1

2

channel 2

4

3 C1

A1

5

7

6 E1

A2

D1

12 9 8 E2 D2 C2 t

1

2 A4

4

3 B1

C1

5 B2

7

6 A2

E1

8 9 10 F1 C2

A3 t

Figure 16.28 Communication cycle of the data transfer. See color section page 1090.

Internal Combustion Engine Handbook | 673

6606_Book.indb 673

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

16.6.7 Electronics for Transmission ECUs

The same applies to the function blocks of the CPU, power supply, driver stages, and interfaces as mentioned above. However, there are a few function modules that were developed specifically for transmission shift controls: (a) Current regulator for electromagnetic valves: In automatic transmissions, the engagement of the various gears is achieved by opening and closing brake and clutch elements in the transmission. These elements are actuated by switching on and off electromagnetic proportional valves in the transmission’s hydraulic control unit. To enable a comfortable gear shift, a precise process must be followed during the actuation of the brake and clutch elements. This is achieved via the regulation of the opening cross sections of the relevant flow control valves in the transmission’s hydraulic control unit. The electrical current proportional to the opening cross section of the relevant valve is used as the measuring property. The microcontroller uses the so-called measuring shunts to calculate this current and to regulate it to the target value. (b) Redundant driver building blocks: The load is operated between a low-side driver stage and a high-end driver stage, for the second driver stage to be able to respond to an error (for example, a short circuit of the low-side driver to the outside) and, thus, to use the electronics to prevent any system states that may be critical for the vehicle. For modern applications, ASIC developments have been made with the goal of further improving the functionality and the diagnostic capability. These building blocks assume many tasks of the microcontroller and those of the driver stages as well. Previously, they had to be given a certain level of “intelligence”; today, simple MOSFETs can be used. Monitoring the driver stages (MOSFETs) and, if necessary, evaluation of any errors, such as short circuit or wire breaks, are now performed only in the ASIC. It uses serial data transmission to transmit the evaluation to the microcontroller. This ensures that an intervention is only required for an

error and decisions regarding critical or noncritical states are made using the system condition. In safety-critical states (such as overvoltage in MOSFET), the shutdown is executed directly in the ASIC without intervention by the microcontroller. Because of the detailed error monitoring in the ASIC, the microcontroller is relieved and can process new functions, such as improving the driving comfort. (c) High-integrated building blocks for applications in transmission management units: Figure 16.29 and Figure 16.30 provide an example. The functions described in b) are realized here. Furthermore, additional functions such as watchdog, interfaces, or voltage regulators are integrated in this example. The integration reduces the number of building blocks of the complete circuit and, therefore, the space requirements on the PCB, and lowers the production costs. Overall, a higher functionality is achieved despite lower costs. (d) Bridge circuits for electric motors with high current demands: In transmissions for all-wheel applications, frequently electric motors are used as actuators for the activation of the differential locking function, the activation of the second drive axle (all-wheel drive), or for shifting gears (off-road gear). A combination of these functions allows the moment regulation in the drive to improve vehicle dynamics. Peak currents far beyond 25 A are required for the operation of these electric motors. Integrated H bridge building blocks are no longer suitable for such loads. Discrete solutions have to be used. For this, an intelligent logic building block with four output transistors is combined to one H bridge circuit (Figure 16.31). This logic building block is responsible for the control and monitoring (diagnostics) of the transistors while the transistors (usually MOSFETs) are used as switches for the current. In newer applications, brushless DC (BLDC) motors are often used. In this case, the rotor is turned by an electromagnetic rotary field. This rotary field is generated by a suitable actuation of three half

Main task

■ Control of 8 solenoid valves via High-side or Low-side driver incl. diagnosis ■ 8 channel current measurement with high CMRR ◊ one wire per valve

Additional functions

■ Voltage supplies for microcontroller and peripherals ■ 2 additional sensor supplies ■ Watchdog ■ SPI Interface ■ ISO9141 (K-Line) or LIN Interface ■ 2 comparators

Figure 16.29 Summary of the ASIC functions.

674 | Internal Combustion Engine Handbook

6606_Book.indb 674

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

RSLx

RSHx

HLSxSOURCE

HLSxDRAIN

NDISDRV

HLSxGATE

NHLSxON

ST_OUT1 SADJ1 ST_OUT2 SAD2

CSDRV CSSNS

MSSNS

MSDRV

16.6 Electronics in the Electronic Control Unit

8x VREFSD

VIN

HLSCONFx

NO_WD_CH8

SSLCT1/2/3

CSSLCT1/2/3

CSSLCT1/2/3

SDCLK

PREADJ PRESNS PREOC PREDRV

OTP bits

WDWx

ENACOMP

V_BG

KLSLCT

RBC

NRSTH NRSTIN NRSTL

TRDx

FAILURE

CLK NCS1 NCS2

SDI SDO

TEST

NWDDIS

NWDOUT

K-Line RxD TxD

CMP_OUTx

CMP_INx

E2

E1

CP1C CP2C CP4C CP5C CP3C

IO_SUPPLY

V_5INT

C_BACKUP

OTP

Figure 16.30 ASIC block diagram.

bridges (see Figure 16.31). The complex, alternating actuation of the half bridges and the monitoring of proper function is executed in building blocks specifically developed for this purpose. Figure 16.32 and Figure 16.33 show an ASIC, converting the speed set point of the microcontroller to a rotary field for BLDC motor. The detection of the direction of rotation of the E motor, the actuation of the half bridges, and (with the microcontroller) the precise

• Supply Voltage Range: 6V…40V • Ambient Temperature Range : -40 … +125°C (package) • 5V µC supply + 2 sensor supplies • Watchdog for µC-Reset + Failsafe function for output stages • Complete Gate control for three half bridges (incl. Reverse Battery MOSFET) • Flexible configuration of half bridges for DC and Brushless motors

Ubat

• Integrated block commutation circuit for Brushless motors • Protection of external MOSFET against over-current on HS and LS • Current limitation during motor start up Forward engine movement OUT1

M

• Detailed diagnosis (Standard SPI interface for configuration and diagnosis) OUT2

Reverse engine movement

• Load current measurement through integrated analog amplifier • Flexible adjustment of over-current limits, current limitation and dead time • 2 small signal low-side-outputs for LEDs and lamp control • 1 small signal high-side-output for battery voltage measurement • Package: QFN64 Figure 16.32 Summary of the ASIC functions.

Figure 16.31 Schematic wiring diagram of an H bridge.

Internal Combustion Engine Handbook | 675

6606_Book.indb 675

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

Control circuit power supply

BLDC-ASIC

- Adjustable pre-regulator (linear) with external MOSFET - 5V main supply 1 x MOSFET

2 Small signal low side outputs with diagnostic (fully integrated)

- 2 sensor supplies - Watchdog

Commutation circuit for Brushless motors

Small signal HS output (switched battery for battery measurement)

Configurable as Full

1 x MOSFT for Reverse Battery Protection

or Half bridge

Half bridge

Half bridge

Half bridge

(Control circuit) (Control circuit) - Current sensing

(Control circuit)

- MOSFET Driver - MOSFET Driver

- MOSFET Driver

- Protection

- Protection

- Protection

- Diagnostic

- Diagnostic

- Diagnostic

2 x MOSFET

2 x MOSFET + 1 x Shunt

2 x MOSFET

Figure 16.33 BLDC actuating ASIC block diagram.

position of the actuator are fully performed in the ASIC. Other functions (polarity reversal protection and voltage regulator, watchdog, and driver) are also integrated in this building block, for the same reasons as given above. This building block can also be configured for brush motors. The control electronics are frequently integrated in the transmission. This produces much higher demands (temperature, vibration, seal, and so on) on the electric components and materials such as the housing and seal elements (see Chapter 16.4).

16.7 Software Structures 16.7.1 Task of the Software in Controlling Engines

Over the past twenty years, the importance of software has increased steadily and dramatically in all areas of automotive electronics, especially in engine management systems. On the one hand, functions have become more economical and effective than were accomplished beforehand with mechanical or electronic solutions, and on the other hand, the potential of a freely programmable computer allows completely new and previously unattainable functions to be added. This includes the comprehensive self-diagnosis functions of modern ECUs or the fine harmonization of the combustion process made possible by software that minimizes emissions and fuel consumption.

The amount of software (Figure 16.34) in a typical engine management system has doubled about every three years in the past. Forecasts indicate that this trend will continue into the future. At the same time, computing power has continuously increased.An 8-bit processor (for example, Motorola 68HC11 and Infineon 80C517) programmed in an assembler was sufficient for typical engine management in 1990. In 1995, 16-bit processors (for example, Infineon C167) were required with the higher programming language “C.” A 32-bit processor (for example, the Motorola Black Oak) was used in 2000. A substantial part of the software (up to 50%) in a modern engine management system is not used for the actual functioning (control of the engines), but deals only with tasks surrounding the engine, such as diagnosing the ECU and peripherals, on-board diagnosis (OBD) II, and so on. The growing amount of software and simultaneously shrinking developmental times have led to a sharp rise in the size of the software teams on a single project from two developers in 1990 to more than ten in 2000. This has, in turn, necessitated strict developmental processes with extensive quality controls.

16.7.2 Demands on the Software

What are the demands on the software structure for the electronic control of the drivetrain? •• to describe the required functions: engine/transmission control, exhaust gas control, comfort functions, engine/

676 | Internal Combustion Engine Handbook

6606_Book.indb 676

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.7 Software Structures

600

120

500

100

400

80

300

60

200

40

100 0

Code [kByte] Data [kByte] 1991

1992

1994 1995 1998 Year of manufacture

transmission protection, self-diagnosis with extensive fault storage (EURO x and OBD), limp-home function, reprogrammability, and communication with other ECUs •• to provide fast reactions in real time: on the I/O level as low as 10 μ s; on the function level, 2 ms–1 s time synchronous or 1.8 ms–1.5 s synchronized with the crankshaft •• to be largely independent of the hardware, especially the microcontroller to support a multisource strategy; to cover different tasks (gas, diesel, automatic transmission, automaticmanual transmission, integrated starter generator, and so on) with a high degree of reusability between the different areas •• to be able to be integrated with software from customers and key-component manufacturers •• to use standard software components (OSEK operating system with communication and network services) •• to be compatible with function packages (aggregates; see Chapter 16.9) that represent a universal solution for the function (such as ignition) from the sensor, via the algorithm describing the function, to the actuator. To economically meet such demands, the function and the hardware must be largely separated. The solution is to use six software layers with specific responsibilities. In addition, a well-defined software development process is required to meet the deadlines and to ensure quality.

16.7.3 The Layer Approach to Software

The relationships between the layers are defined by a few rules: •• Each layer can use all the services of the subordinate layers, but none of the higher layers. •• Each layer can exchange data with the layers directly underneath, but it cannot jump a layer. •• Each layer can have control flows (without data transmission) to other layers. In control flows to higher layers, the layer cannot and may not recognize the recipient of the control flow.

2000

2002

20 0

Figure 16.34 Software of a typical six-cylinder engine management unit in kilobytes.

•• For each layer, only specific data types are allowed. •• Each layer is to be designed independently of the hardware, processor, and compiler. The two lowest layers form exceptions, but they need to follow this rule as much as possible. •• The six layers of software correspond to the respective levels of abstraction of the real world (Figure 16.35). •• The OSEK operating system is a standard software module and offers functions from different layers; whereas the communication and network services primarily act on the hardware abstraction layer, the operating system services are also available on the higher layers. The integration of the layer model is shown in Figure 16.36. Whereas a layered design is desirable from a software vantage point, from the viewpoint of system development, the formation of function packages, termed aggregates, is important. An aggregate includes everything that is required to fulfill a specific function (lambda control or ignition, for example), from acquiring and preprocessing the relevant data in the layer model on the lower layers of BIOS, PAL, and HAL, to the actual regulating and control algorithms (located in CAL and VAL), to triggering corresponding actions. These are sent in the software via HAL, PAL, and BIOS to the actuators. Different function packets use the same sensor data and control the same actuators, which is made easy by the software layer model that provides standardized interfaces. The modules that are assigned specific aggregates and sensors/actuators on a software layer then represent the preferred granularity of reuse. The sought advantages are combined using the presented software layer model: •• A new microcontroller or a new input/output component affects only the BIOS and PAL layers or perhaps the HAL layer as well; the function algorithms remain the same. •• A function can assume the required data as given; the input/output of these data can be implemented or reused independently.

Internal Combustion Engine Handbook | 677

6606_Book.indb 677

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

“real world”

Software model

Coordination of driver, engine, transmission

SUL Demands on regulation strategies

(Supervisor Layer)

Driver Driver command

VAL

Engine/transmission-related models and control strategies

Engine/transmission

(Vehicle Component measurement

Abstraction Layer)

Component measurement (physical measured values)

CAL Component representation Service request by function name

Circuit conversion and diagnosis

Component

(Component Abstraction Layer)

Control unit plug (electric measured values: V, R, I)

HAL (Hardware Service request by Abstraction Layer) logical channel number

μC Pin (electric measurements accepted by μC: V)

Input/output switching

PAL Distribution and adaption for BIOS function

(Processor Abstraction Layer) Service request by logical channel number (in ticks)

Processor Connection to periphery building block (periphery-specific)

BIOS Signal generation and acquisition

(Basic Input/ Output System)

Periphery building block

Figure 16.35 The six layers of drive control software.

•• Externallydeveloped functions that adhere to the interface conventions can be easily integrated. •• The packaging and independence from the hardware is a step in the direction of “software as a product,” that is, the ability to sell and buy individual software functions. •• The model also increases reusability, which allows an exponential increase in software capacity and, hence, keeps the amount of effort involved in developing new software at a reasonable level.

16.7.4 The Software Development Process

Efficient and quality-oriented software development in large teams requires an appropriate, well-outlined developmental process that follows the capability maturity model. The V cycle (see Figure 16.37) is widely used; this does a particularly good job of representing the interaction of task analysis across the levels of abstraction and the associated tests. The V is completely run through for each software delivery, usually at least five per project.

16.8 Torque-Based Functional Structure for Engine Management Today, modern engine management systems fulfill more than just the legal requirements for exhaust gas emissions and fuel consumption; they also increase driving comfort. This expands the task of engine management for an SI engine far beyond the control of injection and ignition. In implementing these requirements, functional quality has been substantially increased by the drive-by-wire system (mechanical separation of the accelerator pedal from the throttle valve). This is accomplished by functionally influencing the cylinder charge simultaneous to the driver’s input. This is especially necessary in SI engines with direct fuel injection and stratified engine operation under a partial load. In a torque-based function architecture (Figure 16.38), all demands that can be formulated as torque or efficiency are defined on the basis of these physical variables.

678 | Internal Combustion Engine Handbook

6606_Book.indb 678

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Coordination Driver – engine – transmission

Engine strategies

VAL

Configuration

CAL

Component regulation, sensor evaluation engine

Component regulation, sensor evaluation vehicle

Hardware abstraction services and function mapping

HAL

PAL

Error treatment and function mapping

Component regulation, sensor evaluation transmission OSEK COM/NM Interaction Layer Network Mgmt Network Layer Data Link Layer

Instrumentation

Abstraction services processor

BIOS

OSEK OS

Transmission strategies

Libraries

SUL

Periphery building block driver

Figure 16.36 Integration of the operating system in the six software layers.

SOFTWARE V-CYCLE

S1

Management Process

Quality

Assurance

Project Planning

Process Project Control Software Configuration Management

Development Process Development Process

System Requirements relevant for SW

S2

Q

Delivery to (internal)Customer Software Validation Plan

Software Requirements Specification Software Requirements Specification S3

(may be combined into one plan) Software Verification Plan

Software Design

Software Validation

S6

Integrated Software Software Integration S5

Verified Modules

SW Design Document

Module Verification (Operation level & Module level)

Coding Code

S4

Q

U

U

A

A

L

L

I

I

T

T

Y

Y

M

C

E

O

T

N

R

T

I

R

C

O

S

L

Result / Document Activity

Figure 16.37 Software development process.

Internal Combustion Engine Handbook | 679

6606_Book.indb 679

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

Start, catalytic converter heating, idle speed regulation

External requirements - Driver - Driving comfort - Speed regulation - Transmission - Driving dynamicsF

M D Injection time

Start

Conversion of torque to available variables

Coordination of torque requests

Idle speed regulating Speed limitation

Fade-out pattern Ignition angle Waste – Gate Valve overlap

Component protection Figure 16.38 Torque-based function architecture.

This means that the interface between individual functions and the (sub)systems are defined as torque or efficiency. This produces a clear function architecture within an engine management system. The goal is to attain the best possible compromise between drivability, fuel consumption, and exhaust gas emissions at all times. Because many functions must be highly dynamic over time, the required nominal torque must be realized by using two paths. In addition to the charging path that undertakes all the charge-influencing actuator judgments of the throttle valve, the crankshaft-synchronous ignition path assumes all interventions that directly influence combustion efficiency. All actuator judgments are undertaken in this path that influence the torque produced by the engine independent of the charge, that is, the ignition and injection timing. By coordinating the two paths, it is possible to quickly increase torque by adjusting the ignition angle. This is important for idle control, starting, transmission adjustments, and drive slip control. Efficiency can also be intentionally diminished by adjusting the ignition, for example, a retard adjustment when heating the catalytic converter to improve exhaust. Also belonging to the fast torque path are the lambda and cylinder shutoff path that can be activated if necessary and coordinated with the charging path and/or ignition path. Fresh gas filling Ignition angle Lambda : :

Motor

Torque from combustion

–

Engine torque

The main influencing variables such as fresh gas charging, the ignition angle, and lambda produce the inner torque (TQI) from combustion that, in contrast to the indicated torque, does not include the entire charge cycle. The torque produced by an engine (Figure 16.39) results from the indicated torque minus the torque lost from friction. The friction and charge cycle losses are combined in the lost torque TQ_LOSS. After subtracting the lost torque from the auxiliary systems, we have the clutch torque TQ_CLU. The drive torque available at the wheels is determined after considering the loss from the clutch and the transmission. The TQI (Figure 16.40) is calculated as follows: TQI = TQ_CLU − TQ_LOSS. The lost torque TQ_LOSS is always a negative torque, because the engine is not fired and must be operated with a trailing throttle. In the equation, it is added to the clutch torque, because the inner high-pressure torque TQI from combustion contains both the clutch torque measurable at the dynamometer and the lost torque TQ_LOSS. This means that the inner torque TQI must be at least as large as the lost torque TQ_LOSS to overcome the friction and charge cycle losses (TQ = 0). Only when the TQI is greater than the TQ_LOSS is there positive torque TQ at the clutch.

–

Clutch torque

Clutch

Transmission

Drive moment

Charge change and friction Auxiliary aggregates Clutch loss and translation Transmission loss and translation

Figure 16.39 Torque transmission.

680 | Internal Combustion Engine Handbook

6606_Book.indb 680

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.8 Torque-Based Functional Structure for Engine Management

TQI = TQ_CLU - TQ_LOSS

Clutch torque (positive or negative)

Induced torque (high-pressure phase only)

TQ_LOSS = 0 [Nm]

TQ = 0

Lost torque (negative)

TQI = 0 [Nm]

Figure 16.40 Calculating the TQI.

The driver command PV_AV is communicated to the system in E-gas systems (drive-by-wire) via a pedal travel sensor and interpreted in the torque structure as a torque request. This torque command can be changed by various actuator judgments such as the cruise control, load-reversal damping, or transmission adjustments. The resulting torque in the torque structure is then simultaneously fed into the two following paths: the slow torque adjustment path (charging path and slow air path), and then the fast torque adjustment path (ignition path and fast ignition path) where the efficiency for lambda (injected fuel quantity) and cylinder shutoff are included in the ignition efficiency. The charging path is, therefore, termed the slow path, because its dynamics depend on the intake manifold volume, throttle valve cross-section, engine speed, and so on. Its time constant lies within the range of a few 100 ms. This is contrasted with the crankshaft-synchronous ignition path in which the time constant depends only on the engine speed. Whereas at an engine speed of 1000 rpm, a time constant of approximately 30 ms results; it is 5 ms at a speed of 6000 rpm. This is one

Air mass sensor Air filter

p im T im , V im

pamb

p thr

R g · Tim · p im = Vim

flows

k

· mk

16.8.1 Model-Based Functions Using the Example of the Intake Manifold Model

Rising demands in the precision of metering the injected fuel mass per cycle, or the precision of the moment of ignition, have led to increasingly complex functions in engine management. The most frequently used multidimensional tables (in which the correction factors for the injection time are saved as a function of various input conditions) no longer suffice for attaining complex goals, because, especially in transient engine operation, the detection of metering the fuel mass or the position of the moment of ignition is too imprecise. This yields deviations from the stoichiometric mixture that can produce problems with emissions and drivability. Therefore, we seek to describe the physical relationships in the so-called models. This allows the set parameters such as the fuel mass for injection per cycle to be derived from physical formulas for the air mass inducted per cycle. A disadvantage

· m cyl

· m kgh

Ideal gas equation:

to two orders of magnitude smaller than the charging path time constant.

p ex T ex , V ex

Exhaust gas system

R g · Tim · · ) (m thr – -m p· im = cyl Vim

Figure 16.41 Intake manifold model (without exhaust-gas recirculation).

Internal Combustion Engine Handbook | 681

6606_Book.indb 681

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

is the increasing demands on the processor performance and the increased memory arising from usually complex algorithms for the models. This is illustrated using the example of the intake manifold charging model in Figure 16.41. In the past, the injection time was calculated directly from the air mass signal (using, for example, a hot-film air-mass meter). In transient operation, that is, when the throttle valve is opening or closing, the intake manifold is filled or emptied as a result of a change in the intake manifold pressure. The arising air mass error produces an incorrect injection time that is manifested by a rich mixture when accelerating (positive load jump) and a leaner mixture upon easing off the gas pedal (negative load jump). Now, with the aid of the intake manifold charging model, we are able to largely eliminate this mixture error. The air mass flowing out of the intake manifold into the cylinder is calculated from the absorption lines saved in the tables that depend on the modeled intake manifold pressure. The intake manifold pressure is calculated from the general gas equation using a differential equation. The air mass signal is initially Air mass flow through throttle valve (equation of St.Venant): m thr = A red · p thr · J · C1

not used. The air mass flowing via the throttle valve into the intake manifold is first determined only from the position of the throttle valve. Later, the parameters of the model, such as the environmental pressure and the reduced throttle valve cross section, are corrected with the aid of adaptive methods. To simplify the formulaic relationship of air mass through the throttle valve (Figure 16.42) as a function of the differential pressure pamb − pim, a so-called ψ function is used (equation by St. Venant). The air mass flowing into the cylinder is saved in the form of line equations as a function of the intake manifold pressure (Figure 16.43). The linearized first-order differential equation for the intake manifold pressure compiled from the individual models can be numerically solved in the segment grid in the engine management system using a method of integration (for example, trapezium rule). The advantages of this model-based functional approach are •• Harmonization of a dynamic model largely by stationary determinable values. The parameter of the intake manifold volume primarily influences the dynamic behavior.

(d) psi function in straight form (offset & slope): J = Jofs – Jslop ·

(b) Reduced throttle valve cross-section: A red = A red (d) (throttle position)

J

p im p thr

( Jofs , Jslop ) j

(c) Temperature-dependent constant: C1 =

2k 1 = C1(Tim ) (k – 1) R g · Tim

section j j

((b), (c), (d) in (a): Mass flow through throttle valve depending on A red, p thr , p im :

p im p thr

m· thr = A red · C1 · (Jofs · p thr – Jslop · p im ) Figure 16.42 Model for air mass over throttle valve.

· m cyl

Volumetric effectiveness n, depending on:

Constant engine speed

– Ambient pressure

hslop

– Engine speed

– Exhaust gas counter-pressure pim

– Intake manifold valve position – Temperature

– Variable camshaft position

hofs

– Variable intake manifold length Mass flow into cylinder, depending on intake manifold pressure im p : m cyl = h slop · p im – hofs

Figure 16.43 Model of the air-mass flow in the cylinder.

682 | Internal Combustion Engine Handbook

6606_Book.indb 682

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

•• Understandability and repeatability of the application, because the function is chiefly based on physical variables. •• Independent structures of pressure and air-mass guided systems. •• Simplified diagnosis. •• Invertible function (important for determining the throttle valve set point in torque-guided systems and for determining the safety strategy).

16.9 Functions 16.9.1 λ Regulation

The three-way catalytic converter with λ regulation has become the most familiar exhaust gas treatment approach for SI engines with external mixture formation. λ regulation ensures that the pollutant components CO, HC, and NO are optimally converted. It is necessary to maintain a stoichiometric composition of the air-fuel mixture (λ = 1) within a very narrow λ range (λ window) (Figure 16.44). 100 NOx 80

Degree of conversion

l - Window 60

40 HC 20 CO 0 0,7

0,9

1,1

1,3

Air-fuel-ratio I

Figure 16.44 Lambda window.

Target value

–

PI regulator

Correction value, injection volume

In a closed-control loop, the air-fuel ratio λ is measured by the λ sensor in the exhaust gas that compares the actual air-fuel ratio with the set point and corrects the fuel quantity if necessary. There are binary and linear lambda sensors. These sensors are described in Chapter 18.3.1. To optimize the function of the three-way catalytic converter, that is, to best oxidize CO and HC and maximize the reduction of NOx, the air-fuel mixture before the catalytic converter must have a certain fluctuation; that is, the internal combustion engine must operate in a specific manner with both excess air and deficient air. This ensures that the oxygen storage unit of the catalytic converter is filled and emptied. When O2 is stored, the NOx level is also reduced; when the oxygen is released, oxidation is supported, which prevents stored oxygen molecules from deactivating sections of the catalytic converter. The control algorithm (Figure 16.45) for binary λ regulation is based on a PI controller. The P and I components in the program maps are saved via the engine speed and the load. With binary regulation, the catalytic converter (λ fluctuation) is implicitly stimulated by the two-step control. The amplitude of the λ fluctuation is set at approximately 3%. A superimposed trimming control via a binary after-converter sensor helps maintain the λ window before the catalytic converter. In linear lambda control (Figure 16.46), forced excitation is necessary to set the λ fluctuation. The diagram provides an overview of the structure of the linear λ regulation including forced excitation and trimming control. At the actual λ set point, the forced excitation (Figure 16.47) modulates a periodic deviation (λ pulse) to optimize catalytic converter efficiency. The obtained signal is directly entered as a precontrol element in the fuel quantity correction; the signal may also be influenced by the secondary air and be further processed as a filtered λ set point, considering the gas travel time and the delay behavior of the linear sensor. The signal of the linear λ sensor is converted via a saved characteristic into a λ value. This characteristic can be corrected by the trimming control (Figure 16.48). The trim controller is designed as a PI controller that uses the after-converter sensor signal, which is less exposed to cross sensitivity (preferably by a binary jump sensor). The control deviation (Figure 16.49) is then calculated from the corrected λ signal and the filtered λ set point as richness (= λ 1) that serves as input to the actual λ controller, which is designed as a PII2D controller and is shown in Figure 16.49. The I2 component serves to balance the oxygen charge of the catalytic converter. The controller output can also be further limited under nonstationary operating conditions.

Engine + lambda sensor

Actual value

Figure 16.45 Control algorithm for a binary lambda sensor.

Internal Combustion Engine Handbook | 683

6606_Book.indb 683

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

measured I Characteristic offset

Nachkatsonden-Signal

(PI-) Trimmregler

corrected I 1

/

Control deviation (Richness)

–

Limitation indication

On/Off Stop Adjustment

Regulator output

(PII2 D-) l-Regulator

Injected volume correction

Limited due to transient conditions

–1 1

/

1 filtered I - target value

Consideration of gas running time and sensor behavior

/ Precontrol path

l - Target value

possibly secondary air flow

l - Window/target value

IPulse

On/Off

Forced excitation

Forced excitation

l -Pulse

Figure 16.46 Lambda control for linear sensor.

0

l -Pulse

Time

l mess

Figure 16.47 Forced excitation.

Trimming control

Characteristic curve without correction

∆l

Characteristic curve correction

Nominal characteristic curve Corrected characteristic curve

l Target I - Window (operating point of trimming regulation)

Figure 16.48 Trimming control.

684 | Internal Combustion Engine Handbook

6606_Book.indb 684

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

Limitation

P-portion Control deviation (Richness)

I-portion

Regulator output

Limitation 2

I -portion

Limitation

D-portion

Figure 16.49 Calculation of control deviation.

The injected fuel quantity correction calculated in this manner is included with the precontrol in the injected fuel quantity calculation. Linear lambda control has the following advantages over binary lambda control: •• Increased control dynamics and reduction of transient λ errors. •• Greater catalytic converter efficiency from settable forced excitation in a closed λ control loop. •• Ability to regulate λ ≠ 1; this allows controlled warm-up or controlled catalytic converter protection.

16.9.2 Anti-Jerk Function 16.9.2.1 Introduction Because of sudden changes in the engine torque that can arise from acceleration or from easing off the gas pedal, the vehicle is excited to oscillate lengthwise. These perceptible changes in acceleration are very uncomfortable to passengers. The effect can be observed in nearly all passenger cars; its intensity

depends on the type of construction of the drivetrain and its parameters (the rigidity of the drivetrain, for example). It is very important to reduce the effect by engine management, because transient handling is an important parameter in the decision to buy a particular vehicle. To develop the functions, a simple physical model is used that sufficiently describes the “jerk effect.” There are two functions in engine management for reducing lengthwise vehicle oscillations: •• load-reversal damping (torque transient) principle: control (driver command filter) •• anti-jerk function (anti-jerk controller) principle: control circuit. The drivetrain can be represented as a two-mass oscillator (Figure 16.50). The mass m1 represents the moment of inertia J1 = m1 ⋅ r2rot (rotating masses)—crankshaft and camshaft, pistons and connecting rod, flywheel, and auxiliary systems. The mass m2 comprises the drivetrain masses (gears, propshaft, and wheel masses) and the remaining vehicle mass. As torque increases in the engine, torsion is exerted on the drivetrain, and the stored energy acts on m1.

(Engine bearing effect) (Play)

Engine torque TQ motor

TQ rad

TQ 2 w1

Clutch

Transmission

m1

Gear

w2

w3 C

m2

Spring Inertia moment J1: • Mass of moved engine components • Flywheel

Damper D Propeller shaft, drive shaft

Inertia moment J2: • Vehicle mass • Propeller shaft • Drive shaft • Gear wheels • Tires

Figure 16.50 Drivetrain as a two-mass oscillator.

Internal Combustion Engine Handbook | 685

6606_Book.indb 685

1/19/16 8:49 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

If TQengine is a step function, the drivetrain oscillates at its natural frequency. The amplitude and frequency of this oscillation and its decay time are gear-dependent. In a low gear, the amplitude and the frequency are higher, and the decay time is longer than in a high gear. To dampen load reversal, the following physical facts are exploited; based on the model of the two-mass oscillator, we can show that the oscillation can be reduced depending on the excitation of the systems. Ramp-shaped signals are particularly suitable. The oscillation amplitude is smallest when the ramp rise time is the same as, or a multiple of, the period duration. This theory can actually be used only when the spontaneity of the vehicle is not impaired. A compromise between the comfort and the dynamics must be reached by shorter ramp times.

When the driver commands are filtered in this manner, oscillations remain in the drivetrain that must be compensated by the anti-jerk function. The output value of load-reversal damping is a torque command of the driver or cruise control filtered by a ramp function. The ramp rise times are determined from the selected gear (Figure 16.51). Because the function is not only for damping oscillations in the drivetrain but also for tipping the engine on its bearings, a distinction is drawn between different torque ranges. The core of the function is a variable calculation of the ramp rise. The speed gradient represents an additional condition for switching the torque range (Figure 16.52). The anti-jerk function and the load-reversal damping work closely together. The control loop fights the remaining

Torque request (filtered)

Torque request by driver

Torque at clutch

Vehicle (two-mass spring system) (Engine bearing effect)

(Play)

Engine torque TQ motor

Load reversal damping

w1

TQ rad

TQ 2 Clutch Transmission w 2

m1

Gear

C

N

w3 m2

Spring Inertia moment J1: • Mass of moved engine components • Flywheel

+

Damper D Propeller shaft, drive shaft

Inertia moment J2 : • Vehicle mass • Propeller shaft • Drive shaft • Gear wheels • Tires

Anti-jerk torque correction Anti-jerk function

Figure 16.51 Torque model.

Torque at Clutch

range

phase INC

CONST

TQ_REQ_CLU

CONST

TQ_REQ_TRA_OLD_CLU 3

TQ_REQ_TRA_CLU

C_TQ_TRA_THD_UP_2 0

CONST

DEC

C_TQ_TRA_THD_DOWN_2

C_TQ_TRA_THD_UP_1

t

2

C_TQ_TRA_THD_DOWN_1 TQ_REQ_TRA_OLD_CLU

LV_TQ_TRA

1

torque transient bit

Figure 16.52 Ramp rise of the torque.

686 | Internal Combustion Engine Handbook

6606_Book.indb 686

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

Reference engine speed calculation

Gear

1800

System traffic activation

1400

Activation Condition

1000 600

2

1

3 4 time [sec]

5

6

I_DIF_Al [rpm]

I_REF_Al [rpm]

Requested Torque

+ –

N [rpm]

1800 1400

200 150 100 50 0 –50 1 –100 –150

2

3 4 time [sec]

Initialization request

5

6

AJ torque request correction Engine speed deviation

Anti Jerk Controller

1000 600

1

2

3 4 time [sec]

5

6

measured engine speed

Figure 16.53 Diagram of anti-jerk regulation.

oscillations from the load-reversal damping. The load reversal damping, hence, is applied for a high degree of spontaneity, whereas the anti-jerk function provides greater driving comfort. The basic idea is to derive a correction signal from speed deviation that enters the torque set point in phase. Because the processes are very dynamic (typical frequencies for drivetrain oscillations range from 2–10 Hz), the torque command must be quickly implemented via the ignition. Oscillations in linear vehicle acceleration can be detected from the engine speed that is optimal for signal detection because of its resolution and updating properties. The oscillations in the drivetrain are expressed as a speed differential based on the deviation of the actual speed from a reference speed. From this difference, the correction signal is derived for the torque command. The phase and amplitude of the correction signal can be influenced by parameters. The correction signal is active only during an applicable time frame. The anti-jerk function (Figure 16.53) is triggered when oscillations in the speed occur.

In certain configurations of the drivetrain, the physical limits of an internal combustion engine are reached. The correction signal can no longer be implemented in phase.

16.9.3 Throttle Valve Vontrol

In a torque-guided engine management system, the set point for the electronic throttle valve results from the intended torque in an “inverse intake manifold charging model” or the reverse path of the intake manifold charging model (Figure 16.54). The target position of the throttle valve is calculated from the torque command via various steps, and it is set by the throttle valve position controller (Figure 16.55). The goal of throttle valve control is to precisely match the actual air with the desired air mass (from the torque model). In the forward branch of the intake manifold charging model, the air mass flowing into the engine results from the throttle valve position and the speed.

Intake manifold charging model Throttle valve angle

α/N system

Rotational speed Load sensor

+

Air mass flow into cylinders

Regulation unit

Engine

Current torque

Torque management … Adaptation

Throttle valve target value

Torque calculation

inverse /N system

Air mass target value

Inverse intake manifold filling model

Safety strategy Air mass target value calculation

Torque target value

Figure 16.54 Consistency of the forward and reverse air-mass paths.

Internal Combustion Engine Handbook | 687

6606_Book.indb 687

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

Knocking is problematic for engine efficiency, because at today’s conventional compression ratios of 10–12°CA, the most efficient moment of ignition is in the knocking range of the ignition characteristic (average pressure as a function of the moment of ignition) (Figure 16.57).

Calculating the:

As function from:

Torque target value

pedal value, EGS, EGR, ...

Filling target value

Speed, torque

Intake manifold pressure target value

Volumetric efficiency

Pressure quotient over throttle valve

Ambient pressure

Target value for throttle valve throughput

Pressure quotient

Target value for reduced throttle valve cross-section

Air mass target value, throttle valve throughput

Throttle valve angle target value

red. throttle valve cross-section

110 90

Figure 16.55 Calculating the throttle valve set point from the target torque.

This relationship must be exactly invertible so that a throttle valve target position can be calculated in the reverse path from the charging set point. Figure 16.56 shows the structure of the throttle valve control loop. The input signal for the position controller is the difference between the actual and desired positions of the valve. Depending on this deviation, a control algorithm calculates a control signal pulse width modulation signal that influences the servomotor on the valve so that the actual throttle valve position is moved to the desired position.

16.9.4 Knocking Control

Knocking is uncontrolled, self-instigated combustion of normally inert gas components usually with high flame speeds around the velocity of sound, and causes high pressure peaks. Continuous knocking combustion damages the engine, primarily the pistons, cylinder head seal and cylinder head. Knocking can be reduced by the following measures: •• later moment of ignition •• higher octane number (RON) of the fuel •• richer mixture •• lower charge pressure •• lower intake air temperature •• reduction of deposits on the piston and valves •• suitable construction of the combustion chambers.

ETC + target position

–

ETC regulating algorithm

PWMsignal

50

Optimal ignition timing Knocking range 55

45

30

Knocking limit

35 25 Ignition timing before TDC

Torque

70

10 15

5

–10

Figure 16.57 Engine torque as a function of the moment of ignition.

If the engine is operated close to this optimum efficiency range, knocking control is required. The goal of engine management is to operate the engine in a closed control loop at the knocking threshold to the extent, that is, before the optimum moment of ignition. The engine management system, therefore, intervenes at the moment of ignition in uncharged engines and influences the charge pressure and the moment of ignition in charged engines. In a knocking control system, we make use of the noise arising from the pressure oscillations in the combustion chamber by tapping the structure-borne noise signals in the crankcase with the aid of a knock sensor. In the knock sensor, a seismic mass acts on a piezoceramic element and induces a charge there proportional to the structure-borne noise oscillation of the installation site. The noise—typically within a frequency range of 5–15 kHz—arises as a resonance of the engine structure with the high-frequency components in the pressure characteristic that arise in the combustion chamber during knocking because of the turbulent flame velocities. Figure 16.58 shows a typical pressure characteristic and the structure-borne noise signal for normal and knocking combustion. The engine management system detects knocking from the electrical knocking signal by first formatting the raw signal in an IC (Figure 16.59).

ETC ETC actual position with actuator

Figure 16.56 Structure of the electronic position control of the throttle valve (ETC).

688 | Internal Combustion Engine Handbook

6606_Book.indb 688

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

Knocking sensor

Knocking IC: Raw signal

Raw signal formating

Knocking IC Programming parameters Formated knocking signal

Knocking Knocking event message event energy

Knocking correction

Ignition angle Correction

Function interruption (when error detected) Bit for error detection Error detection and replacement values

Error correction values (when error detected)

Figure 16.59 Knocking signal processing.

ignition angle is retarded for safety so that the engine operates reliably under all the circumstances outside of the knocking range. Figure 16.60 shows the time characteristic of the knock control adjustments. As shown in Figure 16.60, there is a fast and a slow ignition angle adjustment. This is because of the different phenomena that cause knocking. For example, the deposits on the piston or the fuel quality are influences that change slowly; in contrast, the intake air temperature and the engine operating point are influences that can change from power cycle to power cycle.

The formatted raw signal is further processed in a microprocessor. The knocking event exists in a cylinder-selective form when the formatted raw signal exceeds the previously applied knocking limit adapted in engine operation. This evaluation occurs in a knocking time frame that is established for each cylinder by the crank angle of the engine. The energy calculated from the knocking signal determines, in another block, the extent of the ignition angle correction. In case of error, that is, when, for example, the knocking control cannot work properly because of a sensor error, the

5

2.5

Knock Detection & Correction Correction line

4.5

2

Detection line

4

1.5

3.5

1

3

0.5

2.5

0 KNK_SLOW_COR °CRK IGA_PROP_COR °CRK LV_KNK_DET

2

KNK_FAST_COR °CRK KNK_EGY

–0.5

1.5

–1

1

–1.5

0.5

–2

0

–2.5 0:00

0:04

0:09

0:13

0:17

0:22

0:26

Time

0:30

0:35

Figure 16.60 Time characteristic of the knock control adjustments.

Internal Combustion Engine Handbook | 689

6606_Book.indb 689

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

The position of the knock sensor should be set, so that knocking is easily recognized while the engine is operating and can be clearly differentiated from other influences such as valve gear noise. The engine is extensively investigated in the engine development phase. Knocking is detected with combustion chamber pressure sensors, and the results are compared with the measured structure-borne noise signal. In four-cylinder engines, a knock sensor is usually placed on the inlet side on the crankcase between cylinders 2 and 3. This allows the knocking noise of all the four cylinders to be recognized. In six-cylinder inline engines, two knocking sensors are used. Two knocking sensors (one knock sensor per cylinder bank) are also used in V-6 and V-8 cylinder engines.

16.9.5 On-Board Diagnosis

The air pollution from dense traffic in metropolitan areas led the United States in the 1960s to set legal limits of vehicle emissions. Today, the United States, especially California, has the world’s strictest emission thresholds for passenger cars. This development has caused the exhaust gas purification systems for vehicle engines to become increasingly bigger and more complex. These measures have substantially reduced pollutant emissions in new vehicles, yet simultaneously the portion of emissions has risen to a great degree from vehicles with defective exhaust gas purification systems. According to an estimation of the American Environmental Protection Agency (EPA) in 1990 (for example) approximately 60% of emissions came from uncombusted hydrocarbons from vehicles with faulty exhaust gas purification systems. Because of this problem, the EPA has demanded that engine management systems provide vehicles with self-diagnostic systems that monitor all exhaust-influencing systems, functions, and components, and inform the driver when these components are faulty. The essential component for exhaust gas purification is the three-way catalytic converter. In the catalytic converter, the exhaust gas components carbon monoxide and the uncombusted hydrocarbons arising during engine combustion are oxidized into carbon dioxide and water. At the same time, the nitrogen oxides are reduced to nitrogen. For maximum conversion of all the three exhaust gas components, the engine must operate with a stoichiometric mixture, that is, with an air-fuel ratio of λ = 1. The mixture must, therefore, be precisely controlled. To set the air-fuel mixture, the inducted air-mass and the speed of the engine are measured. In the ECU, these signals are used to calculate the opening time of the electrical fuel injectors and, hence, the fuel mass injected per power cycle to set a stoichiometric mixture. To adjust the mixture as precisely as possible to the required air-fuel ratio of 1, the so-called lambda control is superimposed over this control process. The lambda sensor in the exhaust gas system determines if the mixture is set too rich or too lean. As a function of the sensor signal, a correction factor for the injection duration is calculated in the ECU to produce an average air-fuel ratio of 1. The catalytic converter starts operating only when its temperature rises above the so-called light-off temperature.

Today, this temperature for catalytic converters is approximately 350°C. Fast heating of the catalytic converter during the warm-up phase can be attained by blowing secondary air into the exhaust gas system directly before the exhaust valves. The air-fuel mixture inducted by the engine is adjusted to be rich. The secondary air-mass flow is adjusted so that the air-fuel ratio in the exhaust gas system is slightly lean. This causes the uncombusted hydrocarbons and carbon monoxide to be oxidized in the exhaust gas system. Because this reaction is exothermic, the exhaust gas temperature rises. This, in turn, causes the catalytic converter to quickly heat up. In addition to the three-way catalytic converter, external exhaust gas recirculation is frequently used to lower nitrogen oxide emissions. Combusted exhaust gas is mixed with the combustion air. This causes the combustion temperature to fall, which consequently reduces nitrogen oxide emissions. The returned exhaust gas is metered by a valve in the return line. In addition to exhaust gas emissions that arise from engine combustion, there are additional hydrocarbon emissions from the vaporization of the fuel in the tank. Tank ventilation systems are used to reduce these emissions. These systems have the task of preventing the hydrocarbon vapors that arise in the vehicle tank from exiting into the atmosphere. An active charcoal filter that absorbs gaseous fuel is placed between the tank and the connection to the environment. This fuel is regenerated at specific intervals to prevent the filter from overloading. The tank ventilation valve between the active charcoal filter and the intake manifold of the engine is opened at specific intervals. The rising flow of air from the active charcoal filter causes desorption of the stored fuel. The rising fuel vapor-air mixture flows into the intake manifold and is burned in the engine. 16.9.5.1 Self-Diagnosis Tasks The goal of self-diagnosis is to monitor the functioning of all the exhaust-relevant vehicle components and systems during normal driving conditions. If an error is determined, the problematic components are precisely located, and the type of error and location of the error and environmental conditions are saved in a memory. If the fault causes set exhaust thresholds to be exceeded, the driver is informed via a signal light in the dashboard and asked to bring the vehicle to a repair shop. In addition, suitable measures are undertaken to maintain driving safety, ensure continued operation, and eliminate the subsequent damage. In the shop, it must be possible to read out the fault storage to allow the fault to be quickly found and fixed with the saved data. The first thing that the California environmental authorities developed to attain this goal was a draft of a specific law concerning on-board diagnoses of engine management systems starting in model year 1988. All components that were connected to the ECU of the engine management system had to be monitored. Starting in model year 1994, the expanded OBD II was required by law. For the first time, the monitoring of all exhaust-relevant vehicle components and systems was required. The requirements of the California environmental authorities were partially adopted by the other forty-nine states.

690 | Internal Combustion Engine Handbook

6606_Book.indb 690

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

In particular, the following main requirements exist (Figure 16.61):

Monitoring of all exhaust gas-affecting systems and components no

Error yes

- Detection of error type and location - Storing error information - Driver information - Driver information - Avoiding subsequent damages - Support of shop in repair

Figure 16.61 Tasks of self-diagnosis.

Monitoring •• the catalytic converter system •• the lambda sensor •• the entire fuel system including the fuel injectors, the fuel pressure regulator, the fuel pump, and the fuel filter •• the secondary air system •• the exhaust gas recirculation system •• the tank ventilation system consisting of the active charcoal filter and tank ventilation valve •• other systems held to be relevant to the exhaust that are not directly controlled by the engine management system, such as the transmission shift control for automatic transmissions. Combustion misses should also be recognized.

l -Sensor upstream of catalytic converter

16.9.5.2 Monitoring the Catalytic Converter Monitoring the catalytic converter (Figure 16.62) is one of the most important OBD II tasks. The catalytic converter is displayed as defective when the hydrocarbon emissions exceed a specific threshold in the U.S. FTP75 smog check. The respective threshold depends on the model year and the emissions rating of the vehicle. When the diagnostic threshold is exceeded, the catalytic converter is displayed as defective. For non-low-emission vehicle (LEV)-certified vehicles, the diagnostic threshold is 1.5 times that of the hydrocarbon emissions threshold in the U.S. FTP75 smog test. For transitional LEVs from model years 1996 and 1997, the diagnosis value is twice the exhaust gas threshold. For vehicles starting in model year 1998 and vehicles that are certified according to the low and ultralow emission thresholds, the diagnostic threshold is defined as 1.75 times the emissions threshold. Based on the definition of the diagnostic thresholds, particularly low diagnostic thresholds result for vehicles that are certified according to the strict low emission and ultralow emission thresholds. For example, the maximum permissible HC emissions in the smog test for an ultra-LEV (ULEV) vehicle are 84% lower than for a vehicle not categorized as a low emission vehicle. There are several processes for monitoring catalytic converters that all exploit the oxygen storage capacity of the catalytic converter. This storage ability correlates with the hydrocarbon conversion in the catalytic converter. Even a slight drop in the conversion rate leads to a clear reduction in the oxygen storage capacity of the catalytic converter. The oxygen storage in the catalytic converter can be detected with a lambda sensor. In addition to the sensor in front of the catalytic converter, a second is installed behind the catalytic

Evaluation of signal amplitude upstream of catalytic converter

A

upstream of cat.

Catalytic converter

l - Sensor downstream of catalytic converter

In addition to monitoring these systems, there is a standardized fault light control and a standardized tester interface from which the fault memories can be read in the workshop.

Evaluation of signal amplitude downstream of catalytic converter

A downstream of cat.

Evaluation of amplitude ratio

AV =

A downstream of cat. A upstream of cat.

Figure 16.62 Catalytic converter monitoring.

Internal Combustion Engine Handbook | 691

6606_Book.indb 691

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

•• The engine management system first determines the signal amplitudes of the lambda sensors before and after the catalytic converter. Then, a quotient of the amplitudes is calculated. This amplitude ratio is used to evaluate the conversion rate of the catalytic converter. •• When the conversion rates are low, there is an average amplitude ratio of nearly 1. As the conversion rate rises, the ratio decreases. For transitional LEV vehicles and vehicles not categorized as LEVs, this method represents a reliable way to diagnose catalytic converters. In vehicles that are certified according to the strict LEV and ULEV thresholds, worsening of the conversion rate of a few percent causes the diagnostic threshold to be exceeded. With these conversion rates, however, relatively low amplitude ratios are detected. It is very difficult to reliably distinguish between a defective and a functioning catalytic converter based on the amplitude ratio for these vehicles, especially considering the divergence within the product line. To diagnose the catalytic converter efficiency of LEV and ULEV vehicles, a series of new methods have been developed. Let us consider two. Most future LEV and ULEV vehicles will have a preliminary catalytic converter close to the engine in addition to the main catalytic converter. This preliminary catalytic converter has a relatively small volume, which, along with the fact that it is close to the engine, allows the operating temperature to be reached quickly and, hence, enables good exhaust conversion after a cold start. One way to diagnose these catalytic converter systems is to monitor just the oxygen storage capacity of the preliminary catalytic converter with a downstream lambda sensor (Figure 16.62). The assumption is that the preliminary catalytic converter ages much more quickly than the main catalytic converter. Because the volume of this catalytic converter is relatively small compared with the main catalytic converter, its maximum permissible drop in efficiency is much greater.

Initial measurements show that the value to be diagnosed is a 30%–50% efficiency loss. A problem with this method is that the efficiency loss of the preliminary catalytic converter must correlate directly with the efficiency loss of the overall catalytic converter system. The suitability then strongly depends on the configuration of the catalytic converter system (Figure 16.63). 25

O2 - storage capacity [µmol/g]

converter, and the signals from the sensor behind the catalytic converter are compared with the signals from the sensor in front of the catalytic converter. With today’s standard lambda sensors, there is a low sensor voltage for lean mixtures and a high voltage for rich mixtures. Because of the design of the binary lambda control, there are rich/lean jumps in the sensor voltage with a relatively constant amplitude in the lambda sensor before the catalytic converter at λ = 1. With linear lambda regulation, greater forced excitation is used for catalytic converter diagnostics. In a new catalytic converter with a relatively high oxygen storage capacity, these control fluctuations are strongly suppressed, as shown by the sensor signal behind the catalytic converter. A worn-out catalytic converter, as shown above, has a diminished storage capacity, so that the control fluctuation in front of the catalytic converter influences the sensor following the catalytic converter. The basic procedure for diagnosing the catalytic converter is as follows:

20 15 10 5 0 40

50

60

70

80

90

100

HC conversion [%]

Figure 16.63 Correlation between the oxygen storability and the HC conversion.

In the second method, temperature sensors are before and after the preliminary catalytic converter in addition to the lambda sensors before and after the overall catalytic converter system. These additional sensors monitor the starting behavior and conversion in the preliminary catalytic converter. This exploits the effect that the reactions in the catalytic converter are exothermic, which increases the exhaust gas temperature following the catalytic converter. The increase in temperature, therefore, correlates with the efficiency of the catalytic converter. A disadvantage of this method is that, in addition to the second lambda sensor, precise and, hence, relatively expensive temperature sensors must be used. An overview of the diagnostic methods is shown in Figure 16.64.

16.9.6 Safety Strategy

The law requires protective devices for systems that endanger life and property. When these devices fail, the remaining risk must lie below a tolerable threshold. In complex systems with software, these are termed protective functions. In addition, the law states that safety-relevant systems must be state of the art. For the so-called “drive-by-wire” engine management systems in which the throttle valve is not actuated directly via a Bowden cable but via an electrical drive independent of the gas pedal, a safety strategy is required in the engine management system. This system eliminates hazardous situations for the driver. Such situations can include undesired acceleration (stepping on the gas), that is, undesired starting of the vehicle, or an increase in engine speed. No engine output or only a low engine output is defined as a safe state. A distinction is drawn between individual errors (that is, only one error) and multiple errors.

692 | Internal Combustion Engine Handbook

6606_Book.indb 692

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

16.9 Functions

Monitoring

Technical solution

Catalytic converter system

Comparison of signal amplitude of the sensors up- and downstream of the catalytic converter for LEV and ULEV vehicles and determination of the actuation temperature Determination of regulating frequency, signal range and thermal resistance, superimposed regulation with sensor downstream of catalytic converter

Combustion misses

Calculation of the ride disturbance from the angle speed of the crankshaft

Tank ventilation system

Vacuum check of the tank system

Exhaust gas recirculation system

Determination of the intake manifold pressure with active EGR

Secondary air system

Monitoring the I - sensor signal

Fuel system

Monitoring the I - regulation value

The engine management system must independently recognize individual errors and then be able to restore the vehicle to a safe state within 500 ms. Limited engine output is permissible that allows a “limphome” function. In the case of multiple errors, it is permissible to include the reaction of the driver, for example, brake actuation. To attain this goal, comprehensive changes are required in the engine management system: •• The pedal travel is detected by two independent position sensors.

Figure 16.64 Overview of selfdiagnosis tasks.

•• The throttle valve position is detected by two independent position sensors. •• The engine ECU contains a monitoring unit (usually a second processor) that operates independently of the main processor. •• The engine ECU contains extensive safety functions. The safety functions (Figure 16.65) are divided into several levels that realize different monitoring tasks. A distinction is drawn between the following levels:

Main processor ...

DRI

EMS/ETC funktion

ADC Check

Ignition angle, ...

ECU

DRI

Speed limitation

or

Function and process monitoring

Injection, ... Mot

Disable

λ-sensor

Copy of process monitoring Processor monitoring program execution, instructions, storage Question

Answer

Reset

Processor monitoring

ADC

Monitoring unit Comparison Function (level 1)

Copy of process monitoring (level 2’)

Process monitoring (level 2)

Processor monitoring (level 3)

Figure 16.65 Safety strategy in the engine management system.

Internal Combustion Engine Handbook | 693

6606_Book.indb 693

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 16 Electronics and Mechanics for Engine and Shift Control Transmission Management

•• Level 1: Functions to control and regulate the engine including translating the position of the gas pedal into a throttle valve opening angle. •• Level 2: The process monitoring checks the control and regulation process of the engine (Level 1) with a focus on all functions that could undesirably increase torque in case of an error. •• Level 2¢: A copy of the code of the process monitoring system that is required for Level 3 to function. •• Level 3: Processor monitoring. Level 3 is executed partially by the CPU and partially by the monitoring unit. The monitoring module monitors the program’s execution, the instruction sets, and the memory area. In addition, the monitoring unit gives the CPU arithmetic tasks and checks the function of the CPU by monitoring the responses. In addition, the monitoring unit also detects an analog input signal and makes it available for a plausibility

check in the CPU, which supports the monitoring of the AD converter of the CPU. The process monitoring on Level 2 is designed so that some of the same functions as on Level 1 are calculated, but not using the same data. To obtain consistency, the process monitoring must be precise. This leads to a redundant set point of the reference variable (for example, induced torque). In addition, the actual value of the reference variable is calculated on Level 2 and compared with the redundant calculated set point. Actual values that are too high are, thereby, recognized, and the corresponding error reactions are introduced that put the vehicle into a safe state.

Bibliography

16-1. Braunschweig, M. and Czarnecki, T. 2004. “On-Board-Diagnose bei Diesel-motoren.” MTZ 65, 7 and 8: pp. 552–557.

694 | Internal Combustion Engine Handbook

6606_Book.indb 694

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17 The Powertrain The integrated starter-motor/alternator is examined in this chapter because, in the future, it will play an important role in the conception of the powertrain.

The powertrain’s functions complementary to the combustion engine take the form of

17.1 Powertrain Architecture

•• possibly, the additional importation or recovery of electrical motive force using an integrated starter-motor/alternator, a battery, and/or large electrical capacities

Transmission of engine power output from the crankshaft to the drive wheels makes this output actually effective for the driver in the form of vehicle acceleration and deceleration. The torque-transmitting elements of an automobile powertrain are the following (Figure 17.1): •• the engine •• an integrated starter-motor/alternator (ISG) (possibly) •• the gearbox, consisting of an initial movement element (a clutch) and the actual speed-reduction gearing system •• a power divider gearing system in the case of four-wheel drive (possibly) •• the final drive gearing system(s) (the differential, possibly slip-controlled).

Front-axle differential

Getriebe Transmission

•• balancing engine behavior and vehicle traction requirement (friction or geometrical-locking transmission elements such as a dual-shaft multistage transmission, a continuously adjustable variator, or a torque converter) •• reduction of engine rotation irregularities (e.g., damper elements in the clutch, a multimass flywheel, or a slipcontrolled torque converter) •• distribution of power to the drive wheels (by distribution of torque between the front and rear axles, for example, and by subdifferentials between the left and right sides of the vehicle).

Rear-axle differential

Integrated starter/ alternator (ISG)

Engine Motor

•• initial movement (achieved with elements such as a clutch or a torque converter, or also with an integrated starter-motor/ alternator)

Four-wheel drive: Divider gearing

Kupplung Clutch

Figure 17.1 Automobile powertrain.

Internal Combustion Engine Handbook | 695

6606_Book.indb 695

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

PE nE ME PA Axle drive nA MA

4.

XX

“Ideal” tractive force map

3. 2.

1.

Transmission output torque MA

MA/ME

Engine torque MM = ME

Engine

“Ideal” tractive force map

1.

2. 3. 4. Gang

Engine speed nM = nE

e

Speed ratio nA/nE

Transmission output speed nA

Figure 17.2 Functions of an automobile transmission: matching power requirement and engine output [17-1].

The transmission, in particular, combines the functions of an initial movement element and those of a power adjuster. In the latter case, the gearbox (in combination with the differential) adapts the engine operating characteristics to a significantly larger range of the torque/speed requirement at the gearbox output shaft (Figure 17.2, [17-1]).

This equation can be used to determine the vehicle acceleration capability and maximum vehicle speed (at avehicle = 0), as well as to calculate instantaneous climbing resistance in the context of real-time evaluation.

17.3 Transmission Types Transmissions can be classified into the following types by the design of their transmission elements:

17.2 The Longitudinal Dynamics of the Vehicle If the mass of a vehicle is concentrated at a single point in a model simulation, the acceleration and braking of this vehicle are then derived from the so-called “vehicle resistance equation” (moments listed here refer to summarized half-shafts): mvehicle avehicle = 1/rwheel [iges Mmotor effective − Mvehicle resistance]



(17.1)

where: Mvehicle resistance = Mrolling resistance + Mclimbing resistance + Mair resistance + Mbrake (17.2) ↑ Vehicle mass, rolling resistance coefficient (road surface, tire characteristics)

↑ Vehicle mass, gradient

↑ Vehicle speed, air resistance (air density, bulkhead surface area, air resistance coefficient cW)

Where mvehicle = vehicle mass avehicle = vehicle acceleration rwheel = dynamic tire radius itot

= overall transmission ratio

M

= moments

•• multistage transmissions as distinct from continuously variable transmissions •• transmissions with natural axial eccentricity between the input shaft and the output or takeoff shaft (dual-shaft transmission) as distinct from transmissions with axial arrangement of the input shaft and output shaft (inline transmissions). Multistage transmissions are based on geometrically locking transmission elements (e.g., sets of helical-toothed spur gears and planetary gears), whereas continuously variable transmissions are generally based on friction-locking functional principles. This friction-locking function necessitates additional auxiliary energy, with the result that continuously variable transmissions generally have poorer internal gearbox efficiency. These gearboxes balance out this disadvantage within the overall powertrain by their capability of adapting the engine working point optimally to road situations. A further differentiating factor in automobile transmissions is their level of automation. Manually actuated (“stick-shift”) transmissions (dual-shaft type), for instance, continue to play a significant role in Europe, whereas electrohydraulically actuated automatic transmissions (generally of planetary type) predominate in the United States and Asia. These types are increasingly being augmented with automated dualshaft transmissions

696 | Internal Combustion Engine Handbook

6606_Book.indb 696

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.3 Transmission Types

•• based on an (electric motor or electrohydraulically driven) dry clutch (automated stick-shift transmission) •• based on a hydraulically actuated dual clutch (dual-clutch transmission) and a continuously variable flexible drive transmission mechanism based on a so-called “push belt” or a chain.

Figure 17.3 shows a summary of various transmission types and a number of their characteristic features. Figure 17.4 shows the usual torque ranges for automobile applications. Figure 17.5 shows an example of the Daimler Chrysler W5A 580 five-speed multistage transmission, which incorporates a slip-controlled torque converter and three sets of planetary gears. This transmission is used in nearly all Mercedes-Benz standard (rear-wheel) drive automobiles.

Transmission type

Abbreviation

Ratio

Weight

Noise

Consumption1

Gearshift comfort (ATZ value)2

Manual transmission (5-speed)

5MT

Dual-shaft transmisstion

Low

Low

−10.0 %

—

Manual transmission (6-speed)

6MT

Dual-shaft transmisstion

Low

Low

−12.0%

—

Automatic multistage transmission (5-speed)

5AT

Set of planetary gears

Medium

Low

−0.0 %

9

Automatic multistage transmission (6-speed)

6AT

Set of planetary gears

Medium

Low

−3.0%

9

Continuously variable transmission

S-CVT

Flexible transmission mechanism (based on “push belt”)

High

Medium

−5.0%

9.5

Continuously variable transmission

K-CVT

Flexible transmission mechanism (chain basis)

High

Medium

−5.0%

9.5

Toroidal drive

T-CVT

Friction wheel transmission

Very high

Low

−7.0%

9.5

Automated manual transmission

E-AMT

Dual-shaft transmission with electromechanical actuation

Low

Low

−15.0%

6.3

Automated manual transmission

H-AMT

Dual-shaft transmission with electrohydraulic actuation

Low

Low

−14.0 %

6.5

Dual-clutch transmission

DCT

Dual-shaft transmission with electrohydraulic actuation

Medium

Low

−8.0 %

8.7

Remarks: 1 Approximate consumption average compared to a 5-speed automatic multistage transmission at 300 Nm operation and ungoverned SI engine. 2 The ATZ value is a measure of the quality of the change of the transmission ratio. An ATZ value of 10 indicates an optimum (completely smooth) change of the transmission ration, while a value of 1 indicates an extremely rough transition.

Figure 17.3 Comparative assessment of a number of transmission types [17-2], [17-3], [17-4].

Torque range in Nm

Transmission comparison 900 800 700 600 500 400 300 200 100 0 5MT

6MT

5AT

6AT

S-CVT K-CVT T-CVT E-AMT H-AMT Transmission type

DCT

Figure 17.4 Normal torque ranges for automobile gearboxes.

Internal Combustion Engine Handbook | 697

6606_Book.indb 697

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

1

11 1 2 3 4 5

Torque converter Oil pump Drive shaft Multi-disk brake B1 Clutch K1

2

12 6 7 8 9 10

3

5

4

13 Clutch K2 Multi-disk brake B3 Clutch K3 Multi-disk brake B2 Output shaft

14

6

15 11 12 13 14 15

16

7

17

8

18

9

19

Parking lock gear Intermediate shaft Overrunning clutch F2 Rear planetary gearing Medium planetary gearing

10

20 16 17 18 19 20

Electric-hydraulic control unit Front planetary gearing Overrunning clutch F1 Stator shaft Torque converter clutch

Figure 17.5 DaimlerChrysler W5A 580 five-speed multistage transmission.

17.4 Power Level and Signal-Processing Level 17.4.1 Power Level (Figure 17.6)

This is the level of the actual torque-transmitting components.

17.4.2 Signal Level (Figure 17.6)

Control and regulation of the entire powertrain is based on physical models of the individual components, which are functionally integrated by a torque-based model concept (starting from wheel torque and proceeding up to engine and transmission management).

17.4.3 Links

Modern powertrain architectures are characterized by a clear vertical correspondence between the power level and signal level components and their links. These links exist at the power level in the form of torque-transmitting shafts and in the form of communications channels at the signal level.

17.5 Transmission Management 17.5.1 Functions 17.5.1.1 Overview The following functional groups can be defined for all transmission concepts: •• Gearshift strategy: Determines which target transmission ratio or which gear is selected •• Ratio transition: Management of the actual change of the transmission ratio •• Diagnosis functions: Achievement of a safe condition in case of component failure or emergency operation •• Special functions, such as control of the converter override function, gearshift lever disabler magnet actuation and safety concept in “shift-by-wire” drive systems. These relationships are shown using the example of the automated manual transmission (AMT) in Figure 17.7. The driving strategy in this context determines not only the target

698 | Internal Combustion Engine Handbook

6606_Book.indb 698

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.5 Transmission Management

Signal level

CAN vehicle network

Engine control

Starter motor/ alternator control unit

42 V battery

Transmission management

Power level

Starter/alternator Engine

Clutch

Clutch

Transmission

Figure 17.6 Powertrain power and signal levels.

17.5.1.2 Driving or Gearshift Strategy Electronic transmission control systems, which for the first time enabled the driver to manually select a number of individual gear-changing programs such as “Economy,” “Sport,” and “Winter,” were introduced during the 1980s. This proved not to be an optimum solution, however, since it at all times necessitated manual intervention by the driver in order to adjust the vehicle’s gear-changing behavior to the road situations occurring. Since, in addition, it could not be

gear in automatic mode, but also checks manual gearshift commands by the driver (known, for instance, in the form of the so-called “Tiptronic” and “IntelligenTip®” systems). The subordinate level is responsible for the initiation and overall coordination of the gearshift sequence and, therefore, in the case of the AMT, for control of the engine (torque and speed), clutch torque and logical gear position. The “actuator control” level is responsible for regulation of the corresponding physical manipulated variables (path, pressure, angle, etc.).

Driver interface shift lock

Engine control system

Driving strategy

CAN messages from engine

Confirmed gear

Target gear

Shift process coordinator

CAN CAN messages to engine

Actuator control Clutch

Gear

BIOS/VIOS

IO-HARDWARE (sensors, actuators, switches)

Figure 17.7 Functional groups.

Internal Combustion Engine Handbook | 699

6606_Book.indb 699

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

guaranteed that the driver would actually make a manual selection in every situation, it was ultimately necessary to make compromises in the case of these diverse gear-changing programs, too. For this reason, so-called “intelligent” driving and gearshift strategies, and gearshift strategies that automatically set the correct priorities on the basis of the prevailing conditions, have nowadays become an integral component of every automatic transmission system. SAT (Siemens adaptive transmission control), (Figure 17.8) the Siemens driving strategy for automatic multistage transmissions, is in successful use with a range of vehicle manufacturers, vehicle classes, and driving styles [17-5]. Both global strategy criteria and short-term road situations are defined: •• Driver-type recognition •• Environment recognition: Road gradient •• Adjustment to low-adhesion conditions (ice) •• Manual intervention (IntelligenTip) •• Fast-off detection: Suppression of an upward shift if gas is reduced quickly (indicating the driver’s intention to slow down) •• Bend recognition: Prevention of upward gear changes as a function of lateral acceleration •• Braking response: Additional down-changes if the brake is actuated, taking account of engine speed limits and the road situation. A special feature of the SAT system is its comprehensive use of fuzzy logic, with 30 to 40 rules, depending on the expansion level. This permits achievement of high adaptation logic and dynamics. Thanks to an online learning component,

conceptual future solutions, such as IntelligenTip [17-6] offer the drivers more freedom to generate their own personal preferences in gearshift strategy (Figure 17.8). 17.5.1.3 Automatic Transmissions with Planetary Gears and Torque Converter Present-day conceptual solutions give preference to direct single clutch management (Figure 17.9) using electrohydraulic valves. This eliminates the need for a demultiplexer in the hydraulic system; the individual clutch pressures are calculated in the control unit software. This is also in line with the trend toward replacement of hydraulic functions by software functions, to achieve cost savings. Gear changing with the clutch management system is also made possible in principle. 17.5.1.4 Automated Manual Transmission The basic structure has already been explained in Figure 17.7. Unlike automatic transmissions with planetary gears, explicit gear management is necessary for the dual-shaft transmission with automatic synchronization. The entire gearshift sequence is, thus, more subject to sequential control; that is, clutch operation and gear-changing follow one another. 17.5.1.5 Continuously Variable Transmissions (CVT) In the CVT (Figure 17.10), the variator and the individual thrust forces of the cones require continuous control (flexible drive CVT), whereas the gearshift sequence in multistage transmissions results in changing between discrete conditions. Principal attention is focused on minimization of thrust pressure to achieve the lowest possible fuel consumption and high-ratio change dynamics, but with dependable prevention of belt slipping. In addition, further functions control converter override and the planetary gearing for vehicle reversing.

FP gas pedal FP mean value FP activity

Fuzzification Control basis Inference and defuzzification

FP dynamics

Load value/increase

Vehicle speed

Driver class

Longitudinal acceleration

Shift prevention

Differential torque Lateral acceleration

Curve value SAT Fuzzy

Systeme Systems

Brake delay Brake time

Figure 17.8 Adaptive transmission control (Siemens).

700 | Internal Combustion Engine Handbook

6606_Book.indb 700

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.6 Integrated Powertrain Management (IPM®)

Gear input and output speed

Driver interface

Driving strategy Torque

Target gear

Converter control

Confirmed gear

Engine management

Target pressure

Shift Process Coordinator

Gear speeds

p Clutch 1

p

I

Current control

Clutch 2

p

I

Current control

I

Clutch 3

p

Current control

I

Current control

Figure 17.9 Direct individual clutch management.

17.6 Integrated Powertrain Management (IPM®)

the battery, as an energy storer, and the torque converter (transmission) on the basis of a holistic concept. Because of the many degrees of freedom in such a system, it is important to control and coordinate downstream units optimally on the basis of central driver-intention interpretation and road-situation detection, taking account of the high-level prioritization. Integration in the sense of IPM in this context covers the control and coordination of the entire system, but not design aspects such as system space requirement, installation, etc. (Figure 17.11).

Future powertrain concepts consist of a number of individual subsystems: The engine, an electrical machine, and, in many cases, automated transmissions. IPM (integrated powertrain management) [17-7] has no direct influence on the process of conversion of the energy stored in the fuel (gasoline, diesel fuel, gas, or hydrogen), but instead attempts to optimize the working points for the energy converters (the engine and/or the electrical machine), „+,–“ Gear input and Gas pedal, kick-down Tip output speed

Gear input and output speed

Torque

Driving strategy

Translation regulation and thrust force control

Selector

Remaining gear control (reverse set)

Converter control

Target translation Cone washer speeds

Gear input and output speed

Torque

Target pressure Torque

Current control

p

I

Current control

Target pressure p

I

Current control

PWM-modulated current outputs

Figure 17.10 Function diagram of the continuously-variable transmission (CVT).

Internal Combustion Engine Handbook | 701

6606_Book.indb 701

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

Integrated Powertrain Management IPM®

Battery Management

Transmission Management System TMS

Inteligentip® IT

Siemens Adaptive Transmission SAT Gear ratio control

Integrated Starter Generator ISG

Engine Management EMS

Figure 17.11 Diagram of integrated powertrain management.

An important feature of this concept is the introduction of a higher control level superimposed on the “component control systems.” This guides the torque generator and/or converter through the relevant states and optimizes energy flow. Integrated powertrain management is subdivided into three levels (Figure 17.12): •• Level 1 consists of driver and road-situation detection. Driver recognition includes instantaneous driver intention interpretation and driver-type classification. The state of the powertrain (propulsion, braking, start/ stop, coasting, boosting, recuperating, etc.) is determined at the second level on the basis of signals from Level 1 and other vehiclesensor data. •• Level 2 is referred to as state management in the powertrain and performs the task, on the basis of the inputs from Level 1, of adjusting the powertrain to the state that fulfills the currently prioritized optimization criteria. •• Level 3 supplies target data for the downstream units on the basis of physical variables. The engine and the electrical

Gas pedal Brake pedal

machine can, for example, each be operated at a defined working point by specifying a target torque or a target speed; a target transmission ratio is specified for the transmission system. In addition, certain driving conditions also access the clutch (so-called “freewheeling,” for example, disengaging the clutch if coasting in gear).

17.7 Components for Powertrain Electricification 17.7.1 Overview

With increasing demands on the efficiency of vehicle drivetrains and simultaneous emissions reduction, the electrification of the powertrain assumes an essential role. In the combustion engine, the chemical energy of the fuel is converted into mechanical energy and thus, a torque is generated

IPM® Driver request interpretation

IPM®/ SAT Driver and driving situation detection

IPM® State control (start-stop, boost, coasting, recuperation, etc.) IPM ® Torque management Engine torque calculation

BATTERY (12V/36V)

ISG engine torque calculation

ENGINE MANAGEMENT

ISG MANAGEMENT

Adaptive gear selection

TRANSMISSION MANAGEMENT

Figure 17.12 Integrated powertrain management levels.

702 | Internal Combustion Engine Handbook

6606_Book.indb 702

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.7 Components for Powertrain Electricification

via the pressure. It serves to accelerate or maintain the speed of the vehicle, that is, to maintain or increase kinetic energy. In the thrust phase or during braking, this energy is converted mostly into friction and thus heat. The heat is discharged into the environment and thus lost for driving the vehicle. If so structured, an electric drive can work as both, a starter and an alternator. The combination of the combustion engine with an electric powertrain thus offers the option to, during the vehicle delay, recovering (recuperation) at least some of the energy previously used to increase the kinetic energy and to use if in the next acceleration process, either in combination with the combustion engine or only by the electric drive. The combination provides a significant improvement in the driving dynamics as well. Thus, depending on the type of hybrid drive, both drives, combustion engine and electric motor, can be simultaneously used for the vehicle acceleration (boost), without higher consumption or emissions.

17.7.2 Hybrid and Electric Drive Variants

Various options are possible for the combination of the combustion engine and the electric motor. Both drive systems can be arranged functionally in parallel (parallel hybrid), that is, both can drive simultaneously or separately. They can be also arranged functionally in series, that is, only one system actually drives the drive gears while the other feeds its drive energy serially (serial hybrid) through that drive that directly drives the gears. A further differentiation of hybridization is the capacity or the scope of functionality. The most commonly used hybrid variants are shown below. 17.7.2.1 Microhybrid The microhybrid is usually just a pure stop/start function, that is, the motor is shut down when not required and automatically restarted when needed. By controlling the charge voltage in dependence to the driving state, these systems usually offer a minimum recuperation function. These systems can be represented, for example, with starters and motor control devices modified accordingly, or as a belt-driven starteralternator (designed like a mild hybrid). 17.7.2.2 Mild Hybrid Compared to the microhybrid, the mild hybrid also features a significant recuperation and boost function. Conventional technology (12 V/14 V) is usually no longer sufficient and the specially developed components are used, with operating voltages significantly above the current vehicle on-board systems. 17.7.2.3 Full Hybrid An essential function of the full hybrid, in addition to stop/start, boost and recuperation, is the temporary, purely electric driving. The drivetrain is designed so that the electric motor can drive the vehicle independently of the combustion engine. The required energy is attained by recuperation and stored in the hybrid battery. This requires that the entire electric drive system including energy storage (hybrid battery) must

be significantly more powerful. In addition to the traditional parallel hybrid, we now have the power split concept, in which usually two electric motors are coupled with the combustion engine via a planetary gear.

Battery

Transmission Clutch Electric motor

Inverter

Combustion engine

Figure 17.13 Mild hybrid schematic.

Transmission Battery

Clutch Electric motor Clutch

Inverter

Battery

Combustion engine

Transmission Electric motor

Double inverter

Combustion engine

Figure 17.14 Full hybrid/plug-in-hybrid schematic (parallel hybrid left/ power split right).

17.7.2.4 Plug-in Hybrid The hybrid battery in plug-in hybrids is much larger and, consequently, more powerful. The goal is to achieve a significantly larger range with the electric drive, compared to the full hybrid. As the name indicates, the hybrid battery can be recharged at the socket. As a rule, a plug-in hybrid primarily uses the electric drive and the combustion engine activates only when the battery charge drops below a defined value. A variant of the plug-in hybrid is the range extender which is a major step into the direction of a true electric vehicle. The combustion engine is much smaller and not directly connected to the drive gears. It has the task to, in combination with a small generator, recharge the battery at a critical charge level. The objective is to increase, as required, the range possible with the battery. In this case, it is designed as a serial plug-in hybrid. 17.7.2.5 Electricand Fuel Cell Vehicles An electric vehicle is defined as a vehicle that is driven purely by an electric drive system without containing a combustion engine.

Internal Combustion Engine Handbook | 703

6606_Book.indb 703

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

The required electric energy is taken from a battery (traction battery) or a fuel cell.

Inverter

Transmission Electric motor

Battery Figure 17.15 Electric vehicle schematic.

17.7.3 Components

Electric drive systems as they are currently used in hybrid and electric vehicles consist of a three-phase AC motor, threephase AC inverter and an energy storage system. To supply the 14 V on-board system, a DC/DC converter (three-phase AC converter) is used providing energy from the energy storage of the electric drive system. Three-phase AC inverter and DC/DC converter are frequently combined to one unit and usually called power electronics.

17.7.4 Power Electronics

The name power electronics is used because unlike common control devices, very high currents (450 Arms for the inverter) at high voltages (beyond 700 V) are switched for high output. 17.7.4.1 Three-Phase AC Inverter The electric motors currently used in hybrid and electric vehicles are three-phase AC drives that are activated via three-phase AC, while the energy storage system supplies regular DC. In motor operation, that is, driving or boosting, the inverter converts the DC from the energy storage system into threephase AC for the electric motor. In the reverse case, which is for generator operation, the AC generated by the electric motor is converted into DC. This event of recovery of kinetic energy is also called recuperation. Core components in the converter are the power module converting the DC into a sinusoidal three-phase AC and the control board as the control unit. The control unit in the converter regulates and monitors the adjustable target and limit values for torque, speed, currents and voltages. The losses occurring in the power module are mostly the switching and conduction losses in the power semiconductors. They increase with the switching frequency, the battery voltage and the phase current and cause the semiconductor elements to heat up. Of particular importance is here a good thermal connection of the power semiconductors to the cooling medium and the user of cooler and substrate materials with coefficients of expansion as similar as possible to avoid thermal stress. This thermal stress is the main cause for aging in power electronics modules and, conversely, the decisive factor for maximum switching capacity.

Figure 17.16 Power electronics (inverter and DC/DC in one housing).

Both the cooling water temperature level and the temperature fluctuations directly occurring at the power semiconductors affect the service life of the power end stage. 17.7.4.2 DC/DC Converter The electric drive system, including the energy storage system, work at voltages that are in part significantly above those used in modern on-board systems. Because the available output of the electric drive system is significantly higher than the one of modern generators, it seems reasonable to supply the on-board network from this areas as well, which is, actually, necessary in electric-only vehicles.

Figure 17.17 Power module for direct water-cooling with power semiconductors and cooler (cooling fins visible in the mirror image).

This provides the option to completely omit traditional generators resulting in more space and cost optimization. The DC/DC converter may be integrated with the inverter in a single housing or installed in a separate housing. The actual decision is mostly driven by the space available in the vehicle. An advantage of the integration on a single housing

704 | Internal Combustion Engine Handbook

6606_Book.indb 704

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.7 Components for Powertrain Electricification

is the possibility to further optimize costs by using only one housing or, for example, a single processor for both functions.

Figure 17.20 illustrates possible installation positions of electric engines in a drivetrain. In addition, for each of the positions, a number of different variants is possible which adds to the total number of possibilities.

P0

P1 P2

P3

P4

P = Position E motor K0

K1

K = Clutch

Figure 17.20 Possible installation positions for hybrid engines.

Figure 17.18 Power semiconductor (IGBT) on copper—DCB.

A positive side effect is the significantly higher efficiency of the DC/DC converter compared to a traditional generation.

17.7.5 Electric Motor

17.7.5.1 Technologies Today, mostly, three types of electric motors are used: asynchronous motor (ASM), permanently-excited synchronous motor (PSM) and the separately excited synchronous motor (SM). The relevant literature provides more details about these technologies. For this reason, the advantages and disadvantages of their use in hybrid and electric vehicles will only be discussed shortly here.

Depending on the strategy and application of the vehicle, different numbers of electric motors and technologies are used. The goal is to use an optimized technology for the specific application.

17.7.5.1.1 Asynchronous motor The ASM distinguishes itself by its simple design and ruggedness. Compared with PSM and SM, the simple design is also less expensive. Unlike the PSM, the excitation can be switched off which means that it can be controlled more easily in the event of a fault.

Figure 17.19 DC/DC converter in separate housing.

Figure 17.21 Laterally-installed ASM.

This may be determined by either the system costs, the available space, the required functionality but also by the efficiency for example.

Its disadvantage is the low efficiency in specific operating ranges and, compared to the PSM, larger space requirement. Today, it is most frequently used as a laterally installed variant in mild hybrid systems (Figure 17.20: P0).

Internal Combustion Engine Handbook | 705

6606_Book.indb 705

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

17.7.6.1 Overview Various technologies are today used as energy storage for hybrid applications (with the exception of the micro-hybrid). This includes, among others, double-layer capacitors (DLC) combined with lead-acid accumulators, NiMH (nickel-metal hybrid) and Li-Ion ((lithium-ion) batteries.

Figure 17.22 PSM for integration in the bell housing.

17.7.5.1.3 Separately-excited synchronous motor Because the electric engine is no longer directly places in the powertrain in final drive systems (Figure 17.20: P4 or in electric vehicles), the engine position as a selection criterion is no longer of its usually high importance. Hence, engines with typically low realizable axial lengths (such as PSM) loose one of their crucial advantages. At the same time, the SM may not offer the highest best operating point (as does the PSM) but neither does it have significant weaknesses (as does the ASM) and gains in the area of low output and high speed. Looking at the space requirements, it has a minor disadvantage compared to the PSM due to the space needed by the slip ring system which, however, allows for a simple excitation control in the event of a fault.

17.7.6 Energy Storage Systems

The output capacity of hybrid or electric drive systems depends, amongst other things, on the capacity of the energy storage system. It plays an important role in the fuel-saving potential of hybrid applications and the range of electric vehicles. At the same time, the currently demanded service life of an energy storage is 10 to 15 years and between 160,000 and 240,000 km, that is, the time duration as the vehicle service life.

Figure 17.23 Axle drive with integrated gear.

In the course of the long-time development of numerous electric energy storage systems, secondary battery systems more and more moved into the foreground, that is, battery systems enabling the repeated regeneration of their chemical energy by recharging them after depletion. The essential innovation on the course from the lead battery (with aqueous electrolyte) to modern battery systems such as lithium ions (with organic electrolytes) was the principle of intercalation of ions in the electrodes (Figure 17.24). While the electrodes in a lead battery are built and broken down due to chemical reactions (conversion) and interact with the electrolyte, the lithium in lithium-ion cells is intercalated in solid lattices. The lattice structure is retained and considerably contributes to the service life over charge and discharge cycles.

Service life, energy density

17.7.5.1.2 Permanently-excited synchronous motor PSM offer the major advantage of smaller axial lengths thanks to concentrated windings. This allows the integration in the bell housing without major or no powertrain extension. This variant is called “transmission-integrated” (Figure 17.20: P1/P2, optional P3). Another advantage is their punctually high efficiency. However, the down sides are significant drag losses at high speeds and the comparably high effort protection in the event of fault.

Figure 17.24 Development of battery systems from conversion in lead-acid to intercalation in nickel-metal hydride and Li ions. See color section page 1091.

Nickel-metal hydride (NiMH) batteries are now established in the first-generation hybrid vehicles. For the next generation, the use of Li ion batteries is very probable.

706 | Internal Combustion Engine Handbook

6606_Book.indb 706

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.7 Components for Powertrain Electricification

Safety

Lead-acid

- Low costs but reduced service life expectation - Average overall performance

Costs

Service life

Availability

Output

DLC

- High service life expectation - High costs, low energy retention

NiMH

Process capacity

Cold Start

- Good overall performance

Li-Ions

Environment Lead-acid

DLC

Energy NiMH

Li-Ions

- Good overall performance, in conjunction with high energy densities

Figure 17.25 Assessment of various storage systems for the use in motor vehicles. See color section page 1091.

They exhibit further increased output and energy density, taking into account the required charge and discharge cycles.

and disadvantages relative to output or energy density and safety.

17.7.6.2 Battery System Due to their potential and the probable use in hybrid and electric vehicles, only li ion energy storage will be discussed below. A modern energy storage system on the basis of lithium ions comprises several components. In addition to the cells, that is, the actual energy storage, there are cell monitoring, contactors, switches, and a battery management unit

17.7.6.2.2 Safety Energy density is much higher in lithium-ion batteries, compared to lead-acid or nickel-metal hydride. For the safety issue in particular, it is, therefore, necessary to protect the entire energy storage system, using a number of procedures. As a rule, three levels are differentiated. On the cell level, a high degree of intrinsic safety can now be achieved. New cathode materials, such as lithium–cobalt– nickel–manganese oxide or lithium–iron–phosphate react less exothermically in the case of misuse or defect than the lithium–cobalt oxide conventionally used in consumer batteries.. If these cathode materials are combined with advanced separators with, for example, a ceramic portion and highboiling electrolytes, the safety at cell level can be increased considerably. Because Li-ion cells also allow the combination

17.7.6.2.1 Li-ion cells Lithium ions are to be understood as a generic term for a material combination. At this point in time, most cells are based on lithium–cobalt. Further developments are cells with new cathode materials, such as lithium–cobalt–nickel–manganese oxide or lithium– iron–phosphate. All these combinations have advantages

Figure 17.26 Schematic comparison of round cells (left) and prismatic cells (right).

Internal Combustion Engine Handbook | 707

6606_Book.indb 707

1/19/16 8:50 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 17 The Powertrain

System level Separator Cell level

Cell housing Electrolyte Electrode composition

Electrochemistry (cathode composition)

Lithium Iron Phosphate Lithium Cobalt Nickel Manganese oxide Al-stabilized Lithium Nickel Cobalt oxide

Temperature monitoring Voltage monitoring Current monitoring Cell symmetrization (Li-Ions) Insulation Fuses Contactors and switches Thermal management Figure 17.27 Safety levels on lithium-ion batteries.

of many more but not yet fully researched electrode materials and electrolytes, further increases in safety in combination with attractive energy and output densities are possible. In addition, various safety measures are today integrated on system levels in a battery. The battery is maintained in a safe operating state by constant monitoring the operating states, management of the battery system and targeted cooling. Today, a level of safety of these energy storages can be achieved that was practically unimaginable ten years ago and the systems survive crush tests as shown in Figure 17.28. The high demands on the safety of Lithium-Ion battery system are demonstrated by the numerous passed test scenarios listed in Figure 17.29. 17.7.6.2.3 Future prospects The cell is designed in dependence on the application of the energy storage system. For hybrid vehicles, the cells are optimized for performance, due to the relatively short time during which high performance is demanded during boost and or stored during recuperation. Cells used in electric vehicles and plug-in hybrids are optimized for energy retention. The issue here is to be able to demand energy as much or as long as possible for purely electrical driving. Overall, we are just at the beginning in the development of Li-ion batteries. Just a few facts to clarify: Today’s LithiumIon batteries achieve an energy density of 120 to 150 Wh/kg. Theoretically, 6,000 Wh/kg (Li flour) can be achieved and, in practice, values of 2000 Wh/kg are expected. Due to the clearly higher efficiency chain of the electric drive and the possibility to recuperate energy, one currently assumes that, at energy densities of just approximately 500 Wh/kg, ranges can be reached that are comparable to those of vehicles with combustion engines.

Figure 17.28 Example for the crush test of a lithium-ion cell.

708 | Internal Combustion Engine Handbook

6606_Book.indb 708

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

17.7 Components for Powertrain Electricification

No.

Test

Result

Controlled deformation Penetration with nail Drop test Plunge test Arcing simulation Mechanical shock Thermal stability Simulated fire High-temperature storage test Fast charging/discharging Thermal cyclization Overload/overvoltage Short-circuit Total discharge/undervoltage Figure 17.29 Test scenarios for Li-ion batteries.

Partial short-circuit

Bibliography

17-7. Siemens-VDO Automotive AG.

17-1. Mitschke, M. 1995. Dynamik der Kraftfahrzeuge, Band A, Antrieb und Bremsung. Berlin/Heidelberg/New York: Springer.

17-8. Schümann, U. 2007. Modulare Komponenten für wirtschaftliche Hybridantriebe.

17-2. Förster, HJ. 1991. Automatische Fahrzeuggetriebe, Grundlagen, Bauformen, Eigenschaften, Besonderheiten. Berlin/Heidelberg/New York: Springer.

17-9. Neumann, KT. 2008. Compendium ATZ—Schwerpunktthema Elektromobilität.

17-3. Lechner, G., Naunheimer, H. 1999. Fahrzeuggetriebe: Grundlagen, Auswahl, Auslegung und Konstruktion. Berlin/Heidelberg/New York: Springer, 1994 (Hinweis: Buch ist auch 1999 in Englisch erschienen: Lechner, G.; Naunheimer, H.: Automotive Transmissions: Fundamentals, Selection, Design und Application. Berlin/Heidelberg/New York: Springer.

17-11. Greul, R. 2007. Design of Power Electronic Components for Hybrid Drives.

17-4. Bock, C. 2000. Die ACEA-Vereinbarungen zur Flottenverbrauchsreduzierung und ihre möglichen Konsequenzen auf zukünftige Getriebekonzepte. Vortrag im Haus der Technik anlässlich der Tagung “CVT-Getriebe,” Essen. 17-5. Graf, F., Lohrenz, F., Taffin, C. 1999. Industrialization of a Fuzzy Logic Transmission Controller. VDI Tagung: Getriebe in Fahrzeugen‚ Friedrichshafen.

17-10. Greif, A. 2006. SIA Paris—Design of Power Electronic Components for Hybrid Drives.

17-12. Keller, M., Birke, P., Schiemann, M., Moehrstaedt, U. 2008. ATZelektronik—Lithium-Ionen Batterieentwicklung für Hybrid- und Elektrofahrzeuge. 17-13. Hackmann, W., Märgner, M., Kugland, O. 2009. HdT Fremderregte Synchronmaschinen als Achsantriebe. 17-14. Hackmann, W., Wagner, B., Zwingel, R., Dziedzek, I., Welke, K. Internationaler ETG-Kongress 2007 Karlsruhe—Fremderregte Synchronmaschinen im Einsatz als Achshybridantriebe.

17-6. Heesche, K., Graf, F., Hauptmann, W., Manz, M. 2001. IntelligenTip— eine trainierbare Fahrstrategie. VDI Tagung: Getriebe in Fahrzeugen‚ Friedrichshafen.

Internal Combustion Engine Handbook | 709

6606_Book.indb 709

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 710

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

18 Sensors 18.1 Temperature Sensors Most temperature measurements in the automobile utilize the temperature sensitivity of electric resistance materials with negative temperature coefficient (NTC). The strong nonlinearity enables a large temperature range to be covered (Figure 18.1).

1000000

NTC resistor temperature characteristics

Resistance [Ohm]

100000

type A type B type C type X2 type X1

10000 1000 100 10

–50

For applications with very high temperatures (exhaust gas temperatures up to 1000°C), platinum sensors are employed. The change in resistance is converted into an analog voltage by a voltage grading circuit with an optional parallel resistance to the linearization. The sensors are employed for the following temperature ranges:

1

0

50 100 temperature [°C]

150

200

Application

Temperature range

Intake/charge air

−40° + 170°C

Coolant

−40° + 130°C

Engine oil

−40°+ 170°C

Fuel

−40° + 120°C

Exhaust gas

100° + 1000°C

Figure 18.2 shows various forms of temperature sensors for oil, coolant and air temperature.

Figure 18.1 Typical characteristics of temperature sensors (NTC). See color section page 1091.

Figure 18.2 Various temperature sensor designs.

Internal Combustion Engine Handbook | 711

6606_Book.indb 711

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 18 Sensors

18.2 Fuel-Level Sensors The fuel-level sensor is used to monitor the oil level in combustion engines or transmissions. Today, continuous fuel-level sensors and fuel-level switches are used. Fuel-level sensors, together with a suitable electronic evaluation unit, provide a continuous signal that is proportional to the fuel level. Frequently, a thermoelectric system is used: Thermal conduction of a heated element depends on the filling height. The filling level is calculated from the supplied electric power. The filling-level switch enables the measurement of the filling-level limit. The functional principle used is a float with a magnet that actuates a reed contact or Hall switch. The filling level is usually processed and displayed in the combination instrument. In some cases, the oil-level information is also required by the motor controller.

knock sensors are installed at appropriate positions on the engine block where they can register the vibrations generated by the combustion process. In order to detect knocking for each individual cylinder, several knock sensors are employed on multicylinder engines (e.g., two sensors for six cylinders or four sensors for eight cylinders).

Float Magnet

Reed switch

18.3 Knock Sensors “Knock” is understood to mean an abnormal combustion in spark ignition (SI) engines caused by a spontaneous ignition of the mixture in the cylinder. This undesirable combustion results in a significantly higher mechanical load on the engine. Continuous knock causes damage or even destruction of the piston. When looking for settings ensuring an optimum combustion process, the zones of highest efficiency and of knock are very close together. Knocking generates vibrations with characteristic frequencies. These engine oscillations are registered by the knock sensor and transmitted to the engine controller where the signal is evaluated using corresponding algorithms to identify knocking. The engine controller then regulates the combustion process so that knocking no longer occurs (advancing the ignition timing by a few degrees). Furthermore, “knock control” also permits operation with different fuel qualities. Knock sensors are generally broadband knock sensors. These cover a frequency spectrum from, for example, 3 to over 20 kHz (with an intrinsic resonance of over 30 kHz). The

Figure 18.3 Design of a filling-level switch.

The functional principle of knock sensors is typically based on a piezoceramic ring that converts the engine vibrations into electrically processable signals using a superimposed (seismic) mass (Figure 18.4). Sensor sensitivity is expressed in mV/g or pC/g and is practically constant over a wide frequency range. As demonstrated in Figure 18.5, the transmission behavior of the knock sensor can be adjusted by selecting the proper seismic mass. The resonance frequency can be increased by reducing the seismic mass. Since the tolerance band of the sensitivity of such sensors is approximately ±30%, the use of limit specimen sensors (of the sensitivity tolerance band) is important when tuning the engine controller.

Nut Seismic mass Insulation ring Piezoring Contact ring

Substrate

22 mm

Figure 18.4 Cross section through a knock sensor.

712 | Internal Combustion Engine Handbook

6606_Book.indb 712

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

18.4 Exhaust Gas Sensors

600 550

Sensitivity [m/g]

500 450 400 Seismic mass 2 g

350 300 250 200 150 100 50 0 0

5

10

15

20

25

30

35

40

45

Frequency [KHz]

Knock sensors with integral plugs are increasingly being used. A typical design is shown in Figure 18.6.

Figure 18.5 Influence of the seismic mass on the transmission function of a knock sensor.

18.4.1 Lambda Sensors

A distinction is made between the binary and linear lambda sensors. Binary sensors permit the control of the air–fuel ratio around the stoichiometric point λ = 1 and, hence, the setting of the fuel supply for an optimum conversion in the threeway catalytic converter. Linear sensors monitor the air–fuel ratio continuously between rich and lean and are particularly suitable for the control of lean engines, for example, SI engines with direct fuel injection. With the binary lambda sensor, the Nernst voltage is measured between a catalytically active exhaust gas-side electrode and a reference electrode suspended in air. The voltage changes sharply around λ = 1 (Figure 18.7).

Figure 18.6 Knock sensor design with integrated plug.

18.4 Exhaust Gas Sensors Installed directly downline of the manifold, they serve to control the fuel injection system (λ control) in order to achieve the optimum conversion rate of a catalytic converter; installed downline of the catalytic converter they monitor its function and enable the OBD requirements to be satisfied. Common to all the sensors employed today is that they consist of several layers of zirconium dioxide (ZrO2) that conducts oxygen ions at temperatures above approximately 350°C and use the “Nernst equation”: the voltage available above a ZrO2 layer depends only on the difference in the oxygen partial pressures on each side of the layer.

1,0

0,8 Nernst voltage [V]

In some cases, knock sensors are used today even in diesel engines to control the start of injection and the function of the injection nozzles On-board diagnostics (OBD).

0,6 0,4

0,2 0,0 0,90

0,95

1,00

1,05

l [-]

1,10

Figure 18.7 Characteristic of a binary lambda sensor.

With the linear lambda sensor, the air–fuel mixture is controlled to a Nernst voltage corresponding to λ = 1 in a chamber inside the sensor by connecting an electric current referred to as the pump current. The air reference is generated

Internal Combustion Engine Handbook | 713

6606_Book.indb 713

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 18 Sensors

either via a channel in the ceramics or by a constant supply of oxygen to a cavity (Figure 18.8). The pump currents serve as a measurement signal and depend on the exhaust gas lambda (Figure 18.9).

Exhaust 0.65 £ l £ ¥

UP

Rp

IP

Pump cell

Diffusionszbarriere

Measuring chamber l=1

UG

Ri Nernst cell UN Air reference l=¥

Electrode

ZrO2 - ceramic

Figure 18.8 Principle of a linear lambda sensor.

3

Ip [mA]

2 1 0 -1 0,8 1,0

1,5

2,0

2,5

3,0

3,5 l [-] 4,0

Figure 18.9 Characteristic of a linear lambda sensor.

Main pump electrode (+) P+

Main pump electrode (–) P–

The NOx sensor permits the direct measurement of the nitrous oxide concentration in the exhaust gases of SI and diesel engines. It enables optimum control and diagnosis of NOx catalytic converters by the engine controller (e.g., NOx accumulator, selective catalytic reduction (SCR) catalytic converter) and compliance with OBD requirements for the checking of the three-way catalytic converter for low-emission concepts (SULEV, LEV 2). The most promising functional principle of a NOx sensor is based on the decomposition of nitrous oxides using a catalytically active electrode comprising a mixture of platinum and rhodium. The measurement of the oxygen produced in the process is well known from the amperometric linear lambda sensor. The structure of the multilayer ZrO2 sensor ceramics comprises two chambers (Figure 18.10). First, the oxygen contained in the exhaust gas is reduced (lean exhaust gas) or increased (rich exhaust gas) to a constant partial pressure of approximately 10 ppm by connecting a pump current. The necessary current is proportional to the reciprocal of the air–fuel ratio. The NOx reduction at the measuring electrode takes place in the second chamber. The current necessary to keep the atmosphere around the electrode free from oxygen is proportional to the nitrous oxide concentration and forms the measurement signal (Figure 18.11). A two-step setting of the residual oxygen in the first and second chambers using an additional electrode enables the cross sensitivity of the sensor to oxygen to be reduced. Knowledge of the air–fuel ratios also permits a numeric compensation of the NOx signal. A disadvantage of such sensors is their high ammonia cross sensitivity, caused by the oxidation of ammonia-producing nitrogen monoxide in the first sensor cavity. The necessary currents at the NOx-measuring electrode are a few μ A for a measuring range of several hundred ppm. An electromagnetically safe integration into the engine management system is possible only with electronic control of the sensor in its immediate vicinity. There are two ways of doing this, either with a stand-alone or “smart” NOx sensor (Figure 18.12) with the complete control system (heater control and

Second chamber Ip1

First chamber

18.4.2 NOx Sensor

Ip2

Auxiliary pump electrode M1

Measuring electrode M2

Ip0

Gas

Air channel

Heater

Vref V0

V1

V V2

ZrO2 laminated substrate Reference electrode REF

Figure 18.10 Measurement principle of an NOx sensor (NGK Insulators LTD).

714 | Internal Combustion Engine Handbook

6606_Book.indb 714

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

18.5 Pressure Sensors

pump current control) and digital communication to the engine controller or with a remote pump current controller with analog control.

1,8

Carrier gas Gas temperature Humidity Oxygen

1,6

Pump flow lp2 [µA]

Normal pressure sensors can be subdivided into the following groups: •• MAP: Manifold absolute pressure sensor (intake air pressure sensor)

2,0

1,4

1,2

•• BAB: Barometric absolute pressure sensor (ambient pressure sensor)

Nitrogen 80 °C 3% Volume fraction 0%

•• Turbo MAP: Manifold absolute pressure sensor for turbocharged engines (charge pressure sensor).

1,0

0,8 0,6 0,4

0,2 0,0

18.5.1 Normal Pressure Sensors

0

100

200 300 NOx [ppm]

400

500

Figure 18.11 Characteristic of an amperometric NOx sensor.

The MAP is used to determine the intake manifold vacuum pressure that exists downline of the throttle valve. The typical measuring range here is 0.2–1.1 bar. Together with the temperature, this allows the drawn air mass to be calculated. Depending on the driver’s wishes, this information forms the basis for determining the gasoline injection volume and the throttle valve position. Using the lambda sensor signal, a closed control loop is set up that controls the air–fuel mixture in the range of λ = 1 in order to guarantee minimum exhaust gas emissions. The MAP is often used with an integral temperature sensor to reduce installation work. The BAP is used to determine the ambient pressure.

O-ring Pressure connection Temperature sensor (NTC) Pressure sensor element

Figure 18.12 NOx sensor with control electronics.

Housing with integrated plug Figure 18.13 Structure of a MAP with integral temperature sensor.

18.5 Pressure Sensors Different sensor types are employed in order to meet the various demands of the pressures to be measured. These sensor types can be categorized as follows: Sensor type

Pressure range

Normal pressure sensors

approximately 0–5 bar

Medium pressure sensors

approximately 5–100 bar

High-pressure sensors

approximately 100–2000 bar

Differential pressure sensors

approximately 0–1 bar

Pressure switches (switching function only)

approximately 0–1 bar.

The fields of application and sensor principles for these sensor types are described in the following sections.

The information received serves for the compensation of the air pressure at different altitudes. The typical measuring range here is 0.5–1.1 bar. The Turbo MAP serves to determine the charge pressure in engines with turbocharger. The typical measuring range here is 0.5–2.5 bar. The engine controller optimizes the combustion parameters using the charge pressure information. The charge pressure can also be used to control the turbocharger variable turbine geometry (VTG). The following measurement principles are in use. 18.5.1.1 Piezoresistive Measurement Principle Piezoresistive measuring cells are traditionally used. A piezoresistive measuring cell is a pressure cell consisting of a diaphragm with attached piezoresistors. The pressure acting on the cell causes the piezoresistors to expand. This results in a pressure-dependent change in resistance. These changes in resistance are transformed into analog voltages by a separate

Internal Combustion Engine Handbook | 715

6606_Book.indb 715

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 18 Sensors

electronic circuit. In more recent versions, the pressure cell is integrated into the chip with “volume micromechanics.” 18.5.1.2 Capacitive Measurement Principle A fundamentally new development is the “surface-mounted micromechanical” pressure sensor. Here, the pressure cell and the associated evaluation electronics are manufactured on a chip using standard semiconductor processes (BiCMOS). This eliminates the bond wire connections between pressure sensor cell and evaluation electronics. The pressure is measured by a special capacitor-like pressure cell. The pressure acting on the cell changes the distance between the two capacitor surfaces and results in a change in the capacitance. This capacity change is then transformed into an analog output voltage. The configuration is shown schematically in Figure 18.14.

Evaluation electronics with electrical calibration unit

n-MOS

Pressure cell: Measurement principle capacitive

Pressure application causes diaphragm diversion

p-MOS

Figure 18.14 Configuration of a surface-mounted micromechanical pressure sensor.

The desired characteristic for the different applications and pressure ranges mentioned previously is set by calibration at the end of the sensor production.

18.5.2 Medium-Pressure Sensors

Figure 18.15 View of a high-pressure sensor.

As a rule, media separation is realized by diaphragm for high pressures. The use of high-strength materials for the diaphragm allows measuring ranges above 2000 bar. A burst pressure of 1.5 to 2 times the rated pressure is typically required. The pressure deforms the diaphragm which is picked-up by a strain measuring bridge on the diaphragm. With appropriate design, the signal is proportional to the pressure applied. An electronic circuit (ASIC) amplifies the electrical signal that is compensated with the calibration data for pressure and temperature dependency. By calibrating with pressure and temperature, the high-pressure sensor achieves a typical accuracy of 1% to 2% of the measuring range. In most cases, the output signal is analog and ratiometric.

18.5.4 Pressure Switch

Pressure switches are used to detect pulsating compression stresses in liquid or gaseous media. The switching thresholds are some 100 mbar for oil pressure (Figure 18.16) and up to more than 100 bar in hydraulic applications.

These sensors are used, for example, for engine and hydraulic oil pressure in automatic transmissions and applications outside the powertrain (A/C compressors). For fluids or aggressive media, a configuration is predominantly used that is similar to the high-pressure sensor described below.

18.5.3 High-Pressure Sensors

The pressure range for high-pressure sensors starts at approximately 100 bar. The common design is of hexagonal form with an M12 screw thread connection (Figure 18.15). Plugs with three pins (supply voltage, ground, and output) are generally used. The calibration is also performed via the three pins or additional contacts. The main fields of application can be defined as follows: 100–200 bar high-pressure direct injection (HPDI) gasoline direct injection systems 200–280 bar

brake pressure sensors

1.300–2.000 bar common-rail diesel injection systems Figure 18.16 View of a an oil pressure switch.

716 | Internal Combustion Engine Handbook

6606_Book.indb 716

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

18.6 Air Mass Sensor

The function of a pressure switch is essentially based on a mechanical contact which is opened when the applied pressure is exceeded. A spring presses a metal diaphragm against the metal housing. When the pressure in the medium exceeds a threshold value, the metal diaphragm moves away from the housing; the housing—metal diaphragm—metal spring circuit is interrupted (opening contact). There is also the reverse variant in which the circuit opens at dropping pressure (closing contact).

18.6 Air Mass Sensor In order to be able to determine the air mass flow drawn in by the engine, either a manifold pressure sensor (MAP = manifold absolute pressure) or an air mass sensor (MAF = mass airflow) is employed today. The output signal serves in the electronic engine controller as the basis for determining the load state. In SI engines, the signal serves predominantly for controlling the fuel volume as an input parameter for the ignition map and for determining the exhaust-gas recirculation rate. In the interplay with the lambda sensor, the MAF or MAP forms a closed control loop. Since with diesel engines there is no throttle valve and, hence, the intake manifold pressure is no measure for the drawn fresh air mass, an MAF has to be employed. Here, the signal from the MAF serves as a control variable for the exhaust-gas recirculation (EGR); in newer systems it is also a control parameter for a map-dependent diesel injection pump. Since the exhaust gas provides no feedback with diesel engines, the demands made on the measuring precision of the MAF are higher than with the SI engine. The measurement of the drawn fresh air mass is, therefore, a precondition for reducing the pollutant emissions and for increasing the driving comfort.

18.6.1 Measuring Principle

Apart from the principles for determining mass flows employed earlier, sensor plate and hot wire, the process of hot-film anemometry is predominant today.

Practically all air mass sensors employed in the automobile today follow this principle. A heated element dissipates energy to the surrounding air. The dissipated heat energy is dependent on the airflow and can be used as a measurement parameter.

18.6.2 Mass Air-Flow Sensor

The hands free module (HFM) consists of a sensor module, the plug and, in some cases, a tubular housing. Using a corresponding plug element, the sensor module may be also used as a so-called insert finger. The tube diameter is adapted to the air mass range required in each case (Figure 18.17). Sensors, electronics, and the flow straightener are integrated into the sensor module.

Plug

Sensor module

Tube

w Flo

Figure 18.17 Design of a mass airflow sensor.

Two temperature-dependent metallic-film resistors on a glass substrate (RS and RT) are positioned inside the tube in the direct intake air stream. These two resistors, in combination with R1 and R 2, are linked in a bridge circuit (Figure 18.18). Depending on the drawn air mass flow, RS is cooled more or less strongly. The electronics control the necessary heating current through RS so that there is always a constant temperature difference (e.g., 100 K) at RS to the air temperature measured at RT. The heating current is transformed into a voltage signal at resistor R2.

T(RS) – T(RT) = k

RS RT R1

w Flo

R2 UFLOW Measuring bridge

Figure 18.18 Principle of a mass airflow sensor. The voltage at R2 is a measure of the air mass flow rate.

Internal Combustion Engine Handbook | 717

6606_Book.indb 717

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 18 Sensors

The resistors RS and RT are matched to one another in such a way that the map is independent of the air temperature. The map also exhibits—thanks to the physics—an advantageous nonlinear characteristic permitting a practically constant proportional resolution. Thanks to the use of materials specially adapted to the conditions in the automobile engine compartment, flow control, circuiting technology, and the mechanical configuration, the HFM signal is more or less independent of the temperature, pressure, and soiling. Pulsations in the intake manifold and return-flow compensation: Internal combustion engines with four or fewer cylinders create extreme pulsations in the intake manifold at wide-open throttle or in the case of no throttle valve (e.g., in diesel engines or SI engines with direct fuel injection). At certain engine speeds, at the resonance point, a pulsating return flow occurs that with conventional HFM results in a positive measurement error as the air passes over the sensor three times. This effect can be compensated with the design of the flow tube, the positioning of the sensors within this tube and an additional correction circuit. In many applications, the temperature sensor (NTC resistor) for determining the intake air temperature is also integrated into the HFM.

18.6.3 Secondary Air Mass Sensors (SAF)

During the exhaust gas test cycles, a large proportion of the carbon monoxide (CO) and hydrocarbon (HC) emissions are produced during the start phase of an engine. In the first few minutes after the start, practically no conversion takes place because the temperature is still under the “light-off” temperature of approximately 350°C. In order to achieve the quickest possible heating of the catalytic converter, secondary air is blown into the exhaust gas line and the exhaust gases are enriched with additional hydrocarbons. This can be achieved by enriching the mixture or by subsequent injection of fuel into the exhaust system. The oxygen admitted via the secondary air creates a post-combustion of the rich mixture and, thus, permits a faster heating of the catalytic converter and a consequent significant reduction in the pollutant emissions. This is necessary in order to meet the strictest exhaust emissions requirements. Based on the measurement principle of the main stream HFM, the secondary air mass sensor measures the fresh air mass admitted to the catalytic converter in addition to the exhaust gases during the start phase. The advantages compared with an uncontrolled system are the independence from system tolerances and the possibility of also being able to carry out an extended diagnosis of the system during the secondary air phase.

18.7 Speed Sensors Although one generally speaks of speed sensors; in this case, we are dealing with incremental sensors. The following speed sensors are frequently used in the control of the powertrain:

•• crankshaft sensor •• camshaft sensor •• transmission speed sensor Electronic controllers for internal combustion engines require information on the momentary position of the crankshaft and camshaft for precise control of ignition and fuel injection. For the application on the crankshaft, high precision over the full range of functions (temperature, air gap, speed, mechanical tolerances) is demanded. Furthermore, the sensor should be able to detect even very low speeds in order to permit a rapid detection of the position during the start of the engine. For engines with up to eight cylinders, misfire detection uses the evaluation of the crankshaft sensor signal. A very high repetition accuracy (<0.03°) is therefore demanded. The camshaft sensor is used for synchronization between cam and crankshaft; this means an identification of the first cylinder. To achieve rapid synchronization, either specially encoded sensor wheels for the camshaft or camshaft sensors with a static function are used. In engines with variable valve timing, the camshaft sensors are also needed for control of the camshaft adjuster. One camshaft sensor is required for each adjustable camshaft. For the application as position sensor for the variable valve timing, the accuracy is of paramount importance. Transmission speed sensors are used to measure the speed of the vehicle. Both the input and output speeds are required for control of automatic and continuously variable transmission (CVT). The demands made on transmission speed sensors are considerably lower, but these sensors should be able to detect even very low speeds. Measuring principles The measurement principles are based on passive and active speed sensors.

18.7.1 Passive Speed Sensors

Sensors used as passive speed sensors today are almost exclusively inductive sensors, also known as variable reluctance (VR) sensors. Inductive sensors consist essentially of a coil around a magnetically precharged core. If the inductive sensor comes near a moving ferromagnetic sensor wheel, a voltage is induced. This voltage is evaluated in the electronic controller. Each flank of the sensor wheel induces an electric voltage. With inductive sensors, the level of the induced voltage is dependent on the speed; there is, therefore, a lower speed/frequency limit for the function of the inductive sensor.

18.7.2 Active Sensors

Active speed sensors have integrated electronics for signal processing. Active sensors, therefore, transmit standardized signal levels that can be used in the electronic controller without additional processing.

718 | Internal Combustion Engine Handbook

6606_Book.indb 718

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

18.8 Combustion Chamber Pressure Sensors for Diesel Engines

Active sensors based on the Hall effect are the most widely used, but magnetoresistive (MR) sensors and giant MR (GMR) sensors are also being increasingly employed. Of the Hall sensors, the differential Hall sensor is the most frequently employed (Figure 18.19).

VH2 VH1

VH1 – VH2

H1

H2

18.8 Combustion Chamber Pressure Sensors for Diesel Engines Triggered by the ever more stringent exhaust emission legislation, more precise information about the combustion process is required for future combustion systems. For qualitative information, such as the verification of the presence of preinjection events, knock sensors are currently preferably used as structureborne sound sensors. In most cases, however, a quantitative information about the released combustion energy over each crankshaft angle is required. Combustion chamber pressure sensors are particularly suited. Based on their signal, the following functions can be realized: •• cylinder-selective regulation of the maximum combustion pressure •• detection of pre, main, and postinjection events

Output

•• compensation of tolerances in the injection and intake system •• regulation of the combustion center.

The edge change of a ferromagnetic sensor wheel results in a difference in the magnetic field at the differential Hall element. Thanks to the differential principle, these sensors are essentially insensitive to interferences, such as temperature changes and external magnetic fields and are characterized by high precision. The differential principle allows sensors to be built with a lower speed limit of 0 rpm (zero speed). Because of the differential principle, however, these sensors can be used in only one installation position. “Single-element” Hall sensors are used for static functions. These sensors allow tooth or gap to be detected without the sensor wheel moving (true power on). Thanks to the design as a single element, they can be positioned anywhere between Hall sensor and sensor wheel. Figure 18.20 shows an active crankshaft sensor on the basis of differential Hall sensors.

The use of conventionally functioning pressure sensors with pressure-sensitive diaphragm is extremely difficult to realize. To minimize pressure fluctuation effects in the combustion gas, the diaphragm must be placed as directly as possible to the combustion chamber. Because the cylinder heads of modern passenger car diesel engines usually do not have any or only very little space for an additional bore for a pressure sensor, the use of such sensors is mostly prohibited by the space available. For this reason, the current thrust of development in combustion chamber pressure sensors is focused on the integration in existing components and in glow plugs in particular. With its access to the combustion chamber, the glow plug has an ideal installation position. Figures 18.21 and 18.22 show examples for a combustion chamber pressure sensor with piezoceramic measuring element integrated in a glow plug. The piezoceramic sensor element converts the deformation of the measuring element, built from the glow plug housing and the glow plug electrode, into an electric charge as an output signal.

Figure 18.20 Active crankshaft sensor.

Figure 18.21 Combustion chamber pressure sensor (center). integrated in the glow plug, in combination with knock sensor (left) and glow plug (right).

Figure 18.19 Measuring principle of a differential Hall sensor.

Internal Combustion Engine Handbook | 719

6606_Book.indb 719

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 18 Sensors

Sensor element bracing

Glow element contact

Piezoceramic sensor element Sensor element contacts

Glow plug electrode Glow plug housing

Figure 18.22 Sensor element integration.

with a piezoelectric sensor element, excellent signal noise characteristics [18-1]. If deformations occur at the cylinder head that cannot be traced to the combustion process or if these are too small, the somewhat more involved principle of measuring a sensorinternal deformation must be used. In this case, it is necessary to uncouple the measuring element from the deformations of the glow plug housing induced by the cylinder head. The disadvantage of the higher degree of complexity of this measuring principle is balanced by the advantage of the calibrability at the sensor manufacturer. In addition to the piezoelectric sensor element presented here, deformation-measuring elements with different conversion principles are possible as a rule. The piezoelectric measuring element, however, convinces with its simple structure and the consequent minimum space requirements. A future use of similar concepts in the SI engine by, for example, integrating the combustion chamber pressure sensor in a spark plug, is within reach, thanks to the technologies developed here.

Bibliography

Glow plug-integrated combustion chamber pressure sensors can be classified in two basic measuring principles: On the one hand, the measurement of the combustion pressure by measuring the cylinder head deformations, and secondly, the measurement of the combustion pressure by measuring the sensor-internal deformations (Figure 18.23). Measuring the combustion pressure by measuring the cylinder head deformation represents the simpler solution because the entire glow plug housing can be used measuring element. Research with common-rail diesel engines shows, for this measuring principle, outstanding accuracies and, in combination

18-1. Robert Bosch GmbH (Hrsg.). 2007. Kraftfahrtechnisches Taschenbuch. 26. Aufl. Wiesbaden: Vieweg Verlag. 18-2. Fiedeler, O. 1992. Strömungs und Durchflussmesstechnik. R. Oldenburg Verlag. 18-3. Niebuhr, J., Lindner, G. 1994. Physikalische Messtechnik mit Sensoren. R. Oldenburg Verlag. 18-4. Tränkler, HR., Obermeier, E. 1998. Sensortechnik. Springer Verlag. 18-5. Last, B., Ramond, A., Goretti, S., Burrows, J. 2004. “Integration of a piezo ceramic sensing element in a glow plug in order to get a combustion pressure sensor for diesel engines.” Hildesheim, Adaptronic Congress.

Combustion chamber pressure sensor integrated in the glow plug

Cylinder head

A

Thread Cylinder head Zylinderkopf deformation verformung Sealing seat Sensor-internal Sensorinterne deformation Verformung

Figure 18.23 Combustion chamber pressure measurement with cylinder head versus sensor deformation.

720 | Internal Combustion Engine Handbook

6606_Book.indb 720

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19 Actuators 19.1 Drive Pneumatic and electric actuators are predominantly used as engine management actuators. Figure 19.1 shows a comparison of the advantages and disadvantages of the most widely used drives.

6

19.1.1 Pneumatic Drives

Pneumatic drives (Figure 19.2) as actuators are predominantly used as changeover contacts between two fixed positions. Pneumatic actuators consist of a vacuum unit with a diaphragm that is linked via a control valve to the vacuum supply of the vehicle. The element to be actuated is connected to the actuator either directly or via levers and cables. In commercial vehicles, pneumatic drives with lifting cylinders that are connected to the vehicle pneumatic system are predominantly used. The advantage of the pneumatic drive is its low price in combination with large actuating torques and fast regulating times in relation to its size. A major disadvantage of pneumatic drives is the difficulty of achieving position control, therefore necessitating additional control elements for precise stopping in intermediate positions. This disadvantage has resulted in the widespread change from pneumatic drives to electric drives.

5 4 3 2 1 2 3 4 5 6

1 Resonance valve Carrier lever Vacuum unit End stop Diaphragm Vacuum connection

Figure 19.2 Intake manifold resonance valve with pneumatic drive (source: Continental Automotive GmbH).

Drive

pneumatic Stepping DC motor drive motor with housing Actuating torque ++ – ++ Actuating time + – + Position control –– ++ + Weight + – o Costs ++ + o Service life o + – ++ very positive, + positive, o average, – negative, – – very negative

Torque motor

EC motor

o ++ + – + +

++ ++ + o –– +

Figure 19.1 Comparison of different drives (source: Continental Automotive GmbH).

Internal Combustion Engine Handbook | 721

6606_Book.indb 721

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

19.1.2 Electric Drives 19.1.2.1 Stepping Motor The stepping motor (Figure 19.3) is predominantly used as an actuator where low demands are made on the actuating force. The advantage of the stepping motor lies in its stepwise movement and the corresponding control. The counting of the positioning steps permits a determination of the relative position of the drive compared with its position at the start of the movement and, hence, in a simplified position control of the drive.

The main advantages of the torque motor are the contactfree and, hence, wear-free drive, and its simple construction. Disadvantages compared with the DC motor are the low excess torque and high weight in relation to the actuating torque. The torque drive usually also requires a position sensor for position control.

Rotor Stator housing 1 Coil 1 Coil 2 Stator housing 2

3 1

2

Drive shaft

Figure 19.3 Stepping motor (source: Continental Automotive GmbH).

For simpler demands, no additional sensor is required to determine the actual position, although an absolute determination of the current position is not possible. The main disadvantage of the stepping motor is the low excess torque available to overcome any possible binding in the mechanism, with the related possibility of an error left undetected in the position control as a result of an inability to make a required positioning movement. 19.1.2.2 DC Motor The direct-current motor (DC motor) is used as an actuator preferably in conjunction with gearings. The flexibility of the gearing designs and gear ratio permits here the use of the same DC motor for different actuating torque or actuating time requirements. The main advantage of the DC motor and gearing combination is the high excess torque. This enables short actuating times to be achieved and allows brief binding to be overcome. Position control of the DC motor drive is possible only in combination with a position sensor. Disadvantages of the DC motor are its comparatively more complex construction and the wear behavior of the carbon brushes in the motor and in the gearing as compared with contact-free drives. 19.1.2.3 Torque Motor Torque motors are used in actuating concepts as direct drives without additional gearing (Figure 19.4). Their typical applications are those with low actuating force requirements, in conjunction with the demand for short actuating times.

1 Stator 2 Rotor (2-pole) 3 Coil

Figure 19.4 Torque motor (source: Continental Automotive GmbH).

Some applications use torque drives without feedback. In this case, there is only an actuation between two mechanical end stops or, controlled by a position current-drawing characteristic, against a pull-back spring. The current-drawing control, however, does not allow precise positioning. 19.1.2.4 EC Motor Duty Cycle motors are electronically commuted DC motors. Unlike the classic DC brush motors, they do not have moved contact points. A typical model is the inner rotor with permanent magnets on the rotor shaft and windings in the stator. ED motors require additional sensors in the motor for the rotor position detection and the subsequent control loops for actuating the motor. Hall element sensors are usually used for this purpose. EC motors are preferably used when high actuating torques in conjunction with fast response times or lasting running performance in, for example, utility vehicles or valve poppet alignment. Their advantage is low mechanical wear and high capacity and an essential disadvantage is the comparatively high cost.

19.1.3 Communication with Engine Control Electronics 19.1.3.1 Controlled Actuators Controlled systems are used in simple applications (Figure 19.5). The engine control electronics applies current to an actuator or a vacuum valve and affects in this manner an

722 | Internal Combustion Engine Handbook

6606_Book.indb 722

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.2 Throttle Valve Actuators

actuation movement. A “closed-loop” feedback by the actuation movement to the engine control electronics does not occur at all or only via external systems.

1

2

and a processor unit with software for regulating and controlling the position. The communication of these actuators in the engine control electronics is reduced to signals: the target value signal from the engine control electronics to the actuator and the feedback signal in reverse direction. Standard signals (usually digital) but also BUS systems (CAN bus, LIN bus) are used. The power is drawn directly from the on-board system and power end stages in the engine control are not necessary for smart drives. Based on the feedback received, smart actuators can be used in OBD-relevant functions as well. Due to the integrated electronic components, the structure of the actuators is significantly more complex than those of externally controlled actuators. Depending on the used electronics technology, the permissible temperature may also be limited. An essential advantage of these actuators is their very simple integration in existing engine control systems, even for precise demands on positioning. Actuators with EC motor drives are preferably realized as intelligent actuators due to the complex actuation.

19.1.4 Reset/Default Position 1 Throttle valve actuator 2 Controlled torque actuator to intake manifold extension switching 3 Controlled torque actuator to intake manifold resonance switching

3

Figure 19.5 Controlled torque actuators for switching intake manifold extension and intake manifold resonance (source: Continental Automotive GmbH).

The main advantages of these systems are the low costs and the relatively simple integration in an existing system. Its essential disadvantage is the lack of feedback. For this reason, controlled systems without feedback can usually not used for On-Board Diagnostics-relevant functions. For controlled systems, pneumatic drives, stepped motors or torques motors are preferred. 19.1.3.2 Externally Controlled Actuators Controlled actuators use a position sensor to provide a feedback of the actual system to the engine control electronics. Using this position signal, the engine control can regulate the drive and achieve that the actual position trails the target value as closely as possible. Compared to controlled actuators, externally controlled actuators require an increased effort in hardware and software for the engine control electronics. Essential issues are the power end stage (H bridge) for actuating the drive and the processor’s computing capacity required for the position regulation. DC motor drives with gearing are usually used for externally controlled actuators. 19.1.3.3 Internally Controlled Actuators (Smart Actuators) Intelligent (“smart”) actuators are equipped with integrated electronics that features a power end stage for drive actuation

Depending on the application, it may be necessary to reach a predefined position of the control unit even when the actuator has failed. This is defined as the default position. Such an actuator behavior is achieved by integration of one or more reset springs at the control unit or the drive. The requirement of a default position poses a significantly higher demand on the capacity of the drive and is frequently underestimated. Generally, it necessitates considerably larger drives and excludes the concept of self-locking drive designs.

19.2 Throttle Valve Actuators 19.2.1 Key Function in SI Engines

The power output of a spark-ignition (SI) engine is controlled quantitatively. This requires control of the drawn air mass. The most widely used technical solution for influencing air mass flows is the throttle valve actuator. The position of the throttle valve in the air duct determines the amount of air drawn in by the internal combustion engine and the pressure level in the intake manifold (Figure 19.6).

19.2.2 Key Function in Diesel Engines

The operation of the diesel engine and SI engine with direct fuel injection is regulated by quality control of the air–fuel mixture. Under ideal conditions, no throttling of the air mass flow is necessary. Nevertheless, throttle valve actuators are widely used in diesel and SI engines with direct fuel injection. The main function of these actuators is to create a vacuum to draw exhaust gases into the air taken in by the engine (exhaust gas recirculation) to comply with the strict legal requirements for low pollutant contents in the exhaust gases (Section 19.2.6, “Prethrottling”). In addition, using throttling and

Internal Combustion Engine Handbook | 723

6606_Book.indb 723

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

Typical air mass characteristics in dependence on gas pedal position and air channel geometry Air mass

4

3

Typical differential pressure characteristics at constant pump output and cylindrical air channel geometry Differential pressure

1

2

1 Cylindrical air channel 2 Cylindrical air channel with progressive C cable plate sector 3 Cylindrical air channel with degressive C cable plate sector 4 Air channel with spherical cap geometry Idling

Accelerator pedal position

Full-load Closed

Throttle valve angle

combustion-engineering measures, the exhaust gas temperature of the combustion engine can be increased significantly to support particle filter regeneration.

19.2.3 Additional Functions 19.2.3.1 Idle-Speed Control of SI Engines Apart from its main function of the cylinder charge control, the throttle valve actuator together with its various attachments performs additional functions. The most important auxiliary function of the throttle valve actuator is to control the idle speed of the SI engine (Figure 19.7). The idle speed is normally controlled by influencing the air mass flow. In addition, “fine-tuning” is achieved by shifting the ignition timing. The air mass flow at idle speed can be controlled with an actuator in a bypass to the main air duct when the throttle valve is closed or by direct positioning of the throttle valve in the slightly open working range. 19.2.3.2 Position Signal A sensor on the throttle valve actuator generates a position signal and transmits this to the engine control electronics. Potentiometers are widely used for this function.

Idle control by throttle valve movement

Opened

Figure 19.6 Diagram of air mass and differential pressure characteristics (source: Continental Automotive GmbH).

The signal from the sensor also serves to distinguish between the part-load and idle-speed operation of the internal combustion engine. In some cases, switches on the throttle valve actuator are used in addition to the sensor signal. With electrically powered throttle valves, the position sensor is generally integrated into the drive. To meet rising demands on the reliability of these systems, the potentiometers are increasingly being replaced by contact-free sensor systems. 19.2.3.3 Impact Load Damping The dashpot function provides a slower return of the throttle valve to the closed position when the gas pedal is suddenly relieved. Without the dashpot function, the throttle valve is slammed closed by return springs in such cases. This leads to an impact load and a sharp braking of the vehicle. To improve ride comfort, the impact load is damped. Modern throttle valve actuators, however, perform this dashpot function by opening the bypass actuator for the idle-speed control or by a return of the throttle valve to the closed position at a slower speed and independently of the gas pedal return (Section 19.2.4, “Drive by wire”).

Idle control by bypass valve with closed throttle valve

Figure 19.7 Comparison of idling control systems (source: Continental Automotive GmbH).

724 | Internal Combustion Engine Handbook

6606_Book.indb 724

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.2 Throttle Valve Actuators

19.2.3.4 Cruise Control function The speed control of a vehicle with SI engine (cruise control) is affected by the actuation of the throttle valve independently of the driver. This can be affected by separate cruise control actuators linked to the throttle valve by means of cables or levers. Modern throttle valve actuators perform this function using a direct drive in the throttle valve actuator. The cruise control function requires a power source (e.g., electric or pneumatic drive) that can open the throttle valve independently of the gas pedal position.

1

19.2.4 “Drive-by-Wire”/E-Gas

Unlike the cruise control function, it is necessary for the drive slip control or traction control system (TCS) and for the electronic stability program (ESP) to be able not only to open but also to close the throttle valve independently of the gas pedal position. This function is difficult to achieve with mechanically linked systems. In some cases, a second throttle valve (open during normal operation) that is actuated independently of the gas pedal position is installed up-line of the actual throttle valve to perform these functions. However, an engine drag torque control is not possible with these systems. More widespread here is the use of “drive by wire” systems (also known as E-Gas systems), (Figure 19.8), that permit positioning of the throttle valve totally independently of the gas pedal position. With these systems, a nominal position of the throttle valve is calculated in the engine controller on the basis of various key data and functions that is then implemented by the position control of the throttle valve actuator. The position is controlled by comparing the nominal and actual positions of the throttle valve and corresponding control of the actuator by the engine control electronics. In some cases, the position control is also affected by electronics in the throttle valve actuator. In this case, only the nominal value is formed in the central engine

1 2 3 4 5

5 4 Air channel DC motor Two-stage gearing Position sensor (potentiometer) Pull-back spring

3

2

Figure 19.8 Drive-by-wire throttle valve actuator E-Gas 5 (source: Continental Automotive GmbH).

control electronics and then transmitted to the throttle valve actuator. There are also applications in which the complete engine control electronics is integrated into the throttle valve actuator (Figure 19.9). The positioning of the throttle valve independently of the gas pedal position also enables or simplifies other functions. The opening map of the throttle valve can be accelerated or decelerated in relation to the gas pedal position as desired. The idle-speed control, the cruise control function, and the impact load damping are all performed by the software in the engine control unit and require no additional mechanical parts. This is a significant benefit of the “drive by wire” compared with throttle valve actuators linked mechanically to the gas pedal.

1 2

6 5 1 2 3 4

Air channel Pull-back springs DC motor Two-stage gearing

3 7 4 5 Position sensor (potentiometer) 6 Heating (via cooling water circuit) 7 Engine management electronics

Figure 19.9 Throttle valve actuator (E-Gas 7) with integral engine control electronics. (source: Continental Automotive GmbH)

Internal Combustion Engine Handbook | 725

6606_Book.indb 725

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

Malfunctions or a failure of the throttle valve actuator can be detected by the engine control electronics and are not misinterpreted as drivers’ wishes. For this, a safety concept has been integrated into the software of the engine control electronics. To clearly identify a malfunction of the throttle valve actuator, two redundant signals feed back the throttle valve position.

for exhaust gas recirculation (EGR) feeding. The demands made on such prethrottle actuators with respect to actuating torque and actuating time are typically slightly lower than the demands made on “drive by wire” throttle valve actuators.

19.3 Swirl and Tumble Plates, Resonance Charging

19.2.5 Waste Gate Function

Throttle valve actuators are also used to control or to limit the charge pressure of turbocharged engines. With mechanical turbochargers, another throttle valve actuator is installed in a bypass to the compressor in addition to the throttle valve (waste gate function) that regulates the degree of cylinder charge of the internal combustion engine. If the charge pressure is too high in certain engine operating situations, the throttle valve in the bypass is opened and part of the compressed intake air escapes back into the area in front of the compressor. The charge pressure is reduced.

Due to increasing demands for reducing fuel consumption, valve systems affecting the quality of the air–fuel mixture become more common. Here, it must be differentiated between systems intended for a very high quality of mixture swirling for a particularly homogeneous mixture and systems intended for a targeted splitting into zones with differing fuel concentration (stratified charge). The actuators used for swirl and tumble plates and those for resonance charging are similar. Their differences are essentially the type of drive (pneumatic or electric) and the necessity to approach intermediate positions between the end stops. Electric actuators increasingly replace pneumatic drives in these applications.

19.2.6 Vacuum/Prethrottle Actuators

Throttling of the air mass flow creates a vacuum in the intake manifold of internal combustion engines compared with the ambient pressure. This pressure difference is employed for various functions. It serves the brake booster as an energy medium. Controlled by external actuators (generally valves), the vacuum is used to feed the “blow by gases” (crankcase ventilation) and the air current for the regeneration of the carbon canister and for exhaust gas recirculation. This function of the throttle valve actuator is also employed in diesel engines. The throttle valve of the prethrottle actuators (Figure 19.10) is fully open during normal operation and closes only when a vacuum is required in the intake manifold, for example,

1

2

5

4

1 2 3 4 5

6

3

Air channel Throttle valve DC-Motor Two-stage gearing Pull-back spring

7

9

6 7 8 9

19.3.1 Port Deactivation

Modern multivalve engines (two or more inlet valves per cylinder) are distinguished by an optimal filling of the cylinders at high speeds (Figure 19.11). Frequently, the needed large inlet cross sections at low speeds result, however, in an insufficient turbulence of the airflow and an inhomogeneous air/fuel mixture in the cylinder. To improve this situation, a portion of the inlet cross-section is closed by a valve system under low speeds. This results in high flow velocities and, depending on the arrangement of the inlet channels, to an additional swirling of the mixture, more homogeneous combustion and reduced fuel consumption.

8

Magnet wheel Hall sensor Electronics for position control Device plug

Figure 19.10 Prethrottling valve with integrated electronics (source: Continental Automotive GmbH)

726 | Internal Combustion Engine Handbook

6606_Book.indb 726

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.3 Swirl and Tumble Plates, Resonance Charging

“Position reached” feedback signal

Supply voltage and position demand

ECU ECU

Drallklappen Swirl plate Stellglied actuator

-

Intake manifold, V6 engine

Intake duct

Inlet valves Exhaust valve

Figure 19.11 Port deactivation system for the simultaneous actuation of two valve banks in a three-valve V6 engine (source: Continental Automotive GmbH).

For the control of the port deactivation, actuators that switch between two positions of the valve system are commonly used. Position-controlled systems are not typical. To the majority, pneumatic drives are used. Due to the increased exhaust gas requirements and the OBD relevance of these systems, electrically driven two-point actuators are also used.

19.3.2 Stratified Charge

To generate a stratified charge, valve systems are frequently used that change either the direction or speed of the airflow or both so that the air enters the combustion chamber with a swirling or tumbling motion (Figure 19.12). In some cases, an undeflected airflow and an air flow diverted by a valve system are mixed at a preset angle to also achieve this effect. Because it is necessary to influence the airflow into the combustion chamber, separate valves for each cylinder of the combustion engine are used to influence the flow as closely as possible before the combustion chamber. Because a targeted cylinder-selective control of the valve position is usually not required, the swirl plates of a cylinder bank are actuated via a common shaft that is regulated by an actuator. Due to the very different airflow under different load and speed situations, it is often necessary to reach intermediate positions as well. The design of the drive for a stratified charge actuator with position control is thus similar to the one of

4 2 1 3

5

1 DC-Motor 2 Gear housing with two-stage spur gear drive 3 Device plug 4 Pull-back spring (default position) 5 Push rod 6 Tumble valve shaft

5

Figure 19.12 Actuator with tumble valve system (four cylinders) (source: Continental Automotive GmbH)

Internal Combustion Engine Handbook | 727

6606_Book.indb 727

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

a throttle valve actuator. Due to the actuating torque and actuating time requirements, DC motor and gearing drives are preferably used.

Pneumatic and electric drives are used as actuators. In electric drives, actuators with integrated electronics for position control are preferred.

19.4 Turbocharger with Variable Turbine Geometry

19.5 Exhaust Gas Recirculation Valves

To optimize the turbocharger characteristics at different speeds and to reduce the “turbo lag,” turbochargers with variable turbine geometry become increasingly more common (Figure 19.13). Movable guide vanes are used to adjust the turbine geometry to the different load and speed conditions. These actuators are actuated by drives controlled by the engine control electronics.

Figure 19.13 ATL with variable turbine geometry (source: BorgWarner Turbosystems).

Engine Parameters

Signal Sensor

Engine Computer

Driver

PWM

In the early 1970s, external exhaust gas recirculation was employed for the first time in series-production automobiles in North America. During exhaust gas recirculation, part of the combusted exhaust gas is tapped at the exhaust manifold and returned via a pipe to the intake manifold where the combusted exhaust gas is mixed with the drawn mixture (Figure 19.14). This admixing of the combusted exhaust gas lowers the peak combustion temperature and, hence, reduces the nitrogen oxide emissions. In addition, exhaust gas recirculation can also help to reduce the fuel consumption in the part-load range. As the volume of recirculated exhaust gas has to be varied depending on engine load and engine revs, a corresponding control element is needed—the EGR valve. In addition to external exhaust gas recirculation, there is also an internal exhaust gas recirculation that exists on all four-stroke engines because of the overlap of the intake and exhaust valve systems. Internal exhaust gas recirculation has the same effect on the emissions, while the EGR volumes are relatively small because of the system design and can be influenced depending only on the load and engine revs on engines with variable valve train. Variable valve train systems are fundamentally employed with the aim of optimizing engine power and torque. Exhaust gas recirculation is just an additional benefit that alone would not justify the relatively high costs of these systems, and is, therefore, to be seen only as a bonus. In spite of the limited controllability of the internal exhaust gas recirculation volumes, additional external exhaust gas recirculation

CPU Solenoid Sensor

Intake Manifold

Mini EEGR Oxygen Sensor

Exhaust Manifold Figure 19.14 Schematic exhaust gas recirculation (source: Continental Automotive GmbH).

728 | Internal Combustion Engine Handbook

6606_Book.indb 728

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.5 Exhaust Gas Recirculation Valves

systems are normally not employed on engines with variable valve train. Disk valves with a pneumatic drive (vacuum unit) were used for the first external exhaust gas recirculation systems. Here, the vacuum unit was exposed to the pressure of the intake manifold, resulting in an adjustment of the EGR valve according to the operating point of the engine. Pneumatic deceleration valves, nonreturn valves, and pressure relief valves were installed in the system to limit the functional range to prevent negative effects of inappropriate exhaust gas recirculation volumes. Other control systems also use the exhaust gas backpressure as a control parameter for the vacuum unit. In some cases, electric changeover valves were also integrated into the control line to shut off the exhaust gas recirculation at certain operating points. In the next stage of development, electropneumatic pressure transducers were used with which it became possible for the first time to control the position of the exhaust gas recirculation valve independently of the operating point of the engine. Nevertheless, the range of application of exhaust gas recirculation remained limited to operating points at which the level of the vacuum was sufficient to open the disk valve against the spring force or the pressures acting on the valve. The wish to employ exhaust gas recirculation at higher load points and independently of the intake manifold vacuum triggered the development of electric exhaust gas recirculation valves (Figure. 19.15). At the same time, the demands made on precision increased so that sensors that showed the position of the valve were integrated. By comparison with earlier generations, these

exhaust gas recirculation valves permit a very accurate control of exhaust gas recirculation volumes with reduced actuation times at the same time. The integration of all the modules into one component simplifies the adaptation at the engine and reduces the function-relevant tolerances. Thanks to these functional benefits, electric exhaust gas recirculation valves are more and more widely used and are employed almost exclusively for new engine generations. Apart from stepping motors and lift and turn magnets, DC motors are now also more frequently used today as electric actuators. In addition to the further development of the actuators, the actual control valve has also been changed many times. Apart from disk and needle valves of different forms and dimensions, flap and rotary slide valves are also used today. Essentially, the valve should guarantee a constant function over the service life irrespective of the degree of soiling. Furthermore, the change in the differential pressure through the valve that occurs at every change in position should have the least possible influence on the set valve position. This is particularly important when moving from the closed to the slightly open position as the differential pressure acting here is subject to a large change. At the same time, the precision demands are very high in this working range. To improve the function in this range, there are valve developments with nonlinear opening characteristics. The valve design should also be as insensitive as possible to pulsing pressures. The best compromise at present appears to be flap valves, although, depending on the demanded EGR rate and the engine sensitivity to changes in volume, disk valves can also meet the requirements (Figure 19.16).

Figure 19.15 Electrically-controlled exhaust gas recirculation valve (source: Continental Automotive GmbH).

Internal Combustion Engine Handbook | 729

6606_Book.indb 729

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

Fresh air downstream ETC

Fresh air/ exhaust mixture to intake manifold

Exhaust gas

Figure 19.16 EGR flap valve (source: Continental Automotive GmbH).

For diesel engines, exhaust gas recirculation is a very effective method of complying with the demanded NOx emissions and is used in Europe for all vehicles up to 3.5 t and in some cases even above. The most widespread are controlled systems with pneumatically actuated valves, although electrically controlled valves are increasingly being employed for new applications (Figure 19.17). For conventional SI engines with intake manifold injection, electrically controlled systems are the most widespread, while EGR valves are currently installed in only approximately 50% of the vehicles as the emission values can also be achieved using alternative techniques. For the SI engine with direct injection, a 100% EGR application can be assumed, as the benefits of this engine concept can be fully exploited only together with exhaust gas recirculation. In

Exhaust gas inlet

view of the high precision demands, flap valves with electric motors can be expected to become more widespread.

19.6 Evaporation Emissions, Components 19.6.1 Tank Ventilation Valves

With the tightening of exhaust emissions legislation, the evaporative emissions of the tank system in vehicles with SI engines also came under closer scrutiny, along with the combustion residues. This led to the tank now being vented via a so-called “activated charcoal canister” and no longer

Exhaust gas outlet

Coolant outlet

Coolant inlet

Figure 19.17 Electrically controlled flap valve with water cooling (source: Continental Automotive GmbH).

730 | Internal Combustion Engine Handbook

6606_Book.indb 730

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.6 Evaporation Emissions, Components

directly to the atmosphere. The activated charcoal contained in this canister can bind large quantities of gasoline vapors that can be caused by parking in the sun, and hence, under normal circumstances, no more gasoline vapors can escape into the atmosphere via the tank vent. At the same time, the carbon canister has to be regenerated at regular intervals so that the saturation limit is not exceeded. For regeneration, the absorbed gasoline vapors are drawn in by the engine and combusted. This additional amount of fuel has to be precisely metered, however, so that it does not lead to an overenrichment of the mixture. This is controlled via a so-called “canisterpurge valve.” It is a clocked solenoid valve that is controlled by the engine control unit in relation to the lambda control. Essentially, the amount of fuel admitted to the engine via the injection system has to be reduced by the amount regenerated. The function of the canister-purge valve is determined by the controllability of low throughputs and maximum throughput. The regeneration of the carbon canister should be possible even at engine idle speed, but because of the high-pressure differential and the small overall amount of fuel required, this necessitates a high control precision of the canister-purge valve. At the same time, however, large regeneration volumes are called for in the part-load and full-load ranges. Because of the small vacuum in this engine-operating mode, this necessitates

Tank ventilation valve

Control valve, diaphragm-controlled

a large flow cross-section. Furthermore, the canister-purge valves should be compact and should create the least possible noise radiation. They are installed on the vehicle body, on the intake manifold, or even on the carbon canister, as long as this is located near the engine. Because of the different emissions legislations as well as the function and application demands, many different canister-purge valves have been developed (Figure 19.18). A distinction is made between low-frequency valves (5 to 20 Hz) with pulsating throughput and high-frequency valves (>100 Hz) with continuous throughput. The low-frequency valves are generally cheaper, but the control precision is limited and a high level of noise can develop, particularly at temperatures below freezing. The valves with continuous throughput are more complex with corresponding disadvantages in the dimensions and costs, although fundamental functional and acoustic benefits are achieved. To decrease the sensitivity to pressure fluctuations, pressure-compensating valve seats or nozzles with ultrasonic flow are used in some cases.

19.6.2 Evaporative Emissions Diagnostics

With the introduction of the OBD II legislation (On-board diagnosis, 2nd generation) in North America, leak testing

Pressure-balancing valve

Control valve, current-controlled

Figure 19.18 Types of canisterpurge valves (source: Continental Automotive GmbH).

Internal Combustion Engine Handbook | 731

6606_Book.indb 731

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 19 Actuators

of the complete tank system became a legal requirement for the first time. This requirement was based on the observation that with a further reduction in the exhaust emissions, greater attention had to be paid to the evaporative emissions, as their share of the total emissions of the vehicle is very large. In particular, it was discovered that undetected leaks in the tank system and operating errors (e.g., lost or wrong tank filler cap) resulted in very high evaporative emissions over time. It, therefore, became a legal requirement for a diagnostic system to be installed on the vehicle that would detect all leaks greater than the throughput through a calibrated opening of 1 mm diameter. Here, the system has to be able to distinguish between a normal leak (e.g., leaking hose connection, tank damage) and major leaks (missing tank filler cap). When implementing this legislation in the vehicle, it was discovered that the technical complexity was far higher than had initially been assumed. In particular, the different climatic and operating conditions as well as the respective fuel level in the tank create a broad band of parameters that have to be adapted. In spite of the problems of implementation in the vehicle, the legislation was further tightened by reducing the diameter of the calibration opening from 1 to 0.5 mm. The tank diagnostics can be performed with vacuum and pressure systems. Both of these system types exhibit fundamental advantages and disadvantages irrespective of the components employed. The legislation permits diagnostics both during vehicle operation and at standstill. The pressure systems offer a few benefits during vehicle operation, while the vacuum method is favored for the vehicle standstill with the 0.5 mm legislation. Both technical (e.g., tank volume, tank form, space) and market-related aspects (e.g., the vehicle is sold only with the OBD II system, the vehicle is alternatively available also without the OBD II system, unit price per system in relation to the application costs, etc.) can influence the decision in favor of a particular diagnosis method. Furthermore, the experience gained to date and the strategy of the vehicle manufacturer greatly influences the choice of system. In Europe, on the other hand, leak diagnosis is no longer conducted because the associated cost is regarded as being disproportionately high. The only system that will be used in the future is one that will diagnose a correctly fitted

Tank

19.6.2.1 Tank Diagnosis with Gauge Pressure With the Siemens leakage diagnosis pressure pump (LDP I) (Figure 19.19), a pressure of up to approximately 20 hPa is generated in the tank system using the intake manifold gauge pressure via a clocked three-way valve and a springloaded diaphragm.

Figure 19.19 Tank diagnostics pressure pump (source: Continental Automotive GmbH).

Via the pump diaphragm, a switch detects the change in position, and the corresponding dropout time is compared with the default values stored in the control unit. A simple nominal/actual comparison then allows the leak tightness of the tank system to be evaluated. If a leaking tank system is detected, the diagnosis is repeated to eliminate all ambient influences on the result. Only when

Engine control unit

Intake manifold vacuum

Filter

tank filler cap, where a mechanical or electric switch contact is sufficient.

LDP I

Tank ventilation valve Aktivkohlebehälter

Intake manifold vacuum

Figure 19.20 Schematic of a tank diagnosis with gauge pressure (source: Continental Automotive GmbH).

732 | Internal Combustion Engine Handbook

6606_Book.indb 732

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

19.6 Evaporation Emissions, Components

Intake manifold vacuum

Solenoid valve (N.C.)

Tank ventilation valve

Vacuum valve Air filter

Overpressure valve N.O. vacuum valve

Activated charcoal tank

Tank

Figure 19.21 Tank diagnostics with vacuum (Siemens NVLD system) (source: Continental Automotive GmbH).

NVLD Komponente

the same fault is detected in two consecutive measurements is the OBD warning lamp switched on via the engine control unit. Additional software now makes reliable tank diagnosis possible with the LDP I even under the 0.5 mm legislation (Figure 19.20). 19.6.2.2 Tank Diagnosis with Vacuum The Siemens NVLD (natural vacuum leak detection) system (Figures 19.17 and 19.18), utilizes the ambient temperature influences allowing for the ideal gas law for tank leakage diagnosis (normal leak). The NVLD unit is directly connected to the tank or to the carbon canister. During engine operation, the vent is opened to the atmosphere via an electro-magnetically activated valve. When the vehicle engine is not running, the valve is closed and thus creates a tank system that is sealed to the atmosphere. The different operating states and ambient influences result in differences in temperature of the tank system and the fuel. Since the tank system is completely sealed to the atmosphere, these differences in temperature create changes in pressure in the tank. These changes in pressure also act on the diagnostic diaphragm that, in turn, is connected to a contact switch. When the tank system is leak tight, the changes in pressure trigger a switching signal that is registered by the vehicle electronics. If this switch signal is not received for a given period, the reverse conclusion is drawn that the tank has a leak.

In addition, it is possible to detect major leaks during engine operation. Here, the solenoid valve is closed and a vacuum is built up in the tank via the canister-purge valve. A possible major leak is then detected via the pressure diaphragm and the contact switch. Additional spring-loaded valves integrated into the NVLD unit ensure that certain threshold values for the pressure and/or vacuum level are not exceeded in the closed tank system (Figure 19.21).

Bibliography

19-1. Moczala, H. et al. 1979. Elektrische Kleinstmotoren und ihr Einsatz. Expert-Verlag. 19-2. Richter, C. 1988. Elektrische Stellantriebe kleiner Leistung. VDE-Verlag. 19-3. Kenjo, T., Nagamori, S. Permanent Magnet and Brushless DC Motors, Oxford Science Publications. 19-4. Vogt, K. 1996. Berechnung elektrischer Maschinen. VCH. 19-5. Leonhard, W. 1985. Control of Electrical Drives. Springer. 19-6. Luft, J. 1995. Elektromotorischer Systembaukasten Ansätze zur Gewichts- und Bauraumreduzierung. VDO. 19-7. Mönch, L. Überwachung im Verkehr befindlicher Fahrzeuge—AU— OBD—Wohin geht der Weg, IAV 5th Conference On-Board Diagnostics, April 2011, Braunschweig. 19-8. Netterscheid, M. Konzept zukünftiger Diagnosen im Bereich der Abgasnachbehandlung beim Dieselmotor, IAV 5th Conference On-Board Diagnostics, April 2011, Braunschweig.

Internal Combustion Engine Handbook | 733

6606_Book.indb 733

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 734

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20 Cooling of Internal Combustion Engines 20.1 General Information

20.2 Demands on the Cooling System

The growing demands with respect to fuel consumption, exhaust gas emissions, service life, ride comfort, and packages have led to modern cooling systems for internal combustion engines in the motor vehicle—with few exceptions—exhibiting the following characteristics:

Transient temperature peaks exceeding 2000°C occur inside the cylinder of a combustion engine. Charge cycles, expansion processes, etc., between the ignitions, however, result in far lower mean temperatures. Nevertheless, thermal overloading of the components exposed to the gas has to be prevented and the lubricating properties of the oil film between the piston and cylinder surfaces maintained by cooling. In water-cooled internal combustion engines, roughly estimated around one-third of the admitted fuel energy is discharged by the cooling system, a further third is lost through the exhaust gas, and one-third is transformed into useful work, depending on the combustion process (Figure 20.1).

•• water cooling of the engines with forced circulation of the coolant by a belt-driven centrifugal pump •• operation of the cooling system at a pressure of up to 1.5 bar •• use of a mixture of water and antifreeze, generally ethylene glycol, with a content of 30% to 50% v/v •• aluminum in corrosion-resistant alloys as the dominant radiator material •• the coolants exhibit additional inhibitors for the corrosion protection of aluminum radiators •• plastic as a dominant material for radiators, fans, and fan housings

Energy on the crankshaft 100,0%

•• temperature control using the fan drive and coolant thermostat •• use of intercoolers, engine oil, transmission oil, hydraulic oil, and exhaust gas coolers, depending on engine type, engine output, and equipment features

Exhaust gases 44,0%

•• preassembly of all front-end cooling components in one functional unit, that is, the “cooling module.”

Cooling system 29,7%

Apart from the many development activities for even more compact, lighter, and more efficient components, the electronically controlled cooling system is increasingly gaining in significance for the demands outlined at the beginning.

Radiation and convection 5,5% Fuel energy 20,8%

Figure 20.1 Energy balance in a water-cooled 1.9-l gasoline engine, driving steadily at 90 km/h in 4th gear.

Internal Combustion Engine Handbook | 735

6606_Book.indb 735

1/19/16 8:51 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 20 Cooling of Internal Combustion Engines

Passenger cars

Commercial vehicle (Euro IV)

Maximum amount of heat to be discharged from the coolant (KM) or charge air (LL) with gasoline engine with DI diesel engine

QKM = 0.4–0.6 Pmech QKM = 0.55–0.70 Pmech, QLL = maximal 0.15 Pmech

QKM + QLL = 0.70–0.85 Pmech with exhaust-gas recirculation QKM + QLL = 0.60–0.75 Pmech without exhaust-gas recirculation

Maximum permissible values for the temperature difference between the coolant at the radiator inlet and the ambient temperature.

around 80 K

around 65 K

Maximum permissible values for the temperature difference between the charge air at the radiator outlet and the ambient temperature.

around 35 K

around 15 K

Several thermally critical operating conditions are normally investigated during the design of cooling systems, such as “maximum speed on the flat,” “fast hill climbing,” or “slow hill climbing with trailer.” A distinction is also made between operations in Europe and countries with hotter climates. Travel speed, ambient temperature, heat volumes to be dissipated, and the nominal values for maximum permissible coolant, charge air, and oil temperatures are always given. Typical rules of thumb and nominal values for the main methods of cooling are summarized in Figure 20.2. Ranges for different operating conditions from the smallest car engine up to the most powerful truck engine are as follows: •• Maximum coolant temperature

100°C to 120°C

•• Maximum coolant throughput

5000 to 35,000 l/h

•• Maximum charge air throughput

0.05 to 0.6 kg/s

•• Maximum charge air inlet temperatures (at 25°C ambient temp).

In the radiators used in motor vehicles, heat is transferred from one flowing medium 1 through a fixed wall to a second flowing medium 2 from the higher to the lower temperature level (Figure 20.3). This heat volume is calculated using the parameters shown in Figure 20.3: l Q! = a1 ⋅ A ⋅ ( t1 − t1′ ) = ⋅ A ⋅ ( t1′ − t2′ ) = a 2 ⋅ A ⋅ ( t2′ − t2 ) d Q! Q! ⋅ d Q! (t1 − t1′ ) = a ⋅ A ; (t1′ − t2′ ) = l ⋅ A ; (t2′ − t2 ) = a ⋅ A ; 1 2 (20.1) Q! ⎛ 1 d 1 ⎞ 1 Q! t1 − t2 = ⋅ ⎜ + + ⎟ = ⋅ A ⎝a l a ⎠ k A 1



Air

Water

Heat flow

t1 t1¢

Temperature drops t1 – t2

t2¢ t2

110to260°C

20.3 Principles for Calculation and Simulation Tools

Q! = k ⋅ A ⋅ ( t1 − t2 )

Wall

Figure 20.2 Nominal values for methods of cooling.

2

Finning can be used to increase the heat-transfer coefficients α as compared to those of the flat surfaces. It must be assured, however, that an advantage is actually achieved here in spite of the resulting increase in the media’s flow resistance and the required higher delivery energy.

Figure 20.3 Temperature curve for heat throughput from the water side with the high temperature t1 through a wall to the air side with the low temperature2, t′ 1and t′ 2are the surface temperatures on both sides of the partition wall.

The basic aim when designing a cooling system is to provide the demanded cooling capacities with the most compact, lightweight, and inexpensive radiators possible within the installation space available. This requires an optimization process to be carried out with respect to the arrangement and dimensioning of the heat transfer elements in the module, the choice of the fin or pipe geometry of the radiators, the power consumption of the fan, the matching to the vehicle-specific boundary conditions, and frequently also the air drag coefficient value and crash behavior. Common design tools include analytical programs for heattransfer calculation using the unidimensional flow filament theory. Given the radiator geometry and the heat transfer, heat transmission, and pressure drop conditions as well as the material streams, the parameters’ pressure and temperature at the outlet from the heat transmitter can be calculated from the same inlet parameters. Backed up by empirical data from many years of measurement on a wide range of versions,

736 | Internal Combustion Engine Handbook

6606_Book.indb 736

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20.4 Engine Cooling Subsystems

MÖK

KMK

LU

KM MÖL

MÖK MÖL AC

LU KON

WIn d 111

KON 1 123

AC KON 2 133

KM

NT-KMK KON NT-KMK

KM QMÖ 142 L

SV1 115

KMK 1 114

MÖK 112

QAC 143

KMK 2 124

NT-K M 134 K

DPF Z 117

Air SV2 125 EL1 224

SV3 135

EL2 234

QGÖ L 144

Refrigerant QMOT

Oil

LU 116

154

Coolant

Figure 20.4 Topology model for a one-dimensional simulation of a cooling system in the vehicle.

fin or tube variants of different dimensions and for practically any operating points can be calculated very precisely in advance using these correlations within the framework of the similarity theory. Designs of complete cooling modules with full and partial overlaps of heat-transfer media, fans, and fan housings are almost exclusively demanded today. As a result, “topology models,” Figure 20.4, with several flow paths are created for these modules of which each can be calculated in turn using the flow filament theory. The mutual interaction of the components is taken into account by the calculation codes. Finally, this aid is complemented by elements, such as headwind, fan, and all pressure boosting systems, in the vehicle, such as radiator grille and engine compartment ventilation.

This enables the iterative calculation of the coolant throughput in the vehicle and consequently all the thermodynamic characteristic values of the cooling system. Combined with the broad experience gained from cooling power measurements in the wind tunnel, a quick and reliable simulation tool can be obtained which greatly decreases the need for vehicle measurements. The near future will bring the coupling of analytical unidimensional methods with the numeric three-dimensional computational fluid dynamics (CFD) methods as these can provide the detailed determination of the very complex cooling airflow in the engine compartment Figure 20.5.

20.4 Engine Cooling Subsystems 20.4.1 Cooling by Coolant

Figure 20.5 CFD simulation of the coolant circulation in the front end of a passenger car. See color section page 1091.

The nonferrous metal radiators with copper fins and brass pipes common in early years have disappeared almost completely in Europe. They have been replaced in passenger cars since 1975 and in commercial vehicles since 1988 by continuously further developed Al-alloys, which offer better corrosion resistance and a weight advantage of up to 30% with highpressure resistance due to brazing. Pipes and fins form the “radiator matrix.” A distinction is made between the following: Mechanically assembled rib and pipe systems of round or oval pipes and slot-fitted punched ribs, which are linked to one another by expanding the pipes (Figure 20.6). These systems typically cover the lower power segment but thanks to improved expansion techniques and ever-narrower oval pipes, also achieve the power spectrum of soldered systems of flat pipes and rolled corrugated ribs. Today, these are generally

Internal Combustion Engine Handbook | 737

6606_Book.indb 737

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 20 Cooling of Internal Combustion Engines

3

18

14.8

12 15

26

13

Figure 20.6 Mechanically assembled rib and pipe systems for coolant radiators with round and flat oval pipes.

Ø6

manufactured with only one tube in the system depth; they can be ribbed to increase the strength (Figure 20.7). System depths (extension in the coolant circulation direction) range from the smallest passenger car radiators up to the largest commercial vehicle radiators of 14 mm to 60 mm and in the case of nonferrous metal radiators, even up to 80 mm with cooling air faces of 15 dm 2 to 100 dm2. Aluminum has more or less asserted itself as radiator material in Europe. In the United States and Japan, nonferrous metal systems are also still widespread. As further regional differences, the coolant radiators for cars in Europe are mainly constructed in

cross-flow design with the pipes running horizontally (Figure 20.8), whereas in Japan they are frequently also built in the downdraft design. In commercial vehicles, the arrangement in the downdraft, that is, with the pipes running vertically, inside the vehicle frame is more widespread because the power variants can then be formed simply by using the pipe length with identical pipe shells and expansion tanks (Figure 20.8). The expansion tanks are always made of glass-fiber-reinforced polyamide and are mounted on the radiator block with a gasket and a bead.

10

42

with turbulence sheet without

2

Figure 20.7 Soldered flat pipe system for coolant radiators.

738 | Internal Combustion Engine Handbook

6606_Book.indb 738

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20.4 Engine Cooling Subsystems

Figure 20.8 Coolant radiator for passenger car in cross-flow arrangement and commercial vehicle cooling module with the coolant radiator in downdraft arrangement.

20.4.1.1 Radiator Protector In a liquid-cooled internal combustion engine, the waste heat is dissipated to the environment using a coolant to avoid overheating. Coolants are operating media just like the lubricants and fuels and have to satisfy the following requirements: •• optimum heat transmission properties •• high-thermal capacity •• low evaporation losses •• good antifreeze properties •• corrosion, erosion, and cavitation protection for all metallic materials •• compatibility with elastomers, plastics, and coatings •• prevention of deposits (fouling) and clogging •• temperature stability •• low-maintenance requirements •• long service life •• simple handling •• low-operating medium costs •• minimum environmental impact. The coolant generally consists of a mixture of tap water with a radiator protection medium tested and approved by the automobile and engine manufacturers, normally in a ratio of 50:50% v/v. Depending on the source, the tap water can exhibit significant differences in quality and have a considerable influence on the effectiveness of the coolant. For this reason, certain minimum demands are made on the quality of the tap water (Figure 20.9). Property

Unit of measure

Requirement

Appearance

–

colorless, clear

sediment

mg

0

pH value

–

6.5–8.0

Sum of the alkaline earths

mmol/l

0.9–2.7

Hydrogen carbonate

mg/l

≤100

Chloride content

mg/l

≤100

Sulfate content

mg/l

≤100

Figure 20.9 Minimum demands on the quality of the tap water.

The radiator protection medium consists of approximately 90% monoethylene glycol (1,2-ethanediol), 7% additives, and 3% water. When mixed with tap water, the monoethylene glycol causes such things as a drop in the coolant’s freezing point, which protects the entire engine cooling circuit from freezing in winter; for example, for a typical 1:1 mixture up to around −38°C. The monoethylene glycol is substituted in some products with monopropylene glycol (1,2-propanediol). The additives include substances for corrosion protection (inhibitors) and buffering, antifoaming agents, and pigments. The inhibitors here are of essential importance for the service life of the whole engine coolant circuit and effectively determine the quality of a radiator protection medium. The effectiveness of the inhibitors in the coolant provides additional protection for the materials in the engine coolant circuit against corrosion. Before approval of a radiator protection medium, the corrosion protection capabilities are specially evaluated in extensive laboratory and technical tests. After successful completion of the most important tests such as the glassware test in accordance with ASTM D 1384, the MTU (Motors and Turbines Union) knock chamber test, the Forschungsvereinigung Verbrennungskraftmaschinen e. V. (FVV) Research Association for Internal Combustion Engines) hot corrosion test, the FVV pressure aging test, the FVV vibration test, the water pump test in accordance with ASTM D 2809, and the circulation test in accordance with ASTM D 2570, the vehicle fleet test critical for the product approval is finally carried out by the vehicle manufacturers. During this practical test under the real conditions of road operation, the engine coolant circuits of the test vehicles are normally completely dismantled after running approximately 100,000 km and examined for signs of possible corrosion, erosion, and cavitations, and the results are evaluated. The compatibility with seal and hose materials and with plastics also plays an important role. This information together with the information gained on the test behavior of the coolant gives us a reliable overall view of the suitability of the radiator protection medium. Depending on the operating conditions, the coolant is subject to a natural aging. It is therefore essential to follow the service and maintenance instructions of the automobile and engine manufacturers. The coolant is usually changed

Internal Combustion Engine Handbook | 739

6606_Book.indb 739

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 20 Cooling of Internal Combustion Engines

completely after 100,000 km or after two years for passenger cars or after one year for commercial vehicles. New developments of organic inhibitor-based radiator protection media increase the service life of the coolant and contribute to reducing costs and conserving resources. Their market share is growing at a steady rate.

20.4.2 Intercooling

Turbocharging with cooled charge air has meanwhile become the general standard for the commercial vehicle diesel engine and is almost always used for the passenger car diesel engine with the aim of increasing the power density and of reducing fuel consumption and emissions. It is also finding greater attention than previously in the course of the further development of spark-ignition (SI) engines. The improved cylinder charge makes it possible for the increase in density achieved with the decreasing charge air temperature to be transformed into a higher output. Furthermore, the lower temperature reduces the thermal load on the engine and results in lower NOx contents in the exhaust gas. Intercoolers are preferably soldered aluminum flat-tube radiators, cooled directly by the cooling air. The system depths range from approximately 30 mm to over 100 mm, the face areas from 3 dm2 for cars and up to 100 dm2 for commercial vehicles. Many layouts are common in the car: large intercoolers in front of the coolant radiator, long and slim under or next to the coolant radiator, or totally separate from the module, for example, in the area of the wheel housing, hence, the large bandwidth in the system depths. The air receivers are made almost exclusively of plastic. In the commercial vehicle, large cross-flow arrangements in front of the coolant radiator are most widespread, with the bracket for the whole module being preferably fastened to their air receivers, making the intercooler the supporting element for the module. The previously common cast aluminum construction of the air receiver is being replaced more and more by high-temperature-resistant plastics (Figure 20.10).

Figure 20.10 Flat-tube cooler with air receiver made from hightemperature-resistant plastic for a lightweight commercial vehicle.

A current trend is intercooling using coolants (Figure 20.11). Compared with the air-cooled systems used today, these systems exhibit a smaller pressure drop on the charge airside. Furthermore, valuable space is saved in the front section of the vehicle, and handling dynamics are improved. To date, this technology is mainly being used in small numbers in high-performance engines for luxury class models. It is to be expected, however, that the use of coolants for cooling the charge air will become more prominent in future engine and vehicle developments

20.4.3 Exhaust Gas Cooling

Diesel engines have to comply with ever-stricter emission limits (Figure 20.12). These limits, currently defined as “Euro-4” level, can be achieved with low fuel consumption if the exhaust gas recirculation system familiar from the car is also cooled via an exhaust gas cooler. The cooling of noncombustible exhaust gas constituents and their admixing to the cylinder charge

Cooling air Thermostat

Coolant

Main radiator

Engine

Turbocharger

Pump

Low temperature radiator

Charge air / radiator

Auxiliary pump

Charge air

Figure 20.11 Diagram of a coolant circuit for passenger cars with indirect intercooling in a separate lower temperature circuit.

740 | Internal Combustion Engine Handbook

6606_Book.indb 740

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20.4 Engine Cooling Subsystems

reduce the combustion temperature and, consequentially, the NOx content of the exhaust gas. 0.2 1993

Particle (g/km)

0.15

20.4.4 Oil Cooling

Part of the waste heat from the engine is absorbed by the lubricating oil. With some powerful engines, the cooling by the oil sump is no longer sufficient to maintain the maximum permissible oil temperature and an engine oil cooler must be used. Engine oil coolers in the car are preferably located near the engine, are of a round disk, disk stack, or flat-tube design, and are made of aluminum (Figure 20.14) so that the cooling is achieved solely with coolant.

0.1 1996 0.05

2000 2005

0

0

0.2

0.4

0.6

0.8

HC + NOx (g/km)

1

1.2

Figure 20.12 Emission limits for passenger car diesel engines in Europe from 1993 to 2005.

The exposure of exhaust gas coolers to very high temperatures and extreme corrosion, especially in commercial vehicles, makes stainless steel indispensable here as a cooler material. Laser welding or nickel soldering are the typical joining processes. These coolers are designed as tube banks where the tubes containing the exhaust gas can be simple round tubes or tubes with special performance-enhancing but soilingresistant measures. The performance suffices for legislation level Euro 4 of around 2 kW for passenger cars and up to 80 kW for commercial vehicles. The range of dimensions is, therefore, correspondingly large. The length alone varies from around 100 mm up to around 700 mm. Series production is already underway for passenger cars and its numbers are increasing (Figure 20.13).

Figure 20.14 Oil cooler in disc stack design.

Direct cooling with oil and air coolers is also common, with soldered aluminum flat-tube designs having very high-pressure resistance being arranged in the cooling module. In commercial vehicles, cooling is always performed with coolant, while the coolers are normally installed in an opening in the crankcase where they are exposed to the main flow of the coolant. The most widespread design is with plate-type coolers of stainless steel with turbulence inserts on the inside through which the oil flows. More recently, the use of aluminum coolers with higher capacities and comparable strengths but roughly half the weight has become possible. Transmission oil coolers for cars with automatic transmissions can again be of air-cooled flat-tube design or installed as very slim and long flat-tube coolers in the expansion tank of coolant radiators where they are cooled by the coolant. The latter form predominates today, although disc stack coolers installed in the module are also becoming more and more widespread. Hydraulic oil used in power steering or other servo systems also has to be cooled. This is generally performed in simple pipe coils on a cooling module, and in rarer cases also with long-tube yokes fitted with fin packs by mechanical expansion.

20.4.5 Fans and Fan Drives

Figure 20.13 Double-flow exhaust gas recirculation cooler with corresponding valves for a passenger car V8 diesel engine.

Engine cooling fans are manufactured today almost exclusively of plastic. In addition to the axial blades, hoop rings and inlet guides are also included at the blade tips, depending on the operating states in the vehicle. Additional typical fan characteristics can be curved blades and asymmetric blade pitches (Figure 20.15). Such measures allow the fan efficiency and the noise emissions to be favorably influenced.

Internal Combustion Engine Handbook | 741

6606_Book.indb 741

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 20 Cooling of Internal Combustion Engines

Figure 20.15 Passenger car fan with curved blades and hoop rings for a drive with an electric motor.

Fans in passenger cars are usually installed in a single or double arrangement with maximum fan diameters of around 520 mm. With the exception of the most powerful engines, electric motors are used as fan drives. They consume up to 850 W of electrical power and are equipped with a stepped speed variation using series resistors or a stepless speed variation with brushless electric motors. The upper power segment of cars and the full range of commercial vehicles are equipped with viscous couplings as fan drives (Figure 20.16). Here, a drive speed dictated by the crankshaft or an engine-side gearing—generally that of the coolant pump—is transmitted from the primary side by oil friction to a secondary side connected to the fan. A variable oil filling of the coupling allows the fan speed to be varied from an idle speed to just below the drive speed. The maximum fan diameter installed in commercial vehicles is 815 mm with a power consumption of up to around 30 kW.

20.5 Cooling Modules Cooling modules are units consisting of various components for cooling and possibly air conditioning of a vehicle that include a fan unit with drive (Figure 20.17).

Figure 20.16 Viscosity coupling for commercial vehicle fan drives.

The modular technique that has become more and more widespread since the end of the 1980s because it offers several fundamental technical and economic benefits, which are as follows: •• optimum design and matching of the components •• higher efficiency in the vehicle or smaller and less expensive components possible •• less development, testing, logistics, and assembly costs for the vehicle manufacturer. In normal road vehicles, fixed cooling modules are almost exclusively used, attached to the longitudinal or transverse members of the vehicle. Generally, one of the heat transfer units serves as the module-supporting element, while the other components are snapped, clamped, or clipped to its

Figure 20.17 Cooling module for passenger car usage with coolant radiator, compensation tank, AC condenser, refrigerant collector, and E-fan with housing.

742 | Internal Combustion Engine Handbook

6606_Book.indb 742

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20.6 Overall Engine Cooling System

water tank or air shroud and side parts. The more components a cooling module contains, the more expedient is the use of a supporting frame to hold all the module elements.

20.6 Overall Engine Cooling System The design of the cooling system is dictated by the operating states of the vehicle where cooling is of critical importance, such as driving at high speed or climbing hills, with a large trailer load in the summer with the air-conditioning system switched on. However, these critical cooling states occur only very seldom during a vehicle service life. This means that for the majority of the vehicle-operating life either excessively high fluid flows are pumped for the engine cooling or the temperatures in coolants or oils are too low or too high. This increases the fuel consumption and, consequently, the exhaust emissions. Ride comfort is also impaired and the service life of the engine and attachments reduced. The target for future cooling systems is to control all the fluid temperatures and media flows through a demand-oriented control of the engine cooling in such a way as to minimize energy demand and/or, depending on the priority, to achieve comfort, emissions, or service life advantages. This will require control interventions in the engine cooling to be possible in the future. In present-day cooling systems, the following control options for the fluid flows oriented to the cooling capacity requirements are already implemented: •• A thermostat whose wax element senses the temperature of the coolant flowing around it ensures that the coolant flow passes either through the coolant radiator or circumvents it through a bypass line. Cooling of already very cold coolant can thus be more or less avoided or maximum cooling can be assured at very high temperatures. •• Electrically powered fans are switched on at various speeds or with an infinitely variable speed according to the coolant temperature in the expansion tank. •• On fans with viscous coupling, the oil filling and, hence, the fan speed is controlled depending on the cooling air temperature in front of the coupling. Hot cooling air is produced by flowing through hot heat transfer elements. This is a sign of a high-cooling requirement, and the fan is switched on by a bimetallic element. •• All other systems designed for critical operating conditions, however, are then operated without control. The coolant pump, for example, is driven via a belt by the crankshaft, charge air cooling is then almost always uncontrolled, and the oil cooling is only partially thermostatically controlled. Such cooling systems were sufficient until now and were characterized by their very reliable operation. Future technologies for cooling systems, however, will be based on electronic control as are many other systems of the automobile. A control unit uses a network of sensors monitoring the thermal condition of the engine and cooling system to trigger adjustments of conveyor equipment (fans, pumps) and control elements (valves,

flaps, shutters) on the basis of the stored control concepts to save drive energy in auxiliary systems, to favorably influence exhaust gas and noise emissions, and to shorten heat-up phases in the sense of enhancing comfort and reducing wear by demand-oriented cooling. This requires that all the conveyor and control elements be controllable. This option for the thermostat has been created by the electric heating of the wax element (Figure 20.18). As a result, the thermostat position can be set independently of the current coolant temperature on the basis of set points from an engine map. The possibility of increasing the temperature during part-load operation of the engine reduces the fuel consumption.

Figure 20.18 Coolant thermostat with electrical heating of the wax element.

In present-day vehicles, the coolant flow is generated by a coolant pump that is driven by a belt proportionally to the engine speed. The use of switchable or controllable pumps will be expedient in the future to reduce the coolant throughput with low-cooling requirements, while at the same time being able to supply more coolant to the heater during the warm-up phase. In passenger cars, this can take the form of electric pumps. An on-board supply system of 42 V is required for larger engines. By being separated from the engine belt drive, the electric pump offers new design scopes. Alternatively, the coolant flow can also be influenced by the use of controllable throttle elements or switchable couplings in combination with the mechanical pump. Apart from the control of the coolant flow in the main cooling circuit, there are also concepts for splitting the coolant stream into several circuits. These include the indirect intercooling described in Section 20.4.2 which is designed in separate circuits or on the low-temperature circuits attached to the main circuit. Circuits are also used for transmission oil temperature control in which the temperature transfer medium is supplied either with hot coolant to heat the oil during the warm-up

Internal Combustion Engine Handbook | 743

6606_Book.indb 743

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 20 Cooling of Internal Combustion Engines

phase or with cold coolant from a low-temperature section for cooling the oil. A thermostat ensures the switchover from heating to cooling. The controlled delivery and throttling of the cooling air also offers a great improvement potential for the future. The stepped-speed electric fans in cars will be increasingly replaced by variable-speed fans with EC motors. Viscous couplings for commercial vehicles can meanwhile be electrically controlled because the oil filling is controlled by an electromagnetically actuated valve and no longer by a bimetallic element. This permits a control of the fan speed and a quick switching on and off of the fan. In many driving situations, the fans are switched off. Nevertheless, a high-cooling airflow is required at high travel speeds, hence, an increase in the drag of the vehicle. The use of aerodynamically optimized cooling air shutters here can reduce the fuel consumption and at the same time the noise emissions. In addition, a faster heating of the passenger compartment and of the engine is achieved in winter as heat losses can be reduced by isolating the engine compartment from the cold surrounding air.

Bibliography

20-1. Mollenhauer, K., Tschöke, H. 2007. Handbuch Dieselmotoren, Springer Berlin Heidelberg. 20-2. Knauf, B., Pantow. E. 2005. Auslegung eines Kühlsystems mit elektrischer Kühlmittelpumpe (Design of a cooling system with an electrical coolant pump), MTZ 11/2005, 66. 20-3. Kemle, A., Manski, R., Weinbrenner, M., 2009. Klimaanlagen mit erhöhter Energieeffizienz (Air conditioners with increased energy efficiency), ATZ 09/2009, Seite 650. 20-4. Strehlow, A., Leuschner, J., Scheffermann, J. 2008. CFD-Simulation in der Entwicklung von Hochleistungs-Wärmeübertragern (CFD simulaton in the development of high-power heat exchangers), MTZ 04/2008 69. 20-5. Edwards, S. u. a. 2008. Emissionskonzepte und Kühlsysteme für Euro 6 bei schweren Nutzfahrzeugen (Emission concepts and cooling systems for Euro 6 in heavy-duty utility vehicles), MTZ 09/2008, 69. 20-6. Heinz, M. 2008. Prozessautomatisierung für Motorkühlmodule mit CFD (Process automation for engine cooling modules with CFD), MTZ 12/2008, 69. 20-7. Berger, C., Troßmann, T., Kaiser, M. 2008. Heißkühlung—Kühlmittelzusätze auf dem Prüfstand (Hot cooling—coolant addities on the testbed), MTZ 02/2008, 69. 20-8. Williams, D. J. 2009. Vermeidung von Kavitation in Kühlmittelpumpen (Preventing of cavitation in coolant pumps): MTZ 02/09, 70. 20-9. Thumm, A. u. a. 2007. Hochleistungs-Kühlsysteme als Beitrag zur Erfüllung zukünftiger Abgasnormen (High-power cooling systems as a contribution to meeting future exhaust gas standards), Wiener Motorensymposium.

744 | Internal Combustion Engine Handbook

6606_Book.indb 744

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21 Exhaust Emissions Since the 1940s there have been systematic efforts in California to reduce the effects of mass transportation on air quality. In Europe, alarm was raised over automobiles in the 1960s because of carbon monoxide emissions directly harmful to humans. This led to the restriction of uncombusted exhaust components, such as carbon monoxide and hydrocarbons. Because of the increase in trace gases from combustion and their dissemination to remote areas, trees became damaged in the 1970s and 1980s due to acid rain and photooxidants, among other things. Since nitrogen oxides and uncombusted hydrocarbons contribute to the formation of these substances, it became imperative to limit the amount of these emissions into the atmosphere. This was accounted for in the United States with the introduction of exhaust gas limits for road traffic starting in 1961, in Japan starting in 1966, and in Europe as of 1970. Carbon monoxide (CO) emissions, which are directly harmful to humans, were reduced to a harmless level by legislation in industrialized countries. The drastic restriction of nitrogen oxides and hydrocarbon emissions began in the United States and Japan in the early 1980s. Central European countries were not far behind as they also strongly reduced these trace gases from passenger cars and power plants by the late 1980s. An increase in nitrogen dioxide (NO2) concentrations, however, can currently be noted again in certain European regions. By the beginning of the 1990s, it became clear that other exhaust emissions that were not harmful to humans could influence the earth’s atmosphere. These effects that are summarized by the term “greenhouse effect” led people to focus on carbon dioxide (CO2) emissions. Although the continuous reduction in fuel consumption by individual vehicles in the transportation sector has led to just a slight increase in fuel consumption by individual motorists, the strong increase in energy consumption for heating purposes and to produce electrical energy has led to a substantial increase in the CO2 concentration in the atmosphere. The focus is now on the more economical use of primary energy sources.

21.1 Legal Requirements In this section, we address automotive exhaust emission thresholds for carbon monoxide (CO), hydrocarbons (HC), nitrogen oxides (NOx) and particles (PM) for the European Union, USA, and Japan. Since the statutory exhaust emission thresholds are indicated in different units [(g/km), (g/test), or (g/mile)], they were recalculated to (g/km) for comparison in this chapter. For this reason, a direct comparison of exhaust emission thresholds is only possible when the emissions are measured using the same test cycle. However, this does not usually occur. To measure exhaust emissions of passenger vehicles directly off the assembly line during type approval, there are numerous prescribed procedures worldwide. The most important procedures for passenger cars are listed here: •• U.S. procedure in its 1975 version (FTP 75) with the added test cycles SC03 (with air conditioning), and US06 (aggressive driving): The U.S. highway test cycle •• EG ECE 15/04, EG MVEG-A test cycle •• Japanese 10.15-mode-Test, Japanese 11 mode-cold test, from 2005 to 2011 Introduction of the new test cycle JC08M.

21.1.1 Europe

The European emissions regulations for new passenger cars were originally specified in European Directive 70/220/EEC. This contains the thresholds (ECE R 15) defined by the UN Economic Commission for Europe (ECE). Changes to these regulations are found in the Euro 1 and 2 standards that became valid under directive 93/59/EC. The limit values published in Directive 98/69/EC in accordance with Euro 3 and 4 (2000/2005) were accompanied by an introduction of improved fuel qualities. A minimum diesel oil cetane number of 51 was stipulated, along with a clear reduction of sulfur both in gasoline and in diesel fuel. The Euro 5 Standard currently applies. It will be replaced in 2015 by Euro 6. The changes in the thresholds for gasoline engines were negligible.

Internal Combustion Engine Handbook | 745

6606_Book.indb 745

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

CO limit values [g/km]

35 ECE driving cycle Vehicle mass 1200 kg Displacement 1.4 to 2.0 l

28 21

New European driving cycle

EU

14

Euro II

Euro III

7

Euro IV

(HC+NOx)-limit values [g/km]

0 10 8

CVS process, FID measurement

6 Euro II

4 2 0 1975

1980

1985

1990

Year

1995

However, a particle threshold was added. The thresholds for diesel engines are adapted to those of the gasoline engines. The operation cycle for these regulations is found in ECE R83 (91/441/EEC). The test is carried out according to 98/69/ EC. The development of standards over time for passenger cars up to now is found in Figure 21.1. The currently valid thresholds for gasoline and diesel are in Figure 21.2.

Drive

Emissions category

CO

Diesel

Euro 5

0.50

Euro 6

0.50

Euro 5

1.00

Euro 6

1.00

Gasoline

HC

HC+NOx [g/km]

NOx

PM

–

0.23

0.18

0.005

–

0.17

0.08

0.005

0.10

–

0.06

0.005

0.10

–

0.06

0.005

Figure 21.2 Current exhaust gas standards passenger cars in the European Union.

21.1.2 Japan

The first carbon monoxide emission restrictions for passenger vehicles were introduced in Japan in 1966 and use the fourmode test, which has since been discarded. In 1973, HC and NOx were limited for the first time, and the 10-mode test started to be used. The thresholds introduced in 1975 lowered vehicle emissions of CO and HC by 90%. This 90% reduction was achieved for NOx emissions due to the requirements in 1976 and 1978. Depending on the drive and engine designs, different thresholds apply to vehicles within the same category. For example, a distinction was drawn between vehicles fueled by gasoline and LPG. In passenger cars with diesel engines, a distinction was made according to combustion method (direct fuel injection or indirect injection engine) and the origin of the vehicle (Japan or imported).

Euro III

2000

2005

Euro IV

2010

Figure 21.1 Development of the exhaust gas standards over time for passenger cars with gasoline engines in the European Union.

The emission standards as of 1997 and the new thresholds as of 2009 are listed in Figure 21.3. The current test method is the JC08M cycle that replaces the older 10-mode cycle and 10–15-mode cycle. This test corresponds somewhat to the European ECE + EUDC (European extra urban driving cycle), but at a lower speed level.

21.1.3 Harmonization of the Exhaust Gas Regulations

Efforts are being made to recognize certifications in other countries to prevent development and approval expenses from being excessively high (UN-ECE 1958 Agreement). This is a good idea that can be seen by comparing the most recently valid NOx + HC thresholds from the preceding sections. We note, however, somewhat different requirements for exhaust purification systems since Europe and Japan tend to value a quick catalytic converter light-off after a cold start at a simultaneously low engine load, and more weight is given to transient engine behavior in the United States.

21.2 Exhaust Measuring Technology 21.2.1 Measuring Technology for the Certification of Motor Vehicles

In general, the time and cost of these measuring procedures is very high for the reasons listed below. They all require a chassis dynamometer that must be calibrated to the respective vehicle, climatized test rooms to test specific cold-start conditions, and numerous highly sensitive exhaust emission-measuring devices.

746 | Internal Combustion Engine Handbook

6606_Book.indb 746

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.2 Exhaust Measuring Technology

CO

CO

HC

HC

NOx

NOx

PM

PM

Max

mean

Max

mean

Max

mean

Max

mean

Vehicle weight

Emission category

Diesel <1265 kg

1997

2.7

2.1

0.62

0.40

0.55

0.40

0.14

0.080

2002

–

0.63

–

0.12

–

0.28

–

0.052

2005

–

0.63

–

0,0241

–

0.14

–

0.013

2009

–

0.63

–

0,0241

–

0.08

–

0.005

1997

2.7

2.1

0.39

0.25

0.48

0.25

–

–

2002

–

0.63

–

0.08

2005

–

1.15

2009

–

1.15

Gasoline

[g/km]

–

0.08

–

–

–

0,05

1

–

0.05

–

–

–

0,051

–

0.05

–

0.005

Nonmethane HC

1

Here is a list of generally applicable features of type approval tests: •• conditioning the vehicle at room temperature for approximately 12 h •• cold start and measurement of starting emissions •• dynamic test cycle with speeds from zero to 120 (km/h) •• second test cycle with a hot start (U.S. FTP 75 Test) •• precise measurements of exhaust emissions. Figure 21.4 shows the basic arrangement of a vehicle chassis dynamometer for measuring exhaust emissions under conditions for certification. According to the presently valid laws governing approval, the dilution measuring technique must be used to determine mass emissions. The molecule-specific absorption of bands of infrared light is used as a measuring standard for the exhaust emission components carbon dioxide (CO2) and CO. Fame ionization detectors (FID) are used to measure hydrocarbons (HCs). From a chemical viewpoint, the measuring technique of the flame ionization detector is based on the ionization of

Exhaust gas bag CVS system

Exhaust gas analyzers CO

COx NOx

Heated line

HC

Figure 21.3 Japanese exhaust gas thresholds for passenger cars. The maximum limits apply for the production of less than 2000 vehicles annually. The mean limits for larger production volumes.

oxidizable hydrocarbon compounds in a hydrogen flame. Basically, the detector signal is proportional to the number of supplied hydrocarbon atoms. To detect nitrogen oxides (NO and NO2), measuring devices are used that are based on chemiluminescence detectors (CLD). Particles are measured as mass emissions per kilometer using partial flow filtering and gravimetric evaluation. To measure opacity in the recurring inspections of operated vehicles, partial flow opacimeters are used to determine the k value while operating the engine under “free acceleration.” Particularly exhaust thresholds that correspond to the ULEV or the Euro 5 standard place special demands on exhaust measuring technology since the concentrations in the exhaust must be measured when the engine is running hot and emission concentrations are very low. For this reason, new measuring methods have been suggested that deviate from the earlier used dilution measuring methods [21-1] and directly measure the concentrations in the exhaust emissions. An example of the direct evaluation of the exhaust mass emissions of nitrogen oxide is shown in Figure 21.5.

Indoor air bags

HC heiß

Fresh air Heat exchanger Chassis dynamometer

Dilution tunnel

Filter Exhaust gas

Figure 21.4 Classification of the roller testbed technology for exhaust certification tests.

Internal Combustion Engine Handbook | 747

6606_Book.indb 747

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.2.2 Measuring Technology for Engine Development

The drastically stricter exhaust provisions require a detailed analysis of the origin of the exhaust components and the exploitation of every strategy to further reduce emissions using the latest measuring technology. A primary condition for future engine development is to lower vehicle fleet fuel consumption. In addition to the restricted pollutant components of total hydrocarbons, CO and NOx, there are various other “unrestricted” pollutant components that arise during engine combustion, such as benzene, toluene, xylene, aldehydes, or ammonia. Either these components are already in the fuel and pass uncombusted into the exhaust emissions or they are formed during combustion in the engine. Since specific components such as benzene are hazardous to health and smell bad, it is becoming increasingly important to detect these components. In particular, there is great potential for improving exhaust emission values by optimizing transient engine operating conditions that make it easier to adapt to the operating conditions of the exhaust purification system. Fuel consumption can be influenced to a much greater degree, however, by other measures such as the combustion procedure and drivetrain management. The goal of future research is, therefore, to develop engine management so that changes in load and speed do not cause a noticeable deviation from the optimum lambda characteristic that is determined by the functional principle of the catalytic converter. To accomplish this task, it is necessary to carry out high time resolution measurements in the combustion chamber and at specific points in the exhaust system to determine the precise sources of the emissions. Given the dominance of transient operating conditions, it is also necessary when developing engines to carry out experimental investigations on the dynamic simulation test bench to adapt mixture formation and control systems. These

efforts also alleviate the great amount of effort involved in preparing a complete vehicle for the chassis dynamometer. The goal of simulation on the dynamic engine test bench is to achieve the same speed and torque characteristic for the engine crankshaft as well as to attain equivalent temperatures, fuel consumption, etc., as experienced on the road or when operating the vehicle on a chassis dynamometer. The primary advantage lies in the high reproducibility of the individual test runs. The illustration in Figure 21.6 shows a measuring setup for a modern development test bench that also permits the measuring of the individual combustion cycles. The setup is typically divided into a simulation computer, highly dynamic electrical load machine, a crank-angle-related measured data memory, and high-speed exhaust measuring equipment. The following emission measuring equipment for engine analysis is used on these test benches that can be categorized according to response speed: •• standard measuring devices with a response time in seconds and above •• transient measuring equipment with response times in the 100 ms range •• measuring devices and methods for individual cycle analysis with response times around 1 ms. The measuring devices can also be categorized according to their sites of use on the engine: •• measuring devices that analyze a sampled and conditioned partial flow of the exhaust emissions. The majority of measuring procedures fall under this category. •• sensors and measuring devices that are used in the exhaust system in situ. •• measuring methods for experimentally determining the gas composition in the combustion chamber.

20 18

NO [g/km], v/10 [km/h]

16

NO [g/km]

v/10 [km/h]

14 12 10 8 6 4 2 0

0

200

400

600

800

1000

1200

time [s]

Figure 21.5 Progression of the vehicle speed and the nitrous oxide mass emissions across the FTP 75 test cycle [21-2].

748 | Internal Combustion Engine Handbook

6606_Book.indb 748

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.2 Exhaust Measuring Technology

Driving pattern Road parameters Vehicle parameters Driver parameters

High Speed FID

Electrical loading machine

l-Sensor

Exhaust pressure sensor

Function of crankshaft angle Measurement recording

Speed measurement Torque specification Torque measurement Throttle valve position Simulation recorder

Figure 21.6 Measuring setup for exhaust gas measurement based on the crankshaft angle.

21.2.2.1 Partial Flow Measuring Devices A measuring setup for the engine with the most important exhaust emission measuring devices is shown in Figure 21.7. Please refer to Section 21.2.1 for a discussion of the physical principles of the exhaust emission measuring devices for restricted exhaust components. Furthermore, the very important oxygen measuring method based on the paramagnetic properties of the oxygen molecule is discussed. In addition to the classic devices for measuring restricted exhaust emission components, mass spectrometers and particle size measurements are now used. Mass spectrometry determines the ratio of mass to charge for ions or their components. This is done by diverting the ions in magnetic and electrical fields or determining their movement energy. Theoretically, each exhaust component or several can be determined at the same Mass spectrometer Particle size Mixture formation

←

Air mass measurement

←

l

HC CO

Fuel measurement

NOx

CO2 O2

Figure 21.7 Setup of the exhaust measuring technology for partial flow measurement on the engine testbed.

time by their number of moles. However, two effects stand in the way. On the one hand, there are several relevant exhaust emission components that have the same number of moles, and on the other hand, the simultaneous measurement of several components greatly increases the measuring time. Gas chromatography can selectively be used to detect nonrestricted exhaust emission components. In particular, the detection of size-class-dependent particle emissions represents a very time-intensive test at present. A distinction is drawn between methods that work according to the impactor principle that allows a specific number of size classes to be gravimetrically determined at the same time based on the aerodynamic properties of the particles, and methods that are selective that can measure only one size class as shown in Figure 21.8. These measuring devices combine several measuring principles. They distinguish between individual particle size classes by the variable electrical charge and subsequent aerodynamic evacuation. The particle fraction is then fed to a condensation nucleus counter in which the number of particles per volume unit is determined. 21.2.2.2 In Situ Exhaust Measurement in the Exhaust System To measure the air/fuel ratio at a high response speed both in the lean mixture range (λ greater than one) and in the rich range (λ lower than one), oxygen sensors are used that are mounted in the exhaust emission system. These “wide range sensors” use the oxygen ion pump principle and are available from various manufacturers. However, there are several factors that greatly influence the measuring precision of these sensors.

Internal Combustion Engine Handbook | 749

6606_Book.indb 749

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Neutralizer

Neutralizer

Condensation nucleus counter

Motion analyzer

Figure 21.8 Mobility analyzer to establish size class distribution for particles [21-3].

Figure 21.9 shows the most important errors that can arise when these sensors are used. The typical error from excess exhaust counterpressure can be 20% of the measured value of λ = 2. These sensors are becoming increasingly important since they are also used in engine management as control sensors for lean engines. Exhaust gas temperature

Model formation for diffusion and thermodynamic equilibrium

Exhaust gas pressure

NO concentration in the exhaust

Ambient air Diffusion layer Exhaust gas l-Sensor

Catalytic reaction on the circuit board electrodes

Figure 21.9 Impact of errors on the determination of the air/fuel ratio using a λ sensor.

The basic design and the output signal as a function of the lambda value can be shown in Figure 21.10. For individual cycle analyses with a high time resolution, sensors made of strontium titanate (SrTiO3) are useful. These sensors have a response time of approximately 5 ms and permit the detection of the mixture composition of individual cylinders

from the total engine exhaust. This represents a substantial advance, especially in the cylinder-selective comparison of the fuel injection system while taking into account the charge of the individual combustion chambers. Based on a similar technology as the lambda sensors in Figure 21.10, NOx sensors were made for in situ measurements [21-4] that offer useful support in the development and control of catalytic converter systems for lean engines. 21.2.2.3 Exhaust Measurement in the Combustion Chamber: The gas composition and individual gas components can be measured using different methods as shown in Figure 21.11 that are divided into two main groups: Optical measuring methods that are applied directly in the combustion chamber and are usually used for experimental engines (such as “glass engines”) and methods that withdraw gas directly from the combustion chamber. Optical measuring uses different physical and quantum mechanical properties of molecules or atoms to determine the percentage of a specific gas component. These measuring methods are generally able to determine the distribution of numerous components at the same time. The methods based on gas withdrawal work with either pulsed gas withdrawal valves or capillaries experiencing continuous flow. The measuring devices that are used in conjunction with gas withdrawal valves basically correspond to standard devices that are used for conventional exhaust emissions analysis. To obtain a sufficiently large gas volume for analysis, the average over numerous combustion cycles must

750 | Internal Combustion Engine Handbook

6606_Book.indb 750

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.2 Exhaust Measuring Technology

Exhaust gas

5 4

Pump flow rate lp [mA]

3

O2–

Us ~ lp

Air » 450 mV

2 1 0 –1 –2 –3 –4 –5 0,8

UH

1,0

1,2

l

1,4

1,6

1,8

Operating temperature ‡ 600 °C

IP : Pump flow rate US : Sensor voltage UH : Heater voltage Figure 21.10 Principle setup and output signal of a λ sensor in accordance with [21-4].

be calculated. Individual cycle analyses, especially concerning cyclic fluctuations or transient effects during load changes can be done using only continuous gas withdrawal and gas analyzers that react very quickly for system-related reasons. Today, these are used for the exhaust emission components of hydrocarbons and nitrogen oxide. The access to the combustion chamber is fairly easy when the continuous gas withdrawal method is used (Figure 21.12). Since no mechanical actuation is required in the direct environment of the combustion chambers or the withdrawal site,

the measuring position is less restricted. Another property of this sampling method is the favorable local resolution because of the lack of fuel condensation on the wall at the withdrawal valve and the simple geometric shape of the inflow cross section. These measurements in the combustion chamber can provide much information about mixture formation and combustion up to the point at which pollutants arise. The results can be valuable for emission reduction and for optimizing various constructive details. A very interesting use is in the development

Gas concentration measurement in the combustion chamber

Optical measuring process

Spontaneous Raman Spectroscopy (SRS)

Laser-induced fluorescence (LIF, LIPF)

Laser excitation

Laser-induced fluorescence (LIF, LIPF)

Chemically thermal fluorescence

Light emitted by the flame

Measuring process with gas extraction

Laser excitation

Continuous gas removal

Figure 21.11 Possible measurement methods for determining the gas composition in the combustion chamber [21-5].

Internal Combustion Engine Handbook | 751

6606_Book.indb 751

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Pressure quartz

Exhaust valves

Gas extraction

10 mm

High Speed FID t90 < 3 [ms]

Inlet valves

of combustion systems with direct fuel injection and lean designs in which the focus lies on the variable mixture at the spark plug and, hence, the cycle-selective lambda value. In addition to engine experiments involving combustion, experiments with engines operating without combustion are also carried out. By definition, the HC concentration in the combustion chamber can be used in operation without combustion to calculate the local air/fuel ratio. The measurements also provide information about the process of mixture formation and the residual gas content in the area of the spark plug or the withdrawal site.

21.3 Pollutants and Their Origin An exothermic reaction releases energy as heat during the combustion of fuels with oxygen from the air, which contains 21 [%-Vol.] O2, <1 [%-Vol.] noble gas and nitrogen N2. The release of heat for fuels based on hydrocarbons, such as gasoline and diesel fuel, is determined by a large number of incomplete reactions, dependent on the composition of the fuel’s hydrocarbons. Important fuel components include paraffin, olefin, and aromatic compounds. In theory, the only components formed from the complete combustion of hydrocarbons under ideal conditions or excess air are carbon dioxide and water, and from the oxygen carrier air–nitrogen. For this reason, the air/fuel ratio lambda (λ ) represents the most important characteristic value for the combustion process. Lambda is defined as the ratio of actually existing air volume relative to the ideal stoichiometrically required volume.

(

)

l = ( mL /mK )/( mL /mK )stoich = ( mL /mK )/ mL,th = mL /mL,th (21.1)

mL = the air volume per time unit supplied to the engine

Figure 21.12 Measuring setup for the continuous hydrocarbon indexing on the 4-valve gasoline engine.

1 [CHψ Oφ ] + 4,762 ⋅ (1 + ψ /4 − φ /2) · λ [air] combusted into 1 [ CO 2 ] + (3.762 ⋅ ( 1 + y/4 – j/2 ) ⋅ l – n / 2 [ N 2 ]

+ (( 1 + y/4 – j/2 ) ⋅ ( l – 1) – n/2 )[ O 2 ] + n[ NO ] (21.2) + y/2 [H 2 O]



[  ] = component n

= mol nitrous oxide

ψ = hydrogen/carbon atom ratio of the fuel φ = oxygen/carbon atom ratio of the fuel Along with the main exhaust components, such as CO2 and water vapor, the main pollutant agents are the legally restricted components carbon monoxide (CO), uncombusted or partially combusted hydrocarbons, HC (Aldehyde, Ketone etc.) and nitrous oxide, NOx. Pollutants form primarily from the interruption of the reaction chain due to the short residence time in the combustion chamber. An equilibrium state therefore no longer exists. Inhomogeneities in the mixture due to different air/fuel ratios λ , combustion wall effects and contaminants and additives contained in the fuel also lead to the formation of undesired by-products (Figure 21.13).

N2

Air (N2, O2) Air-fuel mixture Fuel (CxHy)

Reacting elements

O2 C H2

N2 NOx O2 H2O CO2 CO R.CHO CxHy H2

mL = the fuel volume per time unit supplied to the engine mL,th = the air volume theoretically required for the complete combustion of this fuel volume. The following general reaction equation can thus be used for combustion in the operating range with excess air:

Figure 21.13 Reaction mechanisms in the combustion chamber [21-6].

Solids in the form of particle emissions are also possible as a function of the fuel used and the combustion process of solids in the form of particle emissions.

752 | Internal Combustion Engine Handbook

6606_Book.indb 752

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.3 Pollutants and Their Origin

an air/fuel ratio of 1. Figure 21.14 shows the calculated ideal exhaust concentrations.

Unregulated exhaust components which form, for example, from thermal cracking processes of the hydrocarbons and their consequential products are coming increasingly to the forefront, since they either have certain hazard potentials or contribute to unpleasant odors.

21.3.1.1.2 Carbon monoxide CO arises as an intermediate step in the formation of carbon dioxide and from incomplete combustion under a lack of oxygen. This is characterized by the “water/gas equation.”

21.3.1 Gasoline Engine

Combustion in spark-injection (SI) engines possesses the following general characteristics:

CO + H 2 O ⇔ CO 2 + H 2



(21.3)

Essentially, carbon monoxide formation is determined by the local excess air factor and the temperature and pressure. When there is insufficient air, the CO emissions have a nearly linear relationship to the air/fuel ratio. The CO emissions result from a lack of oxygen. When λ > 1 (excess air), the CO emissions are very low and nearly independent of the lambda value. The CO emissions are also largely independent of other parameters, such as the compression ratio, load, moment of ignition, and fuel injection law.

•• externally supplied ignition executed as single or multiple ignition •• a compression ratio of 8 to 14 depending on the used fuel •• 4-cycle and 2-cycle processes are used. An important parameter is the air-fuel ratio λ , which determines the combustion behavior in quite narrow limits. Depending on the approach to combustion and exhaust purification, a constant air/fuel ratio for the entire combustion chamber is selected, or stratified charging with varying ratios in the combustion chamber. Indirect fuel injection into the intake manifold and direct fuel injection into the combustion chamber are possible methods for mixture preparation.

21.3.1.1.3 Hydrocarbons HC emissions arise from uncombusted and partially combusted hydrocarbons and corresponding thermal cracking products. These components can originate from both the fuel and the lubricants. Various mechanisms are responsible for these emissions. For example, incomplete combustion of the hydrocarbons occurs because of partial ignition of the overall combustion chamber volume, and wall deposits of fuel. Other reasons are residual fuel in the dead spaces, such as gaps in the cylinder head seal, valve seats, fire land, piston rings, spark plugs, and squish areas. Misfiring, emissions of hydrocarbons from the lubricant, and absorption of fuel molecules in the lubricant film of the cylinder barrel and at sites with impurities also increase hydrocarbon emissions. If one plots the mass-related HC emissions over the exhaust cycle, then we find increased

21.3.1.1 Restricted Exhaust Components 21.3.1.1.1 Carbon dioxide In Europe, carbon dioxide emissions are restricted exhaust components although they are not toxic. Legal regulations are increasingly restricting CO2 exhaust. Carbon dioxide arises from the complete combustion of the hydrocarbons in the fuel molecules. The CO2 emission essentially depends on the fuel consumption and the fuel composition, and it attains its relative maximum given a complete conversion at

CO2, CO, HC, O2, H2 [%-Vol] dry exhaust gas, 20 °C 15

CO2 CO O2 H2 HC

10

5

0 0,6

0,7

0,8

0,9

1,0

1,1 Air ratio I

1,2

1,3

1,4

1,5

1,6

1,7

Figure 21.14 Calculated exhaust concentrations using the air-fuel ratio lambda for a gasoline engine.

Internal Combustion Engine Handbook | 753

6606_Book.indb 753

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

n = 2500 [1/min], pme = 5 [bar], A/F: 14.7 [-] AÖ

4000

AS

HC3 mass flow [g/s]

0,20

3500

0,15

3000

Exhaust gas mass flow HC3 concentration HC3 mass flow

0,10

2500

0,05

2000

0,00

1500

–0,05

0

90

180

270

360 450 Crankshaft angle [°]

HC emissions in the curve plotted against the crank angle shortly after the exhaust valve opens and before it closes that appear to be primarily determined by the previously cited wall phenomena (Figure 21.15). When the flame front extinguishes as it contacts the cold wall (flame quenching), HCs are released since the mixture at the boundary layer cools to the wall temperature and the reaction is terminated. The origin of partially combusted hydrocarbons essentially depends on the temperature and oxygen content and to a lesser degree on the molecular structure. With air/fuel ratios under 1, HC emissions strongly increase since there is too little oxygen for complete combustion in the combustion chamber. The same holds true as the air/fuel ratio increases and the ignition limit of the mixture is reached, which leads to misfiring when there is homogeneous mixture formation. 21.3.1.1.4 Nitrous oxides This generic term includes the seven oxides NO, NO2, NO3, N2O, N2O3, N2O4, and N2O5. Nitrogen oxides arise from the nitrogen and oxygen in air during combustion. The processes are described by the expanded Zeldovich mechanism (1946). The most important representatives of these oxides are nitrogen monoxide (NO) and nitrogen dioxide (NO2). A general distinction is drawn between two important NO formation processes: the formation of thermal NO is influenced by the parameters of temperature, oxygen concentration, the air/fuel ratio, dwell time, and pressure. The maximum formation of NO occurs at approximately 2200 K to 2400 K and quickly decreases at higher temperatures. Below 750 K, high activation energy is required for the decomposition of NO. NO arises quickly in a side reaction in the flame front from OH radicals that form other compounds with the nitrogen molecules. Because of the high temperatures, nitrogen oxides can also be formed from the nitrogen contained in the fuel. This formation process is, however, less important. The NO/NO2 ratio of the untreated emissions in spark injection engines is over 0.99. The maximum NOx concentration occurs in the slightly lean range of λ = 1.05–1.1.

540

630

720

HC3 concentration [ppm]

0,25

1000

Figure 21.15 Progression of the mass-related hydrocarbon emissions downstream of the exhaust valve above the crankshaft angle [21-7].

SI engines with direct fuel injection and charge stratification, in comparison to intake manifold injection, produce lower NOx emissions due to the low average temperature. Charge stratification yields are more CO and HC emissions from local lean zones. 21.3.1.2 Unregulated Exhaust Components 21.3.1.2.1 Particles All components that can be separated by a filter below 51.7°C are considered particles. The particles consist of solid organic or liquid and soluble organic phases. This includes soot, various sulfates, ash, various additives from the fuel and lubricating oil, and abrasion and corrosion products. Abraded chrome arises as well as nickel aerosols from piston wear. The chrome aerosols have a particle size of 1.6–6.4 μ m [21-8]. Condensated particle emissions play a rather subordinate role in SI engines. However, they are gaining more attention with fuel injection systems that use direct fuel injection. 21.3.1.2.2 Gaseous components Of particular interest are aromates, such as benzene, toluene, xylene, and polycyclic aromatic hydrocarbons (PACs), and aldehydes, such as formaldehyde, acetaldehyde, acrolein, propionaldehyde, hexanal, and benzaldehyde. Aldehydes are intermediate products for the oxidation of hydrocarbons and their formation depends on the temperature [21-9]. The largest agent of the BTEX components in terms of quantity is toluene [21-8]. A direct connection of the fuel composition, the lubricant composition and the quality of the combustion process with the formation of the unregulated components can generally be recognized.

21.3.2 Diesel Engine

Diesel engines have the following characteristics: •• internal mixture formation •• load control via the supplied fuel quantity with unthrottled, inducted air

754 | Internal Combustion Engine Handbook

6606_Book.indb 754

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.3 Pollutants and Their Origin

•• autoignition and a large amount of excess air. From an integral point of view, diesel engines operate with a load-dependent air/fuel ratio of between 1.2 (high load) and 7 (idling) •• compression ratio of 14 to 22 •• higher boiling hydrocarbons as fuel. Indirect injection engines (prechamber/whirl chamber) have favorable untreated emissions and noise emissions, but they are more commonly being replaced by engines with direct fuel injection for passenger cars because of CO2 emissions that are up to 20% higher. In commercial vehicles, diesel engines with direct fuel injection are the main form of propulsion in Europe. Large engines that have the greatest efficiency of all heat engines also use this technique, however, usually with the two-stroke method. Different methods with different pressure generators are used for mixture formation such as inline fuel injection pumps, distributor fuel injection pumps, pump nozzles, pump-rail nozzles, and common rail systems. At present, the most important fuel injection method in passenger car engines uses air-distributing high-pressure fuel injection through multiple-hole nozzles. 21.3.2.1 Restricted Exhaust Components 21.3.2.1.1 Carbon dioxide The clearly improved fuel consumption and the enhanced consumption in the partial load range leads to a real 20% reduction of CO2 emissions per traveled kilometer. 21.3.2.1.2 Carbon monoxide Because of inhomogeneity of the mixture from charge stratification, there are zones with air/fuel ratios that are less than one. In these areas, high concentrations of CO arise during the reaction that largely reoxidized to form CO2. In contrast to SI engines, this leads to substantially lower specific carbon monoxide emissions. 21.3.2.1.3 Hydrocarbons The mechanisms and parameters are similar to SI engines. In general, the HC emissions in diesel engines are much lower. Additional relevant variables are the mixture formation quality of the fuel injection system and metering precision.

Postinjection increases HC emissions. The influence of minimized pollutant quantities with fuel injection nozzles is shown in Figure 21.16 [21-10]. 21.3.2.1.4 Nitrogen oxides The formation processes are also comparable with those of SI engines. The NO-to-NO2 ratio for diesel engines is 0.6 to 0.9 depending on the load. Under a small load, more NO2 is formed. This ratio is primarily determined by the oxygen concentration and the dwell time. NO2 is formed at the flame front. Diesel engines with divided combustion chambers have clearly lower NOx emission than diesel engines with direct fuel injection. Given the extreme lack of air at high temperatures while fuel is injected into the prechamber, the NOx formation rate is much lower. When the prepared mixture passes into the main combustion chamber, the opposite conditions predominate, that is, a large amount of excess air at lower temperatures. Since the compatibility for exhaust recirculation is much higher with diesel engines with direct fuel injection than with indirect injection engines (by approximately a factor of two), the ratios are opposite. 21.3.2.1.5 Particles For the most part, the particles from diesel engines consist of hydrocarbon particles. The remainders are hydrocarbon compounds (some of which are bound to soot), and a few are sulfates in the form of aerosols. In the combustion of different hydrocarbons, there are several intermediate substages in the individual stages, such as cracking, dehydration, and polymerization. The creation of soot is largely determined by the local temperature (800–1400 K) and the oxygen concentration and occurs in two phases [21-9]. The reactions in the primary education phase take place almost exclusively by radical chain mechanisms at the core of the fuel jets and behind the beam peaks. O, H, and OH radicals are formed. Cyclic and polycyclic aromatic hydrocarbons form by polymerization and cyclization. From the addition of other units, relatively stable intermediate products form from aggregation that become increasingly larger particles (so-called primary particles). As the primary particles coagulate to form large units, secondary particles arise. Uncombusted and partially combusted hydrocarbons, especially aldehydes can bond to the secondary Blind hole nozzle

1

500

Engine speed 2500 /min

HC-ppm C1

400 300 Blind hole

200

Zero blind hole nozzle

100 0

0

1

2

3

4

5

6

Effective mean pressure [bar]

7

Figure 21.16 Influence of injection nozzle design on the HC emissions [21-10].

Internal Combustion Engine Handbook | 755

6606_Book.indb 755

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

particles due to their large specific surface. As combustion proceeds, the secondary formation phase is soot reoxidation that is governed by the dwell time and oxygen concentration. The diameter of the particles varies between 1 and 1000 nm. For homogeneous mixtures, soot is found in the exhaust emissions at an air/fuel ratio below 0.5; at a λ above 0.6 under optimum conditions, there is no demonstrable soot formation [21-11]. In addition to soot formation as a source of particles, the lubricant is also an important source of particle emissions. A particular problem area is the conflict between HC and NOx particles. The conditions for low particle formation and low HC emissions contrast with the prerequisites for low nitrogen oxide emissions. Therefore, attention is especially given to the secondary formation phase, soot reoxidation. To support soot reoxidation, generally a large amount of mixture formation energy is required in the last phase of combustion that is attainable by a specific swirl and tumble in the combustion chamber, greater fuel injection pressure, a faster injection rate at the end of the fuel injection process, and more even distribution. These conditions are unfortunately favorable prerequisites for high NOx emissions. Figure 21.17 provides a qualitative summary of pollutant formation in diesel engines [21-13].

HC

NOx HC

Soot

NOx

Figure 21.17 Qualitative representation of diesel engine combustion and pollutant formation [21-13].

21.3.2.2 Unregulated Exhaust Components Important unrestricted components in untreated exhaust from diesel engines are cyanide, ammonia (NH3, sulfur dioxide (SO2) and sulfates. Of the specific hydrocarbons, methane, ethane, ethene, ethine, benzene, and toluene are of particular interest. Of the PACs, phenanthrene, pyrene, fluorene, fluoranthene, and anthracene predominate in descending order. The concentration of these parameters is at least six times greater than that of the other individual PAC substances and forms approximately 90% of the PACs [21-8]. Phenols and different aldehydes such as formaldehyde, acetaldehyde, acetone + acrolein, and propionaldehyde are also being studied more

closely [21-12]. The cited components are formed from trace substances in the fuel, in the lubricant, and, to a degree, from secondary reactions in the exhaust system. If particle emissions are viewed according to their massrelated share of hydrocarbons, we find a ratio of 80% elementary hydrocarbons to 20% organic compounds. Chrome and nickel aerosols originate from abrasion, as is the case with SI engines.

21.4 Reduction in Pollutants Basically, the measures to reduce pollutants can be divided into measures preceding and following the engine. In the first section, we address measures before and in the combustion chamber that play an important role in lowering the untreated emissions and fuel consumption.

21.4.1 Engine-Related Approaches 21.4.1.1 Gasoline Engine A majority of the approaches are both for engines with intake manifold fuel injection and for SI engines with direct fuel injection. In general, it can be noted that a minimum of the untreated emissions generally does not yield the best overall result following exhaust treatment. 21.4.1.1.1 Mixture formation The air/fuel ratio of the mixture in the combustion chamber has the greatest influence on untreated emissions. The emissions of CO and HC are lowest in the slightly lean range of λ = 1.05–1.1; however, untreated NOx emissions are highest in this range. An equivalent air/fuel ratio for all cylinders is another prerequisite for low emissions. This requires highly precise metering of the fuel to all cylinders. A λ spread, in particular, strongly increases CO emissions and to a lesser extent HC emissions. NOx emissions rise with a low λ spread. As the spread increases, NOx emissions fall again. To measure the cylinder-selective A differences for control purposes, refer to Section 21.2.2. For a complete conversion of the fuel in the engine, the fuel must be well prepared. With intake manifold injection, normally the fuel is injected directly before the intake valves. When the intake manifold pressure and temperature are properly exploited, this position yields optimum fuel preparation with minimum wall film formation. Fuel preparation is further optimized by additionally surrounding the injected fuel jet with air, special jet geometries, fuel injection nozzles with a flash boiling effect, and piezoactuated injectors that ensure highly precise metering of very small quantities of injected fuel. The preparation time is substantially shorter with airatomizing direct fuel injection than with manifold injection (a duration similar to diesel DI). In addition, different fuel injection strategies with the fuel injection nozzles must be used for the respective operating modes (homogeneous or stratified charge).

756 | Internal Combustion Engine Handbook

6606_Book.indb 756

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.4 Reduction in Pollutants

Another approach for mixture preparation is to blow the mixture into the combustion chamber. For this, the fuel must be prepared outside of the combustion chamber. Particularly, in the stratified charging of an extremely lean mixture, this method creates favorable conditions for unthrottled engine operation. Most pollutants can be thereby reduced. The ability to run on a lean mixture additionally enables fuel savings over a wide load and speed range. 21.4.1.1.2 Combustion process and combustion behavior The combustion speed is essentially influenced by the fuel, the air/fuel ratio, the pressure and the temperature during conversion, and the flow state in the combustion chamber. The combustion characteristic for manifold injection largely depends on the intake valve lift and the start of fuel injection, the degree of mixture preparation, and the moment of ignition. Short dwell times at high temperatures reduce NO formation. Optimum combustion procedures do not exceed a maximum temperature of 2000 K. Direct fuel injection also allows the fuel injection time to be freely selectable, but the mixture formation time is very short [21-19]. Charge stratification reduces untreated NOx emissions and fuel consumption. However, we need to remember that the highly nonlinear NOx formation does not yield flame front zones that are extremely hot. Directly adjacent to the spark plug, there is normally a rich mixture to ensure ignition. The majority of the surrounding charge is, however, lean. Very homogeneous combustion is to be ensured. 21.4.1.1.3 Valve timing Valve gear/valve control times: The transition from two valves to four valves has a number of advantages, especially for engines with manifold injection. A central spark plug and the symmetrical four-valve combustion chamber form is optimum for low-pollutant combustion. Only higher hydrocarbon emissions can sometimes be determined. By means of the variable control times, fuel consumption and emissions can be influenced over a wide range. With an electromechanical valve gear [21-14] that allows a variation of many degrees of freedom, fuel consumption can be additionally lowered, particularly in engines with direct fuel injection. A small valve overlap, small valve stroke, and late intake time allow emissions to Transition from 2-valve to 4-valve engine

CO NOx be HC

Reinforced exhaust-gas recirculation

CO NOx be HC

be greatly reduced in the partial load range. In addition, the engines can be operated unthrottled under a partial load that substantially lowers fuel consumption. Cylinder shutoff in engines with a large number of cylinders also decreases fuel consumption and consequentially CO2 emissions. 21.4.1.1.4 Exhaust gas recirculation Exhaust emissions are guided from the exhaust section of the engine through an exhaust recirculation valve into the intake tract and replace a portion of the fresh charge. This gas mixture can absorb a large amount of heat under dissociation and, hence, lowers the temperature during combustion to prevent the formation of thermal NO. The related unthrottling of the engine also lowers fuel consumption. Variable inner exhaust gas recirculation (EGR) can be attained by corresponding valve overlap by means of phase shifters. When exhaust emissions are fed to the intake manifold, even distribution to all cylinders must be ensured. EGR rates greater than 15% yield somewhat higher HC emissions and poorer idling. Compression ratio: High compression improves thermal efficiency. The peak combustion temperature is increased, which in turn produces higher NOx emissions. Because of the high-pressure level, HC emissions also rise because of the greater splitting of the combustion chamber. CO emissions tend to fall as compression increases. Variable compression is under development with promising results, at least in terms of fuel consumption. 21.4.1.1.5. Combustion chamber design In addition to the geometry that affects the bore-to-stroke ratio, the surface, volume, and squish area, other parameters influence emission behavior. A central spark plug position for a shorter flame path, compact combustion chambers with a small surface, minimum dead volume of gaps, and specific squish areas especially lower HC emissions and fuel consumption. Measures to reduce the combustion process also somewhat lower NOx emissions. An increase in the compression ratio reduces fuel consumption, but increases NOx emissions. The most important measures influencing quality in SI engines with manifold injection are summarized in the diagram in Figure 21.18, and those for DI engines are shown in Figure 21.19. Numbers were intentionally not used since the individual

Increase in the compression

Variable valve actuation

CO NOx be HC

CO NOx be HC

Figure 21.18 Measures for reducing pollutants in gasoline engines with intake manifold injection. (+) indicates an increase and (–) a decrease in the exhaust gas level.

Internal Combustion Engine Handbook | 757

6606_Book.indb 757

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Better fuel quality

NOx be HC

Reinforced exhaust-gas recirculation

NOx be HC PM

Fully variable valve actuation and injection

NOx be HC

Blowing in the mixture

NOx be HC

Figure 21.19 Measures for reducing pollutants in gasoline engines with direct injection. (+) indicates an increase and (–) a decrease in the exhaust gas level.

measures frequently can lead to substantially different results for different engines, and the intention was to show a generally applicable trend. Other measures to lower untreated HC emissions are variable swirl formation in the intake port and regulated engine temperature control. 21.4.1.1.6 Ignition system The primary method for externally supplied ignition is electrical ignition in the form of single or multiple ignition using spark plugs in the combustion chamber. The selected design influences the shape of the flame front and, hence, also the formation rate of nitrogen oxides. Another basic parameter is the moment of ignition in reference to top dead center (TDC). As we know, late ignition produces low NOx emissions. Moments of ignition optimized by means of adaptive control that covers all necessary parameters are possible with modern engine management systems. For the mixture to reliably ignite, sufficient ignition energy of 0.2–3 [mJ] is required. A long spark duration with stabile, high combustion voltage supports the reliable and stable ignition of the mixture and produces low HC emissions. In other advances, the spark plug is used as a “combustion chamber sensor.” By measuring the ion flow at the electrodes during combustion, the start of ignition (misfiring diagnosis/CH emissions) and the progress of ignition can be measured, and knocking can be monitored. In conjunction with electronic combustion control, this permits effective diagnosis and, hence, lower emissions over long periods. Spatial ignition (to be understood as the simultaneous ignition of the mixture at a theoretically infinite number of locations in the combustion chamber), laser ignition (in which a laser beam spread by a suitable lens system ignites the entire contents of the combustion chamber with sufficient energy), and plasma ignition are under development and have advantages mostly for individual exhaust emission components. In particular, spatial ignition procedures have a very high potential to substantially lower untreated NOx emissions. At the same time, fuel consumption under a partial load can be reduced.

Further improved fuel quality to prevent the coking of direct fuel injection systems produces more stable emissions behavior, especially of jet-directed systems. 21.4.1.2 Diesel Engine Approaches to optimize the emissions of diesel engines generally deal with the classic conflict between fuel consumption, nitrogen oxides, and particle emissions. 21.4.1.2.1 Combustion method and combustion process The most important parameter for conventional nozzles for air-distributed, direct fuel injection is the start of injection in reference to the TDC. The ignition lag represents a relatively constant quantity. Standard approaches based on four-valve cylinder heads use a turbulence and charge duct. There is a trend toward flatter and wider piston recesses that allow unhindered jet expansion. This development has been supported by multiple-hole fuel injection nozzles. This reduces the amount of fuel deposited on the wall. Charge losses can be reduced by increasing the swirl. Homogeneous diesel combustion with lean premixture combustion and a faster injection rate toward the end of the reaction can produce nearly soot-free combustion with minimum NOx emissions [21-11], [21-15]. The practical implementation of this method is hindered at present by insufficient mixture formation and an inhomogeneous mixture distribution as well as the limited operating range in the load and speed program map. In addition, this combustion procedure requires a fully variable fuel injection system with optional preinjection, main injection, and secondary injection that can be varied over wide ranges, in other words, the formation of nearly any fuel injection characteristic. This method competes with combustion chamber ignition in spark-injection engines. 21.4.1.2.2 Supercharging For diesel engines with direct fuel injection, turbocharging still represents the most efficient option for reducing all pollutant components. The most important parameters, such as the charging pressure and charging air temperature, must be varied for each load level. In addition, variable turbine geometries, sequential supercharging, and charge air temperature

758 | Internal Combustion Engine Handbook

6606_Book.indb 758

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

20 15 10

Rail pressure 500 bar 650 bar 800 bar 950 bar 110 bar

5 0

1,5 1,0

SZ [-]

special consumption [g/kWh]

NOx [g/kWh]

21.4 Reduction in Pollutants

0,5 230

0,0

220 210 200 190 –25

–20

–15

–5

–10

5

0

10

Figure 21.20 Impact of the rail pressure and start of injection on consumption, smoke number and NOx [21-16].

Start of injection [°CA]

circumvent the central conflict of particles versus emissions by suitably harmonizing engine control in a common-rail fuel injection system [21-17]. A small preinjected quantity of less than 1 mm3 minimizes NOx and particle emissions when at a suitable duration from the main injection. Secondary injection reduces particle emissions while NOx emissions remain the same. A very short cutoff time especially lowers HC emissions. Fuel injection systems that use piezoinjectors can approach the desired fuel injection curve because of the shorter operating times and permit extensive optimization. Injecting water reduces NOx by up to 25% in engines with an excessively high combustion temperature, but this also worsens CO and HC emissions [21-11], [21-15].

regulated according to load can reduce fuel consumption and especially NOx emissions. Electrical secondary supercharging can substantially reduce particle emissions in transient processes such as starting from idling. 21.4.1.2.3 Fuel injection systems and processes High-pressure fuel injection systems, such as the pump nozzle and common-rail system, can optimize fuel preparation and especially lower particle emissions due to the high fuel injection pressure of 1.500–2.000 bar, (over 2000 bar in the near future), in combination with new fuel injectors. Figure 21.20 shows the possible improvements as the fuel injection pressure rises. The target conflict of particles versus NOx can be managed much better [21-16]. The best strategies must be taken from the test stage for the most important parameters such as the moment of injection, fuel injection law, fuel injection pressure, nozzle shape (jet position, projecting mass, number of nozzles), fuel injection quantity, preinjection, injection spacing, secondary injection, injection time, and cutoff time. Figure 21.21 shows how to best

0,10 0,09

Particles [g/kWh]

n = 2000/min pe = 5.0 bar

EGR variation Variation in the spacing between the preliminary and main injection Rail pressure variation Increasing

0,08 0,07 0,06 0,05

21.4.1.2.4 Valve timing Greater charging is possible using multiple valves. This has a particularly positive effect on fuel consumption and general raw emission behavior.

0,04 0,03

Design point

0,02 0,01 0

0

1

2 3 4 specific NOx emissions [g/kWh]

5

6

7

Figure 21.21 Target conflict of NOx emissions in V8 TDI engines with common rail injection in accordance with [21-17].

Internal Combustion Engine Handbook | 759

6606_Book.indb 759

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Particle concentration dN/dlog(d) [Number/cm3]

Chapter 21 Exhaust Emissions

7.0E+07 EGR: 0%

6.0E+07

EGR: 30%

5.0E+07

EGR: 50%

4.0E+07 3.0E+07 2.0E+07 1.0E+07 0.0E+00

10

100 Particle diameter d [nm]

In particular, four-valve technology now appears to be the best approach for diesel engines because of the injection nozzle position. Variable valve control and phase shifters are now used for passenger car diesel engines since turbocharging is so common, but they are presently not as important as they are in SI engines. 21.4.1.2.5 Exhaust gas recirculation In contrast to SI engines with externally supplied ignition, diesel engines can achieve much higher recirculation rates. However, as we can see in Figure 21.22, there is a notable increase in the number and size of particles as the exhaust recirculation rate increases. Additionally cooling the recirculated exhaust lowers NOx and particle emissions somewhat, but it increases CO and HC emissions. The conflict between NOx and particles is improved by up to 15% with cooled exhaust recirculation [21-15]. 21.4.1.2.6 Combustion chamber design Basically, the same design rules hold true as for gasoline engines. Depending on the injection method and the charge volume, different requirements arise for the combustion chamber geometry. Wide, flat combustion chambers with low turbulence are preferred when there is large piston displacement. In contrast,

Exhaust gas turbocharging

Reinforced exhaust-gas recirculation

NOx be PM HC

NOx be PM HC

High Pressure Injection

NOx be PM HC

1000

Figure 21.22 Particle concentration across the particle diameter as a function of the exhaust gas recirculation rate [21-18].

deeper recesses with higher swirl numbers tend to be used in the combustion chambers of passenger cars with a cylinder volume of 450–550 cm3. Figure 21.23 summarizes the most important measures used to reduce diesel engine emissions. The rapid development of proven technologies can bring about notable progress in the reduction of exhaust pollution. It appears possible that certain families of engines may attain particularly low exhaust thresholds, such as the ULEV or Euro 4, without the use of fuel-consuming additional aggregates. At the same time, new steps should be taken to introduce new combustion procedures for the “pollutant-free automobile that protects natural resources.”

21.5 Exhaust Treatment for Gasoline Engines 21.5.1 Catalytic Converter Designand Chemical Reactions

The basic function of an automobile catalytic converter can be described with the following reaction equations (21.4) to (21.9). The main pollutant components in the exhaust gas

Transition from 2-valve to 4-valve engine

NOx be PM HC

Figure 21.23 Measures for reducing pollutants in diesel engines. (+) indicates an increase and (–) a decrease in the exhaust gas level.

760 | Internal Combustion Engine Handbook

6606_Book.indb 760

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

of SI engines are uncombusted hydrocarbons and carbon monoxide (these must be completely converted by oxidation after incomplete combustion), and nitrogen oxides, which are reduced to form nitrogen. catalytic converter

Oxidation of CO and HC into CO2 and H2O (21.4)



n n C y H n + ⎛⎜ 1 + ⎞⎟ O 2 → y CO 2 + H 2 O ⎝ 4⎠ 2 1 CO + O 2 → CO 2 2

(21.5)



CO + H 2 O → CO 2 + H 2

(21.6) Figure 21.24 Sectional view of a catalytic converter.

Reduction of NO/NO2 into N2 (21.7)



1 NO ( or NO 2 ) + CO N 2 + CO 2 2 1 NO ( or NO 2 ) + H 2 N 2 + H 2 O 2

(21.8)



⎛ 2 + n ⎞ NO or NO + C H → ⎛ 1 + n ⎞ N + y CO + n H O ( ⎟ ⎟ 2 2) y n 2 2 ⎝⎜ ⎝⎜ 2⎠ 4⎠ 2 (21.9)

These reactions are catalyzed in the presence of precious metals Pt, Pd, and Rh. To attain the highest possible conversion rate, the precious metals are dispersed on a substrate oxide with a large surface. These substrate oxides are typically inorganic materials with a complex pore structure (such as Al2O3, SiO2, TiO2), on which the catalytic materials are applied together with promoters. The catalytic substrate is produced in an aqueous solution with a solid content of 30–50%, a substance referred to as slurry. This slurry is coated onto honeycomb-shaped monoliths. Both ceramic and metal monoliths are used. The honeycomb structure ensures the greatest possible surface for the catalytic reaction in a small space. Figure 21.24 shows an example of a catalytic converter that consists of two ceramic monoliths.

21.5.2 Catalytic Converter Concepts of Engines with Stoichiometric Operation 21.5.2.1 Three-Way Catalytic Converter The reactions discussed in the prior section to convert pollutant components in SI engine exhaust produce the conditions under which these reactions are possible. Excess oxygen is necessary to oxidize HC and CO, whereas the presence of reducing components is required to reduce nitrogen oxides. Since all pollutant components must be converted when driving, there is a narrow operating window for exhaust composition and conversion in which combustion can occur. By regulating the air/fuel mixture with the aid of a λ sensor within the narrow range around the stoichiometric ratio of λ = 1, it is possible for the oxidation reactions and reduction reactions to occur with a high conversion rate. Such catalytic converters are termed three-way catalytic converters since all three pollutant components are equally converted. The point at which both CO and NOx are best converted as a function of the air/fuel ratio λ is the optimum operating point for the catalytic converter, termed the crossover point. Figure 21.25

100

HC CO NOx

Efficiency %

80

60

40

20

0 0,96

0,97

0,98

0,99

1,00

1,01

1,02

1,03

l

1,04

Figure 21.25 Conversion of pollutants in relation to the air/fuel ratio.

Internal Combustion Engine Handbook | 761

6606_Book.indb 761

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Ceramic

Metal

Cell density, cpsi

400

600

900

400

600

800

Wall/film stength, mil/mm

 6.5

 3.5

 2.5

 0.050

 0.040

 0.030

   0.025

Geometric surface, cm2/cm3

 27.3

 34.4

 43.7

 36.8

 42.9

 51.6

  56.0

Free cross section, %

 75

 80

 86.4

 89.3

 89.8

 93.7

  91.4

Hydraulic diameter, mm

 1.10

 0.93

 0.79

 0.97

 0.84

 0.72

   0.65

Density g/cm3

 0.43

 0.35

 0.24

 0.77

 0.73

 0.55

   0.61

shows the conversion of pollutant components HC, CO, and NOx as a function of the air/fuel ratio λ . To maintain the presently very strict exhaust laws in Europe and the United States, these three-way catalytic converters are used for λ = 1 regulated SI engines. In addition to conversion under hot operating conditions, the starting behavior (light off) of the catalytic converter is a very important factor for characterizing its behavior. Along with catalytic-converter-specific properties, substrate-specific properties also play an important role in dynamic tests. They determine the thermal mass of the catalytic converter by using geometric parameters, such as cell density and wall and film thickness, and they also influence the catalytic converter’s thermodynamic properties of heat capacity and density. By reducing the thermal mass, the time before light off in a cold start can be greatly reduced. At the same time, a large geometric surface must be made available for the catalytic reaction. Substrates with a high cell density and thin walls are suitable for improving the lightoff behavior when the catalytic converter is heated quickly. Figure 21.26 shows the geometric parameters of selected standard substrates. The temperature-related thermal capacity of ceramics is compared with that of metal substrates in Figure 21.27. 1,2 1,1

Specific heating capacity [J/gk]

Cordierite

1,0 0,9 0,8 0,7

Metal

0,6 0,5 0,4

0

100

200 300 400 Temperature in °C

500

600

Figure 21.27 Thermal capacity.

Different catalytic converter systems are distinguished based on the arrangement of the catalytic converter(s) in the overall vehicle (Figure 21.28). On the one hand, a position close to the engine is possible (underhood main catalytic converter) that

1,000

Figure 21.26 Substrate parameters for high-cell and thinwall substrates.

is subject to restrictions of thermal and mechanical stability and available installation space. Main catalytic converter near the engine

Underbody catalytic converter

Preliminary catalytic converter

Main catalytic converter

Figure 21.28 Catalytic converter concepts.

On the other hand, there is the traditional underfloor position where the lower exhaust temperature is less supportive of catalysis. Frequently, a combination of an underhood preliminary catalytic converter and underfloor catalytic converter is used. This system combines the advantages of a fast heating curve of the underhood arrangement with the required installation space of the underfloor position for larger catalytic converters with the disadvantage of a higher system cost. 21.5.2.2 Oxygen Storage Unit The chemical reactions of the three-way catalytic converter were discussed in detail in the above sections. It was also mentioned that the oxidation and reduction reactions can occur simultaneously with maximum conversion only when the air/fuel ratio is at the stoichiometric point, that is, λ = 1. The engine control system solves this problem by continuously measuring a quantity that is proportional to the air-fuel ratio with the aid of a measuring sensor in the exhaust, the λ sensor, in a closed control loop. If the sensor measures exhaust that is too rich or too lean, the engine management makes a correction in one direction or the other. This means, however, that the air/fuel ratio is only stoichiometric on average over time; at specific engine loads, it can vary widely from one time to the next. This means, however, that the air/fuel ratio is only stoichiometric on average over time; at specific engine loads,

762 | Internal Combustion Engine Handbook

6606_Book.indb 762

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

21.5.2.2.1 The basic component of the oxygen storage unit As already mentioned previously, the air/fuel ratio is never exactly equal to one, but rather oscillates constantly around this point. This means that more oxygen is available in one half of the oscillation than is needed for the conversion, while an oxygen deficit prevails in the other half. Although the amplitude and frequency of the vibration can meanwhile be very precisely controlled, this would be detrimental to the conversion of the exhaust gas. For this reason, an element is included in the catalytic coating which has the property of storing the oxygen. If a half cycle now results, in which an excess of oxygen prevails, then the excess oxygen is stored and can be drawn on for the conversion in the subsequent halfcycle in which the actual oxygen deficit exists. The following reaction equation can be formally set up: Ce2 O 3 + 0.5 O 2 → 2 CeO 2



(21.10)

whereby cerium is the element that forms the oxygen storage unit. The overall special property of the surface chemistry of this element makes this storage possible. Cerium can assume two different oxidation levels, whereby the mechanism operates according to the following intermediate step:

O Ce4+ + PM → Ce4+

∆ + PM-O Ce3+ Ce3+

(21.11)

If a carbon monoxide atom now forms on the surface, it can absorb the stored oxygen for the oxidation and the cerium oxide is thus reduced as illustrated in the following reaction equation:

2 CeO 2 + CO → CO 2 + Ce2 O 3

(21.12)

whereby the precious metal also results in a catalyzed intermediate step here: CO + PM − 0 → PM + CO 2



(21.13)

The CO conversion can thus be directly improved by the oxygen storage unit. The same also applies for the NOx conversion, which progresses in line with the following pattern:

Ce3+

∆ Ce3+ + NO →

Ce4+

O Ce4+ 0.5 N 2

(21.14)

It is generally assumed that the hydrocarbons are converted by the same mechanism. This is not the case, however. The hydrocarbon molecule requires large precious metal surfaces to react to this and is consequentially not catalyzed by the cerium as with CO and NOx. 21.5.2.2.2 The development of the oxygen storage unit In the first three-way catalytic converters, cerium was incorporated in soluble form without special stabilizers. The advantage was a very large surface and, hence, oxygen storage ability when new. However, when these catalytic converters were exposed to high temperatures over a long period, this surface

rapidly decreased. In comparison, cerium has a surface of around 120 m 2/g in its fresh state and after four hours of aging in the kiln at 1050°C, reverts back to less than 1 m 2/g. Given increasingly tougher exhaust laws, catalytic converters are being placed increasingly closer to the engine exhaust outlet to support fast light off. At these installation sites, where consistently high temperatures predominate, the storability can drop very rapidly. One of the groundbreaking inventions in the three-way converter segment was therefore the development of stabilized cerium components. Although the stabilizer largely consists of zirconium, it also contains additional elements of the rare earths. Although the available surface in the fresh state greatly reverted back, to around 80 m2/g, it is still at around 30–40 m2/g after aging, that is, many times greater than with nonstabilized cerium. Figure 21.29 shows the comparison of stabilized and non-stabilized cerium after aging for four hours at 1050°C. The different aging behavior of both materials can be clearly recognized. 100 Standardized BET surface [%]

it can vary widely from one time to the next. Depending on the exhaust, that is, rich or lean, the catalytic converter then produces HC, CO, and NOx peaks in reference to the conversion. The surface chemistry of cerium offers a solution to this dilemma since it has the property of storing and then releasing oxygen.

80 60 40 20 0

Stab. Ceria non Stab. Ceria 800

900 Aging temperature [°C]

1000

Figure 21.29 Comparison of non-stabilized cerium with stabilized cerium. Aging 4 hours at 1050°C in the kiln.

Only after this development could the life of underhood catalytic converters correspond to that of automobiles. The around 30 years of experience using catalytic converters shows, that today’s catalytic converter is no longer the weakest link in the exhaust gas purification chain. 21.5.2.3 Cold-Start Strategies To maintain legally required exhaust standards, it is very important to bring the catalytic converter to operating temperature as soon as possible. The measures to heat the catalytic converters described below have been used in various series applications to accomplish this. A distinction is drawn between active and passive cold-start strategies. 21.5.2.3.1 Electrical heated catalyst One active measure is to electrically heat catalytic converters with electrical heating elements (Figure 21.30). A high electrical output is required before or during the engine start that is supplied by the vehicle electrical system. The fact that this electrical output must be generated with a lower engine efficiency in the cold start and a typically poor generator

Internal Combustion Engine Handbook | 763

6606_Book.indb 763

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

efficiency at a 12 V supply voltage must be considered here. For this reason, this system can achieve strict exhaust gas standards only for large engines.

Figure 21.30 Electrical heated catalyst.

21.5.2.3.2 Secondary air In addition to measures within the engine for increasing the exhaust gas enthalpy, injecting secondary air into the exhaust gas system is an effective variation for quickly heating the catalytic converter. Figure 21.31 illustrates heating catalytic converters upon a cold start with and without secondary air injection. The secondary air is injected by an electrical pump into the exhaust ports of the engine. The additional oxygen promotes exothermic oxidation so that the catalytic converter can be heated to the operating temperature just in a few seconds. Since the fuel/air ratio of the engine can be set slightly rich during hot operation, the drivability is good in a cold start.

21.5.2.3.3 HC-storage catalytic converter Another possibility for reducing HC emissions is to use HC storage catalytic converters (traps) Figure 21.32. The uncombusted hydrocarbons emitted during cold starts are adsorbed by a storage catalytic converter as long as the three-way catalytic converter does not work. After the three-way catalytic converter has light off, the uncombusted hydrocarbons are released and then converted. For this system to work, the desorption temperature of the storage mechanism must be above the light-off temperature of the three-way catalytic converter so that the stored hydrocarbons can also be effectively converted and not just passed into the exhaust gas system at a later time. This requires that sufficient oxygen be available for the oxidation at the moment the hydrocarbons are released. This is made possible by a suitable engine control strategy (precontrolled λ ). At present, the use of storage catalytic converters is limited by the temperature stability of the storage materials. The temperature stability of the zeolites is far below that of a three-way catalytic converter. 21.5.2.4 Deactivation Effects and Their Impact One of the main causes of catalytic converter deactivation is the atmosphere to which the catalytic converter is exposed. Exhaust gas temperatures beyond 900°C is no rarity and is primarily fostered by installation positions close to the engine. Additional deactivation effects are caused by fuel or motor oil in the exhaust and are irreversible with a few exceptions such as thermal aging. For researchers, it is essential to understand the mechanisms in order to develop resistant materials or regeneration methods if possible. 21.5.2.4.1 Thermally caused deactivation One of the most important goals in the manufacturing of catalytic converters is to ensure that the reactants have optimum accessibility to the active sites when the catalytic components are applied to the substrate. In a perfectly dispersed catalyst,

800 700

Temperature [°C]

600 500 400

300 200

Without secondary air

100

With secondary air

0

0

50

100

150 Time [s]

200

250

Figure 21.31 Temperature progression in the catalytic converter bed with/without secondary air.

764 | Internal Combustion Engine Handbook

6606_Book.indb 764

1/19/16 8:52 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

Untreated emissions Emissions as per TWC

HC emissions [a.u.]

Emissions as per TWC + HC Trap

0

10

20

30

40 Time [s]

50

each atom (or molecule) that participates in the conversion reaction is easily accessible as shown in Figure 21.33. Precious metal

60

70

80

Figure 21.32 HC emissions with HC trap.

the crystals grow, the ratio of surface to volume falls, and less catalytically active atoms or molecules are available on the crystal surface for the reactants. In other words, many of the active sites are buried in the crystal, and since fewer active sites participate in the reaction, the performance decreases. Figure 21.34 uses a simple sketch to illustrate the phenomenon. The initially finely distributed precious metal forms crystals or agglomerates under heat.

Al2O3 Washcoat Sintered precious metal

Substrate

Al2O3 Washcoat

Figure 21.33 Schematic diagram of an ideally dispersed catalytic converter on an aluminum oxide substrate.

A few catalysts are formed in this highly active state; however, they are extremely unstable since they easily grow into large crystals under heat. This growth reduces the catalytic surface. Furthermore, the aluminum oxide substrate with its enormous inner surface constructed of a network of pores is also subject to a sintering process. This results in a loss of inner surface. Another deactivation mechanism is described by the interaction of the catalytic species with the substrate material. Alloy formation reduces the active slightly catalytic species. All of the previously described processes are influenced by the nature of the precious metals, by the substrate material, by the exhaust environment, and, above all, by high temperatures. 21.5.2.4.2 Precious metal sintering Highly disperse catalytic species are subject to the natural tendency of combining into crystals under heat. In this process,

Substrate

Figure 21.34 Schematic diagram of stainless steel sintering on a substrate.

The loss of performance due to precious metal sintering in catalytic converters is significant. A focus of catalytic converter research is, therefore, to precisely investigate the relationships of sintering and offer ways to stabilize finely distributed dispersions. Various elements from the group of rare earths have been successfully used in exhaust treatment to stabilize the precious metals. The precise mechanism of stabilization has not been completely researched; however, it appears as if the stabilizers fix the precious metal to the surface and, hence, reduce its mobility.

Internal Combustion Engine Handbook | 765

6606_Book.indb 765

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.5.2.4.3 The substrate sintering: Within a given crystal structure (such asγ -Al2O3), the loss of surface area is related to the loss of H2O and a gradual loss of the pore structure as shown in Figure 21.36. If the sintering process ceases, the pore openings gradually decrease, which raises pore diffusion resistance. A chemically controlled reaction could, hence, be gradually limited by pore diffusion. This phenomenon is decisively influenced by a progressive loss of activation energy of the corresponding reaction. In the conversion/temperature graph in Figure 21.35, the slope of the curve gradually decreases. 100

21.5.2.4.4 Precious metal/substrate oxide interaction The reaction of the catalytically active component with the substrate can be the reason for deactivation when the product has less activity than the originally finely distributed species. Rh2O3 reacts, for example, on the large highly-active surface area of the Al2O3 at high temperatures and lean exhaust gas conditions and forms an inactive mixed oxide. This process describes an important mechanism in the deactivation of the NOx reduction activity. It is assumed that the reaction basically occurs according to the following process:

80

CO-Conversion [%]

Another mechanism for substrate oxide conversion is based on the transformation of the crystal structure. The conversion of γ -AL2O3 into δ -AL2O3 results in a significant gradual loss in the internal surface area of around 150 to less than 50 m 2/g. The same can be noted with a TiO2 which has been converted from an anatase to a rutile structure; the surface decreases by around 60 m2/g to <10 m2/g. The conversion/temperature graph in this case is usually subject to a loss of activity. The sintering process of many substrate materials can be slowed in the presence of certain elements of the 3rd and 4th main groups in oxidized form. It is assumed that they form solid compounds with the substrate and, hence, reduce the surface reactivity largely responsible for sintering.

60

40

20

0 100°C

150°C

200°C

250°C

300°C

Fresh / Regenerated Loss of Active Sites: Sintering Pore Diffusion: Smaller Effective Pores Masking: Surface Sites Covered, Pores Blocked

Figure 21.35 Conversion as a function of the inlet temperature for different deactivation mechanisms.

Precious metal

Sintered poles

Al2O3 Washcoat

Substrate

Figure 21.36 Schematic diagram of substrate sintering.

In an extreme case, the pores become completely closed, and the catalytically active sites inside the pores are no longer accessible to the reactants (Figure 21.36).

Rh 2 O 3 + Al 2 O 3 ⎯800°air ⎯⎯→ Rh 2 Al 2 O 4

(21.15)

Since the activity of the catalyst is impaired, the curve shifts toward a higher temperature with a significant change in the slope. This undesired reaction results in the formation of alternative substrates oxide such as SiO2, ZrO2, TiO2 and their combinations. The negative interaction problem can be resolved by using these alternative substrate oxides; however, they are frequently not that resistant to sintering. 21.5.2.4.5 Deactivation by the effects of poisoning Another cause of catalyst deactivation is harmful substances from the exhaust gas or the machines that apply the catalyst layer. There are two basic poisoning mechanisms: selective poisoning in which a chemical substance directly reacts with the active centers or the substrate material and, hence, impairs or stops activity, and nonselective poisoning in which impurities are deposited on or in the catalyst substrate material and close to the active centers and pores. The result is a decrease in the performance due to scarcely accessible centers. 21.5.2.4.6 Selective poisoning If a chemical species reacts directly with the active centers, the term “selective poisoning” is used. This process directly influences the activity or selectivity of a given reaction (Figure 21.37). Some of these elements or molecules react with catalytic components by forming chemical compounds (such as Pb, Hg, Cd, etc.); an inactive alloy is formed. The process is irreversible and causes the permanent deactivation of the catalyst. Others only adsorb (or more precisely, chemisorb) the catalytic component (such as SO2 on Pd) and block additional reactions. These mechanisms are reversible; desorption occurs by supplying heat, washing, or simply removing the

766 | Internal Combustion Engine Handbook

6606_Book.indb 766

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

harmful component from the process flow, and the catalytic activity is restored. G = Toxin

Precious metal G

G

G

G

G

Al2O3 Washcoat G

G

Substrate

Figure 21.37 Schematic diagram of the selective poisoning of active centers.

If active centers are directly blocked, this always raises the light-off temperature; but the slope of the curve remains unchanged since the functioning of the remaining sites is retained and the activation energy is not changed. The conversion/temperature diagram looks similar to that of precious metal sintering. If the substrate oxide reacts with a component from the gas stream and forms a new compound as is the case with A12(SO4)3, the pores are generally nearly blocked, which increases resistance to diffusion (Figure 21.38). The activation energy falls, and the light-off curve shifts toward higher temperatures with a simultaneously reduced slope, which translates into poorer conversion (Figure 21.35). Precious metal

Precious metal

Al2O3 Washcoat

such a SI engine must be throttled under a partial load. This throttle loss and the subsequent effects of the reduced cylinder charge on the thermodynamic process are the main causes why the efficiency of SI engines greatly decreases at lower engine loads. In SI engines, this is expressed by substantially higher consumption under a partial load in contrast to diesel engines. The efficiency of SI engines under a partial load can be substantially increased by hyperstoichiometric engine operation. To get by without throttling, the mixture must be made very lean; that is, the engine must be operated with a large amount of excess air. The ignitability of a homogeneous mixture formed outside the cylinder poses substantial restrictions to lean adjustment and, hence, unthrottling. Directly injecting the fuel into the combustion chamber together with charge stratification allows further, nearly complete unthrottling. The increase in efficiency related to lean operation directly leads to a reduction of fuel consumption. Independent of whether the mixture is formed externally or internally (direct injection), excess oxygen is in the exhaust of lean-operated SI engines. This makes it substantially more difficult to convert pollutants in lean exhaust gas. In conventional SI engines with stoichiometric operation, there is a nearly complete conversion of pollutant components such as HC, CO, and nitrogen (NOx) by the familiar three-way technology. However, the reaction kinetics of lean-operated SI engines are an obstacle to this. HC and CO are preferably converted in the catalytic converter as a result of the faster reaction speeds. The previously converted reaction partners are missing to reduce NOx. For this reason, technologies are required that allow efficient exhaust treatment, especially nitrogen oxides, in a lean atmosphere. Lower exhaust gas temperatures pose an additional problem to exhaust treatment. 21.5.3.1 Options for Reducing NOx in Lean Exhaust Gas At present, there are different basic approaches to the conversion of nitrogen oxides from which various options can be derived for reducing NOx in lean exhaust gas. The individual technologies can be divided into the following groups and have been discussed in [21-26]. •• direct NO decomposition •• plasma technologies •• selective catalytic reduction (SCR) •• NOx storage catalytic converters.

Substrate

Figure 21.38 Schematic diagram of the nonselective poisoning of active centers.

21.5.3 Catalytic Converter Concept for LeanBurn Engines

Conventional SI engines are operated with a homogeneous fuel/ air mixture that is created outside of the combustion chamber in the induction tract. In principle, the mixture supplied in

21.5.3.1.1 Direct NO decomposition The reaction steps of the direct decomposition of NO into nitrogen and oxygen are shown in Figure 21.39.

N=O + N=O

Catalytic converter

N

O +

N

O

Figure 21.39 Reaction diagram of the NO decomposition.

Internal Combustion Engine Handbook | 767

6606_Book.indb 767

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Catalytic converters that can directly break down NO into N2 and O2 would be the ideal product for use in SI lean-burn engines and diesel engines. The conversion of these technologies into practical use will require a revolutionary invention. Although NO decomposition is thermodynamically preferred and the basic chemistry has been revealed in R&D laboratories, [21-27] transferring it to a real engine or vehicle operation has been unsuccessful to date. 21.5.3.1.2 Plasma technologies In its simplest form, the plasma system operates with AC voltage that is applied between two metal electrodes of which one is coated with a nonconductive material. Silent discharges consisting of microdischarges within the microsecond range occur which cause all rising reactive groups to decompose from chemical bonding and recombination processes. The resulting plasmas have a condition of internal energy unbalance distribution with high electron temperatures between 104–106 K and a low kinetic gas that typically fluctuates in the range of 300–1000 K. The plasma consists of an apposition of electrons and charged radicals and ions as well as photons. Because of the inner energy disequilibrium distribution in these plasmas, chemical reactions can occur via nonthermal channels that permit strongly endothermic reactions [21-28]. The two undesired reactions for reducing NO in plasma occur there in addition to numerous other reactions [21-29]. Reaction partners products

e + N2 → e + N + N

(21.16)



N + NO → N 2 + O

(21.17)

Laboratory prototypes with heterogeneous catalysts in plasma fields were tested in engine exhausts with varying results. The extent to which this technology will become a standard application in SI lean-burn engines is presently uncertain. An important criterion for the success of the plasma method is the required energy for generating the plasma and the related disadvantage in fuel consumption, as well as the reduction of NOx at spatial velocities that are predominate in engines.

21.5.3.1.3 Selective catalytic reduction (SCR) NOx conversion in “lean” atmospheres via specially tailored catalysts are termed “selective catalytic reduction.” The required addition of suitable reducing agents yields the end products N2, CO2, and H2O. The term “passive SCR” stands for catalysts that use components that are exclusively in the exhaust for NOx reduction; that is, no subsequently added reducing agents are required (Figure 21.40, top). To be understood as “active SCR” catalytic converters, the reducing agents must be introduced into the exhaust gas system before the catalytic converter (Figure 21.40, bottom) following actual combustion. Passive SCR catalytic converters: These catalytic converters use the hydrocarbons available in the exhaust gas to reduce NOx to produce the reaction products N2, CO2, and water. The basic work in this area is described in [21-27], [21-30], and [21-31]. The activity of fresh catalytic converters based on Cu-ZSM-5 zeolites is very good. Their durability is problematic [21-32], [21-33]. A deterioration of the NOx conversion can primarily be attributed to the sulfur contained in the fuel and thermal aging in the absence of water. An additional example is a passive SCR iridium catalytic converter with a downstream three-way catalytic converter as schematically illustrated in Figure 21.41 [21-34]. It is worth noting that Ir catalytic converters in their new state have lower NOx conversion rates than storage catalytic converters,

Lean-burn Magermotor engine CO2 N2 H2O

SCR cat.

Three-way cat.

Figure 21.41 Schematic for passive SCR system using NOx.

Passive SCR Reducing agent Engine exhaust with +

NOx

CO2

Catalytic converter

N2 H2O

Active SCR Reducing agent Engine exhaust

CO2

Catalytic converter

N2 H2O

Figure 21.40 Differentiation between passive and active SCR.

768 | Internal Combustion Engine Handbook

6606_Book.indb 768

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

for example. As compensation, the sulfur tolerance is much greater. Furthermore, when a passive SCR catalytic converter is used, a prior catalytic converter close to the manifold to reduce cold-start HC emissions cannot be used since this would also convert the hydrocarbons required to reduce NOx [21-35], [21-36]. The HC emissions in the postcold-start phase must, therefore, be countered with other suitable measures. Reducing the catalytic converter light-off time in the test cycle by putting it closer to the manifold is limited in practice by the temperature stability of the Ir catalytic converter [21-12].

gasoline engine. It breaks down into ammonia and carbon dioxide when injected into the exhaust. Urea has the advantage that no gaseous ammonia has to be stored in the vehicle. Before SCR can be successfully used in passenger cars with lean SI engines, many problems have to be solved [21-13]. Under transient conditions, the proper amount of reducing agent must be provided by a control system without “NH3 gaps.” The injection of the reducing agent into the exhaust gas must be adapted to the strongly fluctuating quantity of NOx, the flow rate, and the temperature, and may not simultaneously contribute to vehicle emissions. The maximum temperature resistance of the catalytic converter is apparently insufficient for use in a lean-burn SI engine. For example, it is approximately 650°C for the cited vanadium–titanium catalytic converter. The cost of the overall system with the nozzles, storage container, tubing, onboard diagnosis, etc., must also be considered, as well as the yet-to-be-developed infrastructure for filling up the reducing agent. The prospects for a successful implementation in lean-burn gas engines are, therefore, rather low to moderate.

 ctive SCR Catalytic Converters: A Active selective catalytic reduction requires an efficient mixture of nitrogen oxides with the additionally introduced reducing agents before the catalytic converter inlet (Figure 21.42). Reducing agents, such as ammonia or urea, are used. This technology is highly efficient in stationary applications, such as energy-generating systems, in which the chemical reactions occur in a narrow window determined by temperature, flow speed, and NOx concentration. In these applications, ammonia is the reducing agent that generates N2 and H2O.

21.5.3.1.4 NOx storage catalytic converters The most promising method for reducing NOx emissions in lean-burn engine exhaust is to use NOx storage catalytic converters, also termed NOx adsorbers or NOx traps [21-14], [21-15], [21-16], [21-17]. Since initial series applications for [21-18], [21-19] exhaust treatment in passenger cars with lean-burn SI engines are based on this technology, a discussion on NOx storage catalytic converters with greater detail is in the following section.

Ammonia (NH3 )

H2O

NOx NH3 NOx NH3

Exhaust gas

N2

Catalytic converter

H2O N2

21.5.3.2 The NOx Storage Catalytic Converter The functional principle is schematically illustrated in Figure 21.43 and can be described by four basic steps for the conversion of NOx into N2. The NO contained in the engine exhaust is oxidized on the precious metal of the catalytic converter by a reaction with oxygen during lean-burn operation and forms NO2.

Figure 21.42 Schematic diagram for active SCR system.

The operating temperature range of the desired chemical reactions depends on the respective catalytic converter. Vanadium–titanium catalytic converters operate most efficiently at temperatures between around 200°C and 450°C. At low temperatures, the catalytic converter is impaired by ammonium sulfates, and at high temperatures, it oxidizes ammonia into NO. The upper temperature limit when using ammonia lies at around 600°C. Urea (an NH3 compound) would be the most promising reducing agent for usage in a

1 NO + O 2 → NO 2 2



The NO2 then reacts with the metal oxides embedded as storage materials in the catalytic converter, which causes the formation of a corresponding nitrate storage material.

LEAN

FAT

NO2

O2

N2

Fuel

NO NO2

NO2

Storage material

Pt Storage

(21.18)

NO2

Storage material

NO

Regeneration

CO

HC

Pt Figure 21.43 Model example of NOx storage and regeneration.

Internal Combustion Engine Handbook | 769

6606_Book.indb 769

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

NO 2 + MeO → Me − NO 3



(21.19)

Since this reaction is not catalytic, but rather stoichiometric, it “consumes” the storage material. The effectiveness of the nitrate formation decreases with increasing NO2 quantity. A saturation state is achieved. The storage material must therefore be periodically regenerated to maintain a high level of storage effectiveness. This requires a brief changeover to understoichiometrical (“greasing”) engine operation. The temperature stability of the nitrates under “greasing” operating conditions is less than in lean operation, which results in decomposition of the nitrates NO and MeO. 1 ME − NO 3 → MeO + NO + O 2 2



(21.20)

The NO released in this process is then converted into N2 with the aid of the reducing agents HC and CO which are also present under “greasing” operating conditions. 1 NO + HC/CO → N 2 + H 2 O/CO 2 2



(21.21)

Requirements for NOx storage catalytic converters can be derived for practical vehicle applications requiring specific properties. The essential criteria for evaluating the quality and usability of NOx adsorbers include the following: •• NOx storability •• NOx regeneration capacity •• operating range for NOx storage/regeneration •• HC/CO conversion during lean operation •• conversions in the Lambda = 1 operation •• maximum temperature stability •• sulfur resistance and regeneration capacity NOx storability, NOx regeneration capacity, operating temperature range, etc., initially represent the properties of

the NOx adsorber with regard to its conversion capacity in the new state. The maximum temperature stability, sulfur resistance and sulfur regeneration capacity are also properties with respect to durability. 21.5.3.2.1 NOx storability and regeneration capacity Figure 21.44 shows the typical storage curve of two NOx adsorber catalytic converters. After the stored material is emptied, a storage process starts with a high degree of efficiency that decreases as the catalytic converter fills. In order to permit low-consumption lean operation between two regenerations for as long as possible, the goal of development is the highest possible NOx storability with a high efficiency. Figure 21.44 shows how catalytic converter B has a higher storage capacity than catalytic converter A. Since efficiency above 90% is necessary to maintain Euro IV exhaust thresholds depending on the application, the NOx storage potential cannot be fully exploited in practice and the storage medium has to be regenerated before this point. As mentioned in the discussion of the mode of operation, the engine is briefly run rich so that the NOx arising during nitrate decomposition is converted into N2 with the aid of the reducing agents HC and CO. To keep the fuel consumption as low as possible during rich operation, developers seek to efficiently exploit the regeneration material. Figure 21.45 shows the NOx efficiency of two catalytic converters in an engine test bench cycle with a 60 s lean phase and 2 s rich phase. An efficiently regenerating catalytic converter representing the current state of development is compared with an older, poorly regenerating catalytic converter. The efficiently regenerable catalytic converter can be completely regenerated with 2 s rich operation despite a high NOx storability, recognizable by the greater efficiency in the initial lean cycle. In contrast, the “bad” catalytic converter that stores NOx in the same manner is incompletely emptied during the rich peak, which decreases efficiency from cycle to cycle.

100 Test conditions: RG = 40 k/h Temperature Cat inlet = 350 °C l = 1.30 NOx Concentration = 330 pm Crystallized sulfur = 50 ppm

90 80

NOx efficiency [%]

70 60 50 40 30 20

Cat A

10

Cat B

0

Stored NOx mass

Figure 21.44 NOx retaining processes of two catalytic converters at 350°C.

770 | Internal Combustion Engine Handbook

6606_Book.indb 770

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

NOx efficiency [%]

90

9

80

8

70

7

60

6

50 40

Can be poorly regenerated

30

Can be well regenerated

5 4

Test conditions: Temperature Cat inlet = 350 °C 60s lean, 2 s fat l = 1.30 / 0.85

l

20

l [-]

10

100

3 2 1

10

0

0 0

100

200

300

400

500

600

Figure 21.45 NOx storage unit regeneration of various catalytic converters.

Time [Seconds]

21.5.3.2.2 Temperature range for NOx storage and regeneration When NOx storage catalytic converters are used in the vehicle underbody, the European emission test can be expected, depending on the application, to show inlet temperatures from under 300°C (ECE range) to over 500°C (EUDC range). To comply with Euro IV laws, a NOx efficiency of more than 90% is required depending on the application. The temperature range in which the cyclic storage and regeneration attains such a level of efficiency in NOx storage catalytic converters is the focus of particular interest. In addition to restrictions posed by the engine, it is restricted by the map range in which the engine can be leanly operated with low consumption; from the user’s perspective, the temperature range should, therefore, be as wide as possible [21-18].

100

Figure 21.46 shows the different NOx efficiency curves of two new catalytic converters (precious metal content: 125 g/ cu.ft.) in an engine test bench test with a 60 s lean and 2 s rich phase plotted against the catalytic converter inlet temperature. The efficiency at low temperatures is limited by the “lightoff” of the catalytic converter in this case, the ability of the precious metals to oxidize NO into NO2. The top temperature limit is essentially restricted by the stability of the formed nitrates, that is, the ability of the storage material to form thermodynamically stable nitrates even at high temperatures [21-16]. Since barium as a NOx storage material does not form stable nitrates such as potassium, the efficiency of Ba catalytic converters drops at temperatures above 400°C, whereas the efficiency of a potassium catalytic converter at 500°C is still above 90% Figure 21.46.

Measurement of average NOx efficiency without preliminary catalytic converter across 8 cycles lean/fat 60/2 s

90

NO efficiency [%]

80 70 60 50

Test conditions: Space velocity = variable Time lean/fat = 60/2 s 60s, 2 s fat I = 1.30 / 0.80 Sulfur content = 30 ppm

40 30 20 10 0

200

250

300

350

Barium cat

Potassium cat 400

450

500

Catalytic inlet temperature [°C]

550

600

650

Figure 21.46 Operating temperature range of NOx adsorbers in their new state.

Internal Combustion Engine Handbook | 771

6606_Book.indb 771

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.5.3.2.3 Three-way properties and HC/CO conversion in lean operation NOx catalytic converters can generally compare well with modern three-way catalytic converters in terms of three-way properties and HC/CO conversions in lean operation. The HC activity is negatively affected, however, when very strongly alkaline materials such as potassium are used as NOx storage components [21-20]. Figure 21.47 compares the conversions of a barium catalytic converter and a potassium NOx catalytic converter with a modern three-way catalytic converter. The conversions in homogeneous lean operation at Lambda 1.5 and a 350°C inlet temperature are displayed in the left-hand side of the image while the right-hand side hows the conversions in the λ = 1 regulated operation at 450°C. Catalytic converters aged 25 h, overrun cut-off 820 °C 100 95 90

Conversion [%]

85 80 75

Three-way cat

70

Ba-NOx trap cat K-NOx trap cat

65 60 55 50

HC

CO l = 1.5 350 °C

HC

CO l = 1.0 350 °C

NOx

Figure 21.47 HC-/CO conversion in lean operation and at λ = 1 operation.

21.5.3.2.4 Temperature stability Figure 21.48 shows the influence of various aging conditions on a NOx adsorber catalytic converter with barium as the storage material. The reference point is 1 h at 650°C, with a catalytic

converter stabilized at λ = 1. After 25 hours of stoichiometric engine testbed aging at 820°C in the catalytic converter bed, a reduction in the NOx activity could be seen across the entire operating temperature range, which did not progress further, however, when the aging was continued up to 50 hours. The deactivation is attributed to temperature-related sintering of the washcoat, the precious metals contained in it, and the NOx storage unit components. Aging under stoichiometric conditions at the same temperature but with periodic overrun fuel cutoff phases substantially increases the deactivation of the catalytic converter that progresses with aging. The causes are the increased sintering of the precious metals under lean conditions and the reaction of the barium with the aluminum oxide of the washcoat that also occurs under lean conditions at temperatures above approximately 700°C. The barium is irreversibly deactivated for NOx storage. The speed of these effects increases with rising temperature [21-21]. One way to increase the maximum temperature stability is to use a NOx storage material that does not produce this interaction with the washcoat. Results from using potassium as the storage material reveal a much greater aging stability in comparison to barium catalytic converters. Figure 21.49 shows a comparison of both technologies in a high-temperature aging process with overrun cutoff at 850°C. The respective storage capacity of the individual technologies when new is set to 100% as a reference quantity. Based on a new state, the Ba technology shows a steady decrease in the remaining NOx storage capacity as the aging increases. The barium catalytic converter is largely deactivated after 50 hours. In contrast, the potassium technology shows a much higher remaining NOx storability. Although the rate notably decreases after 25 h, the most striking factor is the substantial retention of the remaining storage capacity as aging progresses. In addition to this clear advantage of maximum temperature stability, potassium catalytic converters, as mentioned in the section discussion on the temperature range, have a

Measurement without prelim. cat Measurement of average NOx efficiency across 8 cycles lean/fat 60/2 s

100

1 h Stab. 650 °C 25 h stöchiom. 820 °C 50 h stöchiom. 820 °C 25 h overrun cut-off 820 °C 50 h overrun cut-off 820 °C

90

NOx efficiency [%]

80 70 60 50

NOx efficiency [%]: Spatial velocity = 40/60 k/h Time lean/fat = 60/2 s l = 1,30 / 0,80 Sulfur content = 30 ppm

40 30 20 10 0 200

250

300

350

400

450

Catalytic inlet temperature [°C]

500

550

Figure 21.48 Impact of various aging processes on the operating temperature range of a barium NOx adsorber.

772 | Internal Combustion Engine Handbook

6606_Book.indb 772

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

Relative NOx storage capacity at overrun cut-off aging of 850 °C comparison of potassium <> barium catalytic converter

Relative NOx storage capacity [%]

100

Ba-Tech. NS 50%, 350 °C

90

Ba-Tech., NS 50%, 400 °C

80

K-Tech., NS 50%, 350 °C

70

K-Tech., NS 50%, 400 °C

60 50

Test conditions: Inlet temp. = 350/400 °C RG (NOx Trap) = 40 k/h I lean = 130 NOx conc. lean = 350 ppm NOx conc. lean = 350 ppm Sulfur content = 30 ppm

40 30 20 10 0

0

25 Aging time [hours]

50

75

Figure 21.49 Storability of NOx adsorbers with potassium and barium following high-temperature aging.

wider operating temperature window for NOx storage and regeneration at higher temperatures. This contrasts with disadvantages that are to be weighed depending on the system configuration, vehicle package, and cost. The following disadvantages can be cited: •• lower HC conversion (see the section on HC and CO conversion)

21.5.3.2.5 Sulfur poisoning and regeneration The problem with sulfur poisoning of NOx adsorbers results from the fact that all materials that are suitable for NOx storage also tend to store SO2 by forming a corresponding sulfate. The reactions are analogous to those occurring with NOx storage and are schematically represented in Figure 21.50.

•• higher desulfurizing temperature (see the section on sulfur poisoning)

SO3 SO2

•• incompatibility with certain substrate materials. Since the HC conversion of a potassium-containing NOx adsorber can be much less than that of a barium catalytic converter, this fact needs to be taken into account in the corresponding system design. One option is to use larger preliminary catalytic converters that completely take over HC conversion. The higher temperatures required to desulfurize potassiumcatalytic converters due to more thermally stable sulfates than in barium catalytic converters pose additional demands on the engine management system. These systems must be capable of providing catalytic converter inlet temperatures of approximately 750°C for desulfurization even at times in which lower temperatures would exist in normal engine operation. A decisive disadvantage of potassium catalytic converters is the affinity of potassium with ceramic substrates used in series production. Potassium diffuses at temperatures above around 750–800°C into the ceramic substrate where it becomes deposited and forms a compound with the ceramic. Two negative effects arise as a result. On the one hand, the storage component for NOx storage becomes inactive. On the other hand, the mechanical stability of the ceramic substrate increases. This process is accelerated by high temperatures above 800°C. With the aging shown in Figure 21.49, a potassium catalytic converter was used on a metal substrate where this affinity does not exist. Solutions are presently being developed that permit such a coating even on modified ceramic substrates.

NO

NO2

O2

Precious metal

NO2

SO3

SO3

NOx storage material

NO2

• Competing absorption • Sulfur • Decreased NO2 formation • "Consumption" of storage material

Figure 21.50 Reaction diagram of the sulfur poisoning of an NOx storage catalytic converter.

In lean operation, the NOx adsorber first oxidizes the SO2 to form acidic gas SO3. Just as is the case with NO2, the SO3 also reacts with the storage material to produce the corresponding sulfate. The storage material converted into a sulfate is, hence, no longer available for NOx storage. The actual problem of sulfur poisoning is that these sulfates are more thermally stable than nitrates. Classic storage materials are therefore incapable of sulfate regeneration under the same conditions as those used for NOx regeneration. Over time, the sulfate in the NOx storage medium increases so much that the NOx storage capacity falls to an unacceptable level. In Figure 21.51, the reduction of the NOx storability of a barium catalytic converter at 400°C inlet temperature originating from a desulfurized but thermally preaged state can be detected after 10 or 20 hours of sulfurization in engine operation

Internal Combustion Engine Handbook | 773

6606_Book.indb 773

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

100

Initial state

95

10 hrs. sulfurized 20 hrs. sulfurized

90

20 hrs. sulfurized and desulfurized

85

NOx-Efficiency [%]

80 75 70

Sulfurization: Inlet temp. = 400 °C Cycle 56/4 Sec. I = 1.4 / 0.85 Crystallized sulfur = 40 ppm

65 60 55

Desulfurization: Inlet temp. = 650 °C I = 0.98 Time = 15 min

50 45 40 0

200

400

600

Stored NOx mass [mg]

with a 40 ppm fuel–sulfur content. The NOx storability can be completely eliminated after further sulfurization. The catalytic converter was then desulfurized again after being sulfurized for 20 hours. The conditions were as follows: 650°C Cat inlet temperature, Lambda 0.98, 15 minutes constant operation. The NOx storability of the catalytic converter was regenerated to its initial level as shown by the curves in Figure 21.51. The thermal stability of the sulfates is lower under rich conditions in contrast to lean conditions or when λ = 1, which causes the sulfate to decompose and accordingly regenerates the storage capacity. The sulfate decomposition accelerates the higher the temperature and the richer the exhaust. For bariumcontaining NOx storage catalytic converters, temperatures of approximately 650°C are sufficient for sulfate regeneration. When more basic NOx storage components are used that are able to form stable nitrates at higher temperatures than barium, higher temperatures are also required for desulfurization. During sulfur regeneration under constantly rich engine operation conditions, sulfate decomposition produces SO2 that is then converted in the catalytic converter into the undesired secondary emission product H2S. The desulfurizing strategies that avoid the formation of H2S are currently being developed. Because of space restrictions, however, we avoid giving a detailed discussion of the reactions during desulfurization. Basically the loss of activity of a NOx storage catalytic converter from sulfur poisoning accelerates as the amount of sulfur in the fuel increases [21-22]. The introduction of low-sulfur fuels reduces this problem, which in turn reduces the frequency of desulfurizing that is problematic for fuel consumption. To date, 100% protection of the NOx adsorber from sulfur poisoning by an upstream sulfur trap has been unsuccessful. An apparent benefit from sulfur traps is that the time can be

800

1000

Figure 21.51 NOx storability for sulfurization and desulfurization.

increased between sulfur regenerations of the NOx storage medium [21-15], [21-16]. 21.5.3.3 System with Preliminary Catalytic Converter and NOx Adsorber Based on the statements regarding operating temperature range and temperature stability made in Section 21.5.3.2, the NOx adsorber must be installed in the underbody area. This means that an underhood preliminary catalytic converter is required to deal with cold-start emissions. To maintain current and future exhaust laws dealing with lean SI engines, a system should be used with a preliminary catalytic converter and NOx absorber. Figure 21.52 schematically illustrates such a system. In addition to the cited conversion of cold-start emissions, the preliminary catalytic converter also assumes the tasks of three-way conversion under λ = 1 conditions. In addition, HC and CO are converted during lean operation. This feature is very helpful for the NOx storage process in the adsorber. HC and CO molecules reaching the adsorber are converted by the adsorber in a reaction that competes with NOx storage. As a result, less NOx is stored with a higher efficiency, and the effectively exploitable storage capacity is less. Due to the underhood application, the preliminary catalytic converter requires a minimum temperature stability of 950°C. The target value for the temperature stability of NOx adsorbers is 900°C. State-of-the-art technology has not yet been able to achieve this. For this reason, the maximum temperature load for initial series applications needs to be limited by cooling. Possible locations for cooling devices are shown in Figure 21.52. The type of cooling that is used depends on such factors as the available space, the required cooling, and the cost. In conclusion, the effort involved in creating the system must be justified by the attainable practical advantage.

774 | Internal Combustion Engine Handbook

6606_Book.indb 774

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

Possible usage sites for exhaust gas cooling

Otto-DI-Motor

Underbody: NOx storage catalytic converter

– NOx storage/reduction – HC, CO conversion in lean operation – HC, CO conversion in lean operation – Temp. stability 900 °C – Sulfur resistance / desulfurization

Close to the manifold: TWC preliminary catalytic converter

– Cold start of HC conversion – I = 1 conversion – HC, CO conversion in lean operation – Temp. stability > 950 °C

Figure 21.52 Catalytic converter configuration for lean gasoline engine concepts for EU-IV applications.

21.5.4 Metallic Catalytic Converter Substrate

Since the initial development of catalytic converters in the early 1960s, efforts have been made to use metal as a substrate in addition to the cordierite extrusions Figure 21.53. To produce a metal substrate, smooth and corrugated metal foils are wound to form honeycomb bodies and are introduced into a tube (Figure 21.54). Over a period of approximately 20 years, it was rather difficult to maintain the necessary mechanical durability of metal substrates since spiral-wound substrates telescope under dynamic loads at high temperatures. Only with the introduction of a high-temperature soldering process to connect the individual foil layers and with the development of a new winding technique were the obstacles to the use of metal substrate catalytic converters largely overcome. The metal foils currently used have a thickness of 0.05 to 0.03 mm. The aluminum in the foils makes them very corrosion resistant, and the very thin aluminum oxide layer on the metal surface allows the oxide washcoats to adhere well with the substrate material. Figure 21.53 Metallic (left) and ceramic (right) honeycomb structures.

Spiral-Support ~ 1968 ~ 1960

~ 1978 Brazed joint

Attachment cross

Support layer

Figure 21.54 Development of the processes for the manufacture of metal substrates.

Internal Combustion Engine Handbook | 775

6606_Book.indb 775

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Relative exhaust gas counterpressure [%]

Chapter 21 Exhaust Emissions

160 Ceramic support

Mass flow 100 kg/h Temperature 400 °C Pressure upstream of cat 1 bar

140 120

Metal support

100 80 60 40 20 0 400 cpsi, 8 mil

400 cpsi, 6.5 mil

600 cpsi, 4 mil

400 cpsi, 0.05 mm

The very thin metal cell walls only slightly raise the exhaust counterpressure (Figure 21.55), which has a positive effect on fuel consumption and engine performance. The time required for the catalytic converter’s operating temperature to be reached is very important for effective exhaust purification, since around 70–80% of all pollutants formed during a test cycle are emitted during this time. Shortening this time is a focal point in the development of exhaust purification technology. The following constructive features should be incorporated to best exploit the exhaust energy to heat the catalytic converter: •• low-thermal capacity •• large geometric surface of the substrate.

Support surface / heating capacity [(m²K)/J]

Metal substrates are very suitable because of their physical features and large surface. The ratio of the substrate surface to substrate thermal capacity (which strongly influences the heating behavior) increases with the cell density, while the cell-wall thickness decreases as shown in Figure 21.56. Figure 21.57 demonstrates how the use of catalytic converter substrates with a higher cell density and equivalent dimensions decreases hydrocarbon emissions in a cold start.

500 cpsi, 0.04 mm

600 cpsi, 0.04 mm

Figure 21.55 Exhaust counterpressure of catalytic converter substrates [21-26].

The conversion of pollutants can also be enhanced by using substrates with a high cell density even after the operating temperature of the catalytic converter has been reached. This is demonstrated in Figure 21.58 by the representation of the hydrocarbon conversion in bag 1 of the FTP (federal test procedure) cycle as a function of the cell density. The effect of using higher cell densities to increase the conversion rate thus clearly exceeds the effect of enlarging the catalytic converter volume. The increase in catalytic efficiency is not based only on the increase in the substrate surface as the cell density increases. When the channel diameter falls as the number of cells increase, the transfer of material from the gas phase to the channel wall also improves. This effect is illustrated in Figure 21.59. For various reasons, the increase in cell density is only possible to a limited extent. On the one hand, the foil thickness cannot be reduced however desired and power losses due to increased counterpressure are generally not acceptable. On the other hand, the impact of the flow distribution through the catalytic converter cross section on the conversion result increases as exhaust standards become more strict.

0.025

0.02

0.015

0.01

0.005

0 400 cpsi, 0.05 mm

400 cpsi, 0.04 mm

600 cpsi, 0.03 mm

800 cpsi, 0.025 mm

Figure 21.56 Ratio of substrate surface/substrate heating capacity for metal substrates having various cell densities.

776 | Internal Combustion Engine Handbook

6606_Book.indb 776

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

Cumulative HC emissions [g]

0.6

Raw emissions 400 cpsi, 0.05 mm 600 cpsi, 0.04 mm 800 cpsi, 0.03 mm 1000 cpsi, 0.025 mm

0.5 0.4 0.3 0.2 0.1

Figure 21.57 Cumulative hydrocarbon emissions during the first 100 seconds of the FTP cycle (catalytic converter dimensions: ∅ 98.4 × 74.5 mm) [21-27]. See color section page 1092.

0 0

20

40

60

80

100

Time [s]

2

THC Emissions [g] [g] Hydrocarbon emissions

)

100 cpsi

1.5

# 0,21 l 200 cpsi

)

1

Constant cell density (= 400 cpsi) at increasing volume

0,28 l

#

300 cpsi )

0,47 l 400 cpsi # 0,64 l ) 500 cpsi ) # 600 cpsi )

0.5

0,81 l

# ) 800 cpsi

Increasing cell density at constant volume (= 0.47 l)

0,95 l

#

)

1000 cpsi

0 0

1

2

3

1,3 l

#

1200 cpsi

4

Ratio of support surface to hydraulic diameter [m2/mm] Conversion Performance Factor: GSA / dh [m²/mm]

600 cpsi

200 cpsi

Diffusion

HC HC

CO NOx

d1

HC

d2

Diffusion

Diffusion

CO NOx

CO NOx

Material transition of the circulation on the surface of the channel wall due to gas phase diffusion

Figure 21.59 Schematic diagram of the diffusion paths at different hydraulic diameters or cell densities.

5

Figure 21.58 Cumulative hydrocarbon emissions in bag 1 of the FTP cycle as a function of the ratio of substrate surface to hydraulic diameter (GSA/dh).

The usage of perforated foils (PE design) as shown in Figure 21.60 enables flow equalization in the catalytic converter. This leads to a more uniform utilization of the entire catalytic converter volume and to a reduction in the exhaust counterpressure. The loss in the geometric surface area of the substrate can be compensated in this process by the generation of local turbulences during flow equalization between neighboring channels, with the conversion performance remaining constant [21-28]. Structures can be inserted in the channel walls to further increase the catalytic conversion especially of metal substrate catalytic converters. Figure 21.61 illustrates the simplest case of structured channels, the transversal (TS) structure. The micro-TS corrugations are perpendicular to the gas flow and cause a stronger transmission of material from the gas phase to the substrate walls from the formation of local turbulence.

Internal Combustion Engine Handbook | 777

6606_Book.indb 777

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

without lowering the conversion rate, thereby decreasing the required installation space of the catalytic converter and minimizing its output loss.

Figure 21.60 PE design.

Figure 21.62 View of a foil with LS corrugations and a schematic representation of the effect of the LS corrugation.

Figure 21.61 View of the face of a metal substrate with TS channels and a schematic diagram of the TS microcorrugations in the channel [21-29].

An even more intense material transmission in comparison to the insertion of microcorrugations can be achieved using countercorrugations in the channel. Figure 21.62 schematically illustrates the longitudinal (LS) channel structure. By structuring the channels in this manner, the conversion can be improved on in comparison to substrates with identical dimensions but with smooth channel walls. An alternative is to reduce the substrate volume or cell density

Figure 21.63 illustrates various conversion results of various catalytic converter systems on a supercharged 2.7-l V6 diesel engine, which demonstrates the advantages of the structured substrate channels. A constant or even better conversion result could be achieved with a converter volume reduced by around 25%. A special form of channel structures is implemented in the so-called mixer foil (MX). If the corrugated MX foil is combined with a gas-penetrable smooth porous layer, then a PM (particulate matter) filter catalytic converter forms. Its design and mode of operation are illustrated in Figure 21.64. The blades in the corrugated foil cause a part of the exhaust gas flow to be pressed through the smooth sintered metal fleece layer, with the particles carried along in the gas flow being deposited in the fleece. If the PM filter catalytic converter is operated with an upstream oxidation catalytic converter, then a continuous regeneration of the filter can be achieved by NO2 at temperatures as low as around 200°C. As opposed to conventional wall flow filters, this filter has the advantage of making clogging of the gas paths impossible. That guarantees maintenance-free operation with only a marginal increase in the exhaust counterpressure which only negligibly affects the engine output and fuel consumption. Figure 21.65 summarizes the reductions in particle mass and particle count achieved by the usage of PM Metalit® in diesel passenger cars and trucks.

778 | Internal Combustion Engine Handbook

6606_Book.indb 778

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

0.3 FTP-72

Cumulative THC emissions [g]

0.25

US-06

0.2

0.15

0.1

0.05

0 Ceramic 3.8 ltr.

TS metallite 2.8 ltr.

LS metallite 2.7 ltr.

LSPE metallite 2.5 ltr.

Figure 21.63 Reduction in the total carbon emissions of a 2.7-l V6 diesel engines with various catalytic converter systems [21-30]. See color section page 1092.

Sintered metal fleece Mixer shaft foil (MX)

Mixer shaft foil with blades

Gas circulation

The PM-Metalit® reduces the particle count more greatly than the particle mass, which is due to the fact, that primarily small particles with diameters less than 100 nm are retained. Once of today’s greatest challenges in the exhaust gas purification of diesel engines lies in the reduction of nitrous oxide and particle emissions while increasing fuel consumption as little as possible. Upcoming exhaust gas treatment systems should thus have lower weight and allow the exhaust counterpressure to increase as little as possible. Time-consuming regeneration strategies based on fuel post-injection should also be avoided. These requirements

Figure 21.64 Design and function of the PM-Metalit® system.

are fulfilled in the Emitec-SCRi-System, whose design is schematically displayed in Figure 21.66 as an example. The consistent usage of LS and MX structured foils in all substrates allows for a relatively small catalytic converter volume with good conversion performance. Moreover, the urea or ammonia concentration can be homogenized across the cross section by internal flow equalization that results in better utilization of the available converter volume and helps to prevent ammonia penetration.

Internal Combustion Engine Handbook | 779

6606_Book.indb 779

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

TRUCKS

-45% -75%

PM emissions

PM emissions

PASSENGER CARS

-70%

-64%

ESC

Without filter

PM metallite Mass

PM metallite Number

Oxidation catalytic converter Ø 177.8 ~ 114 ~ 101.5 mm, 200/400 cpsi LS hydrolysis coating on the last 20 mm

-80%

ESC ETC

ETC

Without filter

PM metallite Mass

-72%

ESC ETC

PM metallite Number

Figure 21.65 Reduction in the particle mass and the particle count by the usage of PM Metalit® in passenger cars and trucks [21-31].

SCR1 Ø 230 ~ 63.5 mm, 300/600 cpsi LS high-temperature SCR coating

Adblue injection against the circulation on the gas outlet side of the oxidation catalytic converter

PM metallite Ø 174.6 ~ 176 mm, 200 cpsi with hydrolysis coating

SCR2 Ø 230 ~ 110 mm, 300/600 cpsi LS SCR2 low-temperature SCR coating

Figure 21.66 The Emitec SCRi system [21-32].

780 | Internal Combustion Engine Handbook

6606_Book.indb 780

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.5 Exhaust Treatment for Gasoline Engines

800

6

600 500

-66%

400 300 200 100 0

Cumulative NOx emissions [g/kWh]

Cumulative PM emissions [mg]

700

5

4 -80%

3

2

1

0 Raw emissions

Following SCRi

Raw emissions

Converter results with the SCRi system are summarized in Figure 21.67. The use of metals as substrate materials permits the coldstart phase particularly critical for pollutant emissions to be shortened by heating the catalytic converter to the necessary operating temperature. The metal monolith serves as a resistance heat conductor. An EHC (Electrically Heated Catalyst), as shown in Figure 21.68, can be used to decrease the total emissions as compared to an unheated system as Figure 21.69 demonstrates. Because of their construction, metal substrate catalytic converters can be directly welded to the exhaust gas system. The recycling of vehicle components has become increasingly important in recent years. A special method developed for metal substrates allows the utilized materials to be almost completely recovered. The basic progression of this process is shown in Figure 21.70.

Following SCRi

Back-up catalytic converter (eroded)

Figure 21.67 PM and NOx reduction with the SCRi system (ETC) [21-32].

Heating panel with electrical connections

Figure 21.68 Emitec-EHC.

Emissions [g/km]

0.2

HC CO/10 NOx

0.15

0.1

0.05

0 unheated

heated

Figure 21.69 Comparison of the emission results (heated/unheated) in the Euro II-Test for the BMW. See color section page 1092.

Internal Combustion Engine Handbook | 781

6606_Book.indb 781

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Figure 21.70 Schematic diagram of the recycling process developed for metal substrate catalytic converters.

21.6 Exhaust Treatment in Diesel Engines

•• more complex emissions control since gas phase reactions must be dealt with due to the share of particles in the exhaust, and less reducing agent is available due to lower HC concentrations.

21.6.1 Diesel Oxidation Catalytic Converters

For exhaust purification, these characteristics mean that exclusively oxidative reactions are used, and catalytically active components are used that chiefly have to cover the low temperature range (fast light off).

In contrast to SI engines, a diesel system is more thermodynamically efficient. This is expressed in lower fuel consumption and lower CO2 emissions than its SI counterpart. For exhaust purification, diesel oxidation catalytic converters (DOC) have been used for more than ten years in diesel passenger cars. Diesel engines are operated with excess oxygen. Their maximum exhaust temperature is approximately 850°C, and their average is clearly below that of comparable SI engines. This means •• better CO and HC emissions •• higher NOx emissions •• much higher particle emissions

21.6.1.1 Pollutants in Diesel Exhaust 21.6.1.1.1 Hydrocarbons and CO Even under excess oxygen, heterogeneous combustion chamber conditions can cause incomplete oxidation reactions and uncombusted or partially oxidated hydrocarbons in the exhaust in addition to CO. Some of these compounds are responsible for the typical smell of diesel exhaust. All engine parameters that improve the exploitation of the oxygen in the combustion chamber (such as swirling the mixture) or higher combustion temperatures can lower CO and HC emissions.

782 | Internal Combustion Engine Handbook

6606_Book.indb 782

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

21.6.1.1.2 Particles Due to the intermediate steps of acetyls and polycyclic hydrocarbons, locally rich conditions during combustion lead to the formation of graphite-like soot particles. Coagulation and agglomeration processes cause soot particles approximately 100 to 300 nm in diameter (median) to form from these approximately 1–10 nm primary particles. Since the large surface area (up to 200 m2/g) of these particles makes them highly adsorbent, a large share (greater than 50% by weight) of hydrocarbons, sulfates, water, and lubricating oil components can be demonstrated in diesel soot in addition to carbon. 21.6.1.1.3 Nitrogen oxides In oxidation reactions, the nitrogen oxides NO and NO2 form in the presence of nitrogen. Since the concentration of both components and their ratio depend on the reaction temperature and the oxygen concentration during combustion, NOx emissions can be reduced by suitable engine measures such as late fuel injection (gas temperature falls) and exhaust recirculation (oxygen concentration falls). 21.6.1.1.4 Sulfur oxides SO2 is the primary compound forming from the combustion of sulfur-containing fuel. It is then further oxidized at temperatures greater than 300°C by precious metals to form SO3, which reacts in the presence of water to form sulfuric acid (H2SO4). All three compounds can deactivate the catalyst; SO2 and SO3 by specific addition that blocks the precious metal, and H2SO4 by coating the washcoat surface and by producing condensation in the washcoat pores.

21.6.1.3 Deactivating the Catalyst Surface While driving, the catalytic converter surface can be reversibly or irreversibly deactivated by chemical or physical influences. Figure 21.71 offers a few possible examples: •• The washcoat surface is coated by residues from oxidation or the further reaction of hydrocarbons (coking). •• Selective contamination by coating and covering active centers such as the addition of sulfur compounds to precious metals. •• The restriction of pore openings hinders the accessibility of active centers (sintering). As shown in Figure 21.72, “light-off curves” measured in model systems provide information on the corresponding deactivating mechanisms. a

b

c

d

21.6.1.2 Characteristics of Diesel Oxydation Catalytic Converters 21.6.1.2.1 Design Similar to three-way catalytic converters, DOCs consist of the following components: •• honeycombs (monoliths) of ceramic or metal as a substrate for the catalytic coating •• Al2O3 or porous, thermally stable coatings with large surfaces (100–200 m2/g) •• precious metal and promoters as catalytically active centers on whose surfaces the oxidation reactions occur. 21.6.1.2.2 Manufacturing One possible manufacturing method includes the following steps: •• Precious metals and promoters are dissolved. •• This solution is applied to the A12O3 surface (the resulting suspension is termed a washcoat). •• The honeycombs are dipped in the washcoat. •• Subsequent drying and calcining processes remove the water from the honeycombs and set the washcoat.

Figure 21.71 Various types of contamination of precious metals and the washcoat pores, (a) sintering of pores, (b) nonselective surface coating, (c) selective contamination of active centers, (d) condensation of hydrocarbons.

21.6.1.3.1 Deactivation under low-temperature and lowload conditions Temperatures up to approximately 250°C and low loads yield reversible contamination of the surface from components containing carbon. The deterioration of the CO light-off shown in Figure 21.73 after 1 h of idle operation (<120°C, <20 Nm) can be reversed by a transient increase in the exhaust temperature (<1 min, <250°C); this means that the activity can be completely recovered.

Internal Combustion Engine Handbook | 783

6606_Book.indb 783

1/19/16 8:53 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

to λ < 1; this reduces the sulfur compounds adsorbed on the surface to form H2S, which can be desorbed from the surface and detected in the exhaust. If we integrate over the area of the respective H2S peaks, we obtain information on the sulfurization tendency of the respective catalytic converter. To protect from irreversible sulfurization, promoters can be added to the washcoat that suppress the affinity with sulfur compounds. Figure 21.74 shows a comparison of the H2S signals of a standard washcoat and the correspondingly protected version. The corresponding vehicle test (MVEG) can be seen in Figure 21.75. After aging with 1000 ppm sulfur in the diesel fuel the modified washcoat surface in the ECE part of the cycle has clearly less CO [g/km] than the standard version.

100

CO conversion [%]

80

60

40

20

700

0 150°C

200°C

250°C

300°C

fresh / regenerated Loss of active centers: Sintering Pore diffusion: smaller more effective pores Assignment of the surface, active centres, blocked pores

Figure 21.72 Light-off curves provide information about various deactivation mechanisms.

with promoter Hours

600 500

H2S [ppm]

100°C

400 300 200

100

without pre-conditioning with 1 h pre-conditioning in idle

90

100 0

80

250

260

270

CO conversion [%]

70

280

290

300

Time [s]

60

Figure 21.74 A modification of the washcoat surfaces reduces the affinity for sulfur compounds.

50 2,50

40 30

Aging > 750 ppm S, 100-150 h @ 200-300 °C

2,00

20 1,50

0

[g/km]

10

100

120

140 160 180 200 Temp (before Cat) [°C]

220

Figure 21.73 The temperature at which 50% conversion is achieved is referred to as “light off.” Preconditioning in idle worsens the light off.

21.6.1.3.2 Deactivation by sulfurization An increase in the exhaust temperature beyond >300°C leads to the sulfurization of the catalytic converter. Regeneration under excess oxygen is possible; however, temperatures greater than 600C are required. Alternately, the concentration of reducing agents (hydrocarbons) can be increased by enriching

1,00

CO-Std. HC-Std.

CO-with promoter HC-with promoter

0,50

0,00

ECE CO-Std. HC-Std.

MVEG CO-with promoter HC-with promoter

Figure 21.75 Reduced affinity for sulfur compounds also exhibits improved CO and HC performance in the MVEG cycle.

784 | Internal Combustion Engine Handbook

6606_Book.indb 784

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

21.6.1.3.3 Thermal deactivation Higher exhaust temperatures cause the precious metal to be sintered; that is, small particles agglomerate into larger units. This translates into a loss of the available metal surface and thus lowers the oxidative effect. This irreversible process is portrayed in Figure 21.76. Pt dispersion and CO light off are inversely related, that is, the lower the available Pt surface [measured as (%) dispersion], the lower the CO activity.

21.6.1.4 Evaluating Diesel Oxidation Catalytic Converters 21.6.1.4.1 Light off The activity of DOCs on engine test benches is primarily established by their so-called light off. 100

190

35

185

70

30

180

60

25

175

20

170

15

165

10

160

5

155

30' 10h 20h 40h Aging time at 750 °C

60h

HC [%]

50 40 30 20 10 0 100 120 140 160 180 200 220 240 260 280 Temp. [°C] with promoter

Figure 21.76 High-temperature aging reduces the dispersion of the Pt particles. This reduces the activity of the CO oxidation.

Along with the stainless steel, all of the other washcoat components must be checked for their temperature stability. Analyses of 50 h aging ⪢ 700°C reveal both the favorable stability of Al2O3 surfaces (Figure 21.77) and the unrestricted functioning of a zeolite (Figure 21.74) as indicated by its typical HC desorption curve. 110

Fresh

Fresh + idling Aging 10 hr/750 °C + idling

Figure 21.78 Temperature stability of zeolite up to 850°C.

The conversion at specific temperatures and loads at the point at which 50% conversion occurs is termed the light-off temperature. Figure 21.79 shows the corresponding temperature and torque of a measured 1.9 l naturally aspirated engine, and Figure 21.73 shows the associated curve. The CO light off is around 175°C here. 300

100 Temp (before Cat) Volumetric flow Torque

280 90

Temp (before Cat) [°C]

BET Surface Area [m2/cm3]

100

80

70

60 fresh

2h

Al2O3 - 100/5 doped Al2O3 - 70/5 doped Al2O3 - 70

24 h

48 h

Al2O3 - 100/10 doped Al2O3 - 70/10 doped Al2O3 - 90/10 doped

Figure 21.77 Temperature stability of Al2O3 up to 900°C.

260

90 80

240

70

220

60

200

50

180

40

160

30

140

20

120

10

100

2

4

6

Torque [Nm] Vp air [m³/h]

fresh

80

CO-Light Off [°C]

Pt-Dispersion [%]

90 40

0 8 10 12 14 16 18 20 22 24 26 28 Time [min]

Figure 21.79 Activity measurement of catalytic converters by light-off tests, here the temperature ramps and torque progression of a 1.9-l suction diesel engine

Internal Combustion Engine Handbook | 785

6606_Book.indb 785

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

0,7

5

0,6

4,5

0,5

4

0,4

3,5

0,3

3

0,2

2,5

0,1

2

0

2

4

6

8

10

12

0

Sample (Front to Rear) Sulfur - CC Carbon - CC

Sulfur - UF Carbon - UF

HC [g/km]

Figure 21.80 Postmortem analysis: C/S profiles of CC and UF catalytic converters.

In the UF catalytic converter, both gradients are parallel. 21.6.1.4.3 Endurance tests Endurance is determined both within defined driving cycles and in normal driving. Figure 21.81 shows as an example the AMA cycle for 20,000 km of an aged DOC. Primarily low-temperature and low-load stability are inspected in this cycle. The curves of the catalytic converter measured after 5000, 10,000, 15,000, and 20,000 km in the MVEG cycle are horizontal for approximately 5000 km; that is, deactivating ceases after this time. The data in Figure 21.81 originate from a vehicle that was regularly measured in normal operation over 80,000 km. In this case as well, we see a characteristic similar to that in Figure 21.82. After initial aging, the conversion remains stable over the entire duration of the test.

21.6.2 NOx Adsorbers for Diesel Passenger Cars

The removal of nitrogen oxides from lean engine exhaust with the aid of NOx storage catalytic converter technology can be used in SI and diesel engine drive systems. In the following, we address the specific characteristics of exhaust purification of diesel passenger cars. The characteristic differences from SI DI applications are the lower exhaust temperatures, higher soot emissions, and peculiarities of generating rich exhaust conditions that are necessary to regenerate the storage catalytic converter from adsorbed NOx and SOx. 21.6.2.1 Operating Range of Storage Catalytic Converters The low exhaust temperatures of diesel passenger cars mean that catalytic converters have a lower thermal peak load, and

0,200

2,000

0,160

1,600

0,120

1,200

0,080

0,800

0,040

0,400

0,000

CO [g/km]

5,5

Carbon [Weight %]

Sulfur [Weight %]

21.6.1.4.2 Postmortem analyses of deactivated catalytic converters The physical chemistry of aged catalytic converters is investigated in postmortem analyses to determine the deactivation processes during aging. Figure 21.80 shows both the carbon and sulfur gradients of a nearly bent (CC) and an underbody catalytic converter (UF). The characteristics indicate that more carbon-containing deposits are found in front in the cc arrangement, whereas there is strong sulfurization over the entire length of the monolith.

0,000 fresh

10000 km

HC-Bag1: 0–195 s HC-EUDC CO-Bag2: 196–780 s

20000 km

HC-Bag2: 196–780 s CO-Bag1: 0–195 s CO-EUDC

Figure 21.81 Endurance testing under low-load conditions provides information about endurance (here an excerpt of a test series, which was carried out up to 80,000 km).

786 | Internal Combustion Engine Handbook

6606_Book.indb 786

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

Bagresults MVEG EU2000 0,6

HC,CO Emissions [g/km]

HC

CO

0,5

0,3

0,2

0,0 5

10

15

20

25

30

35

40

45

50

55

60

65

70

75

the window of operation is shifted toward lower temperatures. Figure 21.83 shows the typical NOx emissions of a diesel passenger car as a function of catalytic converter bed temperature in the MVEG driving cycle. Below 150°C, there is no NOx conversion because the catalytic converter has not yet achieved light off. Between 150°C and 250°C, NOx can be stored and reduced; however, because of the high spatial velocities and high NOx concentrations, less NOx is stored. The optimum efficiency of the catalytic converter lies between 250°C and 300°C. At temperatures above 350°C, NOx storage can be thermodynamically limited depending on the NOx storage material. As can be seen in Figures 21.83 and 21.84, the majority of NOx emissions in city driving occur from 150°C to 200°C. In this temperature range, the storage and reduction of NOx is kinetically limited. The conversion of NOx in city driving thus poses substantial demands on the low-temperature activity of the catalytic converter.

NOx Emissions in the MVEG cycle, %

Aging [1000 km]

Figure 21.82 Evidence of the durability from 80,000 km field tests.

75 Kinetic limitation of the NO x conversion 50 No NO x conversion

Thermodynamic limitation of the NOx conversion Optimal operating range

25

0

25 °C–150 °C 150 °C–250 °C 250 °C–300 °C

300 °C–380 °C

Figure 21.84 Distribution of the NOx emissions in the MVEG cycle in four temperature ranges relevant to the catalytic converter.

0.1 City operation (ECE)

Highway and Freeway (EUDC)

NOx Emissions, g

0.08

0.06

0.04

0.02

0

0

100

200

300

Temperature, °C

400

500

Figure 21.83 NOx emissions of a diesel passenger car as a function of catalytic converter bed temperature in the MVEG driving cycle.

Internal Combustion Engine Handbook | 787

6606_Book.indb 787

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.6.2.2 Desulfurization Another peculiarity in the operation of NOx storage catalytic converters in diesel passenger cars is the restricted desulfurization temperature when regenerating SOx. It is nearly impossible to drive at λ < 1 under a high load because of increased soot emissions and loss of torque. Typical desulfurization temperatures of diesel exhaust lie within the range of 500–550°C. The desulfurization capacity of the catalytic converter greatly depends on the selected NOx storage component (NSC: NOx storage component). The tighter the NOx is bound to the catalyst, the more efficient the storage of NOx at high temperatures. A high NOx bonding strength also reduces NOx peaks during regeneration phases. The advantages of stronger NOx adsorption are at the cost of higher desulfurization temperatures. Figure 21.85 illustrates the compromise between desulfurization capacity and the efficiency of NOx storage and regeneration.

Efficiency of the NOx storage

NOx breakdown during regeneration

Efficiency of the desulfurization

Figure 21.85 The combined properties of efficiency of the storage and regeneration of NOx and the desulfurization of NOx storage catalytic converters.

Without measures to encourage desulfurization, the NOx conversion rate of storage catalytic converters in the lean/ rich cycle is linear as a function of the mileage or time. The mileage that can be traveled to attain a given NOx conversion rate decreases inversely proportional to the sulfur content in the fuel. Figure 21.86 shows the relationship of the mileage as a function of the fuel sulfur content at a constant NOx conversion rate. The degree to which the NOx conversion rate is reduced when operating with sulfur-containing exhaust also depends on the temperature and frequency of desulfurization. High exhaust temperatures and long periods between desulfurization leads to the formation of sulfates between the SOx of the exhaust and the storage material of the catalytic converter that irreversibly damage the storage catalytic converter. The development of new storage materials has lowered the desulfurization temperature and enhanced the longterm stability of NOx conversion with the aid of periodic

desulfurization in the lean/rich cycle of sulfur-containing synthetic exhaust. However, the required sulfur tolerance in all operating states of real driving remains insufficient. The long periods of constant lean driving at high exhaust temperatures are particularly critical.

Mileage

The development of modern diesel engines with lower fuel consumption also further lowers the exhaust temperature in the ECE cycle.

Sulfur content in the fuel

Figure 21.86 Mileage traveled before reaching a threshold of the NOx conversion as a function of the fuel’s sulfur content.

21.6.2.3 Regeneration Methods Both internal and external methods to regenerate catalytic converters from stored NOx and SOx are being investigated. In external regeneration, a reducing agent, usually diesel fuel, is sprayed in front of the storage catalytic converter. To reduce the oxygen mass flow, a partial stream of exhaust is guided over the storage catalytic converter [21-43], [21-44]. This technique requires the use of exhaust valves that are problematic since they wear quickly. In addition, the injection of diesel fuel when the exhaust is below 250°C causes the fuel to condense. Alternative methods for onboard generation of gaseous reducing agents by the reformation of diesel fuel are being investigated; however, they are complex. For internal or engine regeneration, reducing conditions are achieved in the exhaust by operating the engine with variable injection parameters. The “poorer” combustion of the fuel in the combustion chamber raises the exhaust temperature, and the throttling of the volumetric flow of intake air reduces the volumetric flow of exhaust. Both measures increase the efficiency of regeneration. A disadvantage of internal regeneration is the increased soot emissions that arise during rich engine operation at higher loads and contribute to the deactivation of the storage catalytic converters. As part of an integral approach to purifying the exhaust of NOx and particles, combined exhaust systems with soot filters and NOx storage catalytic converters are under intense investigation.

21.6.3 Particles/Particle Filters

In 1775, Pott reported on cancer in chimney sweeps. In 1868, Tyndall discovered the optical effect for measuring fine particles; in 1936, the newspaper Dust noted the importance of submicron particles in its first issue. In 1958, P.J. Lawther describes the strong increase in lung cancer mortality in airpolluted areas and points out that this increase began around

788 | Internal Combustion Engine Handbook

6606_Book.indb 788

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

1920. In 1959, the Johannesburg Convention determined the size of particles accessible to lungs. In 1980, particle filters were suggested for vehicle use, and, in 1983 (one year after the EPA’s introduction of the first threshold of 0.6 g/mile for the particle mass PM), an SAE Congress was held that addressed this subject. In 1987, WHO declared diesel particles as “probably being carcinogenic for humans.” The subject is, therefore, not a new one. What is new is that after more than two decades of slow technological development with the participation of many researchers and companies, more than 5 million particle filters are now being used—in Europa primarily in passenger cars since 2000 (Peugeot 607), in the USA in SNF with the introduction of EPA 2007, that more than 100,000 particle filters have been retrofitted in construction machinery and forklifts, in locomotives, ships and stationary engines, that in many cities the public transportation buses have been upgraded, that the degree of separation for solid particles (soot) as exceeded 99.9%, that California, Japan and Korea are implementing projects in million quantities, that the introduction of the particle filter is required for new exhaust regulations in Europe, USA and Japan and that OEMs are now consistently including the “particle filter” element in the vehicle concept and the process control of the engine. What is also new is the introduction of a new definition of limit values in Switzerland and the EU which are based on the number of solid particles in as well as their mass. The technical foundation has thus been laid and has set the stage for emission laws worldwide. 21.6.3.1 Particle Definitions and Particle Properties Different definitions of air pollution “particles” can be found in the regulations: •• According to the legally valid definition for street traffic, particle mass is everything that can be filtered and thus weighed at < 325 K (gravimetric method) independent of the size of the particle and its chemical composition. •• With EU Ordinance 715/2007 [21-100], a new limit value will be introduced for passenger cars and lightweight trucks based on the total number of solid particles in the size range 23 nm to 2.5 μ m whereby all particles not evaporating when the gas sample is heated to 300°C are deemed solid particles in accordance UNECE-PMP [21-89], [21-100]. This step is also provided with Euro VI for SNF [21-101]. •• In the workplace, most regulations count the overall mass of elementary carbon EC (soot) that is less than 5 μ m; there are strong tendencies to shift this limit toward less than 500 nm. •• The term “fine dust” is being increasingly used in environmental legislation to quantify the content of airborne particles in the breathing air with potential pathogenic effects. Potential pathogens are everything which can be breathed in. While the settling speed of 10 cm/s (corresponding to the aerodynamic diameter of 57 μ m of a round particle with a density of 1 g/cm3) was formerly considered an upper boundary for inhalable airborne dust, PM10 and PM2.5 are now being used to refer to the aerodynamic

size of the dust particle itself. In accordance with this, PM10 corresponds to a sample of atmospheric airborne particles having passed through an upstream filter with an average separation efficiency of (50% of the mass) at 10 μ m. The separation characteristics of this upstream filter have been standardized to DIN EN 481 since 1993. In line with the gravimetrically defined separation characteristics, particles with a 30 μ m diameter can also quite possibly be contained in a PM10 sample. These definitions are based on the deposition characteristics of particles in the airways [21-93], originate from the field of occupational medicine and date back to the Johannesburg Convention 1959 [21-94]. PM10 is designated as being commonly found in the thorax and PM2.5 as a common substance in the alveole, whereby this classification strictly applies only to hydrophobic particles. The PM10 definition thus solely says something about the maximum size of the particle when the sample is taken [21-95] and nothing about the material composition and the size distribution of the sample and is therefore poorly suited for the classification of health effects. Moisture and temperature are not monitored during sampling. Artifacts are possible. While this definition makes sense at the workplace for known source characteristics, there are many kinds of substances in atmospheric samples from natural sources, secondarily formed aerosols, resuspended dusts, salts and water. This is thus considered a cumulative parameter. Although the subsequent analysis of PM10 samples allows a statement to be made about the material composition and thus a source classification, no conclusions can be drawn about the size distribution. •• The measured emissions value PM for combustion engines cannot be compared with the fine dust definition PM 10 or PM 2.5. With the same unit of measure, such as g/Nm3,thus the mass of the sample per sample gas volume is solely equal to the mass concentration; the material composition and the particle size distribution differ greatly, however, making a correlation impossible. The PM definition formerly set an upper limit for the temperature only, and not for the particle size. A rough size limitation was not established until the PM 2.5 cyclone was introduced with EURO 5 (passenger cars) to prevent errors due to large particles. These measuring guidelines do not provide a satisfactory definition of particles. These measuring procedures are also insufficient for toxicological evaluation since no information is provided on the size distribution of the particles in an aerosol or on their chemical composition and phase (solid or liquid). The particles captured from the aerosol in diluted and cooled exhaust reveal an agglomeration structure under an electron microscope whose basic elements are nearly spherical and quite dense (approximately 1.8 g/cm3) [21-52]. These particles are created every time hydrocarbons are combusted Figure 21.87. These surface-rich agglomerations (BET surface 100–200 m2/g [21-52]) serve as condensation nuclei upon cooling; and capture films of hydrocarbons and sulfurous acid products that, in turn, can bind a large amount of water.

Internal Combustion Engine Handbook | 789

6606_Book.indb 789

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Figure 21.87 Diesel particle agglomerate (Burtscher).

Substances that flow through the filter as a gas and particularly appear only upon further cooling as a result of condensation (droplet formation) are not considered particles according to a physically correct definition of a hot gas filter. The definition must be limited to substances that have the characteristics of a particle when flowing through the filter, that is, are basically solid particles such as soot, lubricating oil ash, abraded material, mineral particles that are not deposited in the intake filter of the engine, and sulfur products such as gypsum that can be formed with the calcium of the lubricating oil. With high-boiling fuels (biodiesel) and a high-lubricating oil content in the exhaust gas (mixture-lubricated two-stroke engine), a high concentration of very small high-boiling hydrocarbon droplets is sometimes also found, usually in the form of bimodal size distributions [21-109], probably with tiny metal oxide cores. Hydrocarbons deposited on soot (OC = organic carbon) such as the polycyclic aromatic hydrocarbons (PAHs), 9,E+13

vehicle 1

vehicle 7

vehicle 2

8,E+13

dn/dlog(Dp) [1/km]

which also remain caught in the lungs upon entry, must also be considered [21-109]. The size of these particles that come in a wide variety of shapes is difficult to describe. It must be defined, however, since the actual geometric shape cannot be detected by any method in situ, that is, as an aerosol. Comparative sizes are commonly used, such as the aerodynamic diameter for particles greater than 500 nm and the mobility diameter for particles less than 500 nm. The particles are therefore not evaluated according to their actual geometric size, but rather according to their characteristics in comparison to spherical particles with a density of one. Evaluating them according to their inertia (aerodynamic diameter) or their diffusion (mobility diameter) yields information on their “diameter.” Since particles from industrial combustion are usually smaller than 500 nm, and nearly, all mechanisms of deposition deep in the lungs are related to diffusion, the definition of the mobility diameter is preferred in this context [21-54], [21-55]. Another parameter frequently used for characterizing the shape is the fractal parameter that is usually far below 3 and frequently around 2, which indicates chain and flat structures. The size distribution (Figure 21.88) of the particles from engine combustion shows a logarithmically normal distribution with average values around 60–100 nm at the engine exit that scarcely changes to the end of the tailpipe. The majority of these solid particles lie within the invisible range (<400 nm). Visible smoke is caused by relatively few but very large agglomerations that primarily arise in older engines from combustion adjacent to the wall, not enough excess air, or deposits in the exhaust system and are then periodically blown out in what is referred to as the “store-and-release” process, a phenomenon which also occurs frequently in open filters [47].

vehicle 8

vehicle 3

7,E+13

vehicle 4

6,E+13

vehicle 6

vehicle 9 vehicle 10

vehicle 5

vehicle 11

5,E+13 4,E+13 3,E+13 2,E+13 1,E+13 1,E+09 10

100

Dp [nm]

1000

Figure 21.88 Size distribution of solid particles in modern diesel car engines [21-56] Dp = Mobility diameter (measuring process SMPS) [21-56].

790 | Internal Combustion Engine Handbook

6606_Book.indb 790

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

Soot particles are largely inert and odorless and are insoluble in water and organic solvents. If large amounts of ash, abraded material, or mineral particles arise, bimodal distributions frequently arise with a distinct second maximum around 20–30 nm. This relatively new knowledge together with the results of the medical impact research regarding penetration of the vascular system and nerve tracks and the meaning of the persistence of inert particles in the organism have led to new approaches for defining particles and determining their limit values. •• In addition to the EC mass, the concentration [particle count/cm3] classified by size in the range of 20–300 nm (mobility diameter) in accordance with the Swiss SNR 277205 standard [21-102] is taken as a basis for the characterization of particle filters. •• In effect since 1 January 2009, the Swiss Clean Air Act (LRV) has been considering the total particle count >23 mm for the type testing of construction machines with particle filters and restricting them to a value of 1012 particles/kWh. •• As part of the UNECE-PMP Program, the determination of the total number of solid particles in the size range 23 nm to 2.5 μ m was selected as a supplementary measuring method for type testing [21-89] (see Section 21.6.3.12). 21.6.3.2 Goals of Particle Filtration The goals of filtration must be oriented around relevance to health and technical feasibility. Relevant to health are particles that penetrate into the depths of the lungs, dwell there a long time, and cannot be fagocyted macrophages or dissolved in bodily fluids. Soot particles fulfill all these conditions. The maximum size of particles that are

deposited in the lung’s alveoli is approximately 10–20 nm; depending on the inspiration volume; due to their very high mobility, smaller particles are deposited in the upper respiratory tract and routed back to the throat by very efficient lung purification mechanisms (mucous layer, cilia). The smaller the particles, the easier they leave the alveoli and pass into the blood vessels and then through the blood and lymph into the entire organism, into the brain and through the placenta into the organism of the unborn child [21-57] (Figure 21.89). In addition, these small solid particles transport adsorbed toxic substances (such as carcinogenic polycyclic aromatic hydrocarbons PAHs) into the organism. The goal must then be to efficiently collect particles in the range of 10–500 nm—preferably with a filtration rate that increases as the particles grow smaller. The deposited substances remain bonded in the filter matrix under all conditions, and particles and adsorbed substances are not released during the regeneration process. The implementation of the “best available technology” (BAT), characterized by Figure 21.90, is generally considered a benchmark for the enforcement with respect to pollutants suspected of being carcinogens. Filters tested in accordance with the VERT suitability test of the Swiss Federal Agency for the Environment (BAFU) [21-59] exhibited rates of deposition for solid particles in the critical range of 99–99.9 %—a common result for modern particle filters. The concentration of the particles in undiluted pure gas downstream from such a filter lies approximately in the same range as today’s atmospheric concentrations and frequently lower. The nature of particle concentrations in environmental air lets us conclude that they are essentially determined by engine emissions into the atmosphere.

10 B

750 ml 2150 ml

deposition fraction

0.80

respiratory tract

0.60

nasal pharyngeal 0.40

pulmonary tracheal bronchial

0.20

0

0.01

0.1

1.0

10

median diameter-micrometers

100

Figure 21.89 Deposition of fine particles in the nose, bronchial tubes and alveoli [21-54].

Internal Combustion Engine Handbook | 791

6606_Book.indb 791

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Number of particles dN/dlogDp [1/cm3)

1.00E+07

Raw emission upstream of particle filter

1.00E+06

1.00E+05

Undiluted pure gas downstream of particle filter

Room air in the laboratory

1.00E+04

1.00E+03

10

100

Particle size [nm]

1000

Figure 21.90 Rate of deposition of a ceramic cell filter on a DI diesel engine utility vehicle [21-58] following field usage of more than 2000 operating hours.

21.6.3.3 Requirements for Filter Media and Technical Solutions The problems that diesel particle filters must overcome are quite challenging [21-60]: •• exhaust temperatures up to 750°C and temperature peaks during regeneration up to 1400°C •• high-thermal and thermomechanical stress during quick temperature changes •• danger of material damage from lubricating oil ash and additives •• ability to store large amounts of soot and ash •• low-pressure loss that only slightly affects the turbocharger and engine •• low-thermal mass (quick light off) •• rates of deposition greater than 99% for particles 10–500 nm •• no formation of additional pollutants—referred to as secondary emissions •• sound absorption when equipping at least identically to the original model with muffler •• insensitive to vehicle vibrations (installation close to the engine preferred)

Figure 21.91 Ceramic monolithic cell filter (Corning SAE 830181).

•• insensitive to damage when cleaning inert ash components In addition to all these requirements, the filter must be economical (for utility vehicles, less than $10/kW; for passenger cars, less than $5/kW), have a small installation volume, and have a service life equivalent to the engine life. Possible filter media are surface-rich structures made of high-temperature resistance materials such as monolithic porous ceramic structures in the form of cells (Wall flow) (Figure 21.91 and Figure 21.92) or as foams (Figure 21.93), high alloy porous metal sintered structures and metal foams (Figure. 21.94), and fiber structures such as fleece (Figure 21.96), yarn windings, or with textile bonds (knitted, woven) (Figure 21.95) using ceramic or metal fibers. The sizes, pore sizes, or fiber diameters decisive for deposition should range around 10 μ m to also achieve the desired separation during and immediately following the regeneration.

500 µm

Figure 21.92 Pore structure of a ceramic filter (Corning) [21-60].

792 | Internal Combustion Engine Handbook

6606_Book.indb 792

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

Figure 21.93 Ceramic foams as filter medium (Alusuisse) [21-103]. Figure 21.96 Fiber rope filter (BUCK), a pleated structure made from hightemperature fiber rope. Filter candle arranged in parallel [21-64].

made of cordierite by extrusion (NGK, Corning). Silicon carbide in various crystal structures (NOTOX, IBIDEN, LIQTEC) and additional ceramic materials appeared later. Intensive development of materials has led to thermoshockresistant structures. A wealth of experience has been gathered worldwide during the past two decades, in particular with the cordierite material. Multilayer wall structures are now being discussed, such as those with an additional ceramic membrane—highly porous, with tiny filter pores, extremely thin—sintered onto the actual filter wall, whereby the deposition can be further increased without changing the reduction in pressure loss. Figure 21.94 Filter made of porous sintered metal plates (SHW; HJS), combined and welded into a cell structure.

Raw gas

Pure gas

Closed Perforated support pipe ceramic twine (Nextel) rhombic wound

Figure 21.95 Filter candle (3M, MANN & HUMMEL), high-temperature ceramic yarn, designed as a thread spool and wound in a rhombic pattern on an internal porous sheet [21-61].

The following examples illustrate the many facets of this technology. Not all of these systems are currently implemented in practical applications: •• Ceramic monolithic cell filter A similar design to a cellular catalytic converter but with alternately sealed cells. This type of filter offers a large filter surface, a low structural volume (1–3 m2/l), a low counterpressure, and a high deposition rate at low gas speeds through the walls (a few cm/s). The filters were originally most often

•• Metal sintered filters With a general structure similar to ceramic monoliths, SHW and HJS developed a filter based on metal materials. The basic element is a thin-sintered plate made of metal powder with a wire mesh or expanded metal support structure (several tenth of a millimeter). These filters are relatively heavy in comparison to ceramics, but they are also very robust. Their heat conductivity is naturally very good. Meanwhile, metal sintered filters have appeared in bag-like structures made of filter plates having a lower weight (HJS, PUREM). •• Fiber-wound filters Yarns made of high-temperature fibers (Mullit) are wound in a special technique to create rhombic channel structures on a perforated substrate tube. Filter candles of this type were developed by 3 M and MANN+HUMMEL. •• Fiber-knit filters Ceramic yarns are processed into round-knit structures and shaped into deep structures by plating. The available macroscopic fiber surface typically reaches 200 m2/l, with the microscopic surface of the fiber itself being 100–200 m2/g. This filter type was developed by BUCK and is also offered in combination with wire knit having a catalytic coating. •• Fiber-woven filters High-temperature fibers also come as plaits and can be used for filtration when mounted on metal substrate structures. Such systems were developed by HUG and 3 M.

Internal Combustion Engine Handbook | 793

6606_Book.indb 793

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

•• Filter papers/filter felts/filter fleeces Paper filters that have a similar structure to intake air filters are used only when the exhaust temperature can be kept reliably low (such as when cooling the exhaust of underground engines). In any case, paper can be used up to temperatures of approximately 300°C (DONALDSON, PAAS, AHLSTROM). These papers are basically used in fiber filters with short fibers that are in a random pattern and whose structure is held by binders. Felts and ceramic fibers can be used for higher temperatures as has been used for a long time in industrial hot gas filtration; fleeces made of resistance-welded metal microfibers are also used (BEKAERT). In contrast to these surface-rich structures that are equipped with mechanical filters, flow-dynamic, electrostatic, and plasma methods have not become widely accepted. Exhaust scrubbers that were initially used frequently are now scarcely used (except in underground coal mining); in any case, they are unsuitable for filtering nanoparticles. This list is not exhaustive; there are numerous other technical solutions being developed. In addition to factors such as filter quality and pressure loss, designers are addressing size reduction, vehicle-tailored design, incorporation in the overall process, and linkage to other exhaust purification methods. 21.6.3.4 Deposition and Adhesion In general, there are three areas in a filter with physically different deposition mechanisms as shown in the following figure with fiber deposition as an example (Figures 21.97 and 21.98).

Soot particles

Trapping

Inertial separation

Filter fiber

Figure 21.98 Deposition mechanisms on a single fiber (3 M).

Larger particles are mostly captured by impaction from inertial force and a bit less from the trapping effects of flow near the wall. Figure 21.97 shows, however, that diffusion is practically the only effective process for submicron particles (nanoparticles), under discussion here. Blocking or sieving effects are practically never seen for such types of small particles. If we compare the deposition characteristics of the lung as illustrated in Figure 21.85, we find a similar deposition minimum of around 1 μ m where impaction becomes weak and the effect of diffusion starts to be seen. Such particles are largely exhaled in contrast to the many larger ones that are deposited and cleaned in the upper respiratory tract and the much smaller ones that tend to be deposited in the alveoli. With these small particles, the ratio of drag force to inertial force in the Stokes’ range (that does a good job describing the conditions in the filter) [21-53], is large enough for the particles to follow the flow lines around any obstacle, even around very fine microscopic structures such as filter fibers.

100

Diffusion

d⋅m⋅v Towing force ~ Inertial force d 3 ⋅ r ⋅ v 2

(21.22)

d = Particle diameter v = Vehicle speed

μ = Dynamic viscosity

Filter efficiency [%]

80

ρ = Density 60

40

ED = Diffusion E1 = Impaction (Inertial effect) ER = Trapped from the flow near the wall E0 = Total ED

These small particles can only be separated by diffusion. Diffusion takes time, however, which means that a sufficient filter depth L and low flow speed are required for successful deposition. The process can be best described with reference to a channel flow (Figure 21.99) where the channel diameter is 10 μ m like the typical pore size of such very fine filters, that is, around 100 times greater than a typical soot particle.

E0

E1 20

ER

L = Filter depth B = Channel width (pore size)

10–2

10–1

100

Particle diameter [microns]

Figure 21.97 Deposition effects in a filter medium as a function of particle size (3 M).

101

v = Flow speed cD = Diffusion speed In order for a particle from the channel center to reach the wall before leaving the channel, thus flow time through the

794 | Internal Combustion Engine Handbook

6606_Book.indb 794

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

Since the time required to flow through the filter wall of a ceramic cell filter generally ranges around 0.01 s, the diffusion path of a 100 nm particle is only 0.3 μ m; deposition in such a channel thus only exists for the particles directly near the wall—and not even at higher temperatures albeit the diffusion speed increases with temperature. This explains when virtually no particles can be separated in typical catalytic converter structures with permeable cells, even when they are very fine. Filter structures, therefore, have to be much better than the channel model to ensure the deposition of fine particles. Two approaches are taken to explain this: The porous walls of the wall flow filter can be described as a flow model, and the fiber depth filter can be described as a circulation model (Figure 21.101 and Figure 21.102).

V

B CD CD

L

Figure 21.99 Channel model for the diffusion deposition of particles.

L must be at least equal to (or shorter) than the v B . diffusion time from the center to the wall t2 = 2 ⋅ cD B⋅v The condition = const. ~ diffusion speed, a function L of the particle diameter d and the temperature T, results for the comparison of various geometries. For typical ceramic cell filters, the filter depth (wall thickness) is 0.5 mm, the pore size is 10 μ m, and the speed is a few cm/s. Fiber filters whose typical pore size is greater and that work at much faster speeds require a greater flow depth as indicated by this condition. Since small particles have higher diffusion speeds or higher mobility b, we can expect that smaller particles are filtered better in such structures. The terms “diffusion speed” and “mobility” are also equivalent in the sense of an Einsteinian relation: D = k ⋅T ⋅b



10 ... 25 pore layers

0.3 ... 0,5 mm

channel t1 =

10 ... 20 µm

Figure 21.101 Porous wall.

(21.23)

D = Diffusion coefficient k = Boltzmann constant 2 ... 8 mm

b = Mobility defined as b = v/F with F as the force exerted on the particle (such as electrical field strength, gravity, or impact forces by molecules) and with v as the resulting particle speed Calculated according to Hinds [21-54], this results in a diffusion speed for particles of different sizes at room temperature in accordance with Figure 21.100. The theoretical settling speed in [mm/h] is also indicated for reasons of clarity. Particle size

Diffusion speed μ m/s

Settling speed mm/h

  10 nm

260

0.2

  100 nm

30

3

1,000 nm

5.9

126

100 ... 500 fiber layers

T = Absolute temperature

Figure 21.102 Fiber bundle.

In porous walls, the flow occurs in channels that run from pore to pore. There are numerous diversions, dwell times in pore caverns, and new walls when channels branch. Diffusion

Figure 21.100 Numerical example according to Hinds [21-54].

Internal Combustion Engine Handbook | 795

6606_Book.indb 795

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

is substantially improved, and impaction is also enhanced. Such filters are distinguished by the superior deposition of coarser particles, but because of the channel-like nature of the passages, the rate of deposition tends to worsen for very small particles when the channel wall is thin, the pores are large and the flow speed is high—the residence time of the particle in the filter structure is therefore short. In the circulation model, new boundary layers are always being formed; that is, the flow channel is continuously being divided so that particles frequently pass directly next to walls where they can be deposited by diffusion. With such pure depth filters, one can expect an increasing rate of deposition of small particles, while the comparatively large pores pose the risk that large particles, particularly including agglomerates forming in the filter, can escape from the filter structure again.



p=

A 6 ⋅ p ⋅ z3

(21.27)

p = Adhesion pressure A = Constant z = Contact distance Small particles whose center of mass is very close to the surface adhere much better than larger particles. Furthermore, since the exposure to the flow of small particles that lie close to the boundary layer is slight, there is not much danger that submicron particles, especially once deposited on a surface, will be entrained by flow. However, other particles can collect on deposited particles so that large agglomerations can form in a filter as illustrated in Figure 21.103.

21.6.3.4.1 Definitions for the deposition rate The rate of deposition can be defined by the overall mass or the number of particles. In the second case, the high sensitivity of the number measuring technique allows the deposition rate to be recorded as a function of the particle size even for very small particle sizes. This provides what are referred to as the filter characteristics or dividing characteristics. This resolution is not possible when the definition according to particle mass PM is used since the detection limit of the gravimetric measuring process is not nearly sufficient for this in the range of the questionable particle sizes. Rate of deposition according to particle mass PMD: PMD =

PM before PF − PM after PF PM before PF

(21.24)

Rate of deposition according to particle count PCFR: (21.25)



PZ before PF − PZafter PF PCFR = = f (d) PZ before PF



Penetration = 1 − rate of deposition

(21.26)

When the sampling and measurement are designed so that only the solid mass is detected or only the solid particles are counted, both these definitions usually produce very similar values. This does not necessarily have to be the case. If a spectrum shifts toward very fine particles, the situation is better described by counting than by mass. Furthermore, in a particle size spectrum that is dominated by very fine particles, measuring by counting is a much more sensitive method [21-104]. For health reasons, the size, surface, and number are to be included in the measurement; the rate of deposition should be defined accordingly. 21.6.3.4.2 Retention of particles Along with the deposition, the reliable retention of particles in the filter matrix, thus the adhesion, is significant as a second component of the filtration process. Putting aside the effects of shape for once, which can largely be neglected, the adhesion force of a particle on a surface under the dry conditions of hot gas filtration is given, according to Van der Waals, by

Figure 21.103 Diesel soot, separated as the finest particles on a ceramic fiber with a 10 μ m diameter and a large agglomerate, which was formed in the filter [21-64].

Such agglomerations offer a large area of attack for the flow. They can then be released and leave the filter, which is typical behavior of large-pore depth filters or what are known as partial-flow filters [21-96]. These are also referred to as agglomerators. Since very fine particles adhere well to surfaces, only filter cakes and large soot and ash agglomerations are removed when trying to clean filters by blowing air through them. The very fine particles can be removed from the filter only by washing, which loosens the Van der Waals bonds, or by burning them out of the filter. 21.6.3.5 Regeneration and Periodic Cleaning The high deposition rate of solid particles causes them to quickly fill filters. Filters become filled with flammable components (soot) within a few hours, and they fill with inert solid particles (ash) after a few hundred to a few thousand hours. These figures can vary widely depending on the untreated emissions of the engine, mode of operation, lubricating oil consumption, fuel and lubricating oil properties, and filter characteristics. In every case, however, the flammable residue consisting of elementary carbon (EC) and organically bound

796 | Internal Combustion Engine Handbook

6606_Book.indb 796

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

carbon (OC) must be removed relatively frequently by combustion. This process is termed regeneration. For filters to be free of residue, regeneration should occur in such a manner that only CO2 and water form. This ideal is frequently not attained. The reasons for this, in addition to the CO/CO2 equilibrium, are effects that occur during the heating phase when the substances can leave the filter by evaporation, and low oxygen during heating that can lead to coking (pyrolysis) and as a result, nearly non-regenerable residues. The complex process of soot combustion that is not influenced by thermodynamic conditions but rather chiefly kinetic conditions can be described according to Lepperhoff [21-63] by a reaction-kinetic model based on the Arrhenius approach as follows: −E dM = k0 ⋅ M m ⋅ pOn 2 e RT dt



(21.28)

M = Relative soot mass

history, on the adsorption of hydrocarbons from the lubricating oil and fuel, and on newly formed substances. To make matters more difficult, the emission of hydrocarbons and CO should not be too high during regeneration, and the stress from the released heat from soot combustion should be controlled as best as possible to not overtax sensitive structures such as ceramic monolithic cell filters. Numerous regeneration processes have been developed to solve these tasks. They can be roughly divided into passive and active processes: •• “Active” if the regeneration is triggered by controlled or regulated interventions for either the purpose of supplying energy, increasing the temperature or increasing the oxygen content. •• “Passive” when catalytic measures lower the activation energy enough for the reaction to occur at the specified operating temperature.

T = Absolute temperature

In special cases (small engines, brief use, and operation in buildings), exchangeable or disposable filters are possible that are externally regenerated or disposed of after becoming filled. Active systems primarily include the following:

E = Activation energy

•• Diesel burner (Figures 21-104 and 21-105)

pO2 = Partial pressure of the oxygen R = Gas Constant

This relationship indicates the primary importance of temperature and sufficient oxygen. The activation energy E is around 140 kJ/mol without the use of regeneration aids. With the involvement of catalyst actions, this value can be reduced to a level of 80–90 kJ/mol. For soot to completely burn, temperatures above 600°C and an oxygen content above 7% are required; that is, the conditions necessary for heating up the filter system that can hardly be attained for many vehicle applications for prolonged periods and when, then only for very short periods of time. The combustion conditions can vary over a relatively wide range depending on the character of the soot and deposition Engine oil temperature Diesel fuel

Engine speed Exhaust gas temperature Pressure upstream of filter

Internal control system

Temperature downstream of filter

Burner power control Burner

Engine

Numerous subvariants are available. In addition to the actual full-flow burners (Deutz, Figure 21.104 (DEUTZ, Figure 21.104/HUG), which are regenerated under all operating conditions—a very difficult technical task—there are twin systems that are alternatingly regenerated under adjustable conditions (Eberspächer, Figure 21.105 Iveco) burners that are accessed during idling (HUG), and burners that heat the filter element from the clean air side (HJS). The task of operating a burner independently of the engine is technically much easier but also requires additional expenditure because the combustion air is provided by an electrically operated blower (HUSS, ERNST, PHYSITRON).

Filter monolith Pure exhaust

Mixing chamber Spark plug

Ignition spark generator

Exhaust gas

Burner air Compressor

alternatively Compressed air

(On-board system) Figure 21.104 Particle filter system with full-flow burner (Deutz).

Internal Combustion Engine Handbook | 797

6606_Book.indb 797

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Burner

Burner

Regeneration phase

Filter phase

Figure 21.105 Double filter system with throttle system (Eberspächer).

Externally operated burners are also used: for the regeneration of exchangeable filters on the one hand, and as hot air suppliers for the regeneration of permanently installed particle filters when the vehicle is idling, on the other hand. Catalytic supported combustion processes, referred to as “flameless” are particularly worth mentioning. They require a very fine distribution of the injected diesel fuel (possibly still in the engine) to enable combustion on the catalytic surfaces with the lowest possible residue (COMELA, PURITECH, GAT, DEUTZ). The use of highly inflammable fuels, which would need to be specially added, have not been considered up to now. •• Electrical heating: Numerous types of electrical systems have been developed: overall heating of the gas stream and thus, the filter, or specific heating of the filter matrix using ohmic heat in electrically conductive materials (SiC), and sequential heating systems in which one filter candle after the other [21-64] or one filter channel after the other [21-65] is heated to the regeneration temperature. If a sizable soot cake has formed, it may be sufficient to merely ignite this coating; the fire then burns through the entire soot cake from the released heat (HJS, EMINOX, PIRELLI); the requirement for this is FBC technology by the addition of fuel. The stored soot is then implemented with tiny metal oxide clusters and ignites thanks to its catalytic effect at temperatures as low as 300°C. The main problem with electrical heating is the limited availability of the electrical energy in vehicles. Electrical methods up to now have become popular only when the regeneration occurs when the engine is at a standstill, and the required electrical energy could be supplied from the outside (HUSS, ERNST, JOHNSON MATTHEY, ECS, DCL, CLEAIRE). The process can then be carried out slowly to spare the filter material. There is extensive experience with this method, especially with off-road vehicles and underground applications.

•• Regeneration energy from engine combustion: Active systems include those in which the required energy is taken from the engine by special strategies during the regeneration periods. The normal measures are late injection, secondary injection, throttling, and exhaust recirculation. These strategies can raise the exhaust temperature by 200–300°C, which in many cases, especially in combination with catalytic measures, is sufficient to regenerate the filter. All of these approaches raise fuel consumption, which, however, is not that problematic since the regeneration phases are short in relation to the operating time between regenerations, typically 1–3%. Measures of this type, however, are usually possible only with original equipment, preferably in engines with electronic fuel injection systems. As an alternative, throttling the gas flow is also a process which can be used for retrofitting and is becoming increasingly popular. Caution must be exercised for fuel injection inside of the engine due to the risk of lubricant oil dilution. The passive regeneration aids are just as diverse. They use catalysts, which can be roughly divided into those which support the reaction of the soot with oxygen and others which promote the reaction with NO2. Combinations can, of course, also be considered. Catalytically active substances are used either by adding fuel or by coating the filter surfaces. •• Regeneration additive (FBC = fuelborne catalysts) [21-105] are substances which are added to the fuel in very low concentrations (10–20 ppm), primarily in organometallic form, which can reduce the catalytic effect to around 300°C. Examples of such substances are cerium, iron, copper, and strontium. The end products of these additives (oxides) surface in the exhaust as extremely small ash particles (around 20 nm) [21-66]; they may, therefore, be used only with corresponding particle filters [21-59]. The advantage is that the additives substantially lower raw soot emissions and, hence, relieve the filter. Additional advantages can be found in the fact that they do not age, that their dosing can be easily adapted to the soot emissions and that the combustion, due to the (usually) high oxygen concentrations in the diesel exhaust and the preferable contact of the catalysts with the soot, progresses very rapidly and completely. These properties distinguish FBC technology from catalytic wall coatings and the really slow NO2 regeneration. FBC generally changes the gaseous composition of the exhaust only a little. It can be observed with some additives that the engine combustion improves and deposits inside of the engine (piston ring package) are reduced. A pronounced reduction of the fuel consumption can even be observed in certain cases. New developments, however, exhibit clear DeNOx effects and an almost complete elimination of the NO2 engine emissions, even in the low idling range where diesel engines often exhibit high NO2 concentrations [21-97]. •• Catalytic coating (Figure 21.106): As with additives, the soot ignition temperaturecan be similarly lowered by coating the filter with transition metals [21-63]. A requirement for this is a very high specific surface (>100 m2/g) and thus a very fine distribution of the active centers.

798 | Internal Combustion Engine Handbook

6606_Book.indb 798

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

Inlet channel plug Exhaust gas, soot loading

• • • • •

Catalytic coating, outlet side

Trapping of soot carbon Oxidation of soot carbon Low temperature regeneration Oxidation of hydrocarbons from unburned fuel and tube oil Principle patented, lowest sulphur content needed (5–50 ppm) P1 catalyst

Outlet channel plug

Catalytic coating, inlet side

Exhaust

Filtered exhaust gas

Porous filter wall

2NO2

2NO + O2

Scource: Johnson Mathey

Figure 21.106 Schematic illustration of a catalytically coated soot filter [Engelhard].

The catalytically supported NO2 reactions are used at considerably lower temperatures, as they were first put into practice for what we refer to as the CRT System [21-67], Figure 21.107, (CRT = Continuously Regenerating Trap based on a patent by Johnson Matthey). The property of a precious metalcoated oxidation catalytic converter connected upstream of the particle filter was used to generate more NO2 from the NO in the engine exhaust. NO2 is not stable at these temperatures, however. The reverse process thus occurs in the downstream particle filter and the released oxygen radical oxidizes the carbon at exhaust gas temperatures starting at 230°C.

CO2 + 2NO CO2

What all of these regeneration processes that are characterized by precious metal coating have in common is that they produce more NO2 and also exhibit a considerable slip in this harmful gas. Various combinations of the regeneration process were implemented. The process developed by Peugeot, Figure 21.108, for usage in passenger cars serves as an example: cerium oxide (later replaced by iron oxide in metal concentrations of less than 10 ppm) used as a fuel additive to reduce the soot ignition temperature by around 200°C—this is not nearly enough however for passenger cars. In addition to this, the exhaust gas temperature was increased by around 100°C by postinjection, with the fuel not completely combusted being

Pressure loss sensor

Software: – Manager – Regeneration strategy – Dignosis – EOBD

SiC DPF

engine ECU

Engine Common rail

Gage

Hydrocarbo Fuel delivery

C + O2

For this to work, however, sulfur-free fuelmust be used to prevent the sulfating reaction (SO2 → SO3) which as the preferred reaction, prevents NO2 conversion. CRT as a passive reaction has caused an actual breakthrough in filter technology for retrofitting, above all in public transportation buses. The process was then varied and launched on the market in many different versions.

Various functional models are available where the catalyst is coated; on the untreated gas side that primarily enhances soot burning, and on the clean gas side where precious metals can be used to reoxidize CO and HC.

CAT

C + 2NO2

Figure 21.107 CRT filter system [Johnson Matthey].

Whereas massive soot deposits can be burned when additives are used (with the danger of generating high-temperature peaks), coated filters avoid the formation of thick soot cakes since this can substantially restrict the effectiveness of the catalysts applied to the wall.

Temperature sensors

Soot filter

Fuel pump

Fuel tank

Dosing system ECU

Additive tank

Figure 21.108 Diagram of the Peugeot soot filter system for passenger cars [21-68].

Internal Combustion Engine Handbook | 799

6606_Book.indb 799

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

800

700 600

Temperature [°C]

implemented in a precatalytic converter increases the temperature further. Exhaust recirculation that lowers the combustion temperature is turned off in the regeneration phase, and the vehicle electrical system is tapped by electrical consumers. All of these elements are required to trigger regeneration even under adverse conditions, that is, at sustained lower load which triggers regeneration when the soot loading limit is reached. The combination of regeneration aids and their processrelated support by electronic means are virtually boundless. An attempt was thus made to use the energy freed by heat exchange during regeneration by means of recuperation [21-105], to form ignition sources, and to inject substances rich in oxygen such as acetyl acetone. The conditions under which the soot can burn with adequate oxygen content are summarized in Figure 21.109. At longer intervals, the filter must be cleaned of inert substances, which form from lubricant oil additive substances, engine abrasion, fuel additives, and mineral substances supplied to the engine with its air intake. These substances particularly include metal oxides from wear-preventing additives of the lubricant oil, such as zinc oxide and calcium, as the main component of anticorrosion additives that can form gypsum with the sulfur from the fuel or lubricating oil. The inert substances can become deposited in the particle filter and plug its pores. Today’s filter is cleaned of these inert substances approximately every 2000 h (100,000 km). The filter must be removed for this. The typical washing process used previously did not prove effective since the fireproof materials of the mats used for packing the filter (canning), primarily ceramic fibers, can be sensitive to moisture. In the machine cleaning process, the filters are first cleaned of all burnable substances at high temperatures, and then, the inert inclusions are blown out by air using a pressure surge process. These ash substances need to be disposed of in an environmentally friendly manner.

500

400 300 200

100 0

CRT DPC Satacen® Eolys® V2O5 PT catalyst Additive

EC

Figure 21.109 Equilibrium temperature during filter generation with various catalytic tools [21-69].

In addition to clogging by the inert metal oxides from the lubricating oil (oil ash), the disadvantageous effects of common lubricating oils on filters include filter damage, for example, from the formation of glass phases. This has led to demands for new lubricating oils that have a low ash content, a low sulfur content, and less phosphorous and alkaline earths for the sake of filters. These are referred to as low SAPS-lubricating oils (low sulfate ash, sulfur, and phosphorous content). The goal is a significant reduction in the emissions of lubricating oil ash to a maximum value of 0.5 mg/kWh [21-70]. A successful selection of the regeneration method largely depends on the knowledge of the operating behavior of an engine, that is, the collective load that arises in typical use. The most important parameter in addition to the oxygen content is the temperature. Figures 21.110 and 21.111 illustrate the problems. 120

4000 3500

100

Frequency

3000 80

2500

60

2000 1500

40

1000 20

500

0

0 75 100 125 150 175 200 225 250 275 300 325 350 375 400 425 450 475 500 525 550 575 600 625 650

Temperature Class

Figure 21.110 Cumulative distribution of the exhaust gas temperature for the bus, 268 kW [21-71].

800 | Internal Combustion Engine Handbook

6606_Book.indb 800

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

30

Frequency

25 20 15 10 5 0 100

200

300

Tem pe

400 ratu re [ °C]

500 600

253

217

Figure 21.110 (below) in which the dwell times are cumulative within certain temperature windows illustrates an apparently sufficient level for several regeneration processes. We note in Figure 21.111 that these temperature episodes are very short; only a filter system with a short response time can use these short phases of sufficient temperature to start regeneration, which indicates the necessity of a lower heating capacity and good heat insulation. 21.6.3.6 Emissions During the Regenerations and Secondary Emissions The loaded particle filter must be understood as a chemical reactor which is equipped with a very large surface area, thus predestined for catalytic processes, and which can operate in broad temperature ranges with pronounced adsorption/ desorption cycles. The formation of new substances can also not be ruled out. The wide variety of reactants from engine combustion enables chemical reactions that can lead to the emission of critical concentrations of toxic substances. By coating or incorporating catalytically active substances, the formation of such substances can be accelerated, and their concentration can increase considerably. There are additional emission risks from the store and release behavior of such systems, as well as reactions that can arise when the soot is burning. Generally speaking, there are three groups of processes: •• Emission peaks of HC and CO can arise during regeneration: HC when adsorbed hydrocarbons evaporate in the heating phase and CO when regeneration is very quick or occurs with low oxygen. •• “Store and release” phenomena are always possible in adsorbing systems with fluctuating temperatures. Pronounced sulfate cycles can be observed when fuels with high sulfur contents are used [21-98]. When large-pore filters and those referred to as partial flow filters are used, this “store and release” behavior becomes a typical feature, which simulates deposition in the “store” phase [21-96], a phenomenon causing

145

181

109

m -Ti

ell Dw

73

37

c.]

Se

e[

1

Figure 21.111 Cumulative distribution of the exhaust gas temperature for the bus according to time episodes, 268 kW [21-71].

rmerly typical testing processes with relatively short test cycles and upstream conditioning to produce completely false particle reduction rates. •• Released substances that did not previously exist in the system are understood as secondary emissions, that is, that are formed in the particle filter. Such reactions arise primarily with catalytic support, and the deposition of lubricating oil ash can produce noticeable catalytic effects. When precious metal coatings are used, a strong sulfate reaction is observed (SO2 → SO3), as well as a substantial shift in the NO/NO2 equilibrium. In conjunction with copper additives, a massive increase in the emission of dioxins and furans of several orders of magnitude has been observed [21-72]. A shift in the PAK spectrum is also conceivable, nitro PAK can be formed, aldehyde. It is therefore necessary to note secondary emissions during the type testing of catalytically coated or catalytically supported systems. In terms of the legally limited gaseous pollutants CO, HC und NOx formed in the engine, the particle filter system generally does not change anything unless catalytically supported processes are involved; a strong reduction in CO and HC by around 90–95% and an increase in NO2 emissions are observed when precious metals are used, but no change in the total concentration of nitrous oxides. The reduction in CO and HC is generally not so great when coatings with transition metals are used; almost no sulfate reactions are noticed, however, and NO2 can be reduced. A generally valuable feature of particle filter systems is that polycyclic aromatic hydrocarbons are usually reduced to the same extent as the rate of deposition of solid particles. This can be explained by the fact that only polycyclic aromatic hydrocarbons are adsorbed during the soot formation phase on the surface-rich structure, they remain fixed to the structure, and they are converted into the end products CO2 and H2O during regeneration. This process is also demonstrated by the time-of-flight analytics [21-66].

Internal Combustion Engine Handbook | 801

6606_Book.indb 801

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.6.3.7 Pressure Loss Because of the unavoidable pressure loss when the exhaust flows through these fine-pore structures, particle filters always negatively influence the engine (supercharged engines more than naturally aspirated engines): Expulsion work increases as exhaust retention rises, combustion is ultimately influenced by increasing counterpressure, and component temperatures rise. The pressure loss of the particle filters continues to increase as the filter becomes filled with soot and ash deposits. Interestingly, surface filters become progressively clogged from the formation of filter cake until completely blocked as the rate of deposition increases. With deep-bed filters, this process is revered according to the fiber growth model [21-73]; that is, a certain threshold load is not exceeded [21-64]. The rate of deposition decreases at the same time. The pressure loss in the fine-pore filter element and soot cake follows the laminar law since the Reynolds numbers in reference to the pore size are less than 1. The pressure loss is typically indicated as follows: For the Jodeit fiber filter [21-73], 1 1− e⎞ v⋅m⋅ 2 ∆p = K1 ⋅ L ⋅ ⎛⎜ ⎝ e ⎠⎟ d



For pore structures according to Ergug and Orning [21-74], ∆p = K 2 ⋅ L ⋅

e3

2

⎛ Op ⎞ ⋅ m ⋅u⋅⎜ ⎟ ⎝ Vp ⎠

(21.30)

Op = Pore surface Vp = Pore volume In addition, there is a significant turbulent component of flows in housing and filter feed channels that are proportional to ρ v2. The pressure loss of new filters at a nominal engine load is usually 20–40 mbar, that is, similar to mufflers in commercial vehicles that are usually replaced by a filter. The conventionally accepted maximum permissible pressure loss of the fully loaded filter is 200 mbar (in reference to the nominal speed and load). This pressure loss has a negative effect on fuel consumption and performance arising from the expulsion work, where a proportional counterpressure of around 300 mbar (at full load and nominal speed) can be assumed. The following applies for the nonsupercharged engine:

(21.29)

∆p ∆b = pe + pr b

L = Filter depth



ε = Porosity = cavity content Pore volume =    Filter volume v = Approach velocity

∆p = Filter pressure loss

(21.31)

b = Fuel consumption pe = Effective mean pressure pr = Mean friction pressure

μ = Dynamic viscosity

In a commercial vehicle operated under a relatively high load, the pressure loss accordingly reduces fuel consumption by approximately 2%. In vehicles with lower load factors, such

d = Fiber diameter

ρ = Density of the flow medium

232

( 1 − e )2

Specific fuel consumption as a function of the exhaust gas back pressure (psurroundings = 957 mbar)

230 228

Specific fuel consumption [g/kWh]

226 224 222 220 218 216 214 212

N = 1380 rpm N = 1750 rpm N = 2120 rpm N = 2400 rpm

210 208 206 204 202 200

0

50

100

150

200

250

300

350

400

Exhaust gas back pressure relative to the surroundings [mbar]

450

500

Figure 21.112 Modeling of the deterioration of the fuel consumption for four full-load speeds resulting from an increase in the exhaust gas counterpressure for a supercharged engine [21-75].

802 | Internal Combustion Engine Handbook

6606_Book.indb 802

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

as passenger cars, the influence is greater, around 5%. If the pressure loss increases beyond this extent, then combustion and supercharging are negatively affected starting at around 400 mbar so that stronger effects occur in a non-linear manner as Figure 21.112 illustrates in the simulation of a supercharged modern DI utility vehicle engine. The effects in this graphic do not consider that only the difference with respect to mufflers may actually be evaluated when this is replaced by the filter as is typical for retrofitting. Mufflers in commercial vehicles are designed with approximately 60 mbar pressure loss; in passenger cars, one frequently finds values above 200 mbar at maximum throughput. A strong effect of the filter pressure loss on the supercharged engine is caused on the one hand by the fact that the pressure loss of the filter for the engine connected downstream of the turbocharger turbine is increased to the extent of the expansion ratio, and on the other hand by the reduction in the relief enthalpy and consequentially by a reduction in the charging pressure and overall efficiency. The limit of the filter’s permissible counterpressure with two-stroke engines and four-stroke engines with high valve overlap is considerably lower than with typical four-stroke engines. Particular care is required with engines having unregulated exhaust-gas recirculation, since an increase in the counterpressure can very quickly lead to an increase in the exhaust gas recirculation, thus to an increase in the particle emissions with a simultaneous reduction in the exhaust gas temperature, whereby the ratios for operating the filter can increasingly worsen. 21.6.3.8 Installation Area and System Integration The installation area required for a particle filter system is approximately 4 to 8 times that of the engine charge volume. It is not always easy to incorporate in the exhaust system such a component whose possible shapes are limited, especially when retrofitting. Certain manufacturers have nevertheless already managed to structurally adapt their systems to the extent that they offer filters in the replacement dimensions of the mufflers, and these even for CRT systems containing both a catalytic converter and a filter element. The problem is easier to solve with original equipment. There are many interesting options for structural and functional system integration: •• Optimum linkage of the filtration + catalysis + noise suppression. •• Combining the functions particle filtration + denitrification—a future option. •• Placing the filter on the high-pressure side before the turbocharger to substantially enhance regeneration and reduce the effect of the pressure loss on the engine in relation to the turbocharger expansion gradient [21-76]. •• When a particle filter is used, there is no longer a need for a trade-off between NOx and particle mass (PM) when harmonizing the engine’s features. The engine can, therefore, be designed for lower NOx emissions, taking into account slightly higher particle emissions.

•• The exhaust after the filter is so particle free that it can even be fed to the intake before the turbocharger. There is no need to fear engine wear and turbocharger compressor soiling from exhaust recirculation. •• The engine management system can briefly increase the exhaust temperature enough to regenerate the filter. The increased particle formation is not externally visible. •• Another facet of system integration is the use of fuels and lubricating oils that are tailored to operate with particle filters. The filtration of the intake air is improved to such a degree that the very fine-pore particle filters are not clogged by inducted mineral dust. 21.6.3.9 Damage Mechanisms/Experience Monolithic ceramic cell filters can easily become brittle. The low mechanical strength of these porous materials and their low thermal conductivity made the structures originally used susceptible to thermomechanical stresses such as those typically occurring during regeneration processes. Since the limit temperature was already around 1400°C, a great deal of damage occurred due to uncontrolled regeneration. Several paths have been pursued to overcome this problem: •• further development of the materials and new development of stronger and more temperature resistant ceramic materials such as porous SiC with a bending strength of around 40 MPa •• the development of ceramic structures with lower crack expansion tendencies (segmented filters) •• optimizing the geometric design of the filter to minimize heat stress •• use of metallic materials as sintered plates or fiber fabric •• measures to control regeneration to limit the temperature gradient and peak temperature. We now have experience with these approaches that have been incorporated into the exhaust systems in large numbers for the retrofitting of buses and construction machinery with annual damage rates below 1%, mileage exceeding 800,000 km in vehicles, and operating times above 20,000 h in production machines. Not to be underestimated is the load on fragile ceramic components from vibrations that can arise especially with underhood installation. The filter elements must be isolated with a ceramic mat because of the large differences in expansion and must be installed under initial stress in a metal housing. Damage is unavoidable if the initial stress slackens. Damage of this type is less of a problem with other filter structures, especially fiber filters or filters made of metal components, such as sintered filters and metal fleece filters. Another class of damage that chiefly arises with monolithic ceramics can be triggered by lubricating oil ash or additive ash: the end product of these substances is, after all, metal oxides such as the ceramic itself. There are many opportunities for these substances to form phases with the ceramic, and this usually causes weakening. Damage of this type is also possible with fibers; however, the fiber structures are less

Internal Combustion Engine Handbook | 803

6606_Book.indb 803

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

sensitive than a monolithic structure because of their high elasticity and redundancy. When additives are used, filter damage causes these substances to be released into the atmosphere. This must be avoided; filter damage must therefore be recognized by the OBD, and the dosing of the additive immediately stopped. For a long time, people feared that fuel additives would negatively influence engine combustion and wear. When added, these substances are generally metalorganic in nature, which allows them to mix on a molecular basis. In fact, no negative effects on the engine wear could be demonstrated up to now; on the contrary, it must be pointed out that the engines have less deposition in the ring area, which should actually have a more positive effect on the wear [21-77]. 21.6.3.10 Quality Criteria In addition to economic criteria, such as investment costs, infrastructure costs, and service, the following criteria must be included in the evaluation of a filter system: •• number-based deposition rate PCFR, or rather Penetration P = 1 − PCFR in the entire relevant particle size range •• pressure loss ∆p, better relative to the average indicated pressure of the engine ∆p/pi for the load case which best characterizes the application •• volumetric throughput in proportion to the size, that is, the spatial velocity V/B = S [1/s] •• thermal response time t1: the time that passes after a sufficient rise in the exhaust temperature until the filter starts to regenerate and the pressure loss decreases •• storage time for inert material until the filter must be cleaned: t2. These quality parameters can be summarized in a multicomponent evaluation as a single filter parameter: 20 t [S] t [h] ⎞ ⎛ P[−] ∆p/pi <1 1/5 ⎜ + + + 1 + 2 ⎝ 0.01 0.02 S[1/s] 100 2000 ⎟⎠



If this value is much greater than one, then at least one or more of the important parameters lie outside of the values that are now attainable. 21.6.3.11 Performance Test, Type Test, OBD, Field Control •• Performance test [21-99] One can generally assume that a filter is capable of collecting a certain percentage of solid particles of a particular size independent of the chemical composition of these particles [21-78], and with particles smaller than 200 nm, not even the density plays a role if their size is characterized by the mobility diameter. The throughput, temperature, and load must be included as boundary conditions since impaction rises as the throughput increases; however, diffusion deposition decreases, the load thus has a varying effect on the deposition rate, and the temperature can influence the diffusion constant and adhesion conditions. It is then sufficient to measure a filter at a maximum spatial velocity and maximum temperature when new (worst case)

to get a good idea of its expected minimum rate of deposition in practical applications. The filter is characterized by a single value for the rate of deposition per particle size class (separation curve); by multiplying this rate of deposition with solid particles in untreated emissions, we can determine the solid particles in the clean emissions of any engine. This test can generally be performed on a filter testing machine with a test aerosol, as this is also handled with many industrial filter applications (DIN 24184, 24185). Priority must be given to the measurement on a representative diesel engine, primarily due to the question of the adhesion condition and the formation of agglomerate in the filter. Transient processes such as the free acceleration of an engine from low idling to high idling, to cite an extreme example, do not produce any new information as expected since the very low speeds (extraordinarily small Reynolds and Mach numbers) in the filter medium do not yield any flow-dynamic effects. A filter can then be characterized by a simple test in reference to its basic function, namely, the retention of solid particles. If, in addition to filtration, the filter must fulfill other functions such as catalysis, it must undergo additional tests. •• Type test of a vehicle with filter The type test is done in driving cycles under transient conditions. The emissions factor (g/km, g/kWh, particle count/ km) of the corresponding vehicle is measured as opposed to the rate of deposition. For regenerating systems, the type test must also be run during regeneration, and this result weighted over time must be included in the overall result (ECE/324/Add.82/Ref. 2/Amend.1). Since the particle count in pure gas downstream of modern particle filters can be magnitudes lower than the detection limit of gravimetric methods, particularly with regard to small particular sizes and since the effects on human health correlate much more strongly with the particle count than with their mass, supplementary test procedures are being sought. A process was developed as part of the UNECE-PMP program and proposed for future type testing, according to which the total number of solid particles ranging in size from 23 nm–2.5 μ m is integrated and recorded during the entire driving cycle and an average limit value defined for the passenger car as [particle count / km] and for the commercial vehicle as [particle count/kWh] (www.unece.org/trans/doc/2003/ wp29grpe) [21-100]. •• On-board diagnostics (OBD) As essential input for the system control, the clogging of the filter with respect to pressure loss is continuously measured. This measurement is used to monitor three critical pressure levels: an upper limit for initiating the regeneration in active systems, a maximum level which draws attention to the necessity of cleaning inert dusts from the filter, and a lower level which signals damage to the filter. It has meanwhile become common practice with retrofits to store such measurements over several weeks time and possibly to also supplement them with temperature and speed measurements in order

804 | Internal Combustion Engine Handbook

6606_Book.indb 804

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

to monitor the interface with respect to the particle filter in cases of engine damage. •• Period field inspections With utility vehicle engines up to the emission generation Euro III, control can be provided by the free acceleration method measuring opacity to yield reliable information on the proper functioning of the system. The sensitivity of the simple opacimetric methods is no longer sufficient for engines with much lower untreated emissions and carefully controlled smoke emission. This applies for engines without particle filters. With particle filters, the opacity of the purified exhaust low in a way that makes measurements with typical opacimeters no longer meaningful. Sufficiently sensitive measuring processes are well known but still not approved for robust field applications [21-79]. 21.6.3.12 Particle Measuring Technology Since the harmful effects of very fine combustion-generated particles is related to their size, surface area, concentration, and substance, appropriate sampling and measuring techniques must be chosen to reliably characterize these quantities. The measuring technology must additionally be suited to record the particle content in pure gas downstream of particle filters with similar precision and resolution in terms of time even during transient driving cycles as is currently already state of the art for the measurement of gaseous substances. The gravimetric process for the determination of the total particle mass PM, as it is currently used for type testing worldwide, does not meet these criteria, and especially not when the volume of the finest particles under 50 nm would need to be adequately considered as well, and at most even classified by size [21-82]. Conventional in situ measuring technology that has been used for quite a while in aerosol physics is much more sensitive and offers highly developed procedures for •• separating according to phase (solid or liquid) during sampling •• classifying according to mobility diameter or aerodynamic diameter •• measuring the pollutant concentration for each size class as particle number or active surface The methods are described in [21-53], [21-54], [21-83], [21-84]. The SMPS (scanning mobility particle sizer) method is frequently used today for measuring in the engine exhaust that can be combined with upstream thermodesorbers or thermal dilators or selectively separating the volatile substances from solid particles. The measuring sequence is as follows: •• sampling probe, with isokinetic extraction not being mandatory. •• heated line made of electrically conductive material for reducing thermophoretic effects, subsequent condensation, and electrostatic deposition. •• high dilution directly after sampling to prevent changes in the aerosol from agglomeration and recondensation.

•• generation of a statically uniform electrical charge state •• classification into a differential mobility analyzer (DMA) (Figure 21.113). •• count of the particles by size class with a condensation nucleus counter (CPC) (not shown). Sheath-air in aerosol in 85

Kr-charger

monodisperse excess air out aerosol out Figure 21.113 Differential mobility analyzer DMA.

The electrically charged particles in the DMA drift along a trajectory that is determined by the ratio of aerodynamic drag to electrical field force in the annular area to the center electrode. At a specific throughput and set field strength, only a specific, very narrow-band class reaches the exit slot. By varying the voltage, approximately 60 size classes can be scanned in 1–3 minutes. For systemic reasons, this method is not suitable for dynamic measurements. The option remains of measuring one size class after the other or connecting several devices with set size classes [21-85] or of combining this type of parallel measurement into a single device type which is already offered by manufacturers such as TSI, GRIMM and CAMBUSTION [21-86]. Aerosol physics offers other measuring principles. A particularly interesting alternative for the size classification is the electrical diffusion battery, which has a much flatter separation curve that the DMA, however, but instead delivers an online signal and can provide information on the particle composition as a function of the type of electrical charge (Figure 21.114). A distinction must be made here between ion charging by classical corona discharging, which evenly charges all particles of the same size, and photoelectric charging by energy-rich UV radiation, which due to differing photoelectric properties of the particle surfaces, enables differences between soot particles, ash particles, and liquid droplets to be detected.

Internal Combustion Engine Handbook | 805

6606_Book.indb 805

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

Amplifier Insulator

Stage 1

Filter

Screens

Stage 2

Stage N

Backup-Filter

The particles are deposited in capture grids; the finest particles are deposited in step 1 with the lowest grid count, where they discharge their electrical charge. The grid geometry physically determines the separation characteristic, that is, the mobility diameter range of particles that are deposited in a specific grid geometry. The number of electrical elementary charges is determined by the particle size; it is a measure for the active surface of the particles. The measured current per step is therefore a transiently available signal for the overall surface of the deposited particles, and since the average diameter per step is known and the average number of elementary charges is known, the number of deposited particles per step can be derived. Another measuring procedure that is frequently used is the electrical low-pressure impactor (ELPI) shown in Figure 21.115. This device also provides online information and is suitable for measuring during dynamic driving cycles. However, the

Figure 21.114 Electrical diffusion battery.

classification, as in every impactor, is made according to aerodynamic diameters. Compared with SMPS, ELPI offers a lower resolution for the smallest particle size, but the measuring range for large particles is greater. Following careful analysis of all available methods [21-88], the efforts to introduce a measuring technology for type testing which focuses on the numeric concentration as the most important parameter for health effects, have led, as part of the UNECE-PMP project [21-87], to a method of measurement without size classification. The goal is to record the total number of all solid particles (per km or per kWh or per m3) in the size range of 23 nm to 2.5 μ m. The solid character is defined by the fact that the sampling device must be able to evaporate volatile substances up to C40 to at least 99% when heated to 300°C and to keep them in the gas phase, but at the same time without uncontrolledly losing more than 10 % of the solid particles (Figure 21.116).

High voltage source

Corona charger

Ion trap voltage source +5 kV

Electrometers

400 V

Impactor with Insulators and contact needles

A/D

Computer and control electronics RS-323 serial External computer, data logging, user interface

Vacuum pump Figure 21.115 Electrical low-voltage impactor ELPI (DEKATI).

806 | Internal Combustion Engine Handbook

6606_Book.indb 806

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

it regenerates everything it has separated and not only stores it temporarily [21-98]. Thus, the usage of a wall flow filter in the proportion 667/1 offers a much greater benefit for the environment than that shown by the deposition rate alone.

CVS tunnel

54321 CPC Dilution

Heating

Counting

Figure 21.116 PMP measuring setup for recording the total numeric concentration for the type testing of diesel passenger cars.

The sample for the passenger car was taken from the CVS tunnel. Samples for the truck or other engines can also be taken directly from the exhaust gas system. The exhaust gas sample is immediately diluted and then heated. Condensate which has formed in the CVS tunnel is evaporated in this process [21-89] and as a result of the dilution at a ratio of around 1:100, partial pressure is simultaneously reduced so strongly that recondensation is no longer possible. Once the gas has been cooled (to protect the instrument), which can include additional dilution, the particles are then counted in the condensation nucleus counter, a commonly used instrument. The particle size is limited to a maximum of 2.5 μ m by an impactor not shown here and to a minimum of 23 nm by the counting characteristics of the condensate nucleus counter CPC. Along with the rotating diluting agents shown here, other processes such as ejector diluting agents are also allowed for dilution. The dilution rate, the quality of the evaporation, the total dilution loss and the counting quality must be tested to calibrate the measuring sequence. Standard aerosols and even the CAST (Combustion Aerosol Standard) can be used for these calibration tasks which provide particles of a certain size distribution and concentration that are very similar to particles from the diesel engine combustion [21-90]. 21.6.3.13 Penetration or Deposition Rate Modern particle filters are generally evaluated by their deposition rate or efficiency. A filter with a 99.9% deposition rate, a common value for modern wall flow filters, is evaluated at around 3× better than an “open” filter, which may separate 33.3%. The deposition rate is not relevant for the environment, but rather the penetration, that is, the question of how many particles reach the breathing air. Penetration = 1 − deposition rate (more precisely 1 − deposition rate × regeneration rate [21-98]) 1 out of 1000 particles reaches the atmosphere through a wall flow filter, as opposed to 667 through an open filter—if

21.6.3.14 Global Warming due to Soot Particles The rapid increase in the percentage of diesel engines in the European passenger car fleet is primarily caused by the fact that due to their improved thermodynamic efficiency and resulting lower CO2 emissions, diesel vehicles contribute less to the greenhouse gases than gasoline engines and thus help to achieve the goal of the automobile industry of reducing CO2 emissions. This does not consider the fact that soot particles in the atmosphere also have a very high potential for heating the atmosphere by absorbing sunlight and transmitting it to their surroundings in the infrared range, that is, heating up the atmosphere. The GWP of BC = black carbon particles finely distributed in the atmosphere is, according to [21-107], around 600,000 times higher for each kilogram of substance than the potential of CO2 for increasing the atmosphere by reducing the earth’s radiation. Diesel engines in their modern design must be equipped with particle filters having a deposition rate of at least 99% to make them equal to gasoline engines with regard to their overall greenhouse gas potential. Without filters or with filters below this value, diesel engines contribute more to global warming than gasoline engines, even when their efficiency is further increased. 21.6.3.15 Cost/Benefit Consideration for the Retrofitting of Particle Filters Retrofitting measures for reducing emissions have no direct economical justification; on the contrary, they require considerable investment and operating costs. It is therefore common practice to express the ratio of these costs to environmental benefits, that is, the invested money in proportion to the reduction of particle emissions as €/kg soot. When retrofitting a passenger car with a remaining mileage of 100,000 km, the emission of 1.2 kg soot can be prevented when the average emission is 0.04 g/km and the deposition rate of the partialflow filter 30%. If the investment costs for the filter together with its installation and operation are set at 1000 €, then the cost/benefit efficiency is 833 € per kilogram soot. The cost/benefit consideration for the utility vehicle with a full-flow filter provides much more favorable figures: The emission of 0.2 g/kWh, for example, is completely eliminated; with a remaining service life of 5000 operating hours and an average output of 100 kW, this corresponds to the prevention of the emission of 100 kg of soot. The retrofit costs and operating costs of 8.000 € result in a cost-benefit factor of 80 €/kg. This factor for older vehicles in the off-road area is calculated at 30–50 €/kg [21-106]. Since the external health damage caused by 1 kg of fine particle soot is calculated at around 300 € [21-108], this is a considerable economic benefit. In terms of maximum environmental benefit, it is therefore recommended that high-efficiency filters be retrofitted in utility vehicles with long service lives, while full-flow filters should be included in the initial equipping of passenger cars instead of subsequent retrofitted filters with low separation rates.

Internal Combustion Engine Handbook | 807

6606_Book.indb 807

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21.6.3.16 Catalytic Particle Filter One method to improve soot regeneration is to use catalytic particle filters that are an alternative for diesel passenger cars in addition to the systems discussed in Section 21.6.1 based on fuel additives. Figure 21.106 shows a schematic representation of a wall-flow particle filter. The catalyst is applied to the particle filter by a coating method similar to that used for diesel oxidation catalytic converters or three-way catalytic converters, and it permits catalytic reactions at the boundary layer between the solid soot particles and the catalytic coating. Differences in coating methods, however, result from the alternately sealed channels that prevent linear flow as it occurs in uniflow structures and consequentially increase the manufacturing effort. It is possible to coat the inlet and outlet channels with different catalytic materials to obtain different functions. Regeneration occurs from the oxidation of soot. The available oxidants are the residual oxygen from the engine exhaust as well as the oxygen bonded to the nitrogen as NO2 that is in the engine exhaust as the pollutant NOx. Figure 21.117 lists the most important chemical reactions that are relevant to soot burning. The NO2 available in small concentrations in the engine exhaust can be generated by a diesel oxidation catalytic converter before the filter as well as the catalytic coating in the soot filter. Another technical variation does not use an upstream oxidation catalytic converter. In this case, the reduction of the carbon monoxide and hydrocarbons required to satisfy exhaust laws occurs by means of the catalyst in the soot filter. The catalyst in the soot filter can also oxidize several times the NO that forms when the soot burns, which is also termed NO2 turnover. The schematic representation in Figure 21.118 illustrates such a cyclic process. This multiple use of the NO can clearly improve the soot burning given the right system design. Another important function of the catalytic soot filter is the oxidation of the CO that rises in addition to the NO from the incomplete burning of the soot. If the CO is not converted, it can lead to substantial pollutant emissions.

O2

NO O

O

2CO + O2 2NO + O2

C NO2

Catalytic converter

CO2 CO

Figure 21.118 Cyclical process of NO oxidation during soot burning [Engelhard].

The relevant quantities in catalytic soot regeneration are the exhaust temperature and the concentration of the O2 and NO2. However, the soot regeneration can also be influenced by the following parameters: •• exhaust gas temperature •• oxygen content of the exhaust (O2, NO2) •• flammable residual components of the exhaust •• exhaust mass flow •• particle composition such as the mass of the deposited HCs •• particle characteristic (such as the ability to form active O2 centers) Figure 21.119 illustrates the influence of the temperature and the use of hydrocarbons deposited on the soot to increase temperature. The figure shows that the previously described burning to form NO2 functions from only approximately 250 to 450°C. The bottom temperature threshold is set by the light-off behavior of the catalyst that catalyzes the NO oxidation. Between 250 and 450°C, the NO2/C reaction determines the soot-burning rate. At a NO2∷NO ratio of 1∶1, the NO2/C mass ratio in the exhaust must be at least eight corresponding to the stoichiometry of the NO2/C reaction to attain quantitative soot burning. At 450°C, the NO2/C and O2/C reactions occur at the same speed (isokinetic point). Above 450°C, the more active O2/C reaction dominates, and the soot burns independent of the NOx concentration.

Diesel oxidation Dieseloxidations catalytic converter katalysator

HC (SOF) + O2

NO2

Catalytic converter

CO2 + H2O 2CO2 2NO2

Catalytic Katalytischer Soot Filter Rußfilter

C (soot) (Ruß + O2 2NO + O2 C (soot) (Ruß + NO2 2CO + O2

CO/CO2 2NO2 CO/CO2 + NO 2 CO2

Figure 21.117 Chemical reactions for soot burning through a catalytic soot filter [Engelhard].

808 | Internal Combustion Engine Handbook

6606_Book.indb 808

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

10

C+O2 Thermal

9

C+O2 catalytic

8

Temperature increase II

7

NO oxidation

6

C+NO2

5

Temperature increase I

4

Engine oil oxidation

3

HC engine oil - liquid

2

Fuel oxidation

1

Liquid fuel 0

100

200

300

400

Exhaust gas temperature °C

500

In the operation of a diesel passenger car, these conditions (NO2:C, T) cannot always be maintained: Soot accumulates in the filter, the exhaust counterpressure rises, and the soot must be burned with oxygen. The required temperature can be reduced by approximately 150 K by the catalytic soot filter. The required temperature range of 450 to 600°C cannot be sufficiently ensured in practical driving so that active engineside measures must be taken. The complete regeneration of the soot may occur at different speeds, taking into account the state of the soot in the filter, the O2 concentration in the exhaust, and the volumetric exhaust flow. The speed of the soot burning rises with the temperature, the quantity of deposited hydrocarbons (“moist soot”), the O2 concentration, and the decrease in the volumetric exhaust flow. Figure 21.111 shows, in particular, how available HCs that may not be in the liquid phase can enhance soot burning by exothermic reactions. In this case, the catalytic combustion of the SOF (soluble organic fraction) component provides the necessary activation energy for igniting the soot. The catalytic soot filter must be protected from uncontrolled burning that can arise when the quantity of flammable components, that is, soot and deposited HCs, becomes too great in the filter since the coating cannot withstand temperatures greater than 1000°C. However, the materials of the wall-flow filter are also subject to restrictions in view of potential local temperature differences that can easily arise from uncontrolled burning and that may cause the material to crack. Another consideration in the use of catalytic soot filters is the exhaust counterpressure. This requires coordination of the filter material with the catalytic coating: The primary conflict is between high filter efficiency with simultaneously effective regeneration and a minimum exhaust counterpressure. The filter efficiency can be influenced by the filter parameters of pore size, porosity, and pore structure as well as the type and mass of the applied catalytic coating. During operation, the catalytic filter is loaded with ash from the engine oil. A positive factor promoting filter efficiency in comparison to fuel-additive-supported systems is that there is

600

700

Figure 21.119 Temperature ranges of soot regeneration [Engelhard].

no ash from the fuel additive. Nevertheless, the system must be designed so that engine oil ash does not restrict the function of the catalytic converter over the desired operating time.

Bibliography

21-1. Staab, J. 1997. Automobil-Abgasanalytik bei niedrigen Grenzwerten (Automobile exhaust analyses at low limit values) In: MTZ Motortechnische Zeitschrift, 58(3): 168–172. 21-2. Tauscher, M. 2000. Optimierung eines On-board Abgasmesssystems (Optimization of an on-board exhaust gas measurement system) Diplomarbeit an der Technischen Universität Wien (Thesis at the University of Vienna). 21-3. Mohr, M. 1998. Feinpartikel in Verbrennungsabgasen und Umgebungsluft (Fine particles in combustion exhaust and ambient air). Internet publication, http://www.empa.ch/deutsch/fachber/abt137/motor/ partikel.html. 21-4. Neumann, H., Hötzel, G., Lindemann, G. 1997. “Advanced Planar Oxygen Sensors for Future Emission Control Strategies.” SAE-Paper 970459, Detroit. 21-5. Pucher, E., Weidinger, Ch., Holzer, H. 1998. Kontinuierliche HC-Indizierung am 4-Ventil Ottomotor (Continuous HC indexing on the 4-valve gasoline engine). Conference transcript of the 3rd International indexing Symposium, AVL Deutschland GmbH, Mainz. 21-6. Pachta-Reyhofen, G. 1985. Wandfilmbildung und Gemischverteilung bei Vierzylinder-Reihenmotoren in Abhängigkeit von Vergaser- und Saugrohrkonstruktion (Wall film formation and mixture distribution with four-cylinder serial engines as a function of carburetor and intake manifold design. Dissertation at TU Vienna. 21-7. Pucher, E., Schopp, G., Klawatsch, D. 1995. Fast Response Cycle-toCycle Exhaust Gas Analysis for S.I. Engines. In: 5th International EAEC Conference Strasbourg. 21-8. Puxbaum, H et. al. 1998. Tauerntunnel Luftschadstoffuntersuchung 1997—Ergebnisse der Messkampagne vom 2.–5. Oktober 1997 (Tauern tunnel air pollutant investigation 1997—Results of the measuring campaign from October 2–5, 1997). Land Salzburg, Department 16 Environmental protection, Salzburg. 21-9. Klingenberg, H. 1996. Automobile Exhaust Emission Testing: Measurement of Regulated and Unregulated Exhaust Gas Components, Exhaust Emission Tests. Berlin: Springer Verlag. 21-10. List, H., Cartellieri, W.P. 1999. Dieseltechnik, Grundlagen, Stand der Technik und Ausblick—10 Jahre Audi TDI-Motor.(Diesel technology, state of the art and outlook—10 years of Audi TDI engines), In: Special edition of the MTZ, S. 10–18.

Internal Combustion Engine Handbook | 809

6606_Book.indb 809

1/19/16 8:54 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21-11. Spindler, S. 1992. Beitrag zur Realisierung schadstoffoptimierter Brennverfahren an schnelllaufenden Hochleistungsdieselmotoren, VDI Fortschrittberichte Reihe 6 Energieerzeugung (Contribution for the realization of pollutant-optimized combustion processes on quick running powerful diesel engines, VDI progress reports Series 6 Power generation) No. 274. 21-12. Kohoutek, P et. al. 1998. Status der nichtlimitierten Abgaskomponenten bei Volkswagen (Status of the non-limited exhaust gas components at Volkswagen). VDI progress reports, Issue no. 348, Düsseldorf. 21-13. Boulouchos, K et al. 1997. Verbrennung und Schadstoffbildung mit Common-Rail Einspritzsystemen bei Dieselmotoren unterschiedlicher Baugröße (Combustion and pollutant formation with common rail injection systems for diesel engines of various size ). International conference “Common Rail injection Systems—Present and Future Potential,” ETH Zürich, S. 111. 21-14. Langen, P., Cosfeld, R., Grudno, A., Reif, K. 2000. (BMW Group): Der elektromechanische Ventiltrieb als Basis zukünftiger Ottomotorkonzepte (the electromagnetic valve train as the basis for future gasoline engine concepts) 21. Internationales Wiener Motorensymposium. 21-15. Wirbeleit, C., et al. 2000. Können innermotorische Maßnahmen die aufwendige Abgasnachbehandlung ersetzen? (Can internal engine measures replace the extensive exhaust gas post-treatment?) VDI progress reports, Issue no. 420, Düsseldorf. 21-16. Härle, H. 1997. Anwendung von Common Rail Einspritzsystemen für NKW-Dieselmotoren, Internationale Konferenz “Common Rail Einspritzsysteme—Gegenwart und Zukunftspotenzial” (Usage of common rail injection systems for commercial vehicle diesel engines, international Conference “Common Rail injection Systems—Present and Future Potential” ) ETH Zürich, S. 42–43. 21-17. Bach, M., Bauder, R., Endres, H., Pölzl, HW., Wimmer, W. 1999. Dieseltechnik, Der neue V8-TDI-Motor von Audi, Teil 3: Thermodynamik (Diesel technology, The new Audi V8 TDI engine. Part 3, thermodynamics. 10 years of Audi TDI engines. In: Special edition of the MTZ, S. 40–46. 21-18. Guber, M., Klawatsch, D., Pucher, E. 1999. Comparative Measurements of Particle Size Distribution: Influences of Motor Parameters and Fuels. Proceedings Second International ETH-Workshop on Nanoparticle Measurement, ETH Zürich Laboratorium für Festkörperphysik. 21-19. Fraidl, K., Kapus, P., Piock, W., Wirth, M. 2000. Fahrzeugklassenspezifische Ottomotorkonzepte (Vehicle classes—specific gasoline engine concepts). VDI progress reports, Issue no. 420, Düsseldorf. 21-20. Hammer, J., Kufferath, A., Herynek, R. 2010. Ottomotor 2015—Anforderungen und Systemlösungen zur Erreichung künftiger Emissionsziele (Diesel engine 2015—Requirements and system solutions for achieving future emission goals), 5th Emission Control 2010, Dresden. 21-21. Liebl, J. 2010. BMW Efficient Dynamics—unser konsequenter Weg zum emissionsfreien Autofahren (Our efficient path toward emissions-free driving), 5th Emission Control 2010, Dresden. 21-22. Carberry, B., Balenovic, M., Chigapov, A., Dubkow, A., Roemer, D., Reichert, M., Schneider, M., Ukropec, R., Yacoub, Y. 2010. NOX Aftertreatment Technologies for Future European Emission Standards, 5th Emission Control 2010, Dresden. 21-23. Pfeifer A., Schlager, G., Jaussi, F. 2010. Emissionskonzepte für zukünftige Off-Highway Industriemotor-Anwendungen (Emission concepts for future Off-highway driving, 5th Emission Control 2010, Dresden.

eines 1.8 l 5V-Motors” (Catalytic converter concepts for future exhaust gas legislation using a 1.8 l 5 V engine as an example), 17. Internationales Wiener Motorensymposium (international Vienna Engine Symposium) 4/25–26/1996, Vienna. 21-27. Maus, W., Brück, R., Hirth, P., Hodgson, J., Presti, M. “Potential von Katalysatorkonzepten zum Erreichen der SULEV-Emissionsgrenzwerte” (Potential of catalytic converter concepts for achieving the SULEV emission limits) 20. Internationales Wiener Motorensymposium (international Vienna Engine Symposium) Friday, May 6–7, 1999, Vienna. 21-28. Bollig, M., Liebl, J., Zimmer, R., Kraum, M., Seel, O., Siemund, S., Brück, R., Diringer, J., Maus, W. 2004. “Next generation catalysts are turbulent: development of support and coating,” SAE 2004-01-1488. 21-29. Held, W., Rohlfs, M., Maus, W., Swars, H., Brück, R., Kaiser, F.W. 1994. “Improved cell design for increased catalytic conversion efficiency,” SAE 940932. 21-30. Kent Dawson, E., Kramer, J. 2006. “Faster is better: The effect of internal turbulence on DOC efficiency,” SAE 2006-01-1525. 21-31. Konieczny, R. 2006. “The PM Metalit™: Experience with the partialflow particulate trap with regard to the reduction of particulate number and mass,” Lecture at the CTI Munich. 21-32. Brück, R., Hirth, P., Rice, M. 2009. “NOx aftertreatment for passenger cars and heavy duty truck applications for EU 6 and EU VI/US 2010 legislation,” SAE 2009-01-0846. 21-33. Hanel, F.J., Otto, E., Brück, R. 1996. “Electrically heated catalytic converter (EHC) in the BMW Alpina B12 5.7 Switch-Tronic,” SAE 960349. 21-34. Geringer, B., Hofmann, P., Holub, F. 2010. “Verbesserung des Hochlaufs und des Emissionsverhaltens im Kaltstart und Warmlauf bei Ottomotoren” (Improvement in the run-up and the emission behavior for cold start and warm-up with gasoline engines), MTZ 05.2010, Wiesbaden. 21-35. Kannapin, O., Guske, T., Preisner, M., Kratzsch, M. 2010. Particle reduction—New challenge for gasoline engines with direct injection, MTZ 11.2010, Wiesbaden. 21-36. Gaiser, G., Seethaler, C., Eberspächer, J. 2010. The latent heat accumulator catalytic converter—a new concept for an immediate lights-off of the catalytic converter during cold start, 5th Emission Control 2010, Dresden. 21-37. Heck, Farrauto. 1995. Catalytic Air Pollution Control, Van Nostrand Reinhold. 21-38. Farrauto, Voss. 1996. Monolithic Diesel Oxidation Catalysts, Applied Catalysis B: Environmental 10 29. 21-39. Bond. 1990. Heterogeneous Catalysis, Oxford University Press. 21-40. Bode (Ed.). International Conference on Metal-Supported Automotive Catalytic Converters (MACC 97), Wuppertal Germany 1997, Werkstoff -Informationsgesellschaft, Frankfurt. 21-41. Kruse, Frennet, Bastin (Eds.). 5th International Congress on Catalysis and Automotive Pollution Control (CAPOC 5, Brussels/Belgium, April 2000), Vol. 1 and 2, Universite Libre de Bruxelles. 21-42. Guyon, M., Blanche, P., Bert, C., Philippe, L., Renault, Messaoudi, CIRCA. Segime, NOx -Trap System Development and Characterization for Diesel Engines Emission Control, SAE2000-01-2910, Baltimore. 21-43. Patent DE 196 26 835 A1, Patent DE 196 26 836 A1.

21-24. Weigel, C., Schäffner, G., Kattwinkel, P., Viehweg, P., Hehle, M., Bergmann, D. 2010. Abgasnachbehandlungs-Technologien—Erprobung im realen Betrieb (Exhaust gas treatment technologies—trials in real operation) MTZ 11.2010, Wiesbaden.

21-44. Beutel, T., Dahle, U., Punke, A. 1999. “Euro 4—Abgasnachbehandlungstechnologien für Magermotoren (Otto/Diesel) (Exhaust treatment technologies for lean-burn engines (gasoline/diesel),” VDA Technical congress IAA Frankfurt/Main.

21-25. Brück, R., Konieczny, M., Brugger, M. 2010. Aktives Temperaturmanagement in SCR-Systemen—Anwendungsmöglichkeiten und Betriebsstrategien des elektrisch beheizbaren Katalysators EmiCat (Active temperature management in SCR systems—application options and operating strategies of the electrically heatable EmiCat catalytic converter) 5th Emission Control 2010, Dresden.

21-45. Cooper, B.J., Thoss, J.E. 1998. Role of NO in Diesel Particulate Emission Control, SAE Technical Paper no. 890404.

21-26. Faltermeier, G., Pfalzgraf, B., Brück, R., Kruse, C., Maus,W. “Katalysatorkonzepte für zukünftige Abgasgesetzgebungen am Beispiel

21-46. Mul, G. 1997. Catalytic Diesel Exhaust Purification, Proefschrift, Technische Universiteit Delft. 21-47. Neeft, J.P.A., Makkee, M., Moulijn, J.A. 1996. Diesel particulate emission control, Fuel Processing Technology 47.

810 | Internal Combustion Engine Handbook

6606_Book.indb 810

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

21.6 Exhaust Treatment in Diesel Engines

21-48. Winkler, A., Ferri, D., Dimopoulos Eggenschiler, P., Aguirre, M. 2010. Analyseverfahren zur Alterung von Dieseloxidations-Katalysatoren, (Analysis procedure for the aging of diesel oxidation catalytic converters) MTZ 06.2010, Wiesbaden. 21-49. Fiebig, M., Schönen, M., Gütering, U., Pischinger, S. 2010. Einflüsse motorischer Betriebsparameter auf die Reaktivität von Dieselruß (Impacts of eingineoperating parameters on the reactivity of diesel soot), MTZ 07.– 08.2010, Wiesbaden. 21-50. Schröder, C. 2010. Neuronales Netzwerk zur NOx -Sensierung im SCR-System (National network for NO sensing in SCR systems), MTZ 02.2010, Wiesbaden. 21-51. Maroteaux, D., Beaulieu, J., D’Oria, S. 2010. Entwicklung der NOx -Nachbehandlung für Renault-Dieselmotoren (Development of the NO post-treatment for Renault diesel engines), MTZ 03.2010, Wiesbaden. 21-52. Pauli, E. 1986. Regenerationsverhalten monolithischer Partikelfilter (Regeneration behavior of monolithic particle filters), Dissertation, Mechanical engineering department of the Rheinisch-Westfälischen Technical University of Aachen. 21-53. Burtscher, H. 2000. Partikelemissionen und Partikelfiltertechnik (Particle emissions and particle filter technology), Seminar Haus der Technik, Munich. 21-54. Hinds, W.C. 1989. Aerosoltechnology John Wiley. 21-55. Siegmann, K. Soot Formation in Flames, J. Aerosol Scie. 31, Suppl. 1. 21-56. ACEA-Programme on the Emissions of Fine Particles from Passenger Cars, December 1999. 21-57. Wichmann, E., Peters, A. 2000. Epidemiological evidence of the effects of ultrafine particle exposure, The Royal Society, 10.1098/ rsta.2000.0682, Phil. Trans. R. Soc. Lond. A 358, 2751–2769. 21-58. Mayer A., et al. Particulate Traps for Retro-Fitting Construction Site Engines VERT: Final Measurements and Implementation, SAE 1999-01-0116. 21-59. Geprüfte Partikelfiltersysteme für Dieselmotoren, Vollzugsunterlagen (Tested particle filter systems for diesel engines, implementation documents) “Suva/BUWAL Switzerland, http.//www.BUWAL.ch/Projekte/ Luft/Partikelfilter/d/Index.htm. 21-60. Ebener, S., et al. 20001. Parallele Reduktion von Partikel und NOx — ein neues Abgasnachbehandlungskonzep (Parallel reduction of particles and NOx—a new exhaust treatment concept) Wiener Motorsymposium. 21-61. Hardenberg, H. Wickelrußfilter für Stadtomnibusse in der Erprobung im Verkehrsbetrieb (Wound soot filters for city buses in the testing in traffic operation) Der Nahverkehr 4/86. 21-62. Baraket, M. Das dynamische Verhalten von Faserfiltern für feste und flüssige Aerosole (The dynamic behavoir of fibre filters for solid and liquid aerosols, Dissertation ETH Zurich Nr. 9738/192. 21-63. Mayer, A., et al. Passive Regeneration of Catalyst Coated Knitted Fiber Diesel Particulate Traps, SAE 960138. 21-64. Buck., et al. 1995. Gestrickte Strukturen aus Endlosfasern für die Abgasreinigung (Knitted structures made from infinite fibers for exhaust purification), MTZ Motortechnische Zeitschrift 56. 21-65. Dürnholz, M., Krüger, M. 1997. Hat der Dieselmotor im Pkw eine Zukunft? (Does the diesel engine in passenger cars have a future?) 6. Aachener Kolloquium Fahrzeug und Motortechnik.

on the emissions of diesel engines with exhaust treatment), Wiener Motorsymposium. 21-71. Mayer, A., et al. Particulate Trap Selection for Retrofitting Vehicle Fleets based on Representative Exhaust Temperature Profiles, SAE Paper 2001-01-0187. 21-72. Heeb, N. 2000. Sekundäremissionen durch Abgasnachbehandlung (Secondary emissions due to exhaust post-treatment), Seminar Haus der Technik/Essen, Partikelemissionen Mai. 21-73. Jodeit, H. Untersuchungen zur Partikelabscheidung in technischen Tiefenfiltern (Investigations for particle deposition in technical deep filters), VDI-Fortschrittsberichte Nr. 108. 21-74. Rausch, W. 1988. Untersuchungen am Sinterlamellen-Filtermedium (Investigations on sintered fin filter medium), Aufbereitungs-Technik Mineral Processing. 21-75. Partikelfilter für schwere Nutzfahrzeuge (Particle filters for heavyduty utility vehicles), Issued from Switzerland Bundesamt für Umwelt, Wald und Landschaft, UM 130/12.2000. 21-76. Mayer, A. Pre-Turbo Application of the Knitted Fiber Diesel Particulate Trap, SAE 940459. 21-77. Fanick, E.R., Valentin, J.M. Emissions Reduction Performance of a Bimetallic Platinum/Cerium Fuel Borne Catalyst with Several Diesel Particulate Filters on Different Sulfur Fuels, SAE 2001-01-0904. 21-78. Mayer, et al. Particulate Traps for Construction Machines Properties and Field Experience, SAE 2000-01-1923. 21-79. 4. ETH Conference “Nanoparticle Measurement” Zurich August 2000, Proceedings, Switzerland Bundesamt für Umwelt, Wald und Landschaft. 21-80. Hüthwohl, G., et al. 2001. Partikelfilter und SCR, Abgasnachbehandlungstechnologien für Euro 4-Anforderungen (Particle filters and SCR, exhaust treatment technologies for Euro 4 requirements), 4. Dresdner Motorenkolloquium. 21-81. Toshiaki, Tanaka, et al. 2001. Parallele Reduktion von Partikel und NOx —ein neues Abgasnachbehandlungskonzep (Parallel reduction of particles and NOx—a new exhaust treatment concept) Wiener Motorsymposium. 21-82. AVL Forum on particle emissionsions, 2000, Darmstadt. 21-83. Matter, U. 1999. Probleme bei der Messung von Dieselpartikeln (Problems with the measurement of diesel particles), Seminar Haus der Technik, “Feinpartikelemissionen von Verbrennungsmotoren.” 21-84. Kasper, M., Matter, U., Burtscher, H. 1998. NanoMet: OnLine Characterization of Nanoparticle Size and Composition, SAE. 21-85. Gruber, M., et al. 2001. Partikelgrößenverteilung im instationären Fahrzyklus (Particle size distribution in the transient driving cycle), Vienna engine symposium. 21-86. ETH Conference “Combustion Generated Nanoparticles,” Zurich 8/2004. 21-87. Draft Amendment to Regulation No. 83, GRPE-48-11, 4 June 2004, www.unece.org/trans/doc/2004/wp29GRPE. 21-88. Comparison Study of Particle Measurement Systems for Future Type Approval Application, GRPE-PMP CH6, M. Mohr, EMPA-Bericht 202779, Mai 2003 (www.empa.ch).

21-66. Kasper, M. Ferrocene, Carbon Particles, and PAH, Dissertation No. 12 725/1998 ETH Zürich.

21-89. The Number Concentration of Non-Volatile Particles—Design Study for an Instrument According to the PMP Recommendations, M. Kasper, SAE 2004-01-0960.

21-67. European Patent CRT EP 0835684.

21-90. CAST: ME company documents: www.matter-engineering.com.

21-68. Salvat, O., Marez, P., Belot, G. Passenger Car Serial Application of a Particulate Filter System on a Common Rail Direct Injection Diesel Engine. SAE Paper 2000-01-0473, PSA Peugeot Citroen.

21-91. Mayer, A., et al. 2004. Minimierung der Partikelemissionen von Verbrennungsmotoren (Minimization of particle emissions from combustion engines), Expert Verlag, ISBN 3-8169-2430-1.

21-69. Herzog, P. 2000. Exhaust Aftertreatment Technologies for HSDI Diesel Engines. Giornale della “Associazione Tecnica dell’Automobile,” Torino, ATA vol. 53, 389.

21-92. Schommers, J., Enderle, C., Binz, R., Duvinage, F., Ruzicka, N. 2004. Das neue Mercedes-Benz Dieselpartikelfilterkonzept in Verbindung mit der Abgasstufe EU-4 (The new Mercedes Benz diesel particle filter concept in connection with exhaust level EU-4). 25. Int. Wiener Motorensymposium, Band 2. Fortschritt-Berichte VDI, Reihe 12, Nr. 566.

21-70. Jacob, E. 2001. Einfluss des Motorenöls auf die Emissionen von Dieselmotoren mit Abgasnachbehandlung (The impact of engine oil

Internal Combustion Engine Handbook | 811

6606_Book.indb 811

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 21 Exhaust Emissions

21-93. Aerosole—Stäube, Rauche und Nebel (Aerosols—dusts, smoke and mist); MAK, 24. Delivery 1997. 21-94. Proceeding of the Pneumoconiosis Conference, Johannesburg 1959; J & A Churchill Ltd., London. 21-95. Feinstaub: Definition und Messverfahren (Fine dust: Definition and measuring processes); Kolloquium Feinstäube der VDI-Kommission Reinhaltung der Luft, Düsseldorf; Staub-Reinhalt. Luft 34 (1974) 9. 21-96. Mayer, A., et al. Properties of Partial Flow and Coarse Pore Deep Bed Filters Proposed to Reduce Particle Emissions of Vehicle Engines; SAE 2009-01-1087. 21-97. Czerwinski, J., et al. Diesel NO/NO2/NOx Emissions—New Experiences and Challenges; SAE 2007-01-0321. 21-98. Mayer, A., et al. Particle Fitler Properties after 2000 hrs Real World Operation; SAE 2008-01-0332.

21-108. Nachrüstung von Baumaschinen mit Partikelfiltern; Kosten/NutzenBetrachtung (Retrofitting of construction machines with particle filters; cost/benefit consideration); Schweizer Bundesamt für Umwelt BAFU:, Umweltmaterialien Nr. 148. 21-109. Czerwinski, J., et al. Combinations of Technical Measures for Reduction of Particle Emissions & Toxicity of 2-S Scooters; SAE 2009-01-0689. 21-110. Felchner, B., Bösswetter, A., Eder, M., Frambourg, M. 2010. CO2-Neutrale Partikelreduzierung im Dieselpartikelfilter (CO2-neutral particle reduction in the diesel particle filter), MTZ 05.2010, Wiesbaden. 21-111. Ferkel, H., Bachmann, M., Volpp, H.R., Stöwe, K., Hensgen, L. 2010. Edelmetallfreie Nanokatalysatoren für Dieselpartikelfilter (Nanocatalytic converters without precisous metals for diesel particles), MTZ 02.2010, Wiesbaden.

21-99. Mayer, A., et al. Qualitätsstandards und Prüfverfahren für Partikelfilter zur Nachrüstung von Nutzfahrzeugen (Qualtiy standards and test methods for particle filters for the retrofitting of utility vehicles); MTZ 01/2009 Jahrgang 70.

21-112. Hirth, P., et al. 2010. SCR- und Partikel-Abgasnachbehandlungssysteme für Heavy Duty EU VI und NRMM Stufe IV—Die Zukunft auf dem Prüfstand (SCR and partcile exhaust post-treatment systems for heavyduty EU VI and NRMM level IV—The future on the testbed). 10. International Stuttgart Symposium.

21-100.Euro 5/6 for passenger cars: EC-Regulation No. 715/2007 for the Euro-pean Parliament and the Council of 20 June 2007 on type approval of motor vehicles … http://ec.europa.eu/enterprise/ automotive/index_en.htm.

21-113. Eggenschwiler, P.D., et al. 2010. Rußaufbau- und -abbauphänomene sowie Ascheakkumulation in Dieselpartikelfiltern (Soot formation and decomposition phenomena and ash accumulation in diesel particle filters). 10. International Stuttgart Symposium.

21-101. Euro VI for utility vehicles Type Approval of Motor Vehicles and Engines with Respect to Emissions from Heavy Duty Vehicles Euro VI … ; Proposal for a Regulation of the European Parliament and of the Council; Brüssel 21.12.2007 http://ec.europa.eu/enterprise/ automotive/index_en.htm.

21-114. Linhart, J., et al. 2010. DPF-C—Ein Partikelfilter mit neuem Design und erweiterten Freiheitsgraden (DPF-C—A particle filter with new design and expanded degrees of freedom). 10. International Stuttgart Symposium.

21-102. Prüfung von Partikelfiltersystemen für Verbrennungsmotoren (Testing of particle filter systems for combustion engines); Swiss Standard SNR 277205. 21-103. Mizra, T., et al., Open-Pore Ceramic Foam as Diesel Particulate Filter, SAE 890172. 21-104. Mohr, M., et al. Conventional and New Methods of Particle Measurement … Expert-Verlag ISBN 3-8169-2552-9. 21-105. Mayer, A., et al. 2008. Particle Filter Retrofit for all Diesel Engines; Expert-Verlag, ISBN 978-3-8169-2850-8. 21-106. U.S.EPA:2007. The Cost Effectiveness of Heavy-Duty Diesel Retrofits and Other Mobile Source Emission Reduction Projects and Programs; EPA 420-B-07-006. 21-107. Jacobson, M.Z. 2007. Control of Fossil Fuel Particulate black Carbon and Organic Matter Possible the Most Effective Method of Slowing Global Warming; Journal of Geophysical Research; April 2002 and Testimony for the Hearing on Black Carbon and Global Warming of the United States House of Representative.

21-115. Maus, W., et al. 2010. PM-Metalit Advanced—Der innovative Partikelfilter zur Reduktion der Nanopartikel. (PM-Metallit Advanced—The innovative particle filter for the reduction of nano-particles) 31. Internationales Wiener Motorensymposium (31st international Vienna Engine Symposium), Vienna. 21-116. Pflaum, S., et al. 2010. Wege zur Rußbildungshypothese (Approaches to the soot formation hypothesis) 31. Internationales Wiener Motorensymposium (31st international Vienna Engine Symposium),Vienna. 21-117. Predelli, O., et al. 2010. Kontinuierliche Einspritzverlaufsformung in Pkw-Dieselmotoren—Potenziale, Grenzen und Realisierungschancen (Continuous injection process forming in passenger car diesel engines—prospects, limits and implementation opportunities). 31. Internationales Wiener Motorensymposium (31st international Vienna Engine Symposium), Vienna. 21-118. Sorger, H., et al. 2010. Herausforderung CO2: Aggressives Downsizing für HSDI-Diesel—Motorkonzeptdefinition. (The CO2 challenge : Aggressive downsizing for HSDI diesel engine concept definition) 31. Internationales Wiener Motorensymposium (31st international Vienna Engine Symposium), Vienna.

812 | Internal Combustion Engine Handbook

6606_Book.indb 812

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22 Operating Fluids In automotive technology, the term operating fluids is used as a generic term for fuels, lubricants, coolants, and hydraulic fluids. In this chapter, we use it specifically in reference to application engineering. We will not, therefore, be discussing the exploration, recovery, or processing of mineral oil and synthetic products.

22.1 Fuels (For fuel chemistry, see also chapter 14.) Let us first briefly note a few basic properties of fuels that do not appear in standards because these properties refer to the fundamentals of combustion. Nonetheless, an understanding of the related context is not without merit.  /H ratio, air requirement, and air–fuel ratio C Fuels essentially consist of hydrocarbons composed of the elements C and H. The minimum amount of air (L) required for their complete combustion—the theoretical air requirement—can be calculated when the masses of carbon, hydrogen, and, possibly, oxygen are known from an elementary analysis of the relevant fuel. This is identified as L and is indicated in kg/kg. Calculate the sulfur-free hydrocarbon mixtures according to the following relationships:

L=

O 0.23

(22.1)

O = 2.67 ⋅ 0.01 C + 8 ⋅ 0.01 H 2 − 0.01 O 2

Example of a calculation of SuperPlus:



O = 2.67 ⋅ 0.01 C + 8 ⋅ 0.01 H 2 − 0.01 O 2 O = 2.67 ⋅ 0.847 + 8 ⋅ 0.133 − 0.02 O = 2.261+ 1.064 − 0.02 O = 3.305 kg L = 3.305/0.23 = 14.369 kg air/kg fuel.



(22.2)

(see Figure 22.1). Figure 22.1 lists the percent mass of C, H2, and O2 for a few important hydrocarbons and fuels, the resulting C/H ratio, and the theoretical air requirement. % (m/m)*

kg/kg

Fuel

C

H

O

C/H

L

Methane

~75.0

~25.0

—

~3.0 

~17.4

Propane

~81.8

~18.2

—

~4.5 

~15.8

Butane

~82.8

~17.2

—

~4.8 

~15.6

n Heptane

~84.0

~16.0

—

~5.25

~15.3

i Octane

~84.2

~15.8

—

~5.33

~15.2

Cetane

~85.0

~15.0

—

~5.67

~15.1

Xylene

~90.6

~9.4

—

~9.64

~13.8

Toluene

~91.3

~8.7

—

10.5 

~13.6

Benzene

~92.3

~7.7

—

12.0 

~13.4

Gasoline

~85.5

~14.5

—

~5.9

~14.9

Super gasoline

~85.1

~13.9

~1

~6.1

~14.6

SuperPlus

~84.7

~13.3

~2

~6.5

~14.4

Diesel fuel

~86.3

~13.7

—

~6.3

~14.8

*% (m/m) responds to mass percent.

Figure 22.1 C/H ratio and air requirement [22-1].

The ratio of the actual air supplied for combustion to the theoretically required air is called the air–fuel ratio (λ ). When there is excess air, that is, λ > 1, the engine operates at a lean setting; given an air deficiency of λ < 1, the engine operates at a rich setting. When λ = 1, the air–fuel ratio is stoichiometric. For high performance, a spark-ignition engine operates with a fuel-rich air mixture around λ = 0.85–0.90. For low fuel consumption, we can lean off up to λ = 1.1; in a spark-ignition engine with direct injection, the air–fuel ratio can be above λ = 2. In principle, diesel engines operate with excess air. Because of their technical principle, diesel engines operate with excess air. Under a full load, they operate at λ ≈ 1.2 and at λ > 8 while idling.

Internal Combustion Engine Handbook | 813

6606_Book.indb 813

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Today, a distinction is drawn between fuels that convert energy in internal combustion engines, aviation fuel for generating thrust in air travel, and fuels for heating purposes. Fuels can be liquid or gaseous. The energy that is chemically bound within the fuels is first converted by combustion into heat and then directly converted in the same machine into mechanical work. The term “sprit” frequently used in German media and in the German vernacular is insufficient because it actually refers only to ethyl alcohol. It originated from the economic crisis after World War I since fuel alcohol (Kraftspiritus) consisting of ethanol from potatoes (Sprit) produced by the administrative monopoly for distilled spirits had to be used increasingly as an additive to alleviate the drastic gas shortage. To further expand sales, “Reichskraftsprit GmbH, Berlin” was founded in 1925 whose products were called Monopolin. These were mixed with gasoline and/or benzene in quantities up to 65%.

Product Description

Density (kg/dm3)

Boiling Range (°C)

Cetane Number

Kerosene

805

150–260

45

Light gas oil

840

210–320

55

Heavy gas oil

860

200–400

55

Vacuum gas oil

870

250–400

56

Figure 22.2 Diesel fuel components from distillation [22-1].

The properties of typical diesel refinery components that arise from today’s cracking procedure are shown in Figure 22.3. Density (kg/dm3)

Product Description

Boiling Range (°C)

Cetane Number

Hydrocrackers

860

170–400

52

Thermal crackers

857

180–400

40

Catalytic crackers

953

195–410

40

Figure 22.3 Diesel fuel components from cracking [22-1].

22.1.1 Diesel Fuel (DK)

The boiling ranges of diesel fuels extend from approximately 180–380°C. Diesel fuels are used in high-speed diesel engines, especially automotive diesel engines (passenger cars and commercial vehicles). They consist of approximately 300 different hydrocarbons that are obtained in refineries using various methods for processing petroleum from a wide range of sources. Whereas earlier they consisted of relatively simple distillation products, in recent years they have become highly complex because of the much higher demands of engine manufacturers and developments in the mineral oil industry. Additives have been needed to greatly enhance the properties of the basic products that affect the engine. Since 1987, a few branded fuels have offered super diesel. From then on, diesel fuels and heating oils have developed in different directions. Since 2004, so-called “designer fuels” have been offered, some containing synthetic components suitable for highly developed and future engine concepts in particular. The previously frequently used term, “gas oil,” is now dated in the field of application engineering, although it is still used in refineries for middle distillates. 22.1.1.1 Diesel Fuel Components and Composition Diesel belongs to the group of light middle distillates of petroleum. It is a mixture of primarily paraffinic hydrocarbons (alkanes) whose respective percentages influence engine behavior. Whereas primarily fractions from atmospheric distillation were used in an earlier period, today cracking components are more frequently used because of the continuously increasing demand for diesel fuel. Figure 22.2 lists the properties of typical refinery components of diesel fuel from distillation.

The continuously increasing demand for diesel fuel has led to a shift in the types of fuel. Figure 22.4 shows the historical ratio of diesel consumption to gasoline consumption in Germany. The percentage of diesel in total mineral oil consumption in Germany is now approximately 55%. In addition to conventional petroleum-based diesel fuel, there is a series of other, synthetically made substances that can be used for combustion in diesel engines. In 1925, Fischer– Tropsch synthesis was discovered in which synthetic gas was, for example, obtained from coal or natural gas. Synthetic hydrocarbons can be made from this gas with the aid of catalysts, and the hydrocarbons can be refined into gasoline or diesel fuel. This relatively inefficient method is rarely used today. Two synthetically manufactured products are interesting as mixed components for diesel fuel, even if they are not readily available, that is, Shell middle distillate synthesis from natural gas, and extra high viscosity index that occurs in slight amounts as a byproduct from the production of synthetic lubricants. Both products have a very high cetane number of greater than seventy (see “Ignitability”) and are practically sulfur-free. Because of the high production costs and low availability, they can be used only as blend components for diesel fuel. For some time, “biodiesel” (FAME from fatty acid methyl ester) has been produced from biomass, mainly rapeseed oil, by esterizing with methanol. Because of bio-quota legislation in the EU and further national efforts, approximately 5% (V/V) FAME have been added to diesel fuel since 2003 and 7% since February 2009. This will be discussed further in Section 22.1.1.4. For diesel fuel, the demands of application technology and production

Fuel

1975

1980

1985

1990*

1995

1998

2000

2001

2002

Diesel

10,333

13,099

14,556

21,464

26,208

27,106

28,922

28,545

28,631

Gasoline

20,174

24,463

23,131

31,779

30,306

30,281

28,807

27,948

27,195

Diesel/gasoline

0.5120

0.5400

0.6290

0.6970

0.8650

0.8950

1.0060

1.0210

1.0530

*Since 1990 in all of Germany.

Figure 22.4 Ratio of diesel fuel consumption to gasoline consumption in Germany in millions of tons [22-2].

814 | Internal Combustion Engine Handbook

6606_Book.indb 814

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

are highly contradictory. Figure 22.5 shows how, for example, the paraffin component, density, final boiling point, crack component, and sulfur content are advantageous or disadvantageous either in the engine or in production. 22.1.1.2 Characteristics and Properties The minimum requirements for diesel fuel are found in DIN EN 590. They primarily concern density, ignition quality (cetane number), boiling curve, and resistance to cold and sulfur content. The standard characteristics of diesel fuels and their practical importance are shown in Figure 22.6.

The compromise arrived at in 1998 between the automobile and mineral oil industries in the EU commission for the so-called car-oil program EPEFE, see also Section 22.1.2.2) produced the following change to some environmentally relevant characteristics of DIN EN 590 (Figure 22.7). The aspirations of the German government for an increased portion of fuels from renewable sources required, among others, the approval for a higher quantity of FAME as permitted by EN 590. Before this background, the German parliament adopted in early 2009 a national diesel fuel standard (DIN 51628) permitting up to 7.0% FAME in diesel. It is assumed that

Characteristic

Demands on Production

Advantages for Use

Disadvantages for Use

Disadvantages for Production

Paraffin component

High

Ignitability

Low-temperature behavior

Cost

Density

Low

Exhaust emissions

Engine performance; Fuel consumption

Yield; cost

Final boiling point

Low

Exhaust emissions

Low-temperature behavior

Yield; cost

Crack components

Low

Ignitability Aging

Sulfur content

Low

Emission

Characteristic bandwidth

Narrow

Harmonization

Yield; cost Pump wear

Figure 22.5 Contradictory demands between application technology and production [22-3].

Yield; cost Yield; cost

Characteristic

Unit

Requirements

Influence on Vehicle Operation

Density at 15°C

kg/m3

820–845

Exhaust, performance, fuel consumption

Cetane number CI

— —

min. 51.0 min. 46.0

Starting and combustion behavior, exhaust and noise emission

 Up to 250°C

% (V/V)*

<65

 Up to 350°C

% (V/V)

min. 85

 95% point

°C

max. 360

Viscosity at 40°C

mm2/s

2.00–4.50

Evaporability, atomization, lubrication

Flash point

°C

Above 55

Safety

°C

max. 0

Distillation

Exhaust emissions, deposits

Filterability (CFPP)  15. 04. to 30. 09.

Low-temperature behavior max. −10

 01. 10. to 15. 11 and  01. 03. to 14. 04.

max. −20

 16. 11. to 28.(29.) 02. Sulfur content

mg/kg

max. 10**

Corrosion, particles, catalytic converter

Polycyclical aromatic hydrocarbons

% (m/m)

max. 11

Exhaust emissions, deposits Lubricity

FAME content

% (V/V)

max. 5

Carbon residue

% (m/m)

max. 0.30

Combustion chamber residue

Ash content

% (m/m)

max. 0.01

Combustion chamber residue

Water content

mg/kg

max. 200

Corrosion

Lubricity (WSD 1.4) at 60°C

μm

max. 460

Wear

*% (V/V) corresponds to percent by volume. **Sulfur content in Germany since early 2003 maximal 10 mg/kg (preferential tax treatment); from 2009 mandatory for the entire EU.

Figure 22.6 Minimum diesel fuel requirements, according to DIN EN 590, and their importance (excerpt) [22-1].

Internal Combustion Engine Handbook | 815

6606_Book.indb 815

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Characteristic

DIN EN 590 (1993–99)

Unit

Euro III (from 2000)

Euro IV (from 2005)

Sulfur (max.)

mg/kg

500

350

50

Cetane number (min.)

—

 49

 51

51

Density (max.)

kg/m3

860

845

845

T95 (max.)

°C

370

360

360

Polyaromates (max.)

% (m/m)

—

11

11

Figure 22.7 Result of the EU Auto/ Oil program for diesel fuel [22-1].

at the end of 2009, the European standard will be harmonized and then will replace the national standard. In addition to these standards, the global automotive industry has developed the so-called Fuel Charter (worldwide fuel charter—WWFC) defining the requirements for fuels in four quality levels. In meeting the top quality level (category 4), the automotive industry sees potential for the development of future engine concepts.

production costs. A helpful way out of this conflict would be to introduce a density sensor in the fuel tank to meter the fuel based on the measured density. The density of winter diesel fuel is lower than summer diesel fuel between five and ten units. The reason for this is discussed in the section on low-temperature behavior. In modern engine management systems, a density correction system is provided depending on, at least, the temperature.

 ensity D Density is an essential parameter. As the density increases, the energy content increases per unit volume. For example, the calorific value in the standard permissible density range is 34.8–36.5 MJ/L. Given an unchanging injected quantity of fuel, the energy supplied to the engine increases with the density, which increases engine performance. However, the exhaust emissions and, especially, the particles increase under a full load because of the richer mixture. On the other hand, the volumetric fuel consumption increases as density decreases. Engine manufacturers would, therefore, like a further restriction of the density range in the standard. However, this would greatly restrict the use of crack components that are basically heavy, which would restrict fuel availability in the face of continuously rising demand and thereby increasing

I gnitability This is characterized by the cetane number (CN). Now, it is set at a minimum of 51.0 in the standard. Engine manufacturers are requesting an increase to at least fifty-five. In the market, it presently lies between fifty-one and fifty-six and individually beyond sixty (since 2004), with a tendency toward higher values in summer fuel. In winter fuel, some of the higher boiling components must be discarded to ensure sufficient cold resistance. Essentially, the CN of the individual fractions rises with the boiling temperature. A certain amount of time is required (ignition lag) to start the combustion of the fuel injected into the hot air. This variable depends on the engine construction and the operating conditions and, very significantly, on the ignition quality of the diesel fuel. The primarily influential cetane number is the volumetric percentage of cetane C16H34

Improved diesel combustion due to improved ignitability Ignitability = essential diesel property • Dimension is the time between injection start and autoignition

low

Cetane number

Cylinder pressure

• Expressed as cetane number, measured as relative comparison in standardized test engine high

Ignition lag

Injection start

• Good ignitability (low ignition lag) = high cetane number

• Better energy conversion • Better cold start • Cleaner exhaust gas • More quiet combustion

• Minimum requirement DIN EN 590 CZ 49, from 2000 min. 51 • CN (ignition lag) can be improved by, amongst other things, adding an ignition accelerator

Autoignition Time lapse [°CA]

Figure 22.8 Combustion behavior as a function of ignition quality.

816 | Internal Combustion Engine Handbook

6606_Book.indb 816

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

(n-hexadecane), the paraffinic reference fuel where CN = 100 (in a mixture with α methylnaphtalene C11H10, an aromatic double-ring compound), and the reference fuel where CN = 0. The CN has a major influence on the combustion process and, hence, on exhaust and noise emissions. Figure 22.8 illustrates the improvement of combustion behavior from increasing the ignition quality. A high CN also has a positive effect on the starting behavior and the emissions of uncombusted hydrocarbons (HC). Because the natural CN is frequently insufficient, it must be increased by the addition of organic nitrates, such as amyl nitrate or ethylhexylnitrate (EHN). The required dose is usually under 0.1% (V/V), and an improvement of up to five units, depending on the basic fuel, can be attained. Figure 22.9 shows the effect of EHN in different diesel fuel samples with different responses. The cetane index (CI) indicated in the standard, in addition to the cetane number, is calculated from the density and boiling behavior as an alternative. It has only a limited correlation with the CN calculated in test engines because it cannot represent the universally used ignition accelerators. The CN is determined in a coordinating fuel research (CFR) or BASF test engine by changing the compression ratio ε , or by varying the throttling of the intake air. A high CN means that the compression ratio must be lowered or that the air flow rate must be reduced. For the standard value, the testing must be done with a CFR engine. The BASF engine evaluates 1.5 units higher than the CFR engine so that the measured values must be correspondingly corrected. Diesel Fuel Sample

CN Without EHN

CN With EHN

Gain

1

48.5

51.0

2.5

2

49.0

53.5

4.0

3

50.0

53.3

3.3

4

51.3

53.0

1.7

5

52.5

56.6

4.1

6

55.4

58.0

2.6

too great a percentage of low boilers causes evaporation directly at the injection nozzle, and this disturbs the intended distribution of the fuel in the combustion chamber. A restriction of the boiling curve sought by the automotive industry, for example, in Sweden for “Class 1” (200–290°C) would seriously limit the availability of diesel fuel, in Germany by approximately 40% (V/V). The desired lowering of the final boiling point would be primarily responsible, although it would understandably ease a few problems incurred by the engine manufacturers.  iscosity V The viscosity or inner friction of diesel fuel generally increases with the density. It must be prevented from falling below the set minimum value to ensure that there is sufficient lubrication between the sliding parts of the fuel injection system. If it is too high, the droplet size rises at the provided injection pressure. This results in poorer mixture formation and, hence, energy exploitation, lower performance, and higher soot emissions. The viscosity initially increases quickly as the temperature rises and then gradually decreases at a slower rate. Therefore, the diesel fuel in the fuel tank, fuel lines, and fuel filter should, if possible, be prevented from becoming too hot by using constructive measures.  lash point F The flash point is the temperature at which fuel vapors can be ignited by externally supplied ignition. It is important for determining the fire hazard and the subsequent safety measures in the storage and distribution systems. In the danger classification for this purpose, diesel fuel is rated A III—that is, less hazardous (gasoline is A I)—and must, therefore, have a flash point over 55°C. Even slight mixtures with gasoline can cause a forbidden drop below this threshold. For branded fuels, measures are taken to ensure that even a slight mixture with gasoline is impossible in storage and transport. The seriousness of a mixture with gasoline is illustrated in Figure 22.10. When producing diesel fuel, the flash point also restricts the use of highly volatile components.

Figure 22.9 CN increased by EHN [22-15]. Mixing with gasoline results in fall below flash point 0

Flash point lowering °C

 oiling curve (distillation) B Because fuels are a mixture of many hydrocarbons, they do not have an actual boiling point like pure hydrocarbons; rather, they have a boiling range. Diesel fuel starts to evaporate at approximately 180°C and stops at approximately 380°C. This behavior is not very important in comparison to gasoline because the mixture is prepared directly in the combustion chamber in diesel engines. The three points established in DIN EN 590, that is, 250 and 350°C and the 95% (V/V) point characterize only the top boiling range. Too large a portion of high boilers, especially aromates, that is, an overly high final boiling point, enlarges the droplets in the injection jet. The resulting longer ignition lag has a negative influence on the progression of combustion, which in turn increases noise and soot. On the other hand, a slightly high volatility is advantageous for cold starts, whereas

10 20 30 40

0

0,2 0,4 0,6 0,8

1

1,2 1,4 1,6 1,8

% gasoline in diesel fuel

Figure 22.10 Effect of gasoline in diesel fuel on the flash point.

Internal Combustion Engine Handbook | 817

6606_Book.indb 817

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

 ulfur content S Petroleum naturally contains more or less sulfur depending on its origin, as illustrated in Figure 22.11.

Geographic Location

Origin

Sulfur Content % (m/m)

North Sea

General Brent

0.6–2.2 0.4

Middle East

Iran heavy Arabian light Arabian heavy

1.7 1.9 2.9

Africa

Libya light Nigeria

0.4 0.1–0.3

South America

Venezuela

2.9

Russia

Siberia

1.5

Figure 22.11 Typical sulfur content of a few crude oils [22-15].

The sulfur is chemically bound, and more than 95% is converted into gaseous sulfur dioxide (SO2) upon combustion. The remainder largely passes into the particle mass of the exhaust that contains sulfurous acids and sulfates. Corrosion and exhaust pollution arises. In addition to the soot in diesel exhaust that is suspected of being carcinogenic, there are also polycyclic aromatic hydrocarbons (PAHs). According to analysis specifications, not only these critical substances are measured as “particles” but

also the sulfates formed from the sulfur and their adsorbed water. The goal is to reduce this component of the particles by further reducing the sulfur content. For this reason, the European standard EN 590 has been further tightened over the last two decades. In contrast to the prior Euro II threshold valid since 1996 of a maximum 0.05% (m/m), the European standard requires a reduction to 350 mg/kg (350 ppm) in Euro III (which began in 2000) and to 50 mg/kg (50 ppm) in Euro IV starting in 2005. Given the increasing use of oxidation catalytic converters, the percentage of sulfur converted to SO3 (sulfate) has strongly increased so that particle emissions have also increased. For this reason, among others, the European automotive industry (European Automobile Manufacturers Association—ACEA) seeks a further reduction of sulfur. Modern fuel production had to be further adapted for these tightened regulations, resulting in additional expensive plants and production processes for the refineries. Thanks to the rigorous desulfurization measures taken by the refineries in the last 30 years, SO2 emissions caused by diesel fuel are no longer an environmental issue. Particles can only be clearly reduced by corresponding exhaust treatment systems such as particle filters because the critical percentage of particles cannot be effectively lowered just by limiting the sulfur content, as can be seen in Figure 22.12. "Critical" particle portions are not reduced by low sulfur content approx. 5%

Sulfate Water

Particle emissions

 ow-temperature behavior L This describes the flowability and filterability of diesel fuel. The paraffinic hydrocarbons that are particularly suitable for diesel fuel because of their favorable self-ignition behavior unfortunately form crystals as the temperature falls. They precipitate and clump into “slack wax.” They accordingly impair the pumpability of the fuel and can plug the fuel filter. If this occurs, the engine can no longer be operated. The low-temperature behavior is substantially influenced by the fuel properties, the technical vehicle features, and the driving conditions. In DIN EN 590, the filterability in the cold filter plugging point (CFPP) test is a criterion for the resistance to cold of diesel fuel. In addition to the CFPP, brand manufacturers also use the criterion of the start of paraffin precipitation to determine the CP (cloud point, earlier also termed begin of paraffin precipitation). A distinction is, therefore, drawn between summer diesel fuel and winter diesel fuel. To produce a suitable winter quality, “tailored” additives are used. A combination of flow improvers and “wax antisettling additives” (WASAs) has proven to be particularly effective. WASAs are also effective in storage and distribution systems where wax crystal clumping can be prevented. In winter traffic, super diesel achieves CFPP values at up to −33°C from the optimum combination of boiling range and additives, which is far below the limit of −20°C required in the standard. All over Europe, the cold resistance of fuels is adjusted to seasonal requirements (see Figure 22.6). An all-season, particular cold-resistant premium fuel has been offered in the German market since 2007. If vehicles have an installed fuel/filter heating system, further substantial improvements can be attained.

Others PA H

considered to be carcinogenic

Soot

max. 0,20

max. 0,05

Sulfur in diesel fuel

Figure 22.12 Further desulfurization does not reduce a critical percentage of particles [22-15].

Another problem is that the hydrogen treatment in the refinery required for desulfurization yields a welcome increase in the CN, but at the cost of decreasing density with the abovedescribed consequences. The selective catalytic reduction (SCR) catalytic converters (with urea = AdBlue as reducing agent) used in modern concepts for exhaust gas aftertreatment are highly sensitive to

818 | Internal Combustion Engine Handbook

6606_Book.indb 818

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

large differences in density, as is the case with fuels B and C, there is no notable difference in the calorific value.

sulfur. For this reason, engine manufacturers globally demand practically sulfur-free fuel qualities with an S content of less than 10 mg/kg as they have been voluntarily offered in all of Germany since early 2003. EU regulation demands this quality only since January 1, 2009. In this context, it may be appropriate to recall the experimental “smoke suppressors” that were introduced a few decades ago. These additives, primarily barium compounds (as well as manganese and calcium), could not reduce the particle emissions that were not measurable at that time, but gave only the visual impression of smoke reduction by brightening (masking) the particles. As an addendum to discussing the characteristics cited in DIN EN 590, let us briefly address a few other interesting characteristics.

 arbon residue C This is determined by the last 10% of the distilled diesel fuel from low-temperature carbonization. It basically contains organic and a few inorganic components and provides information on the tendency of diesel fuel to coke the injection nozzles. Because ignition accelerators may slightly increase the carbon residue, it makes sense to measure it only in diesel fuel without additives. Whereas DIN EN 590 permits a maximum of 0.3% (m/m), we find clearly lower values in commercial diesel fuels. The average is 0.03% (m/m). 22.1.1.3 Additives for Diesel Fuel Additives are agents that improve the properties of fuel and lubricants and are normally added at concentrations in the ppm range. The goal of developing them, which is usually very expensive, is essentially to achieve a marked effect in the desired direction at the lowest dose without undesirable side effects. The additives useful for diesel fuel have been already discussed with occasional details when describing the individual characteristics and their practical relevance. Some additional information is appropriate, however. Figure 22.15 shows different problems associated with diesel vehicles that can be solved with additives.

 alorific value C A distinction can be drawn between the top calorific (TC) value that describes the combustion heat of the fuel including the condensation heat of water, and the bottom calorific (BC) value that indicates the actual useful amount of heat. In practice, only the BC is important (termed just the “calorific value” in the following). It provides information on the energy density. Whereas for scientific purposes the BC is generally expressed as MJ in reference to the unit of mass (kg), in practice the calorific value is expressed as MJ/L which refers to the volume. Also of interest is the calorific value of the combustible air–fuel mixture that depends on the calorific value of the fuel and the air–fuel ratio. It is not the BC that primarily influences the performance of the engine; it is, rather, the calorific value of the combustible air–fuel mixture. Figure 22.13 compares the BC of diesel fuel with super gasoline, methanol, and rapeseed methyl ester (RME).

 etergent and dispersant additives D Detergents are soap-free, surface-active wetting, and cleaning agents that reduce surface and interfacial tension. Usually they have a dispersant effect and can keep foreign materials in a liquid from clumping. A series of organic substances are suitable and proven as diesel fuel detergents and dispersants. These are amines, imidazolines, amides, succinimides, polyalkyl succinimides, polyalkyl amines, and polyether amines. Their task is to reduce or prevent deposits on the injection nozzles, especially throttling pintle nozzles, and in the combustion chamber. They are essential to ensure the operation of particularly fine direct injection nozzles and exactly maintain the pilot injection phase over a long period. Their efficiency with reference to the needle lift is particularly important. Also of interest is the positive effect of these additives on particle emissions during operation.

Net Calorific Value (CB) Fuel

MJ/L

MJ/kg

Diesel fuel

35.7*

43.0*

Rapeseed oil methyl ester

32.7*

37.2*

Gasoline (super)

30.8*

41.0*

Ethanol

21.17

26.80

*Mean value.

Figure 22.13 Comparison of calorific values of diesel fuel, super gasoline, ethanol, and RME [22-15].

We can see that diesel fuel has approximately 15% more energy than super gasoline, whereas RME has approximately 9% less energy than diesel fuel. In the case of methanol, we know that nearly twice the volume is consumed to produce the same amount of energy. It is also interesting to compare the calorific values and elemental analyses of three different diesel fuels as in Figure 22.14. We can see that even with more-or-less

 orrosion inhibitors C These use oxidation inhibitors and metal deactivators to ensure the aging stability of the diesel fuel that can vary widely depending on the type of crude oil and manufacturing procedure used. Oxidation inhibitors (antioxidants) prevent the corrosive attack of atmospheric oxygen. With metal deactivators, Calorific Value

Elemental Analysis % (m/m)

BC

BC

C

H

O

MJ/kg

MJ/kg

MJ/L

829.8

86.32

13.18

—

45.74

42.87

35.57

B (5% FAME)

832.31

85.86

13.12

—

45.55

42.70

35.43

C (7% FAME)

833.31

85.68

13.10

—

45.51

42.65

35.38

Diesel Fuel Sample

Density/15°C (kg/m3)

A (no FAME)

TC

Figure 22.14 Calorific values and elemental analysis of commercial diesel fuel [22-15].

Internal Combustion Engine Handbook | 819

6606_Book.indb 819

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Problems in diesel vehicles that can be resolved with additives

DIESEL

Low CN of diesel fuel disturbs the combustion process

Corrosion in fuel system affects operating behavior

Foam formation affects tank-filling process

Ignition accelerator for high cetane number requirement of the engine

Corrosion inhibitors prevent corrosion

Foam preventers (anti-foam) prevent foam formation

Nozzle coking disturbs mixture formation

Sulfur-free diesel may increase pump wear

Detergents reduce nozzle coking

Wear protection lubricates fuel pumps

Figure 22.15 Additives solve problems in diesel vehicles [22-7].

they form with the aid of organic compounds a catalytically inactive protective film that physically or chemically adheres to the metal surface.

However, there is no unanimity regarding the effectiveness of such measures. The increasing desulfurization of diesel fuel has resulted in the availability of some very mild-smelling fuels.

 ubricity additives L These are lubricity improvers that are added to the diesel fuel when, because of the strong drop in the sulfur content, the lubrication of the parts of the fuel injection pump under high mechanical stress is no longer ensured by the fuel itself. Without an additive, high pump wear occurs after a short period of operation. This affects, in particular, the distribution injector pumps, pump-nozzle and common-rail systems, long-term wear occurred even with a threshold of 0.05% (m/m) sulfur valid up to 1999. The high-frequency reciprocating wear rig (HFRR) test is used to measure wear protection. It simulates sliding abrasion in the fuel injection pump: a sphere with a 6 mm diameter is rubbed under constant pressure on a polished steel plate under a liquid. In DIN EN 590, a threshold (WSD) of 460 μ m wear to the sphere diameter is prescribed at 60°C test temperature. Polar compounds are used as the high-pressure additives. A lubricity additive can be omitted when a diesel fuel contains FAME. The HFRR is significantly improved by adding small quantities of FAEM (less than 1%).

 egeneration aids for particle filters R For the regeneration of modern particle filters, some systems use a metal-containing additive to facilitate the burning of the particles accumulated in the filter. In experiments, the iron compound ferrocene has proven to be particularly effective. A summary of the most important additives for diesel fuel and their purpose is provided in Figure 22.16.

 ntifoaming agents A The annoying foam of diesel fuel that arises while filling up can be largely suppressed by antifoam additives. They change the surface tension of the foam bubbles; that is, they loosen or destroy the boundary layers between them. These are usually liquid silicones that are added to the diesel fuel in very slight amounts (~0.001%).  dor improvers O To reduce the penetrating smell of diesel fuel, especially the annoying smell when filling up diesel passenger cars, aromatic substances are used.

 uper diesel additive package S Since 1987, leading market producers use additive packages for quality improvement which are being constantly adapted for new requirements. In addition to the so-called detergent improvers that keep the injection nozzles clean, the use of so-called ignition accelerators is essential as they raise the cetane number in some cases significantly above the minimum threshold of 51.0. These additives also improve operation in cold starting, during warmup and during typical daily driving and lower noise emissions. Furthermore, there is also greater protection against wear of the fuel injection system. Diesel fuel from different manufacturers can be mixed, but the balanced effect of the additives may be lost. 22.1.1.4 Alternative Diesel Fuels Every plant is a renewable raw material that is termed biomass. Some contain a particularly large amount of exploitable energy such as sugar beets, sugar cane, and rapeseed. By means of suitable conversion processes, these sun-fed primary energy sources can yield liquid secondary energy such as ethyl alcohol (ethanol) and rapeseed oil. In addition, biogas can also be generated. There are several reasons for the current interest in such “biofuels” that can be used in engines. The primary demand is to reduce dependence on fossil energy sources

820 | Internal Combustion Engine Handbook

6606_Book.indb 820

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Diesel Fuel Additive

Active Ingredient

Improved Characteristics

Advantage in Use

Ignition accelerator, combustion improver

Organic nitrates such as ethylhexyl nitrate

Cetane number

Cold start, white smoke, combustion noise, exhaust emissions, fuel consumption

Detergents

Amines, amides, succinimides, polyetheramines

Flow improver

Ethylvinyl acetates

Low-temperature behavior

Reliable operation at low temperatures to allow the use of paraffinic components with a high CN

Low-temperature behavior

Start, cold operation, storage

Clean nozzles, fuel consumption

Wax-antisettling

Alkylaryl amides

Lubricity

Fatty acid derivatives

Pump wear

Antifoaming agent

Silicone oils

Filling up

Corrosion protection

Oleamides, petroleum sulfonates, amino compounds

Protection of the fuel system in storage and in the vehicle

Figure 22.16 Summary of the most important diesel fuel additives and their purposes [22-7].

such as crude oil. With the use of biofuels, the CO2-discharge into the atmosphere via the closed carbon dioxide loop can be reduced. In addition, the growth of energy plants can be an alternative use of the fallow fields in Europe which was caused by overproduction in agriculture. The EU is providing corresponding tax incentives for biofuels to support this process. The focused reduction of greenhouse gas emissions is now the driving factor. These gases (primarily carbon dioxide— CO2) are considered to cause climate changes. In addition to the natural emission sources for CO2 that have existed forever, CO2 emissions from the combustion of fossil energy carriers are under scrutiny. A biofuel that is fundamentally suitable for operating diesel vehicles is rapeseed oil as an initial product for biodiesel.  io diesel B Its use is based on the idea that it produces only as much CO2 during combustion as is taken from the air when the rapeseed plant is growing. This is spoken of ideally as a closed CO2 loop that does not raise the CO2 concentration in the atmosphere. However, we must not forget that farming and converting the biomass also requires energy. In addition, biofuels are very expensive. When calculating the cost per ton of CO2 reduction for different biomass fuels or preventive measures, it becomes clear that other measures such as heat insulation and wind energy are much more economical. The suitability of rapeseed oil as a starting material for use in engines has been demonstrated in extensive experiments. However, it was quickly proven that pure rapeseed oil cannot be used without further modification. The fuel systems, engines, and engine oil have to be more or less extensively reconfigured. Investigations sponsored by the German Federal Ministry for Research and Technology have shown that most diesel engines in Germany cannot be directly operated with rapeseed oil. The technical problems primarily arise from the high viscosity. This causes the injection nozzles and piston ring grooves to coke, makes operation difficult at low temperatures, and worsens the atomization of the injected fuel. The poorer combustion causes a substantial increase in exhaust pollutants except just a slight increase of NOx. In addition, there is the familiar “fritting” odor from the exhaust that can be reduced by catalytic converters. Furthermore, the emission of aldehydes

and PAH is greater than with diesel fuel. Other problems include insufficient stability, low resistance to cold and poor elastomer compatibility. In addition, the inbuilt glycerides and glycerins can produce substantial deposits on the injection nozzle and in the combustion chambers. Rapeseed oil needs to be substantially modified for it to be useful in modern engines. This can be done either by esterifying it into RME (rapeseed oil methylester), by hydro-cracking in the refinery in a mixture with hydrocarbon refinery products or by the hydration of the plant oils themselves [hydrated vegetable oils (HVOs)]. The most important general minimum requirements for diesel fuel made of vegetable oil methyl esters are presented in Figure 22.17. The transesterification of rapeseed oil is carried out with methanol. Basically, this conversion improves cold resistance, viscosity and thermal stability. In addition, undesirable minor constituents are removed. RME is thus a clearly better alternative fuel for diesel engines than pure rapeseed oil. However, the conversion takes additional energy, resulting in a lessfavorable energetic balance of RME compared with pure rapeseed oil. The RME used as fuel must meet the requirements of DIN EN 14214. In addition, elastomer compatibility must be ensured for the vehicle. In the relevant emissions tests, RME has lower particle, PAK, HC, and CO emissions than diesel fuel. The NOx and aldehyde emissions are unfortunately higher. Further disadvantages include low performance, higher volumetric fuel consumption, and clearly higher production costs. Accordingly, RME requires large state subsidies for it to be cost competitive at the pump. Without subsidies, the production costs of RME are presently two to three times higher in comparison with diesel fuel. The most important characteristics of RME in comparison to typical diesel fuel are listed in Figure 22.18. The high cetane number and, compared to pure sulfur-free diesel, subsequent better lubricability are positive characteristics. The characteristics can be notably improved to approach those of diesel fuel using the cited, alternative path of processing rapeseed oil by hydro-cracking in the refinery mixed with hydrocarbon refinery products. Figure 22.19 shows the properties of pure rapeseed oil and diesel fuel in comparison to three fuels produced by different mixtures of rapeseed oil with vacuum gas oil with subsequent hydration in the hydro-cracker.

Internal Combustion Engine Handbook | 821

6606_Book.indb 821

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Properties

Unit

Threshold Value Minimal Maximal

Density at 15°C

kg/m3

860

900

EN ISO 3675

Kinematic viscosity at 40°C

mm2/s

  3.50

  5.00

EN ISO 3104

Flash point

°C

120

CFPP fitrability  15. 04. to 30. 09.  01. 10. to 15. 11 and  01. 03. to 14. 04.  16. 11. to 29. 02.

°C

Sulfur content

mg/kg

Cetane number

—

Ash content (sulfate ash)

% (m/m)

  0.02

ISO 3987

Water content

mg/kg

500

EN ISO 12937

Neutralization number

mg KOH/g

  0.50

EN 14110

Methanol content

% (m/m)

  0.20

EN 14110

Phosphorous content

mg/kg

10.0

EN 14107

Test Process

prEN ISO 3679 DIN EN 116

  0 −10 −20 −10 10.0

prEN ISO 20846

51.0

Composition % (m/m)

EN ISO 5165

Figure 22.17 Minimum requirements on FAME in diesel engines Extract from DIN EN 14214 [22-15].

Fuel

C

H

O

Density at 15°C kg/m3

Calorific Value BC MJ/l

Cetane Number (CFR)

RME (typical)

77.2

12.0

10.8

880

32.8

51–54

Diesel (typical)

86.6

13.4

0.4

835

35.5

51–55

R10, R20, and R30 indicate the rapeseed oil percentage in the end product. Of note are the gradual reduction of the sulfur content and the improvement of the CN. According to diesel requirement standard DIN EN 590, 5% FAME may be added to conventional diesel fuels. Before the backdrop of an EU Biofuels Directive that demands the addition of bio-generated components to fuels and a corresponding subsidization via tax breaks, up to 5% FAME have been added to diesel fuel in Germany since late 2003, first sporadically but, since 2005, as a rule. The German Bio Quota Act in force since 2007 bears major responsibility for this development. Compared to the EU directive, it demands higher bio portions for all fuels. Since

Figure 22.18 Comparison of characteristics of RME onrapeseed oil basis with diesel fuel [22-15].

2009, compliance with an overall quota was demanded which could not be met with the maximum 5% FAME content specified in EN 590. For this reason, the national diesel standard DIN 51628 was put in force in early 2009 which permits the admixture of up to 7% FAME. Likewise, rapeseed oil can also be added in the middle distillate desulfurization system (MDE). Figure 22.20 shows the properties of such fuels with 10%, 20%, and 30% rapeseed oil. In this case also, there are advantages for the sulfur content and the CN, as well as disadvantages for low-temperature behavior. We can see that the transesterification of rapeseed oil into RME produces an overall better end-product.

Characteristic

Unit

Diesel

DK-R10

DK-R20

DK-R30

Rapeseed Oil

Density

kg/m3

841.5

835.7

830.5

824.9

920.0

Sulfur content

% (m/m)

  0.19

  0.13

  0.09

  0.04

  0.01

CFPP

°C

−9

−7

−5

−2

16

Cetane number

—

54.5

59

63

66.5

41

Net calorific value (CB)

MJ/kg

42.82

42.98

42.84

43.23

37.40

Viscosity/20°C

mm2/s

  4.90

  4.99

  5.01

  5.01

73.5

Characteristic

Unit

Diesel Fuel R10*

Diesel Fuel R20*

Diesel Fuel R30*

Density

kg/m3

836.7

832.1

827.5

Sulfur content

% (m/m)

  0.13

  0.09

  0.04

CFPP

°C

−5

−4

−2

Cetane number

—

58

63

 69

Net calorific value (CB)

MJ/kg

42.92

43.06

43.11

Figure 22.19 Mixtures of diesel fuel and rapeseed oil (rapeseed oil in vacuum gas oil/hydrated) [22-15].

Figure 22.20 Diesel fuel– rapeseed oil mixtures after conversion in [22-15].

*The % rapeseed oil in middle distillate desulfurization.

822 | Internal Combustion Engine Handbook

6606_Book.indb 822

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

A further variant of the conversion of plant oils into fuel is direct hydration. The final products, so-called HVOs exhibit excellent properties for the applications and are, therefore, not limited in their admixture to conventional diesel fuel. It should also be mentioned that dimethyl ether (DME) (CH3)2O is a suitable component for diesel fuel. It arises in another processing step from methanol or, more recently, directly from natural gas or synthetic gas from other primary energies. DME is presently used as a liquid under pressure especially to replace FCKW as a propellant gas in spray cans. In summary, the technical feasibility of rapeseed oil as an alternative diesel fuel has been proven, although it requires substantial effort. Production costs are, however, prohibitively high. It can become affordable only if it is supported by large state subsidies. In addition, the limited availability of vegetable oil-based fuels makes their full replacement of fossil diesel fuel improbable.  iesel racing fuel D Today, special racing fuels are produced for high-performance diesel engines which are, because of their composition, capable of increasing the engine output. In endurance racing such as the 24-h race in LeMans or marathon rallies, it is important to attain maximum performance values even in the event’s final phase, in addition to keeping the injection system very clean. Very challenging in this context is the use of selected alternative or biogeneous components which are now available to a limited extent, such as gas-to-liquid (GTL), HVOs, or refined fatty acid methylesters.  lcohol–diesel fuel mixtures A Methanol or ethanol alone as an alternative “diesel fuel” has basic substantial disadvantages and requires substantial, expensive adaptations to the engine and fuel. To adapt diesel engines to pure alcohol requires, for example, a second injection system for dual-fuel operation. Diesel is used for cold starts, idling, and warmup, whereas alcohol is increasingly added as the load and speed increase. Other options are ignition assistance with glow plugs or spark plugs. Chemical ignition quality improvers as fuel additives have also been tested. They are expensive, however. Because alcohols have lower ignition quality and high evaporation heat, they must be correspondingly adapted as a fuel. A disadvantage for operation is the clearly lower calorific value (see Figure 22.10), which translates into inferior performance and higher fuel consumption. Particularly advantageous are the lower particle and NOx emissions. Mixtures of methanol or ethanol with diesel fuel are easier to use. Because methanol and ethanol are nearly impossible to mix with diesel fuel at room temperature, a large amount of solubilizers must be used, such as ethyl acetate in this concept. The stable mixture ranges can be determined from three-phase solubility diagrams for methanol with diesel fuel. Alcohols are technically feasible; but with today’s cost structure and tax load, they are not competitive. Another aspect that should not be underestimated is the change of the danger classification when diesel fuel is mixed with alcohol as the flash point and

the explosion limits are significantly less favorable. It must be further noted that the entire infrastructure, including the vehicles, is not designed for such a product which may result in, among other things, leaks because of adverse elastomer compatibility [22-40].  iesel–water emulsion D There are basic advantages to introducing water into the combustion process. In particular, nitrogen oxide formation is reduced from the decrease in the peak temperature as a result of internal cooling from water evaporation. The water can be introduced either through a second fuel injection system or by diesel and water emulsions. Whereas the first approach requires a substantial redesign of the engine and vehicle which would be the more effective path, diesel and water emulsions are much easier to realize in the vehicle. Tests have shown that with an increase in water content, NOx emissions and black smoke sharply decrease as expected; however, while HC and CO emissions increase. The higher HC emissions are particularly notable in the lower load range, which more than offsets the advantages gained in particle emissions. To attain realistic advantages for emulsions, a variable ratio of diesel fuel to water is required as a function of the working point, and this takes quite a bit of effort to achieve. Hence, the use of emulsions is more successful in stationary engines. Diesel and water emulsions are also more costly because additional wear protection is required for the fuel injection system; modern high-pressure fuel injection systems are particularly sensitive. The manufacturers of modern highpressure injection systems, in particular, oppose the use of emulsions in their systems. Furthermore, the insufficient long-term stability of the emulsion, especially at low temperatures, must be compensated for by additives, and the attack of microorganisms must be countered. The required emulsifiers mean higher fuel costs that, until now, have inhibited the widespread use of such products. Today, this technology is used occasionally in stationary machines and large engines (marine applications, for example).  ompressed natural gas in diesel engines C Compressed natural gas CNG (methane) is natural gas compressed to 200 bar for vehicle use. Figure 22.21 shows the physical characteristics of CNG in comparison with diesel fuel. [22-6] Physical Property

Diesel

CNG

Aggregate state in the tank

Liquid

Gaseous

Pressure in the tank

Atmosphere

200 bar

Density

830 kg/m3

170 kg/m3

Calorific value (CB) volume

34.7 MJ/L

7.2 MJ/L

Calorific value (CB) mass

42.0 MJ/kg

47.7 MJ/kg

Figure 22.21 The physical characteristics of CNG in comparison with diesel fuel [22-4].

One can see that even at 200 bar, the energy density in the gas tank is low. The low ignition quality of methane means that energy must be supplied to the diesel engine for ignition.

Internal Combustion Engine Handbook | 823

6606_Book.indb 823

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

The jet ignition system with two fuels can be used for this. Nevertheless, people prefer to convert diesel engines for city buses into spark-ignition engines to exploit the advantages of easier fuel storage (monofuel). The cylinder head and pistons are correspondingly altered, the injection nozzle is replaced with a spark plug, and high-tension ignition is used instead of the fuel injection pump. The compression ratio is reduced from 17.5:1 to 11.0:1. The use of such commercial vehicle approaches in metropolitan areas makes sense, given the emissions advantages. It is questionable if this approach will gain wider acceptance given, the higher energy consumption in comparison to diesel engines.

22.1.2 Gasoline

Gasolines have a boiling range of approximately 30–210°C and are provided for driving spark-ignition engines, primarily in the automotive sector. They consist of many hydrocarbons found in basic gasoline that is obtained in refineries by various processing methods from petroleum from a wide range of origins. They also contain slight amounts of other organic compounds and additives. The term used for many decades, CF for carburetor fuel, is outdated given the general use of fuel injection.  xplosion limits E The explosion threshold is always a factor in the consideration of gasolines. It describes the limits within which sudden combustion of an air–fuel vapor mixture occurs when an ignition source is activated. A distinction is drawn between a bottom threshold (slight amount of fuel vapor) and a top threshold (great deal of fuel vapor). At concentrations outside of these thresholds, no combustion can occur after ignition. Component

Density

Unit

Kg/m3

Octane Numbers MON

RON

Gasoline–air mixtures have a bottom explosion threshold of approximately 1% (V/V) fuel in air, and a top one of approximately 8% (V/V) fuel in air. When gasoline is stored, a very rich fuel–air mixture normally forms above the fuel that is far above the top threshold. Investigations have determined that fuels may exceed the top threshold given a minimum vapor pressure, low volatility, and low environmental temperature, which makes the fuel vapor–air mixture in the gas tank ignitable. 22.1.2.1 Gasoline Components and Composition Gasoline is one of the low-boiling components of the petroleum. It is a mixture of reformates, crack gasolines (olefins), pyrolysis gasolines, isoparaffins, butane, alkylates, and so-called replacement components such as alcohols and ethers. Figure 22.22 presents the basic characteristics such as density, octane number, and boiling behavior of the gasoline components in use today. The components methyl tertiary butyl ether (MTBE) and ethyl tertiary butyl ether (ETBE) are particularly important because they are required to manufacture SuperPlus. The alcohol mixture from methanol and tertiary butyl alcohol (TBA) used in the 1980s has been replaced by the use of bio ethanol. We will not address the antiknock lead compounds that occupied fuel research departments for decades, because they are now forbidden in nearly every part of the globe. Figure 22.23 shows the elementary composition of the portrayed components according to their paraffins, olefins, and aromates determined by fluorescence indicator absorption analysis. Figure 22.24 provides information on the quantity of the individual components in a typical gasoline from German refineries. The percentages of reformates and crack gasolines are approximately equivalent. The percentages of all the other components are much less, although they must all be present. E 70*

E 100**

% (V/V)

% (V/V)

Distilled gasoline

680

62

64

70

100

Butane

595

87–94

92–99

100

100

Pyrolysis gasoline

800

82

97

35

40

Crack light gasoline

670

69

81

70

100

Catalytic crack light

685

80

92

60

90

Catalytic crack heavy

800

77

86

  0

  5

Hydro-crack light

670

64

90

70

100

Full range reformate 94

780

84

94

10

40

Full range reformate 99

800

88

99

  8

35

Full range reformate 101

820

89

101

  6

20

Isomerizate

625

87

92

100

100

Alkylate

700

90

92

15

45

Polymer gasoline

740

80

100

  5

10

MTBE

745

98

114

100

100

ETBE

751

105

118

−10

120

Methanol/TBE 1:1

790

95

115

50

100

Ethanol

789

96

115

0

100

Figure 22.22 The basic gasoline components [22-15].

*Evaporated quantity at 70°C. **Evaporated quantity at 100°C.

824 | Internal Combustion Engine Handbook

6606_Book.indb 824

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Component

Paraffins*

Olefins

Aromatics

Unit

% (V/V)

% (V/V)

% (V/V)

Distilled gasoline

 94

 1

 5

Butane

100

—

—

Pyrolysis gasoline

approximately 20

approximately 10

approximately 70

Crack light gasoline

approximately 57

approximately 40

approximately 3

Catalytic crack light

 61

26

13

Catalytic crack heavy

 29

19

52

Hydro-crack light

100

 0

 0

Full range reformate 94

 45

—

55

Full range reformate 99

 38

—

62

Full range reformate 101

 29

 1

70

Isomerizate

 98

—

 2

Alkylate

100

—

—

Polymer gasoline

  5

90

 5

Figure 22.23 Elemental analysis of gasoline components [22-15].

*Including naphthene.

 lcohols and esters as replacement components A To compensate for lower octane numbers following the ban on lead, various ethers were discovered as essentially new gasoline components in addition to the further developed high-octane, classic components and alcohols. These are oxygencontaining hydrocarbon compounds in which a CH2 group is replaced by an oxygen atom. Ethers with at least five C atoms are suitable for gasoline. Figure 22.25 shows the most important physical characteristics of the alcohol components compared with those of super gasoline. Methanol and ethanol have been used at different times and different places in the history of motorized travel. Their use as alternative fuels is discussed in detail in Section 22.1.2.3. Ethers are distinguished by their favorable miscibility with gasoline without an azeotropic increase of the volatility and their being less sensitive to water. The high octane numbers and low vapor pressure are noteworthy. Because of the lower oxygen content in comparison to methanol and ethanol, the reduction of the calorific value is within tolerable limits in contrast to normal fuel components. Today, MTBE is manufactured on an industrial scale. Figure 22.26 presents the most important physical characteristics for

Typical component portions in gasoline from German refineries Replacement components (alcohols, esters)

Crack fuels Reformate

Pyrolysis gasoline Alkylt

Butane

i-pentane

Figure 22.24 Gasoline components in Germany [22-7].

Designation

Abbreviation

Boiling Point

Density 20°C

Vapor Pressure

°C

kg/m3

Kpa

RON

MON

Net Calorific Value (CB)

Evaporation Heat O2 Content

MJ/l

kJ/kg

% (m/m)

Methyl alcohol

Methanol

64.7

791.2

32/81*

114.4

 94.6

15.7

1170

49.93

Ethyl alcohol

Ethanol

78.3

789.4

17/70*

114.4

 94.0

21.2

 880

34.73

Isopropyl alcohol

Isopropanol

82.3

775.5

14/72*

118.0

101.9

23.6

 700

Sec. butyl alcohol

SBA

99

806.9

Unit

27.4  90.1

26.63 21.59

Isobutyl alcohol

IBA

107.7

801.6

4/63*

110.4

26.1

 618

21.59

Tert. butyl alcohol

TBA

82.8

786.6

7/64*

approximately approximately 26.8 105 95

 589

21.59

Super gasoline

Super gasoline 30 to 210 720775

S 45.0–60.0 W 60.0–90.0

95

 85.6

approximately 380–500 41

0–2.7

*As a mixture component in gasoline (10%).

Figure 22.25 Most important physical characteristics for alcohol components in comparison to super gasoline [22-15].

Internal Combustion Engine Handbook | 825

6606_Book.indb 825

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Designation

Abbreviation

Unit

Boiling Point

Density 20°C

Vapor Pressure

°C

kg/m3

kPa

RON

MON

Net Calorific Value (CB)

O2 Content

MJ/kg

% (m/m)

Methyl tertiary butyl ether

MTBE

55.5

740

48

114

 98

26.04

18.15

Ethyl tertiary butyl ether

ETBE

72.5

742

28

118

102

26.75

15.66

Diisopropyl ether

DIPE

68.5

725

24

110

100

26.45

15.66

Tertiary amyl methyl ether

TAME

85.5

770

16

111

 98

27.91

15.66

Isopropyl tertiary butyl ether

PTBE

88.5

740

20

27.46

13.77

Super gasoline (typical 1999)

Super gasoline

30–215

725–780

60–90

approximately 41

0–2

 95

 85

Figure 22.26 Most important physical characteristics for ether components in comparison to super gasoline [22-8].

ether components with those of super gasoline. The other cited ethers except ETBE and TAME are rarely used as fuel components because of their high production costs. The use of particularly valuable, oxygen-containing replacement components, such as ethers and alcohols, is restricted to a relatively low amount given the EU limitation of the overall O2 content in gasoline of 2.7% (m/m). For example, the ethanol content is limited to 5% (V/V), and the methanol content is restricted to a maximum 3% (V/V) if additional suitable solubilizers—usually TBA—are added to the methanol. Figure 22.27 shows permitted use of oxygen-containing components according to the EU directive and, in brackets, the planned expansion. This limitation was to exclude undesirable side effects on elastomers, prevent separation at low temperatures, and prevent excessive lean adjustment that can disturb driving behavior.

Component % (V/V)

Permissible Components According to DIN EN 228

Methanol

3

Ethanol

5 (10)

IPA

10 (12)

TBA

7 (15)

IBA

10 (15)

Ether*

15 (22)

Others**

10 (15)

Oxygen content % (m/m)

2.7 (3.7)

*MTBE, TAME and ETBE as well as others with min. 5 C atoms. **Other monoalcohols (intentional increase).

Figure 22.27 Maximum concentration of O2-containing components (EU) [22-15].

MTBE/ETBE has proven itself especially in SuperPlus as a replacement for lead-based antiknocking agents to increase the octane number. Now in Germany, SuperPlus and even

higher quality types contain an average of approximately 10% MTBE/ETBE.  asoline types G Now in Germany, there are three types of unleaded fuel: Regular since 1985, (Euro) Super since 1986, and SuperPlus since 1989. For older vehicles without a catalytic converter, leaded super was available up to 1996, whereas leaded regular gasoline was removed from the market in 1988. In nearly all European countries, regular gasoline cannot be found since all automobile manufacturers were pressured to offer minimum fuel consumption and CO2 emissions. Almost all of their engines are, therefore, designed to run on Super or SuperPlus. Figure 22.28 shows the percentages of the three types of gasoline on the German fuel market. One can see that SuperPlus, after an initial success, probably because of greater cost, has not been widely accepted. In addition to conventional gasoline based on petroleum, a series of other synthetic substances are suitable for combustion in spark-ignition engines. In addition, there are several possibilities for using alternative gasolines. These will be discussed further in Section 22.1.2.3. 22.1.2.2 Characteristics and Properties The minimum requirements for the three cited lead-free gasolines are found in DIN EN 228. As can be seen in Figure 22.29, they primarily concern density, antiknock quality, the boiling curve, vapor pressure, benzene content, and sulfur content. Because of the environmentally relevant importance of some fuel characteristics in practice, a compromise was reached to bridge the different perspectives of the automobile and mineral oil industries on a European basis (EU Commission) in 1998 within the framework of the Auto-Oil program that further stiffened the standards for environmentally relevant fuel characteristics in a first step in 2000 and in a second step in 2005. For gasoline, this primarily concerned the sulfur,

Fuel Type

1994

1995

1996

1997

1999

2000

2002

2005

2008

Gasoline

39.4

38.4

37.6

 36.9

 34.3

 33.4

 30.9

 28.0

 10.6

Super

46.9

50.7

54.4

 47.3

 61.1

 62.8

 65.3

 68.1

 85.4

SuperPlus

 6.0

 5.4

 5.3

  5.8

  4.6

  3.8

  3.8

  3.9

  4.0

Lead-free overall

92.3

94.5

97.4

100.0

100.0

100.0

100.0

100.0

100.0

Super leaded

 7.7

 5.5

 2.6

—

—

—

—

—

—

Figure 22.28 Percentages of gasoline types consumed in Germany (%) [22-5].

826 | Internal Combustion Engine Handbook

6606_Book.indb 826

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Requirements According to DIN EN 228 Characteristic

Unit

SuperPlus

Density at 15°C

kg/m3

720–775

Super

Normal

Antiknock quality RON

min. 98.0

min. 95.0

min. 91.0

MON

min. 88.0

min. 85.0

min. 82.5

Lead content

mg/L

Boiling curve *

% (V/V)

max. 5

Vaporized quantity (Class A)  at 70°C, E70

20.0–48.0

 at 100°C, E100

46.0–71.0

 at 150°C, E150

Min. 75.0

Vaporized quantity (Class D/D1)  at 70°C, E70

22.0–50.0

 at 100°C, E100

46.0–71.0

 at 150°C, E150

Min. 75.0

Final boiling point FBP (Class A/D/D1)

°C

Max. 210

Volatility index VLI** (VLI = 10 × VP + 7 × E70) Class D1

Index

max. 1150

Distillation residue

% (V/V)

max. 2

Vapor pressure (DVPE)

kPa

 Class A

45.0–60.0 (summer)

 Class D/D1

60.0–90.0 (winter)

Evaporated residue

mg/100 ml

max. 5

Benzene content

% (V/V)

max. 1.0

Sulfur content

mg/kg

max. 10

Oxidation stability

min.

Min. 360

Copper corrosion

Degree of corrosion

max. 1

Figure 22.29 Gasoline characteristics according to DIN EN 228 [22-4].

*Sulfur content in entire EU since early 2009 maximal 10 mg/kg.  Class D: 16.11.–15.03. (winter).  Class D1: 16.03.–30.04./01.10.–15.11. (transition). **Vapor lock index.

benzene and aromatic contents. The exhaust thresholds were established correspondingly as Euro III starting January 1, 2000 and as Euro IV starting January 1, 2005. Figure 22.30 shows the changes in gasoline. In addition to these standards, the global automotive industry has developed the so-called fuel charter (WWFC) defining the requirements for fuels in four quality levels. In meeting the top quality level (category 4),

Characteristic

Unit

DIN EN 228 until 1999

Euro III from 2000

Euro IV from 2005

Sulfur

mg/kg

500

150

50

Benzene

% (V/V)

  5

  1

 1

Aromatics

% (V/V)

—

 42

35

Vapor pressure

kPa

 70

 60

60

Olefins

% (V/V)

—

(21) 18

18

Figure 22.30 Result of the EU Auto/Oil program for gasoline [22-4].

the automotive industry sees potential for the development of future engine concepts. The changes to the European exhaust laws are shown in Figure 22.31.  ensity D The ranges of the density for all three unleaded gasolines are set to a uniform 720–775 kg/m3 at 15°C. Figure 22.32 shows averages and the ranges of density of commercial German gasoline for summer and winter. As the density increases, the volumetric energy content of the fuel also generally rises, which is related to falling volumetric fuel consumption. Based on experience, a rise in density of 1% equals a drop in volumetric fuel consumption of 0.6%. One can see that the values in the summer are all higher, and that there are advantages in fuel consumption for Super and especially for SuperPlus.

Internal Combustion Engine Handbook | 827

6606_Book.indb 827

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

91/441/EEC Euro I

Pollutant

94/12/EC Euro II

98/69/E* Euro III

98/69/EC* Euro IV

Engine

in g/km

from 1992

from 1996

from 2000

from 2005

Gasoline

CO HC + NOx HC NOx

3.16 1.13

2.20 0.50

2.30

1.000

0.20 0.15

0.100 0.080

CO HC + NOx particles

3.16 1.13 0.18

1.00 0.70 0.08

0.64 0.56 0.05

0.500 0.300 0.025

Diesel

Figure 22.31 Development in European exhaust laws (passenger cars) [22-4].

*Changed (tightened) test process.

Density in kg/m3 Range Average

SuperPlus

Super

Normal

Summer

733–756

736–630

729–758

Winter

732–754

724–758

721–748

Summer

748

745

743

Winter

741

735

729

Combusted mixture

Uncombusted mixture

Source: Market monitoring, Winter 2007/2008 and Summer 2008 in Germany.

Normal combustion Targeted ignited flame burns constantly due to normal combustion

Figure 22.32 Density of German commercial gasoline types [22-4], [22-15].

Cylinder pressure [bar]

Cylinder pressure [bar]

 ntiknock quality A The antiknock quality of gasolines is their ability to prevent undesired combustion, that is, combustion not triggered by the spark plug or uncontrolled combustion of the uncombusted residual exhaust gas before it meets the flame front. Depending on the fuel composition and design, the flame front passes through the charge at a propagation velocity of more than 30 m/s. Figure 22.33 schematically compares normal combustion with knocking combustion. In knocking operation, the combustion speed is approximately ten times faster, causes steep pressure peaks and cavitation-like pressure fluctuations, and is accompanied by a substantial increase in the combustion chamber temperature. A simplified comparison of the corresponding pressure/ time diagrams is shown in Figure 22.34.

Ignition system OT Time Normal combustion

°CA

Combusted mixture

Undesired autoignition

Knocking combustion Several flame fronts, that is, pressure waves, meet and "knock"

Figure 22.33 Normal and knocking combustion [22-8].

Knocking

E

Ignition system OT Time Knocking combustion

°CA

Figure 22.34 Pressure/time diagram [22-12].

828 | Internal Combustion Engine Handbook

6606_Book.indb 828

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

If knocking is continuous, the spark plugs, pistons, cylinder head seals, and valves become damaged or even destroyed, especially when preignition occurs. In Figure 22.35, we see a piston that was destroyed from continuous knocking.

Figure 22.35 Piston destroyed by continuous knocking [22-8].

Modern engines are largely protected from such mechanical damage by the use of knock sensors—structure-borne sound sensors or ionic current meters. They delay the moment of ignition when knocking starts, reduce the charge pressure during charging, or throttle the intake air. In vehicles with knock control, the ignition map is electronically adapted to the fuel in the tank. However, when the antiknock quality is lower than that specified by the manufacturer, the later ignition setting causes a drop in performance, higher fuel consumption, and a higher thermal load on the catalytic converter. Conversely, a transition from Super to SuperPlus can increase performance with an earlier ignition setting in conjunction with lower fuel consumption and emission advantages. In determining the required antiknock quality of an engine, a distinction is drawn between acceleration knock and high-speed knock. Whereas acceleration knock as a transient condition is not that dangerous at a low speed and load, continuous high-speed knocking at a high speed and full load can be hazardous enough to cause engine damage.  ctane number O The octane number is a measure of the antiknock quality of a gasoline. A distinction is drawn between the minimum requirements of research octane number (RON) and engine octane number (MON). Both terms are based on traditional names from American fuel research that do not have a logical correspondent. In practice, the street octane number (SON) is also relevant. For earlier carburetor engines, the front octane Engine Speed (min–1)

Intake Air (°C)

Mixture Preheating (°F)

number (FON) or the RON 100 (corresponding to the RON of the fuel components boiling at 100°C) was used. Whereas RON and MON are measured in special CFRs by changing the compression ratio, the SON is determined in production vehicles by advancing the moment of ignition. The test according to the MON method is based on speed, moment of ignition, and mixture preheating under harder conditions; the MON is, therefore, always lower than the RON. In practice, this means that engines under high thermal stress—which today is practically every single one—have a minimum requirement for the MON of a fuel in addition to the RON. The RON–MON difference is termed the “sensitivity” and should not exceed the value of 10. Figure 22.36 portrays the operating conditions when determining the RON and MON in the CFR test engine.  ctane number scale O The octane number scale extending from 0 to 100 is dimensionless. The 0 represents the particularly knock-susceptible reference fuel, regular heptane (C7H16), and 100 is the particularly knock-resistant reference fuel, isooctane (C8H18), also termed 2,2,4-isopentane [C5H9(CH3)3].The ON of a fuel is determined in a comparative test between the fuel sample and i-octane/n-heptane mixtures. The compression ratio is first increased in the CFR test engine until the sample starts knocking. Then the associated ON is calculated by maintaining a constant compression ratio and changing the mixtures of i-octane and n-heptane until the engine starts knocking. The knock limit is determined with the aid of an electronic knock sensor. For example, RON 95 means that the gasoline measured in the CFR test engine using the research method acts like a mixture of 95% i-octane and 5% n-heptane when it reaches the knock limit.  ixed octane number M The range of the ON scale ends by definition at 100 (isooctane). For fuels with an ON above 100, the following procedure is used to calculate the ON. The high-octane fuel is mixed with a portion of 10% or 20% (V/V) in a gasoline with significantly lower and known ON. The ON of this mixture is then measured and the “Mixture ON” of the high-octane added portion is calculated from the attained improvement of the low-octane fuel using this formula: Mixture ON = ( M − (K ⋅b/100)) /(a/100)



(22.3)

where M = ON of the diluted mixture K = ON of the low-octane gasoline a =%M b = % K.

Ignition Timing [°CA before TDC] Compression Ratio

RON

600

51.7 ± 5

—

13

Variable 4–16

MON

900

38

Variable 285–315

Variable 14–26

Variable 4–16

Figure 22.36 Operating conditions of the CFR test engine [22-4].

Internal Combustion Engine Handbook | 829

6606_Book.indb 829

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Example: If one mixes 90% (b) of a gasoline with an ON of 85.5 (K) with 10% (a) of the unknown high-octane fuel and obtains an ON of 88.3 (M) during measurement, the mixture ON of the high-octane fuel is 113.5. Practice has shown that it is useful to perform several mixtures with different admixture ratios (10%, 20%, and 50%) to obtain an estimate as precise as possible of the actual octane number of the high-octane fuel. However, this method returns reliable results only when hydrocarbons of the same type are mixed which limits the application. Since 1956, the Wiese scale has been used as a practical method (DIN 51788). Starting with i-octane, increasing amounts of TEL (tetra ethyl lead) are added. The method corresponds to the performance number used for airplane fuels. Figure 22.37 shows the numeric values for the relationship between octane numbers > 100 and the respective TEL added to i-octane.

opportunity to assume a high temperature because of reduction processes without absorbing heat. The combustion that occurs at a nearly constant volume is also advantageous for the quality of thermal efficiency. The general return to longstroke engines improves the exhaust and further increases the compression ratio. For the combustion chamber design to have a very low octane number requirement with a high efficiency, the following points are observed: •• compact combustion chamber with a very low surface-tovolume ratio (spherical cap or roof-shaped) •• central location of the spark plug in the combustion chamber to attain equally long flame paths (four valves) •• large squish area from piston overlap with minimum thickness (generates turbulence) •• energetic charge movement •• strong cylinder head cooling.

ON

% (V/V) TEL

ON

% (V/V) TEL

100

0.0000

111

0.0399

101

0.0020

112

0.0468

102

0.0042

113

0.0546

103

0.0066

114

0.0634

104

0.0092

115

0.0734

105

0.0124

116

0.0850

106

0.0158

117

0.0963

107

0.0195

118

0.1133

108

0.0238

119

0.1308

109

0.0285

120

0.1509

110

0.0338

Figure 22.37 Wiese scale for ON > 100 [22-4].

 ctane number requirement O The octane number requirement of an engine is measured on an engine test bench within the overall speed range under a full load. A knock limit curve map arises in which is entered the spark advance characteristic determined by the manufacturer. The octane number requirement then results from the intersections of the knock limit curves with the ignition map, which allows the maximum to be read immediately. Usually it lies around the maximum torque, that is, the maximum average pressure.  ngine design and octane number requirement E From the vantage point of the engine, the octane number requirement is chiefly determined by the compression ratio. Given geometrically similar combustion chambers, an increasing piston displacement corresponds to a decrease in the knock limit compression. Hence, larger cylinders are more knock-sensitive. To a certain degree, an oversquare cylinder (s/D < 1) has a higher octane number requirement with otherwise equivalent dimensions than an undersquare engine (s/D > 1). In both cases, the path traveled by a flame during combustion plays a role. The connecting rod ratio r/l is important because a higher r/l maintains the efficiency of the piston overlap (squish area) while the overall combustion is approximately maintained. The end gas, therefore, has no

In summary, it can be stated that the best results are obtained when the area filled by the mixture upon the moment of ignition is as close to the spark plug as possible. The valve timing also influences the relative knocking sensitivity of an engine. For example, a large valve overlap reduces knock because of its influence on the residual exhaust gas and mixture temperature. Early closing to increase torque in the lower speed range can raise the octane number requirement. Today’s wide use of light metal and the practically uniform lack of air cooling have a positive effect.  perating conditions and octane number requirement O The octane number requirement largely depends on the operating conditions. These are largely influenced by the state of the intake air, excess-air factor, speed, moment of ignition, volumetric efficiency, load, and the coolant temperature. When the pressure and temperature of the intake air rise, this raises the octane number requirement, whereas increasing humidity decreases the octane number requirement. The octane number requirement is highest at the stoichiometric air–fuel factor. A richer or leaner mixture does not produce the pressure and temperature required for knocking because of the lower velocity of combustion. The temperature of the uncombusted fuel–air mixture is analogous. A rising speed generally translates into a rapid decrease of the octane number requirement because the piston is already moved away from the top dead center at a time critical to the end gas; the combustion chamber volume therefore increases, and the compression of the end gas correspondingly decreases. Furthermore, at high speeds, combustion is quicker because of the large turbulence generated in the combustion chamber. The throttling loss and lower compression end pressure also have the same effect. The moment of ignition naturally directly influences the octane number requirement. The earlier it occurs (far before TDC), the earlier combustion begins in the piston travel of the compression cycle which compresses the end gas. In general, spark-ignition engines tend to knock, especially when the throttle valve is fully open, that is, under a full load, because it is at this point that the maximum combustion pressure arises from the greatest cylinder charge. The maximum octane

830 | Internal Combustion Engine Handbook

6606_Book.indb 830

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

number requirement is usually at the speed where there is maximum torque (average pressure) because the volumetric efficiency, moment of ignition, and excess-air factor interact to promote knocking. As the coolant and oil temperature rise, the octane number requirement naturally increases because the critical conditions for spontaneous, undesired combustion of the residual exhaust gases are enhanced. On average, the octane number requirement increases by 1 for each 5°C rise in coolant water temperature. The influence of the oil temperature is somewhat less. Combustion chamber deposits and octane number requirement While an engine is operating, combustion chamber deposits form which cover its surface, the piston head, the valve head, and the spark plug. They arise both from the fuel and the lubricant. Soot arises from the fuel because of incomplete combustion in the idling and warm-up phases. Cracked or coked oil components that remain on the top piston ring in the combustion chamber come from the lubricant or via the valve guides. Ash-forming additives can also cause deposits. The deposits raise the octane number requirement by reducing the combustion chamber volume; that is, they increase the compression ratio and insulate against heat. The knocking tendency rises quickly in a new, clean engine to a maximum until the deposit equilibrium is reached. In practice, this rather unstable equilibrium arises after approximately 10,000–20,000 km. The rise of the octane number requirement from deposits in city traffic has been greatly reduced because of the transition to unleaded fuels. Of course, driving style also plays a large role. When all the factors combine, the octane number requirement can rise by seven from a new engine to the deposit equilibrium, even when state-of-the-art operating fluids with additives are used. For example, an engine designed for regular gasoline operates without knocking only when it uses super.  treet octane number S Although the antiknock quality of a fuel provides information on the practical knocking behavior expected in a given automobile by determining the RON and MON, it is difficult to assign the laboratory octane number to actual street behavior. For example, different fuels with the same RON can produce widely different knocking behavior in the same automobile because of the many cited factors that can influence the octane number requirement. To precisely monitor these conditions, mineral oil researchers use test methods to determine the

octane number that actually occurs on the street, the SON. In this case as well, the produced fuels are compared with the familiar reference fuels. The measurements are carried out either on suitable automobile test benches or on engine test benches. In comparison to the CFR test engine, the measuring range is greatly limited. Meaningful values can be measured only from approximately 10° to 15° crank angle around the basic spark adjustment set by the automobile manufacturer. This approximately corresponds to a bandwidth of 5–6 ON. For earlier carburetor engines, the CRC F-28 method was used, the “modified Uniontown method,” that is basically the same as determining the octane number requirements for acceleration knock. As expected, the limit curves rise sharply because the knocking tendency falls quickly, and the octane number requirement drops rapidly with increasing speed. If the laboratory octane numbers RON and MON provide only a limited amount of information on the actual behavior of the fuel in practice, they are still a useful yardstick for indicating the interaction between the engine and fuel. The SON usually lies between the RON and MON. At low speeds, it tends toward the RON; at high speeds and with a large amount of residual exhaust gas, it tends toward the MON. The SON–RON difference has proven useful for purposes of comparison. It is termed the street evaluation number (SEN). The advantage of this nomenclature is that the SEN is positive when the SON exceeds the RON which is usually the case, and it is negative when the SON is less than the RON. The positive and negative signs then can be used to directly evaluate a fuel in a given automobile or engine. A positive SEN also indicates that the relevant engine is evaluating a fuel with less “severity” than the CFR test engine in the RON method and vice versa: when the SEN is negative, the engine is providing a more severe evaluation than the CFR engine. By establishing a minimum MON in addition to the RON in the standards, many earlier used mixture components with a low MON were excluded. In addition, the general use of multipoint fuel injection has eliminated the earlier predominant sensitivity of engines to an uneven distribution of the octane numbers over the boiling range of the fuel. In general, this has eliminated the necessity of determining the SON, and it remains relevant only for research purposes. The influence of a few fuel components on the SON is shown in Figure 22.38. It can be seen that light distillate and light and heavy crack gasoline negatively influence modern fuel injection engines.

Property Octane Numbers Component

RON

MON

Boiling Characteristics

Influence on SON Acceleration Carburetor Engine

High Load and Speed

Light distillate

Low

Low

Highly volatile

Negative

Negative

Butane i-pentane/i-heptane

High

High

Highly volatile

Positive

Positive

Light crack gasoline

High

Low

Highly volatile

Positive

Negative

Heavy reformate

High

Middle/High

Nonvolatile

Negative

Positive

Heavy crack gasoline

Medium

Low

Nonvolatile

Negative

Negative

Figure 22.38 Influence of a few fuel components on the SON [22-8], [22-12].

Internal Combustion Engine Handbook | 831

6606_Book.indb 831

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

 oiling curve B When conducting a boiling analysis according to DIN EN ISO 3405, the fuel specimen is evaporated and then condensed with variable heat output and a fixed temperature increase of 1°C/min. The resulting boiling curve contains a great deal of information for application engineering. Well-balanced boiling behavior is an essential prerequisite for operating automobiles with spark-ignition engines under every condition. The meaning of the boiling curve and its individual sections is shown in Figure 22.39. For example, light, that is, low-boiling, components are responsible for quickly starting cold engines, good response and low exhaust emissions during the warm-up period. Too many of them, however, can lead to vapor bubble formation and greater evaporation loss in the summer. In wet, cold weather, throttle valve icing can occur. Too many high-boiling components, on the other hand, can lead to condensation on the cylinder walls in cold operation and dilute the oil film and oil supply. Too few components in the middle boiling range impairs drivability and may cause “jolting” during acceleration. After a hot engine is turned off and quickly restarted, the demands on the fuel are precisely the opposite. Under unfavorable conditions, parts of the fuel system can become so hot that a large portion of the fuel evaporates, which causes vapor bubbles in the fuel pump or vapor cushions in the fuel injection lines. In hot starting in particular, the opening of the injection nozzles and resulting sudden pressure drop may cause gas

Consumption/emission 200 180 160 140

Hot driving

120 100

Oil dilution Residues – in oil – Spark plugs – Combustion chamber

Evaporation losses

80 60 40 20

Cold starting

Cold start

 oiling curve (distillation) B The boiling behavior or volatility is determined by the boiling curve and the vapor pressure. In addition to the antiknock quality, it is the most important evaluation criterion for gasolines that change into a vaporous state between 30 and 210°C.

Boiling curve and its influence on engine behavior

Boiling temperature in °C

 ront octane number F For the sake of completeness, we mention the FON that is no longer relevant today. It provides information on the RON of the components of the gasoline that boil below 100°C. It was especially useful for carburetor engines with long intake ports. Because only the light components reach the combustion chambers when the throttle valve is quickly opened, it had to be ensured that within this boiling range there were enough knock-resistant components available. In the low boiling range, except butane, the other light components such as distillate and reformate gasoline generally have a moderate ON. The antiknock quality of the front boiling range was, hence, too low in comparison to overall fuel. To deal with this fuel problem, high-octane light components such as isomerisates, catalytic crack gasoline, and alcohols were used. The earlier used lead compounds were also adapted by introducing highly volatile lead tetramethyl instead of lead tetraethyl. Given the general transition from carburetor to individual cylinder fuel injection with the precise metering and preparation of the mixture under transient conditions, the FON became irrelevant and was, therefore, withdrawn from the standard.

0

20

40 60 Distillate in Vol.-%

100

80

Figure 22.39 Boiling curve and its influence on engine behavior [22-8].

bubble formation which makes starting the engine difficult or even impossible. Figure 22.40 shows the opposite requirements for cold starting and hot driving. Influence of fuel volatility on cold-start and hot-driving behavior better Hot driving

Driving behavior Cold starting worse falling

Fuel volatility

rising

Figure 22.40 Influence of volatility on cold-start and hot driving behaviors [22-8].

The EN standard (see Figure 22.29) defines six different volatility classes that cover geographic and yearly changes in the weather. Figure 22.41 shows examples of typical German ON winter values for E70, E100, and E150. In addition, Figure 22.42 provides a comparison of ON values for the end of boiling of German gasoline.

832 | Internal Combustion Engine Handbook

6606_Book.indb 832

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

SuperPlus % (V/V) E70 E100 E150

Super % (V/V)

Regular % (V/V)

Standard Range Class D % (V/V) 22–50

Average

36

35

37

Range

29–46

30–47

29–48

Average

55

54

58

Range

48–62

50–63

50–67

Average

87

86

87

Range

78–93

79–94

76–98

SuperPlus

Super

Regular

Range

176–210

172–210

162–208

Average

194

193

190

Figure 22.42 Final boiling point of German gasoline [22-8].

 apor pressure V The pressure that arises in a sealed container as a function of temperature from the evaporation of fuel is termed the vapor pressure. It influences (sometimes in connection with other volatility criteria) cold and hot starting, cold driving behavior, and evaporation loss. It is basically determined by the light components such as butane toward the initial boiling point. Until 1993, the “wet” Reid method was specified in standard DIN 51754 (RVP = Reid Vapor Pressure) at a test temperature of 37.8°C (100°F) and a vapor–liquid ratio of 4:1. Because the “wet” Reid method returned too low (noncritical) vapor pressure values for alcohol-containing fuels, the testing method was changed, within the European standardization process, to the “dry” determined RVP according to DIN EN 12 to be used until 1999. With the amendment of DIN EN 228 on February 1, 2000, the method used to determine the vapor pressures also changed. The Reid method was replaced with the generally applicable DVPE (Dry Vapor Pressure Equivalent) according to DIN EN 13016-1. The DVPE is calculated from the ASVP (Air Saturated Vapor Pressure), determined with a Grabner test setup for example, with a vapor–liquid ratio of 4:1. By changing DIN EN 228, the volatility classes also changed. For the first time, two transition periods were provided between winter and summer quality (see Figure 22.29). In addition to the vapor pressure, a vapor lock index (VLI) value limits, as additional parameter in four of the six volatility classes of the EN standard, the fuel volatility correlating with hot starting and hot driving behavior. This value is calculated from the formula 10 ⋅ RVP + 7 ⋅ E70 and has proven its viability in carburetor engines in particular. Because the fuel in modern fuel injection engines is exposed to high temperatures, particularly before and in the nozzles, an additional measuring method, also on the basis of Grabner's test setup, was worked out within a broader measuring range: +40 to 80°C (DIN EN 13016-2). With this "dry" method, the vapor–liquid ratio is 3:2.In particular, it indicates the azeotropic increase in vapor pressure in the measurement of the vapor pressure of methanol-containing gasoline hotter than >38°C.

46–71 min 75

Figure 22.41 Distillation values for German winter gasoline [22-8].

The VP in this instance is clearly higher than that of gasoline without alcohol. This method is used predominantly in the development of fuels. A limit value in the requirement standard restricting volatility at high temperatures (80°C) has not yet been established. In general, a vapor pressure that is too low, that is, a fuel that evaporates too slowly, results in unsatisfactory starting and cold driving behavior, whereas a vapor pressure that is too high produces problems in hot starting and hot driving behavior. In addition, the formation of an air–vapor mixture when storing fuel above the top explosion point requires a sufficiently high vapor pressure to be safe. The “true” vapor pressure at 50°C is found in the transport regulations. It applies to a vapor–liquid ratio of 0:1 and is calculated from the RVP.  enzene content B Benzene (C6H6) is the basic element of aromatic hydrocarbons. Because of its high ON (RON and MON > 100) and availability from coke manufacture, it was earlier used as an important component in super gasoline. However, this was engine benzene, a mixture of benzene, toluene, and xylene (see Figure 22.1), the secret of the first super fuel of the world “ARAL” (aromates/aliphates) marketed in 1924, a product whose antiknock quality and other qualities was clearly superior to gasoline. After introducing catalytic reformers in the 1950s, the use of engine benzene generated from coke production in Germany became increasingly less important. After the health risks from handling benzene became known, benzene largely stopped being used as an additive, especially because one became aware of other ways of replacing undesirable leadbased antiknocking agents. However, other aromates continue to play a large role in modern gasoline. In EU standard 228 for gasoline, the benzene content was limited for a long time to a maximum 5% (V/V). In the market, it was an average of 2% (V/V) and in SuperPlus (since 1995) even 1% (V/V). Since January 1, 2000, the established maximum threshold for all gasoline qualities is 1% (V/V). Figure 22.43 shows the changing benzene content in German gasoline from 1986 to 2008. However, many other aromates are also used in fuels. Figure 22.44 gives an overview of the aromates used in gasoline. Aromates already exist in petroleum, but most are produced by catalytic reformers with the release of hydrogen. Figure 22.45 provides information on the aromate content in German gasoline.

Internal Combustion Engine Handbook | 833

6606_Book.indb 833

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Year

SuperPlus

Super

Regular

Super Leaded

Average

1986

—

2.8

2.4

2.8

2.4**

1988

—

2.6

2.2

2.8

2.6****

1990

2.6

2.8

2.2

2.7

2.4

1992

2.4

2.5

1.8

2.5

2.2**

1994

2.0

2.1

1.6

2.3

1.9**

1996

0.9

1.9

1.5

1.6**

1998

0.8

1.6

1.4

1.5**

2000***

0.6

0.8

0.8

0.8**

2003

0.5

0.7

0.8

0.7**

2008

0.6

0.7

0.7

0.7**

Figure 22.43 Change of the benzene content [22-15].

*% (V/V). **Ban of leaded Regular. ***Only maximal 1.00% (V/V) benzene permitted in fuels.

Product

Total Formula

Boiling Point Boiling Range

Mixture ON RON

Mixture ON MON

Toluene

C7H8

110°C

124

112

Ethyl benzene

C8H10

136°C

124

107

Xylenes

C8H10

138–144°C

120–146

103–127

C9 aromates

C9H11

152–176°C

118–171

105–138

C9 + aromates (small quantities)

C10H12 C11H13

169–210°C

114–155

117–144

% (V/V)

SuperPlus

Super

Regular

Toluene

13.1

10.5

9.8

Xylenes

12.7

11.0

11.4

C8 + Ar

12.7

12.2

12.8

Figure 22.45 Aromate contents of German gasoline (average 1994) [22-15].

It is interesting to compare the olefin content in German gasoline as shown in Figure 22.46. We can see that it strongly decreases as the MON requirement increases. Olefin Content % (V/V)

SuperPlus

Super

Regular

Range

0–17

1–22

1–37

Average

4

10

18

Figure 22.46 Olefin contents of German gasoline [22-15].

 ulfur content S Sulfur occurs in petroleum almost exclusively in bound form as mercaptan sulfur, disulfide sulfur, thiophene sulfur, and so on. Mercaptans (thioalcohols) are sulfur derivatives of alcohols in which the oxygen of the hydroxyl group OH is replaced by S. In petroleum, the S content ranges from 0.01% to 7.0%. In fuels, a high S content has always been undesirable, and sulfur has, therefore, been removed in the refineries as costs allow. Apart from SO2 emissions, a few catalytic converters, especially uncontrolled catalytic converters, tend to convert sulfur into hydrogen sulfide (H2S) that smells under certain operating conditions. In addition, the catalytic converter efficiency decreases as the S content of the fuel raises, which

Figure 22.44 Aromates used in gasoline [22-15].

correspondingly increases the emission of CO, HC, and NOx, which can have serious consequences, particularly in storage catalytic converters. According to the quality standard DIN EN 228, gasoline in Europe can contain only 10 mg/kg sulfur as of January 1, 2009. A tax incentive in 2003, ensured that all fuels, gasoline and diesel, were converted to maximum 10 mg/kg (sulfurfree) in Germany.  eformulated fuel R This is to be understood as a change in the composition and/ or physical characteristics to reduce pollutant and evaporation emissions. Within the European Auto-Oil program (EPEFE), the influence on emissions of all essential gasoline parameters was investigated. Figure 22.47 shows the qualitative options and consequences of the different measures. Apart from economic disadvantages, some of the possible measures have contrary effects on the individual types of emissions. As can be seen, the only measure that reduces all types of pollutants in the exhaust gas is a far-reaching reduction of the sulfur contents. Some words to the term “designer fuels” that has surfaced recently. These are tailored special fuels for the automobile industry that are used, for example, to meet special requirements for first tanking or for research purposes. They are not generally defined but are rather individually composed to achieve the desired special properties. In addition, fuel components generated by the Fischer– Tropsch synthesis were also called “designer fuels” for some time, because a multitude of fuel parameters over a wide

834 | Internal Combustion Engine Handbook

6606_Book.indb 834

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Possible Refinery Method

Economic Disadvantage

Technical Disadvantage

—

Expanded system, hydrodesulfurization

Increased costs

More CO2

—

⇓

No investments

Low availability, less profitable

Low density results in higher volumetric consumption

—

—

Only some investments Low extra costs required

Endangers MON level

—

⇓

No investments

Evaporation losses, hot behavior

Fuel Parameter

CO

HC

Benzene

NOx

SO2

CO2

Sulfur reduction

⇓

⇓

⇓

⇓

⇓

⇑

⇓

⇑

Lowering the final boiling point Higher volatility within average boiling range Higher volatility within lower boiling range

⇓

⇓

Reducing aromate content

⇓

⇓

⇓

Lowering benzene content

—

—

Alcohols and/or esters as gasoline components

⇓

⇓

⇓

⇓⇑

⇑

⇓

—

⇑

Isomerization Alkylation

Substantial extra costs

—

—

Various levels of investment

Moderate extra costs

⇓

⇓

Component tanks

Higher product costs

Endangers RON level, Refinery CO2 rise

Higher volumetric consumption, hot behavior, corrosion

Figure 22.47 Reformulated gasoline [22-15].

range can be set with the process parameters and the potential further refinement of the raw products.  dditives for gasolines A There is no longer any need to discuss the most important antiknock agent of an earlier era, lead, because it is no longer used for reasons of toxicity. For this reason, halogen-containing combustion chamber residue converters (scavengers) are no longer necessary. However, wear has arisen in a few older engines with “soft” valve seats when unleaded fuels are used under continually high loads (mostly high speeds), and this can be countered with special additives on the basis of potassium or sodium. In addition, additives against carburetor icing—earlier essential—are no longer used. Throttle valve icing that occasionally occurs can be countered with surfaceactive detergents. Today’s additive packages in gasolines primarily prevent system-related deposits in the fuel and mixture formation systems, primarily on intake valves. It has been demonstrated by many investigations that the use of tailored additive packages are essential for the endurance and cleanness of the engines and their fuel systems, the maintenance of exhaust gas values of new engines, and attaining and maintaining overall favorable operation. They are also economical over the long term. These problems are not new. Modern and future high-performance engines have given rise to new problems. For example, different temperature and flow conditions exist at the intake valve. There is practically no flow of oil that earlier had a certain rinsing effect. The result is increased deposits on the rear of the valves. With regard to combustion chamber deposits, the crowded normal multivalve arrangement can substantially worsen emissions. A reduction of these deposits by 1 g can reduce NOx emissions by 18%–19%. Depending on the engine and operating conditions, deposits of 4–8 g arise in practice without additives. The goal is to limit combustion chamber deposits to 1.3 g per cylinder. The problem in developing additives is optimizing a package to

deal with the contradictory requirements of intake valve cleanness (requires thermally stable components) and combustion chamber deposits (lowest possible thermostability). Because oil consumption is nearly zero today, fuel is increasingly finding its way into the engine oil from fuel and additive condensate. This phenomenon can be very pronounced because today’s oil-changing intervals are much longer than were previously experienced. Fuel and oil vapors accordingly pass through the enclosed crankcase ventilation system into the air intake system, into the combustion chamber, and up to the catalytic converter which can damage or destroy it. To reduce such problems, additive components should be prevented from entering the oil pan. This illustrates yet another conflict with the otherwise necessary thermally stable additives. The transition to extremely sulfur-low fuels worsens the natural lubrication properties of the gasoline because surface-active components are removed in desulfurization. The resulting higher pump wear must be counteracted by special antiwear additives. As a positive side effect, this can reduce fuel consumption by up to 3.5%. The spark-ignition engines with direct injection (DI engines) that are becoming more common have a series of particular problem zones that can be dealt with only by special additives. The anticipated consumption and emission advantages especially rely on the precise formation of a mixture cloud in time and space. This sensitive system can be destroyed by the slightest deposits with correspondingly negative effects. Deposits must, therefore, be avoided, particularly on the nozzles. The cleanliness of the intake ducts is important to the generation of the required swirl. Fuel additives, however, may not enter the intake ducts of DI engines. The injection pressure is also much higher, and this increases high-pressure fuel pump wear. The required protection from wear must be assumed by new “lubricity improvers” or “friction modifiers.” Figure 22.48 provides an overview of the additives that are required today.

Internal Combustion Engine Handbook | 835

6606_Book.indb 835

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Component

Active Ingredient

Improved

Remarks

Antioxidants

Paraphenylene diamine Hindered alkyl phenols

Storage stability Polymerization

Improved stability of crack components

Metal deactivators

Disalicylide Propane diamine

Stops the catalytic effect of metals

Improved stability of crack components

Corrosion inhibitors

Carboxyl amine, ester amine compounds

Corrosion protection

Usually used with detergents

Detergents

Polyisobutene amine Polyisobutene polyamide Carboxylic acid amide Polyether amine

Cleanliness of intake and fuel system, prevents throttle valve icing Drivability Exhaust emissions

Most important gasoline additive used with carrier oils

Lubrifier/Friction modifiers

Polyisobutene amine among others

Service life of injection pumps

Offsets lubricability losses in sulfurfree gasoline

Wear protection

Organic potassium, sodium compounds

Protects exhaust valve seats

Lead replacement for old vehicles, usually as a separate additive

Figure 22.48 Overview of gasoline additives [22-10].

Gasolines from different manufacturers are all miscible as a rule, but the balanced effect of the additives can be lost and, in certain circumstances, damage may arise. 22.1.2.3 Alternative Gasolines Although many alternative fuels are mentioned in the media, only a few of them pose real alternatives to the familiar fuels based on fossil energy. To draw clear distinctions, we first need to define nonrenewable and renewable or regenerative energy sources. The generated primary energies are not directly useful for powering vehicles in their normally available form and must first be converted by suitable methods into appropriate secondary energy. In addition, certain minimum requirements need to be posed on the secondary energies suitable as alternatives to today’s gasoline such as technical feasibility and storage life in the distributor system, transportability, use of the gas station infrastructure, and storability with sufficient energy density in the automobile. Given these basic requirements, the following secondary energies are suitable in addition to today’s gasoline types: •• LPG Liquefied Petroleum Gas. Pressurized, liquefied auto gas based on propane and butane. •• CNG Compressed natural gas based on methane. •• LNG Liquefied Natural Gas. Gas liquefied at low temperatures based on methane. •• MEOH Methanol. Alcohol, usually from natural gas (methane), also termed wood spirit. •• ETOH Ethanol. Alcohol from sugar-containing plants. Also termed spirit or “sprit” in German. •• GH2 Gaseous Hydrogen. Can be made from water and all hydrogen-containing energy carriers. •• LH2 Liquefied Hydrogen. Hydrogen that is liquid at low temperature. Strictly speaking, only those fuels that are not produced from the primary energies petroleum, natural gas, or coal can be considered true alternative fuels. They must also be available in a sufficient amount to supply a continuously

growing percentage of the world’s automobile population that is continuously growing itself. Accordingly, only hydrogen remains a real alternative fuel that is available over the long term. Because the solution to the many related problems will take a great deal of time, we need to closely consider supplements to classic gasoline, that is, LPG, CNG/LNG, methanol, and ethanol for the transitional period. It may be added that the Fischer–Tropsch technology developed during WWII for producing synthetic fuels can be once again looked at more closely. While this technology was used to produce gasoline from coal in war times, today the focus lies on the conversion of natural gas to liquid hydrocarbons. Accordingly, the “synfuels” are generally called GTL. Because the process level of fuel synthesis in the Fischer–Tropsch technology is preceded by a vapor reformation or gasification stage, all carbonaceous resources can be used as a rule. This means that, in addition to coal and natural gas, it is possible to convert biomass into liquid fuels. According to the resource used, this fuel is generally called “biomass-to-liquid“ in standard nomenclature and demonstrates the sustainability of this technology. Regarding the product characteristics, the Fischer–Tropsch products are very similar to conventional diesel and, compared to other potential fuel alternatives, could be introduced in the market relatively easily. Because of the enormous investigation effort of the Fischer–Tropsch technology and the high logistical expenses, this technology is economically feasible only under high-price scenarios for crude oil, as long as no further technical leaps are being made.  as fuels LPG/CNG/LNG G Under the name of propellant gas or liquid gas, mixtures of the refinery gases propane and butane were used as emergency fuels, particularly in the initial period after World War II. They were primarily used in commercial vehicles with thin-walled steel tanks that were exchanged at filling stations. Today, LPG is sometimes used in dual operation with gasoline as a gas in pressure tanks that are filled at special LPG pumps. Special quality requirements are established in EU standard EN 589. The details are shown in Figure 22.49.

836 | Internal Combustion Engine Handbook

6606_Book.indb 836

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Threshold Value Property

Unit

Minimal

MON

—

89

Content of 1,3-butadiene

Molar %

Maximum

Test Process

     0.5

ISO 7941

Calculated

Hydrogen sulfide

Mg/m3

  <4

ISO 8819

Overall sulfur

Mg/kg

 200

ISO 24260

Cu corrosion

Cor. Degrees

Evaporated residue

Mg/kg

 100

NF M 41-015

Absolute vapor pressure at 40°C

KPa

1550

ISO 2456

Absolute vapor pressure min. 250 kPa at temp. Class A Class B Class C Class D

°C  –10   –5    0  +10

ISO 4256

1

ISO 6251

It can be seen that the vapor pressure is much higher than with gasoline. The vapor pressure can be set by the ratio of propane to butane in the mixture. The maintenance of the vapor pressures in classes A–D is required to ensure cold starts. Liquid gas is gaseous at normal pressure and temperature. Because propane and butane in relation to their volume have substantially less energy than gasoline, they are liquefied under pressure for storage. At room temperature, they are liquid at 25 bar. Figure 22.50 shows a few interesting characteristics for liquid gas. Characteristic

Unit

Propane

Butane

Total formula

—

C3H8

C4H10

50/50 —

Gas density at 15°C

kg/m3

  1.81

  2.38

  2.06

Liquid density at 15°C

kg/m3

510

580

540

Boiling point

°C

−42

 −0.5

−20.7

Volumetric calorific value

MJ/m3

 93.45

108.4

101.9

Mass calorific value

MJ/kg

46.1

 45.75

 45.8

RON

—

111

 94

100

MON

—

 96

 89.6

 95

Figure 22.50 Characteristics of liquid gases [22-15].

The use of liquid gas in spark-ignition engines has a few advantages such as clean combustion with high performance and low fuel consumption, and improved untreated emissions in the exhaust. Unfortunately, these advantages can be attained only in monovalent gas vehicles when the engine and automobile are set up for gas operation. Likewise, their high antiknock quality cannot be exploited without a clear increase in compression. Because of the more favorable C ratio, less CO2 is generated during LPG combustion than with gasoline at the same vehicle running performance. Disadvantages are the increase in weight and reduced trunk area from the pressure tank. An important factor for economical use is the country-specific tax burden. In the past, the Netherlands and Italy have used excess liquid gas from refineries as a supplemental fuel with attendant preferential taxation. Until the end of 2018, LPG will be taxed in Germany at a reduced taxation rate as demanded by the Energy Tax Act. For this reason, the

Figure 22.49 Quality requirements of liquid gas (excerpt from DIN EN 589) [22-15].

difference from tax incentives and the costs of retrofitting to dual operation in Germany is large enough that the increased costs can be offset by high-mileage drivers. Because of the increased demand of LPG, the number of the gas stations providing this fuel has significantly increased in Germany. In respect of the usage of LPG, there are restrictive ordinances such as the ban on the use of underground garages and multistorey car parks. It is also difficult for engines to maintain the further-tightened exhaust thresholds in dual operation. In summary, auto gas may gain increasing importance as a supplemental fuel, particularly in view of the lower CO2 emissions compared to gasoline. CNG or natural gas—mostly methane—under high pressure (300 bar) was first used after World War II in the Ruhr for powering heavy commercial vehicles with spark-ignition engines from the Benzene Association of that period. Given the close collaboration with the mining industry at that time, methane from mines was compressed to 300 bar in a highpressure ring main built in 1950, and sent via compressor stations to different distribution sites where high-pressure bottle batteries in heavy commercial vehicles were filled. This pioneering effort that was somewhat daring for its time was terminated in 1953 since a sufficient amount of gasoline again became available, and heavy trucks with spark-ignition engines fell out of favor. Based on the present consumption of natural gas, easily accessible supplies would last 60–65 years. CNG is natural gas compressed to 200 bar for use in automobiles. It can be brought to this pressure at appropriately setup gas stations. Figure 22.51 shows the design of such a station. Natural gas that has a different composition depending on its origin, typically approximately 90% methane and approximately 10% ethane, is primarily suitable for correspondingly adapted spark-ignition engines in monovalent and bivalent operation because of its high antiknock quality. Despite corresponding tax incentives, it is not yet used in relevant quantities in the passenger car sector in Germany. Furthermore, its use is found predominantly in commercial vehicles and pickup trucks because of the expensive and large tanks required. The substantial advantages in emissions are particularly favored in city buses. Figure 22.52 compares the values of CNG with gasoline.

Internal Combustion Engine Handbook | 837

6606_Book.indb 837

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Design of a CNG gas station

Natural gas connection Compressor

Natural gas drying

High-pressure storage tank

Dispenser Figure 22.51 Design of a CNG gas station [22-11].

Physical Characteristic

Super Gasoline

CNG

Aggregate state in the tank

Liquid

Gaseous

Pressure in the tank

Atmosphere

200 bar

Density

751 kg/m3

170 kg/m3

Calorific value (CB) volume

30.8 MJ/L

7.2 MJ/L

Calorific value (CB) mass

41.0 MJ/kg

47.7 MJ/kg

Figure 22.52 shows the physical characteristics of CNG in comparison with gasoline [22-15].

In earlier gas engines, the mixture was formed in a mixer that is similar to the earlier carburetor. The functional principle is that of a Venturi tube. Because of the vacuum that predominates at the narrowest site, the required amount of natural gas is sucked in through holes at the constriction and mixed with air. Modern CNG, however, use a spark-ignition engine in which the injection system has been fitted with additional injectors for CNG operation. Because CNG is stored in automobiles under a pressure of 200 bar, a gas pressure controller is necessary to expand the natural gas to a lower system pressure and supply it to the injection system. In comparison to pure gasoline operation, a drop in performance of approximately 5% must be taken into account. The reason for this is the lower air intake air volume corresponding to the amount of gas and, in older systems, throttling arising from the throttle valve and Venturi mixer. Empirical values taken from municipal bus operations show a 22%–35% higher consumption depending on the use. However, a particular advantage in this type of use is the complete freedom from soot at high torques and low speeds, which makes the exhaust gas practically particle-free, and the engine is substantially quieter. The standard regulated three-way catalytic converter also ensures extremely low emissions. CNG can also be stored at −160°C at 2 bar, while natural gas is then in liquid form as LNG. The liquefaction of the gas requires additional energy and is done beforehand on a large industrial scale. A perfectly insulated cryotank must be used in the automobile in conjunction with the required control system. The expense is substantial. This technique has never been used in practice except for test purposes. CNG can

likewise be used only to a limited extent as a supplemental fuel. The essential product requirements on natural gas as a fuel are defined in DIN 51 624.  ydrogen H This secondary energy can be obtained from many hydrogencontaining substances such as water, natural gas, methanol, or biomass with the expenditure of energy to release the H2. The importance of hydrogen as an environmentally friendly cyclical system can be seen in Figure 22.53. Cycle of hydrogen technology Infrastructure

H2

hydrogen Solar energy

lys

ro

ct

e El

is

Figure 22.53 Cycle of hydrogen technology [22-11].

Ideally, but still extremely expensive, hydrogen is obtained by being regenerated, for example, from water with the aid of solar energy, water power, wind energy, or electrolysis. When combusted in spark-ignition engines, NOx arises from the air, but practically no pollutants or CO2 occur, just water in the form of vapor that can return into the cycle. Figure 22.54 lists the physical characteristics. In terms of its mass, liquid hydrogen has approximately three times the energy of hydrocarbons and much wider ignition limits in air. There are theoretically three options for transport, storage, and the distribution system: but also the vehicle: high-pressure storage tanks, metal hydride storage tanks, and liquid storage tanks. High-pressure storage tanks at 350 bar (up to 700 bar under development) are now commonly

838 | Internal Combustion Engine Handbook

6606_Book.indb 838

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Designation

Unit

Physical Characteristic

Liquid density at 20.3 K

kg/m3

  70.79

Gas density at 20.3 K

kg/m3

   1.34

Gas density at 273.15 K

kg/m3

   0.09

Evaporation heat

kJ/kg

445.4

Bottom calorific value (CB)

MJ/kg

 119.97

Lower ignition limit in air

% (V/V)

4.0–4.1

Upper ignition limit in air

% (V/V)

75.0–79.2

*At 1.013 bar (a).

Figure 22.54 Physical characteristics of hydrogen [22-15], [22-16].

used as storage mediums in the first generation hydrogendriven passenger cars. Metal hydride storage tanks in which the hydrogen is adsorbed on metal alloys offer a high safety standard, but the amount of hydrogen they can store is limited despite the present level of development. For a passenger car range of 200 km, a tank weighing several hundred kilograms is required. In a liquid storage tank, the hydrogen is cooled to −253°C and stored in a cryotank using high-performance insulation (LH2). The large amount of energy required for cooling to such low temperatures is substantial, however. In addition, when automobiles stand for a long period, hydrogen is lost from leakage. At the present level of development, a loss of up to >2% per day must be anticipated. Filling-up requires an enormous effort because all moisture and air must be evacuated in addition to dealing with the low temperature. By now, the challenges for the tank engineering are resolved and fully automated filling by robot and conventional filling by the driver has been successfully tested in the field. The long time that it initially took to fill up has been reduced to an acceptable level. Further developments can be anticipated for commercial use. Possible energy converters in cars are both the spark-ignition engine and the fuel cell. The method of mixture preparation and combustion control would have to be redeveloped for the spark-ignition engine. Combustion occurs in a very lean range, which is optimum for controlling the combustion process and for NOx emissions. This is associated with a performance loss, however. With less excess air, it may be necessary to inject water into the intake manifold because backfiring in the intake system could arise. It would be ideal to directly inject the liquid hydrogen into the combustion chamber. In addition, because of the high ignitability of hydrogen, the spark-ignition engine compression cannot be very high, which worsens the thermodynamic efficiency.

Boiling Point

Density 20°C

Vapor pressure

The NOx problem should take a while to solve. Given the worldwide success in developing fuel cells over recent years for electric engines in motor vehicles, one can assume that this system is more suitable for the use of hydrogen than the spark-ignition engine. However, the production methods for fuel cells must still be optimized. The generation of hydrogen based on fossil energy is not expedient. Only photovoltaics seems to be sensible as the energy source for H2 production, considering the environmental issues involved. Finally, a suitable infrastructure must be developed for filling up such automobiles. The use of an electric motor in vehicles with fuel cell technology has introduced a further long-term alternative for sustainable mobility: the electric drive. But similar to the hydrogen technology, fundamental issues such as the increase in energy storage density of the batteries, reduction of battery weight and volume, installation of a charging infrastructure and, last but not least, sufficient electricity on a regenerative basis must be resolved before electric mobility will be take its place in the market.  lcohol fuels A Alcohols are hydrocarbon-oxygen compounds (so-called oxygenates) whose particular feature is an OH group in the molecule instead of a hydrogen atom. Short-chain primary alcohols (one OH group) are, in principle, well-suited for vehicles with spark-ignition engines. The techniques for producing them are known and well developed. They can be transported, stored, and distributed essentially in the existing system. The physical characteristics of the most important alcohols have already been portrayed in Figure 22.24. In this discussion, only methanol and ethanol are of interest as supplemental fuels. Figure 22.55 again provides a comparison of their physical characteristics with super fuel. Serious differences in practice are found in the antiknock quality, the calorific value, and the evaporation heat. A particular advantage is the high antiknock quality that can be used to improve efficiency through a correspondingly high compression ratio. Alcohols also burn faster, which means that the ignition map must be correspondingly adapted (see also “racing fuels”). The substantially lower volumetric calorific value correspondingly yields higher fuel consumption. The significantly greater evaporation heat causes greater cooling of the fuel–air mixture which, because of the superior internal cooling, improves the charging and, hence, performance. The substantial increase in volume of the fuel–air mixture after fuel evaporation allows higher average pressures than gasoline and offers greater thermodynamic engine efficiency. In addition, the ignition range of an alcohol–air mixture is greater than that Net Calorific Value CB

Evaporation Heat

O2 Content

Designation

Abbreviation

°C

kg/m3

Hpa

RON

MON

MJ/L

kJ/kg

% (m/m)

Methyl alcohol

Methanol

64.7

791.2

32

114.4

94.6

15.7

1100

49.93

Ethyl alcohol

Ethanol

78.3

789.4

17

114.4

94.0

21.2

 910

34.73

Super fuel

Super gasoline

30–215

725–780

S: 60–70 W: 80–90

95

85

approx. 41

380–500

0–2

Figure 22.55 The physical characteristics of methanol and ethanol in comparison with super gasoline [22-11], [22-15].

Internal Combustion Engine Handbook | 839

6606_Book.indb 839

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

of gasoline, which allows more excess air under a partial load. The effects on untreated exhaust emissions are also positive. Special measures, such as preheating the air intake system, are required, particularly at low temperatures, to deal with the higher boiling point in comparison to the initial boiling point, and the lower vapor pressure in conjunction with the stronger cooling from the high evaporation heat. In comparison to gasoline, the drop in temperature of the theoretical mixture without preheating is 120°C for methanol and 63°C for ethanol. The aggressiveness of alcohols against metals and elastomers in mixtures with conventional gasoline, in particular, requires the use of special materials and special additives. Alcohols can be used in their pure form as well as in mixtures with hydrocarbons. Low concentrations as described in Section 22.1.2.1 do not require changes to the vehicle, except for the fact that the compatibility of elastomers in the fuel system must be assured. Higher concentrations such as 15% (V/V) methanol in the gasoline require corresponding adaptations. The basic approaches for this were worked out 20 years ago in a joint project of the German automobile and mineral oil industries subsidized by the Federal Ministry of Research. The following facts are relevant concerning the production and use of methanol and ethanol: The simple alcohol molecule CH3OH (methanol) arises from the synthetic gases CO and CO2 that can be obtained from every carbon-containing primary energy carrier, today preferably from natural gas. It has a large amount of H2 (C/H ratio 4:1) which also makes it interesting as a starting product for fabricating hydrogen using the infrastructure of fuel cell technology, at least in the initial phase. It can be used as a gasoline–methanol mixed fuel, for example, M 15 (15% MEOH in gasoline) or as methanol fuel (M 100). Because the danger of separation upon the addition of water increases as the methanol content decreases, the stability of methanol–gasoline–water mixtures must be monitored. The methanol fuel M 100 must contain HC components for cold starts and warmup, as well as other substances such as additives to allow problem-free use in automobiles. The addition of a certain amount of gasoline is also required for safety reasons because methanol burns with an invisible flame. The essential points of possible specifications for methanol fuel are shown in Figure 22.56. The advantage of the high octane number, fast combustion, and larger volume expansion of the fuel–air mixture is contrasted with a consumption disadvantage of approximately 70% in an optimized methanol engine. In addition, methanol tends to preignite much more strongly than gasoline, which requires special modification of the engine such as cold spark plugs. When methanol fuels are used, special engine oils without ash-free dispersants are required because they can form a sticky residue upon contacting methanol. In addition, they require additives against corrosive engine wear.  thanol E C2H5OH is the second in the homologous series of alcohols characterized by the hydroxyl OH group. It can be obtained from biomass by fermenting agricultural products. Suitable starting products are all sugar, starch, and cellulose-containing

Characteristic

Unit

Summer

Winter

Methanol

% (m/m)

Min. 82

Min. 82

HC total*

% (m/m)

Min. 10, Max. 13

Butane

% (m/m)

Max. 1.5

Density 15°C

kg/m3

Max. 2.5 770–990

Vapor pressure RVP

kPa

55–70**

Water content

ppm

Min. 2000, Max. 5000

75–90**

Higher alcohols

% (m/m)

Max. 5

Formic acid

ppm

Max. 5

Overall acid

ppm

Max. 20

Evaporated residue

mg/kg

Max. 5

Chlorine

ppm

Max. 2

Lead

ppm

Max. 30

Phosphorous

ppm

Max. 10

Sulfur

ppm

Max. 10

Additive

%

Max. 1

*Type of hydrocarbon, boiling behavior and quantity depending on use. **Example from Central Europe. ***With corrosion inhibitor. ****Measured as acetic acid.

Figure 22.56 Specification for methanol fuel [22-16].

raw materials. Figure 22.57 shows the possibilities for ethanol generation. Sugar-Containing Sugar cane Sugar beet Sorghum

Cellulose-Containing Grain Corn Cassava Potatoes

Forest scrap wood, quickly growing trees Hemp, kenaf Bagasse, straw Salks, husks, hulls Used paper

Figure 22.57 Vegetable raw materials for generating ethanol [22-11], [22-15].

Glucose is converted into alcohol with yeast. To date, the generation of ethanol for fuel from sugarcane has the greatest economic importance in Brazil. To increase the yield and reduce the competition between foods and fuel production, it would be more advantageous to use primarily cellulose-containing plants. The fermentation must be preceded by a conversion process that transforms the different cellulose types into glucose depending on the plant type. From the user’s perspective, cold-start problems arise unless special constructive measures are taken. The use of alcohols in diesel engines is discussed in section 22.1.1.4 (Alcohol-diesel mixtures). In summary, the alcohols methanol and ethanol are not viable alternative fuels, but they can serve a role as supplemental fuels in the (very) long transition period to hydrogen. Before this backdrop, all essential requirements on ethanol as mixture component are specified in DIN EN 15376 and as “E 85” ethanol fuel in DIN 51625.  acing fuels R Of course, beyond the special constructive measures to attain maximum specific performance, advantages have also been

840 | Internal Combustion Engine Handbook

6606_Book.indb 840

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.1 Fuels

Density at 20°C RON

Evaporation Heat

Net Calorific Value (CB)

°C

kJ/kg

MJ/L

 56 

 524

Component

kg/m3

Acetone

 791

Diethyl ether

 714

 35 

 487

24.3

Ethanol

 789

114.4

 94.6

 78.5

 910

21.2

Methanol

 792

114.4

 94.0

 64.7

1100

15.6

Benzene

 879

99

91 

 80 

 394

34.9

Toluene

 867

124 

109

110 

 356

34.6

Nitrobenzene

1200

208 

 397

Water

1000

100 

2256

–

MON

Final Boiling Point

–

sought from the fuel. As long as existing laws do not provide restrictive regulations, such as for commercial Super, performance enhancements from fuel have been substantial. The main goal is to combine the maximum antiknock quality at a maximum compression ratio, the highest internal cooling for the best charging, and a high rate of combustion for maximum speeds with the greatest possible energy supply. Furthermore, the fuel volatility must be adjusted so that it meets the requirement for the greatest possible volumetric efficiency. Figure 22.58 presents the mixture components that are relevant for racing fuels with their particularly important characteristics (some of which are no longer used). Performance can be increased directly by fuel with components with a high energy content and low stoichiometric air–fuel ratio. With a given air supply, this combination allows an increase in the actually supplied amount of energy. A typical example is nitromethane whose calorific value is clearly lower than gasoline; however, the much lower stoichiometric air–fuel ratio allows a more than twofold increase in the energy supply (specific energy). Its use is, however, limited by the high thermal and mechanical load on the engine-transmission assembly. Figure 22.59

Figure 22.58 Components of racing fuels [22-4].

–

provides information on these relationships for nitromethane and methanol in comparison to iso-octane. Prestressed ring-shaped compounds such as quadrocyclane and diolefins such as diisobutylene yield measurable, direct increases in performance. Although they usually have a low ON and are unsuitable according to a conventional evaluation, they are much less knock sensitive than their ON seems to indicate because of a clearly higher rate of combustion with a corresponding adaptation of the ignition map. Another advantage of a high rate of combustion is that the extremely high speeds tend to shift the energy conversion toward the top dead center, which improves efficiency. This advantage also holds true for olefin-containing, conventional crack components. Given the sometimes contradictory properties of the fuel components, harmonizing the engine with the fuel is a delicate and, hence, involved procedure. The availability of the discussed special fuels is restricted, and they are all very expensive. Figure 22.60 compares quadrocyclane with toluene and diisobutylene with i-octane. Special racing fuels were already being used in the 1930s for Grand Prix racecars. Figure 22.61 presents the contents of

Characteristic

Unit

Nitromethane

Methanol

Total formula

—

CH3NO2

CH3OH

Iso-Octane C8H18

O2 content

% (m/m)

 52.5

  49.9

  0

Evaporation heat

kJ/kg

560

1170

270

Net calorific value (CB)

MJ/kg

 11.3

  19.9

 44.3

Stoichiometric air–fuel ratio

—

  1,7:1

   6.45:1

 15,1:1

Specific energy*

MJ/kg

  6.65

   3.08

  2.93

Figure 22.59 Specific energy of nitromethane in comparison to i-octane [22-15].

*Quotient from BC and stoichiometric air–fuel ratio.

Characteristic

Unit

Quadrocyclane

Toluene

Diisobutylene

Total formula

—

C7H8

C7H8

C8H16

Iso-Octane C8H18

Density

kg/m3

919

874

719

699

RON

—

 54*

124*

 98*

100

MON

—

 19*

112*

 78*

100

Net calorific value (CB)

MJ/kg

 44.1

 40.97

 44.59

 44.83

Stoichiometric air–fuel ratio

–

 13.43

 14.70

 13.43

 15.05

Specific energy

MJ/kg

  3.28

  2.79

  3.32

  2.98

Figure 22.60 Specific energy of quadrocyclane and diisobutylene [22-15].

*Mixture ON.

Internal Combustion Engine Handbook | 841

6606_Book.indb 841

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

the then strictly secret racing fuels of the prewar competitors Auto-Union and Mercedes-Benz with Alfa Romeo and Maserati.

Component

Auto-Union Mercedes-Benz

Alfa Romeo Maserati

% (V/V)

% (V/V)

Ethanol

10 

49.5

Methanol

60 

34.5

Denaturation

—

 0.5

Benzene

22 

—

Petroleum ether

 5 

—

Water

—

0.5–3

Res

 3*

12–15**

as multivalve technology, hydraulic valve lash compensation, camshaft timing, and supercharging represent substantial and, partially new challenges. The new generation of sparkignition engines with direct injection can, in contrast to existing engines with manifold injection and catalytic converters, be operated in the lean range of combustion. This has produced a new set of problems for the engine oil. In addition, we are confronted with demands to reduce fuel consumption by reducing friction, not impair the exhaust purification systems in spark-ignition and (in the future) diesel engines (Euro IV), and maintain environmental compatibility [22-9].

22.2.3 Lubrication Types

Lubricants are design elements without which the internal combustion engine and transmission could not function. They were developed along with the automobile and continue to be developed because of a number of interrelated factors. The complex composition of modern lubricants allows them to satisfy even the highest requirements.

A distinction is drawn between full and partial lubrication (liquid friction and mixed friction). Liquid friction, the ideal state, predominates, for example, in plain bearings at a certain speed, or after the external application of oil pressure. However, mixed friction must be assumed when, for example, the crankshaft and bearing directly contact each other upon starting without the lubrication film that must first build up. It is also unavoidable that, in certain component assemblies such as the valve gear with tappets and cams, or at the point where the pistons reverse in the cylinders, mixed friction predominates over long periods of operation. It is important to provide the lubricant with antiwear and antioxidation additives and to have the lubricant be sufficiently viscous to minimize wear.

22.2.1 Lubricant Types

22.2.4 Lubrication Requirements

The term “automobile lubricants” includes the following subcategories:

Among the most important requirements for engine oil are the following.

•• engine oils for four-stroke spark-ignition and diesel engines

 ransmitting forces T From the connecting rod, the entire combustion pressure exerted on the piston is transferred to the crankshaft via the piston pin and connecting rod bearing only with the aid of the slight amount of oil in the lubrication gaps. The arising pressure in the thin lubrication gap can be as large as 10,000 bar.

*Toluene/nitrobenzene/castor oil. **Information not provided.

Figure 22.61 Grand Prix racing fuels used before 1939 [22-4].

22.2 Lubricants

•• engine oils for two-stroke engines used in motorcycles, scooters and mopeds, for example •• universal oils for tractors •• transmission oils •• hydraulic oils •• greases. In this section, the two first categories will be discussed.

22.2.2 Tasks of Lubrication

The lubricant is to reduce friction between contact bodies, reduce their wear, remove any wear particles from the lubrication site, and prevent the penetration of foreign materials into the lubrication gap. They must also transfer force, for example, from the piston to the conrod, cool by transferring heat to the oil pan or oil cooler, provide a seal, for example, of the annular gap between the piston and cylinder, protect against wear, prevent deposits and corrosion, neutralize acidic combustion products, be compatible with the elastomers of the seals, be stable over time to allow long change intervals, have a low evaporation loss for low oil consumption, and manifest optimum viscosity temperature behavior to ensure easy cold starts and reliable hot operation. Recent design elements such

 ooling C The engine oil plays a relatively minor role in removing heat, but its specific task is quite important, namely, piston cooling. On the one hand, the oil flying about within the engine conducts heat from the hot piston; on the other hand, particularly in highly supercharged diesel engines, usually additional oil is sprayed from below against the piston head or guided into a separate cooling channel to cool especially the upper piston ring area.  ealing S Engine oil has the important task of providing a fine seal between the piston, piston rings, and cylinder barrel surface to transfer the high pressure from combustion with minimal loss to the piston head surface. Even when there is an optimum seal, approximately 2% of the combustion gases pass by the piston and enter the crankcase (blow-by gas). This gas attacks the engine oil with aggressive reaction products from combustion.

842 | Internal Combustion Engine Handbook

6606_Book.indb 842

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

 rotection from deposits P During combustion in spark-ignition and diesel engines, oilinsoluble residues necessarily arise in solid or liquid form. The residues must be prevented from agglomerating and collecting in the engine, for example, in the piston ring grooves or in the oil pan. Oil-insoluble residues can form sludge under special circumstances. This problem is solved with detergent and dispersant additives.  orrosion protection C Mineral oil per se offers a certain amount of corrosion protection against slight amounts of water. This protection is, however, insufficient in the additional presence of aggressive combustion products. After the engine is turned off, humidity can condense into water inside the engine. Water also arises during combustion as a reaction product. One liter of fuel yields approximately one liter of water depending on the HC ratio. Water leaves the hot engine largely in the form of steam with the exhaust gases. A small part enters with the blow-by gases into the crankcase and oil pan where it condenses when the engine cools. The engine oil can absorb only a certain amount of water so that corrosion can appear on unprotected metals if not counteracted by corrosion inhibitors.  ear protection W Mechanical and corrosive wear must be prevented primarily in the cylinder barrel, on the piston and piston rings, the bearings, and valve gear such as the cam, tappet, and rocker arm. In diesel engines, a special load arises in areas with mixed friction from soot formation. This especially includes the cylinder barrels. Mechanical wear can be effectively reduced by EP/AW additives (extreme pressure/antiwear), whereas corrosive wear can be kept under control by the neutralizing effect of corrosion inhibitors.  eal compatibility S The properties of radial shaft sealing rings used in engines, valve shaft seals, and other seals made of elastomer materials (elastic plastics) may not be altered by fresh or used oil. They may not become brittle, soften or shrink and form cracks under stress. To ensure a lasting seal, a certain amount of swelling is desirable. To prevent or compensate for drying, that is, the exchange of softener with polyalphaolefins (PAO) (for example) in the elastomers when certain synthetic basic liquids are used, “seal swell agents” are used.  ging stability A This is particularly important in view of the continual extension of the oil change intervals. At high operation temperatures, engine oils tend to “age” because oxygen bonds to the hydrocarbon molecules yielding acids, and resinous or asphalt-like components can form. The oil is continuously mixed with air as a thin film as it flows off, drips off, is flung off and is sprayed inside the engine. This can make the engine oil thicker with the assistance of the blow-by gases. To prevent this, antioxidants are used.

Evaporation loss The evaporation loss strongly depends on the viscosity and the type of basic liquid. Earlier, when only mineral oil raffinates were used as the basic oil, a general rule was the thinner the basic oil, the greater the evaporation loss The basic liquids in high-performance engine oils of the “new technology” such as special raffinates, hydrocrack oils, synthetic hydrocarbons, and esters have the same viscosity yet different levels of evaporation loss. These are essential because of the increasing dwell times of the oil in the engine.  iscosity–temperature behavior V Today’s assumptions that oils should be as thin as possible when cold and as viscous as possible when hot can be achieved only by multigrade oils with a high viscosity index (V.I.). In this case as well, modern basic liquids are far superior to earlier basic oil raffinates.

22.2.5 Viscosity/viscosity Index (V.I.)

Viscosity is a measure for internal friction or the resistance a liquid offers to deformation. Assuming a laminar flow, the shear stress arising between two flowing layers is proportional to the speed gradients perpendicular to the direction of flow, according to Newton’s shear stress law. The arising proportionality factor is termed dynamic viscosity or the absolute viscosity of the relevant liquid. It represents the force that counteracts the movement of a liquid layer in reference to 1 cm2 that flows at the speed of 1 cm ⋅ s–1 parallel to a resting liquid layer at a distance of 1 cm. The unit of measure of the dynamic viscosity is 1 P (Poise) = 100 cP (centipoise; 1 cP = 1 mPa ⋅ s; millipascal second). The term viscosity in the Newtonian sense is limited to the range in which the proportionality is retained independent of the gap width and shear speed. In lubricating oils, this proportionality can, for example, be lost by cooling when the original Newtonian liquid no longer follows the proportionality law from the precipitation of solid particles such as paraffins and the formation of a mixture of solids. This is particularly the case with artificially thickened oils such as multigrade oils. Instead of measuring dynamic viscosity, the easier solution in practice is almost always to measure the kinematic viscosity or relative viscosity. This results from the ratio of the dynamic viscosity to the density. The unit of measure is 1 St (Stokes) = 100 cSt (centistokes; 1 cSt = 1 mm 2 ⋅ s–1). The viscosity is primarily influenced by the temperature and pressure and, in the case of nonNewtonian liquids, also by the shear speed. 22.2.5.1 Influences of Temperature on Viscosity As the temperature rises, the distance of the molecules in the lubricating oil increases so that they move away from their mutual range of influence This reduces inner friction and, hence, the viscosity. The dependence on temperature of the viscosity is particularly important in lubrication technology. It is evaluated and rendered comparable in an internationally uniform manner with the aid of the viscosity index (V.I.). The greater the V.I., the lower the temperature sensitivity of an oil.

Internal Combustion Engine Handbook | 843

6606_Book.indb 843

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

This relative identification is carried out using two extremely different temperature-sensitive reference oils. Both have the same viscosity at 100°C. With Pennsylvania reference oil, the viscosity increases very slowly as the temperature falls; with Gulf Coast reference oil, it rises quickly. The value V.I. = 100 has been assigned to the first reference oil, and V.I. = 0 to the second. The viscosity values of both reference oils are defined according to DIN ISO 2909 for the range of 2–70 mm2 ⋅ s–1 at 100°C. For values above 70 mm2 ⋅ s–1, equations are provided in the standard for calculation. The V.I. cannot be measured directly. It is calculated for a given oil by comparing it with the reference oils using the relationship in Figure 22.62. V.I. = 100 ⋅

L− P L− H

H (High) ⇒

Viscosity of the reference oil with V.I. = 100 at 40°C

⇒

L (Low)

Viscosity of the reference oil with V.I. = 0 at 40°C

P (Probe) ⇒

Viscosity of the oil to be determined at 40°C

Figure 22.62 Calculation of the viscosity index.

Viscosity index (V.I.)

Figure 22.63 shows an example of for an oil with a viscosity of 8 mm2 ⋅ S –1 at 100°C. For the L reference oil, 97 mm 2 ⋅ s–1 at 40°C was measured, and 57 mm2 ⋅ s–1 at 40°C was measured for the H reference oil. These two values were assigned 0 and 100 V.I., and the differential range of 40 mm 2 ⋅ s–1 was divided into 100 increments. Since, for the oil sample P, a viscosity of 100

0 20

90

40

80

60

70

80

L

60

100

P

Kinematic viscosity [mm2/s]

50

H

40 30 20

V.I. = 100

L–P L–H

10 0 40

Temperature [°C]

100

Figure 22.63 Determining the V.I. in a graph. Pressure

Paraffin-Based Oil

61 mm2 ⋅ s–1 at 40°C was measured, the value 90 can be read on the V.I. scale. With the development of multigrade oils with a V.I. > 100, a fundamental problem in determining the V.I. arose in the form of overvaluing the oils with a low viscosity. To solve the problem, a new calculation method was introduced that produced the V.I.E (expanded V.I.). When the V.I. is particularly high, the V.I.E allows a clear differentiation in evaluating the efficiency of V.I. improvers. The fully synthetic basic liquids in the engine oils of the “new technology” have a very high V.I.E that is additionally raised by new V.I. improvers to cover the wide Society of Automotive Engineers (SAE) ranges (see Section 22.2.8.1). The relationship in Figure 22.64 is used to calculate the V.I.E. V.I.E = 100 +

G−1 0.0075

G=

lgH − lgP lgY

H (High) ⇒

Viscosity of the reference oil with V.I. = 100 at 40°C

P (Probe) ⇒

Viscosity of the oil to be determined at 40°C

Y (Probe) ⇒

Viscosity of the oil to be determined at 100°C

Figure 22.64 Calculation of the V.I.E.

22.2.5.2 Influences of Pressure on Viscosity When a lubricating oil is under very high pressure, its viscosity rises sharply as under the influence of temperature because the now-denser molecules generate greater inner friction. In calculations of plain bearings, the influence from pressure is usually dismissed because it is assumed that the rise in viscosity from a rise in pressure is approximately compensated from the drop in viscosity because of the increase in temperature that always occurs. Accordingly, we know from substantial experience that a rise in pressure of approximately 35 bar has the equal but opposite effect as an increase in temperature of approximately 1°C. At very high pressures, for example, in roller bearings or geared transmissions (up to 15,000 bar), the rising pressure has to be taken into account. Roughly speaking, at room temperature the viscosity of most petroleum products doubles when the pressure rises by approximately 300 bar. Furthermore, the same rise in pressure within a high-pressure range increases the viscosity more than within a low-pressure range. Highly fluid oils are influenced less by a rise in pressure because of their viscosity than are viscous oils. It is interesting to note that a rise in pressure can also increase the V.I. Naphthene-based oils respond more readily than paraffin-based oils. In general, similar to the influence of temperature, a change in pressure has less of an effect on paraffin-based oils than on naphthene-based oil. Figure 22.65 shows the influence of pressure on the viscosity and VI. Naphthene-Based Oil

Bar

cSt*/40°C

cSt/100°C

V.I.

cSt/40°C

CSt/100°C

   1

  52.5

  6.8

 90

   55.4

 5.8

V.I.  16

1400

 810 

 43.5

100

   21.9

53.5

 54

2500

8700 

195 

125

91,000  

454

115

Figure 22.65 Influence of pressure on viscosity and V.I. [22-4], [22-12].

1 cSt = 1 mm2 ⋅ s–1.

*

844 | Internal Combustion Engine Handbook

6606_Book.indb 844

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

22.2.5.3 Influences of Shear Speed on Viscosity Because of V.I. improvers that can be used for a wide temperature range (see Section 22.2.8.3), Newton’s shear stress law no longer applies to multigrade oils because the proportionality, that is, the viscosity, is now determined by the gap thickness (film thickness) and the shear speed. This is a non-Newtonian liquid. Whereas with a Newtonian liquid, the viscosity remains constant as the shear speed rises, it falls with a non-Newtonian oil. Because, in this case, an infinite number of viscosities are conceivable depending on the shear speed at which they were measured, the term “apparent viscosity” is used to draw a distinction; for the absolute viscosity in poise, the corresponding shear speed is given in reciprocal seconds. One speaks of a shear gradient in this context. Under a high shear load, both the V.I. and viscosity can fall since long-chain polymer V.I. improvers can be broken up and lose some of their effectiveness. A loss in viscosity and V.I. with a high shear gradient can be permanent because of the mechanical or chemical breakup of the large polymer molecule into smaller molecules or temporary since the long-chain polymer molecules tend to align corresponding to the direction of flow that results in less flow resistance. When the V.I. loss is temporary and the shear stress decreases, the oil resumes its original viscosity. Permanent V.I. loss is undesirable in practice. In automobile engines, shear speeds of 50,000–1,000,000 s–1 arise. Earlier polymer V.I. improvers with a high molecular weight experienced both high temporary and permanent V.I. loss. Today’s V.I. improvers are stable even under maximum shearing stress because of their moderate molecular weight and special structures, and they maintain the viscosity temperature behavior engineered for the fresh oil; that is, the multigrade oil remains in the given SAE ranges (stay in grade).

22.2.6 Basic Liquids

Engine oils always consist of a basic liquid or a mixture of basic liquids and an additive package that has been harmonized in extensive experiments without which today’s requirements cannot be met. The basic liquids, also termed basic oils, are mineral oils, synthetic oils, or a mixture of both (partially synthetic oils). The basic liquids influence important properties of engine oil such as viscosity, evaporation loss, and additive response. Different basic liquids react differently to the effect of the additives and can produce different engine test results. For this reason, they are divided into five groups according to Association Technique de l’Industrie Europeenne des Lubrifiants (ATIEL) as shown in Figure 22.66. In addition to the listed properties, there are other important criteria for selecting basic oils depending on the use.

22.2.6.1 Mineral Basic Oils Mineral basic oils are still used today in most conventional lubricants; however, they are increasingly being replaced by synthetic basic liquids because of steadily increasing requirements. The raffinates from petroleum are obtained by atmospheric distillation, vacuum distillation, solvent refinement, deparaffination, and hydrofinishing. They consist of large molecules with infinite possibilities for branching, even with the same number of C and H atoms. Despite involved refining methods, there is no uniformly structured mineral basic oil from different crude oils. There are basic oils with different viscosities, from highly fluid spindle oil to highly viscous Brightstock for engine oils, depending on the desired viscosity. In practice, generally mixtures of at least two basic oil components are used that lie between spindle oil and Brightstock. Adjacent distillation cuts are preferably used. The sulfur content of the crude oils suitable for producing lubricants is between 0.3% (m/m) (North Sea) and 2.0% (m/m) (Middle East). A distinction has always been drawn between paraffinbased and naphthene-based basic oils. The paraffin-based oils are preferred because of their better viscosity temperature behavior. Their V.I. is generally high, ranging between 90 and ≤100 (see Section 22.2.5). The viscosity of the mixture components at 100°C is between 3.7 and 32 mm2 ⋅ s–1. For the high-performance oils in demand today, the properties of the best raffinates are no longer sufficient. 22.2.6.2 Synthetic Basic Liquids Synthetic basic liquids are essential in engine oils when highperformance multigrade oils are required with minimum oil consumption, minimum residue, maximum wear protection, high fuel economy, and the potential for flexible oil change intervals. The starting materials for synthetic engine oils are largely based on raw gasoline that exists in the form of ethene (ethylene) after cracking. The synthetic hydrocarbons PAO (polyalphaolefin) and PIB (polyisobutene) are made from this using various catalytic processes. If ethene is reacted with oxygen and hydrogen in the presence of a catalyst, synthetic estersor polypropylene glycols and polyethylene glycols arise in different steps. Another possibility for producing synthetic basic liquids is from vacuum residue. A hydrocrack oil and light gas or gasoline fractions arise from catalytic hydrocracking. For “new technology” engine oils, PAO, PAO plus esters, or PAO plus hydrocrack oil are used; the other synthetic basic liquids such as the above-cited polyglycols are used for hydraulic and industrial gear oils. Synthetic hydrocarbons such as PAO and hydrocrack oils have a very special molecular structure that does not exist in the starting products; they are more or less a tailored product.

Group

Composition

Sulfur Content

Viscosity Index

I

<90% (m/m) saturated hydrocarbons

>0.03% (m/m)

≥80 < 120

II

≥90% (m/m) saturated hydrocarbons

≤0.03% (m/m)

≥80 < 120

III

≥90% (m/m) saturated hydrocarbons

≤0.03% (m/m)

≥120

IV

Polyalphaolefins (PAO)

V

All others not contained in group I, II, III, or IV (ester, for example)

Figure 22.66 Categorization of the basic liquids according to ATIEL [22-4].

Internal Combustion Engine Handbook | 845

6606_Book.indb 845

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Component

Advantage

Reason

Extra Costs (%)

Hydrocrack oil

High V.I. >110 Low evaporation loss Good viscosity-temperature behavior

Molecular structure Uniform composition Low pour point

Approximately 30–100 (depending on V.I. and quality)

Polyalphaolefin Polyisobutene

Very high V.I. <150 Very low evaporation loss Very good viscosity-temperature behavior

Molecular structure Uniform composition Low pour point

Approximately 250–350 Approximately 200

Figure 22.67 Vorteile Advantages and extra costs of PAO and hydrocrack oils in contrast to raffinates [22-13], [22-15].

Mineral oils, esters, PIB, and PAO are all miscible with each other. In the transition period from mineral-based oils to today’s fully synthetic engine oils, partially synthetic engine oils were introduced. These are still used today as a low-cost alternative for average loads. The advantages and extra costs of synthetic basic oils compared with raffinates are shown in Figure 22.67.

22.2.7 Additives for Lubricants

Engine oils always consist of one or a mixture of several basic oils (see Section 22.2.6) and additive agents. Additives are oil-soluble agents of various types that are added to the basic liquids to produce properties that do not exist (sufficiently) in the basic liquid, to reinforce positive properties, and to minimize or eliminate undesirable properties. Not all properties of the engine oil can be influenced by additives such as thermal conductivity, viscosity pressure dependence, gas solubility, and air releasability. Additives almost always work as mixtures and can have both synergistic and antagonistic effects. The additive content extends from a few ppm to 20% (m/m) and more in modern high-performance oils. Many lubricant additives are surface-active or interface-active materials whose structures can be compared with that of a match. The “head” is a functional chemical group that, for example, is “ignited” by water, acids, metals, or soot particles. It is also termed the “polar group” in which the actual active ingredients are concentrated. They can be organic (ashfree) or metal–organic (ash forming). The “stem” consists of a nonpolar hydrocarbon residue (radical), the oleophile (that is “drawn” by oil). It is primarily responsible for dissolving the additive in the oil. Many additive types have several stems on their polar group. Another important group of oil additives consists of high-molecular hydrocarbons with a special molecular structure that can also contain oxygen. Figure 22.68 lists the additive types for engine oils. In principle, transmission oils have the same composition. As with the basic liquids, the additives also need to be considered in terms of environmental compatibility. For example, chlorine-containing compounds are rarely used nowadays. A measure of the amount of added metal-containing substances is the sulfate ash content in the fresh oil. Modern exhaust gas aftertreatment systems in diesel engines (diesel particulate filter = PDS, for example) to reduce principle-dependent particulate emissions demand engine oils with a reduced content of ash generators. Because of combusted engine oil, sulfate ashes travel via the exhaust gases into the fine-pore particulate filters and thus reduce the flow down to a blockage because, unlike soot, they cannot be burnt off by

regeneration. By using relatively large, robust filter systems, it is possible to omit low-ash engine oils, however, compact systems close to the engine are becoming more popular. The requirements regarding sulfate ash content are defined in the ACEA specifications (new subgroups C1 to C4) and the OEM specifications issued by some manufacturers. The sulfate ash content of low-ash or low SAPS (sulphated ash phophorous sulfur) for passenger car engine oils is between 0.5% and 0.8%, for commercial vehicles up to 1.0%. In “classic” engine oils for spark-ignition and diesel passenger cars, it is 1.0%–1.5%, and for European diesel commercial vehicles, 1.5%–2.0%. Passenger car low-ash engine oils are have been developed with the focus on diesel engines, but they can be used in most gasoline engines as well. 22.2.7.1 V.I. Improver Today’s high-performance engine oils for passenger cars and commercial vehicles must function perfectly under all driving and weather conditions with extremely long oil change intervals. This means sufficiently low viscosity for reliable cold starts (even when the outside temperature is very low), immediate energy-saving lubrication, and sufficiently high viscosity for reliable lubrication under high thermal and mechanical loads. Only multigrade oils (see Section 22.2.8.3) meet these requirements. The given viscosity temperature dependence of the basic liquids is insufficient even when the V.I. is very high; suitable V.I. improvers must therefore be used. These are polymers with a high molecular weight. Their effect can be explained with reference to their dissolving behavior. At low temperatures, V.I. improvers are tightly bunched in the oil. Because they require little space, they have only a slight effect on the viscosity. As the temperature rises, the required space increases; the bunches unravel and, hence, reduce the thinning. When selecting a V.I. improver, we need to determine its sensitivity to high shear stress (for example, between the cam and tappet, in roller bearings, or between the gears of the oil pump) so that the desired V.I. is maintained even under high shear stress and after a long operating time. The response of basic oils to V.I. improvers is less favorable in basic oils with a high V.I. than in those with a low V.I., and it quickly falls as the added amount increases. Depending on the type, V.I. improvers usually have pour-point-lowering properties. Because of their molecular size, they form a disrupting site for crystal growth when forming paraffin crystals and thus give rise to small, separate crystals. Because of the marked influence of temperature on the viscosity of an oil (see Section 22.2.5.1), only multigrade engine oils with a V.I. ⪢ 100 are used. The top products that are designed as fuel economy

846 | Internal Combustion Engine Handbook

6606_Book.indb 846

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

Additive Type

Active Ingredient

Function

V.I. improvers Dispersing or nondispersing

PMA Polyalkylstyrenes OCPs Star polymers PIB Styrene ester polymers

Improvement of viscosity–temperature behavior

Detergents (basic)

Metal sulfonates Metal phenolates Metal salicylates (metal = Ca; Mg; Na)

Keeping the engine interior clean Neutralization of acids Prevention of lacquer formation

Dispersants (ash-free)

Polyisobutene succinimides

Dispersion of, aging products and other foreign materials Prevention of deposits and lacquer formation

Oxidation inhibitors

Zinc dialkyldithiophosphates Alkylphenols Diphenyl amines Metal salicylates

Prevention of oil oxidation and thickening

Corrosion inhibitors

Metal sulfonates (metal = Na; Ca) Organic amines Succinic acid half-esters Phosphorous amines, amides

Prevention of corrosion

Nonferrous-metal deactivators

Complex organic sulfur and nitrogen compounds

Prevention of oxidation and oil thickening

Friction modifiers

Mild EP additives Fatty acids Fatty acid derivatives Organic amines

Reduction of friction loss

Wear reducers (EP additives)

Zinc alkyldithiophosphate Molybdenum compounds Organic phosphates Organic sulfur and sulfur-phosphorous compounds

Reduction or prevention of wear

Pour point depressant

Polyalkyl methacrylates

Improvement of flow properties at low temperatures

Foam inhibitors (defoamant)

Silicone compounds Acrylates

Reduction or prevention of foam formation

Figure 22.68 Typical engine oil additives [22-14].

and long-lasting oils for extremely long oil change intervals are all fully synthetic, and their basic liquids have a very high V.I., such as 130. These also receive additional VI improvers to create SAE classes 10W-60 or 0W-40, although in comparatively lower quantities. Because V.I. improvers are, because of their size, subject to strong shear stresses, synthetic engine oils with less V.I. improvers are generally more shear-stable and can, therefore, maintain their viscosity properties over extended oil change intervals. Syrene butadiene copolymers, polymetacrylate (PMA), or olefin copolymers (OCPs) are added predissolved to POA or mineral oil. The amount of additive in market-ready oil is generally 1%–10% (m/m). 22.2.7.2 Detergents/Dispersants Many combustion products occur that damage the engine oil during combustion in spark-ignition and diesel engines. This includes oil-aging products, partially combusted and uncombusted fuel residue, soot, acids, nitrogen oxides, and water. These largely oil-insoluble solids or liquid foreign materials enter the oil circuit and have an undesirable or damaging effect. Resin- and asphalt-like oil aging products cause deposits on metal surfaces, oil thickening, and sludge deposits on engine parts. Acidic combustion products generate

corrosion, catalyze oxidation, and can break down wear protection additives. Carbon- and lacquer-like deposits cause the piston rings to cake tight in the ring grooves so that more blow-by gas enters the crankcase and further burdens the oil. In addition, seized piston rings cause bore polishing on the cylinder walls that can lead to performance loss and increased oil consumption. Sludge deposits can plug oil lines and the oil filter and cause scoring at the valve train and on the piston and cylinder barrels as a result of insufficient lubrication. Detergent and dispersant additives are essentially soaps. They encase solid and liquid foreign particles and keep them suspended in the oil to prevent them from depositing on engine parts and agglomerating, which could lead to sludge formation. In addition, acidic products are neutralized. The mechanisms of the action of detergent and dispersant additives can be generally categorized as follows: •• encase and wash •• keep clean and hold solid foreign particles in suspension •• encase liquid foreign particles and hold them in suspension (interface active) •• chemically neutralize acidic components.

Internal Combustion Engine Handbook | 847

6606_Book.indb 847

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Several specially harmonized, multifunctional agents are used as detergents and dispersants. They also protect against corrosion and neutralize acids. A long-lasting basic reserve (TBN = total base number) must be ensured, considering the very long oil change intervals. Usually metal–organic compounds such as phenates, phosphates, sulfonates, salicylates, and naphthenates are used that are made basic with excess metal carbonate. When they participate in combustion, they form ash in which calcium, magnesium, sodium, and zinc can be found, depending on the additive type. For this reason, low-ash engine oils have a relatively lower TBN because of their basic principle. Polyisobutene succinimide is also an ash-free organic detergent that works well against cold sludge deposits that tend to arise during “stop-and-go operation.” The amount of additive is generally 1%–5% (m/m). 22.2.7.3 Antioxidants and Corrosion Inhibitors Even the most superior lubricating oil tends to oxidize under the influence of heat and oxygen, that is, age or become rancid. Acids form as well as lacquer deposits, resin deposits and sludge-like deposits that are largely oil insoluble. The addition of antioxidants substantially improves the aging protection. Initially, aging is very slow and the oil scarcely changes. After the antioxidants are used up, the oxidation speed increases and a rise in oil temperature of 10°C doubles the reaction speed and significantly reduces the oil service life up to 50%. This process can be accelerated by trace amounts of metals, in particular, copper and iron (whose activity increases as they grow finer) that enter the oil from abrasive and corrosive wear and substantially reduce the reaction temperature with oxygen. Water can also have this effect. Without highly effective antioxidants, today’s conventional oil change intervals would be inconceivable. The mode of action of antioxidants primarily originates from free-radical scavengers and is supplemented by peroxide replacers and passivators. Radicals are hydrocarbon molecules in which free, highly reactive valences arise on the carbon from chain breakage. Oxygen or another radical immediately seeks to bond to the carbon. The free-radical scavengers saturate the free valence by transferring hydrogen from the additive. The peroxide replacers work only when oxygen-containing aging products have formed. They react with the oxygen and form nonreactive compounds. The nonferrous metal passivators are chemical substances of the triazole type that weaken the otherwise catalytic effect of copper and iron particles by encasing the metal ions in the oil. They accordingly also protect the surface of bearing materials from the corrosive attack of active sulfur, for example. 22.2.7.4 Friction and Wear Reducers (EP/AW Additives) When parts gliding against each other under high pressure and temperature are no longer fully separated by the lubricant, the surfaces of the friction partners contact and increased wear occurs or, in an extreme case, seizure or even welding. EP/AW additives (extreme pressure/antiwear) can be of assistance in this situation. They form very thin coatings on the sliding surfaces of the friction partners and are continuously renewed as needed. They are solid under normal conditions but can slide upon wear and prevent direct metal-to-metal

contact. These are interface-active materials that contain, among other things, zinc, phosphorous, and sulfur in the polar group. The most famous is zinc dialkyldithiophosphate that has been most successful in the area of mixed friction of the cam, tappet and rocker arm. Transmission oils use phosphoramines, thionates, and various phosphorous–sulfur compounds, causing the typical sulfur odor, in addition to sulfurized esters and hydrocarbons. 22.2.7.5 Foam Inhibitors Air or other gases can be present as finely distributed bubbles or surface foam in the lubricating oil. The primary contributing factor is air from the swirling in the crankcase, as well as pressure and temperature. Surface foam can be dispersed by a special agent that reduces the surface tension between the oil and air. Foam suppressors must be largely insoluble in the oil and have a lower surface tension than the oil. Very low concentrations (0.01 g/kg oil) of silicone oils such as polydimethylsiloxane have been successful as the active ingredient. However, the removal of the dispersed air can become more difficult because the silicone oil prevents the recombination of small air bubbles into larger, easily rising bubbles.

22.2.8 Engine Oils for Four-Stroke Engines

The operating conditions to which engine oils are exposed extend from extremely short-distance use—50% of all travel in passenger cars is over distances less than 6 km—to extremely long-term loads over long distances. In addition, we now have engine oil change intervals of up to 30,000 km or a maximum two years for passenger cars with spark-ignition or diesel engines, and up to 150,000 km for commercial vehicles on long-haul routes. The top-up requirement is also very low. This means, that only small amounts of fresh additives are provided to the engine. The oil volume in the oil pan of modern engines does not increase on par with the power density of the engine. Downsizing is a principle that is followed very strictly by engine designers by reason of consumption and emission issues. The specific oil load, hence, rises continuously. Specific average outputs of more than 60 kW/l are not rare in two-stage supercharged diesel engines—at average pressure above 20 bar. The engine oil is, hence, subject to substantial thermal and mechanical stress. In addition, the oil as a hydraulic fluid needs to efficiently accomplish many tasks in the engine such as hydraulic valve lash compensation, camshaft timing, and chain tightening under every operating condition during its entire time in the engine. 22.2.8.1 Society of Automotive Engineers Viscosity Classes for Engine Oils In 1911, the SAE (USA) introduced a binding classification for engine oil viscosity that is still valid today after many adaptations. In its present version, a total of twelve classes are defined: Six for the winter (0–25W) and six for the summer (20–60). Figure 22.69 shows the viscosity grades for engine oils according to SAE J300 of 4/2002. They inform the user that he is using oil with the right viscosity stipulated by the engine manufacturer.

848 | Internal Combustion Engine Handbook

6606_Book.indb 848

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

Low-Temperature Pump Viscosity SAE Viscosity Class

Maximum Apparent Viscosity at °C cP/°C

cP/°C

 0W

Kinematic Viscosity at 100°C

Max.

Maximum Pump Threshold Temperature °C

Min.

Max.

6200/−35

6,000/−40

−35

 3.8

—

 5W

6600/−30

6,000/−35

−30

 3.8

—

10W

7000/−25

6,000/−30

−25

 4.1

—

15W

7000/−20

6,000/−25

−20

 5.6

—

20W

9500/−15

6,000/−20

−15

 5.6

—

25W

13,000/−105

6,000/−15

−10

 9.3

—

cSt

HTHS1-Viscosity in cP at 150°C and 106 s–1 Shear Gradient Min.

20

 5.6

< 9.3

2.63

30

 9.3

< 12.5

2.93

40

12.5

< 16.3

2.92

40

12.5

< 16.3

3.73

50

16.3

< 21.9

3.73

60

21.9

< 26.1

3.73

High-temperature high-shear viscosity. for 0W-40, 5W-40, 10W-40. 3 for 15W-40, 20W-40, 25W40, and 40. 1 2

Figure 22.69 SAE viscosity classes for SAE J300 engines-lubricating oils, as of April 2002 [22-4].

22.2.8.2 Single-Grade Oils Single-grade engine oils meet the viscosity requirements of only the individual SAE grades 0W–60 shown in Figure 22.69. These oils, therefore, have a low V.I. and are accordingly suitable only for undemanding engines that run primarily under unchanging conditions at practically the same temperature, such as stationary engines for power generators. Single-grade oils must be frequently replaced according to season and deployment conditions which prevents their use as all-year engine oil and makes them, therefore, technically less interesting. Today, they are mostly used in some specific gear drives and retarder brakes. 22.2.8.3 Multigrade Oils The term “multigrade oil” means that such an oil covers the viscosity requirements of several SAE grades, for example, 5W-30. It covers the low W class in the low temperature range and ends with the high-temperature viscosity class at 100°C. Figure 22.70 shows in bold the most common combinations in Central Europe with other possible combinations that, however, are technically or geographically meaningless in this context. 0W-40

5W-40

10W-40

0W-30

5W-30

10W-30

15W-40 15W-30

0W-20

5W-20

10W-20

15W-20

The viscosity combinations in bold are the most common.

Figure 22.70 Multigrade oil viscosity classes [22-4].

The com​bination 0W-40 can meet maximum demands on the viscosity temperature behavior and requires a fully synthetic basic liquid and a particularly shear-resistant V.I. improver (see Section 22.2.5.3) to maintain the physical requirements. The combination 15W-20 meets the lowest requirements and is made with a mineral oil-based basic oil and a relatively low

V.I. improver, but it is technically irrelevant. In multigrade oils, the low-temperature viscosity is determined by the basic liquid, whereas the high-temperature viscosity is set by the effect of the V.I. improver. The very high performance of modern multigrade oils derives primarily from combining synthetic basic liquids with highly effective additive packages plus temperature and shear-resistant V.I. improvers, without which today’s extremely long oil change interval would not be possible. 22.2.8.4 Fuel Economy Oils Multigrade oils whose low-temperature viscosity lies between SAE grades 0W or 5W are termed LL engine oils in German (Leicht-Lauf oils) or FE engine oils (fuel economy oils). Fuel consumption is clearly reduced in two ways: •• by lower viscosity during full lubrication (hydrodynamic lubrication) •• by friction-reducing additives for boundary lubrication (mixed friction). Reducing viscosity has the greatest influence since primarily hydrodynamic lubrication predominates in the engine. The effect of friction modifiers for mixed friction is narrowly restricted. Overall friction losses are shown in Figure 22.71. Engine

Full Load

Partial Load

Gasoline

3%–5%

11%–18%

Diesel

7%–9%

13%–14%

Figure 22.71 Friction loss [22-13], [22-15].

It turns out that the largest reduction of consumption is within the partial load range close to idling. The first FE oils were SAE grade 10W-X, and then came 5W-X and 0W-X.

Internal Combustion Engine Handbook | 849

6606_Book.indb 849

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

However, all other requirements for hot operation must be covered, and the evaporation loss must remain low enough to keep oil consumption at a minimum. For example, ACEA specification A3 or A5 demand a loss of less than 13%. According to the generally valid fuel consumption test according to CEC L-054-96, the FE oil is compared to a reference oil with a viscosity of SAE 15W 40. In addition, tractive torque measurements, for example, using the Mercedes-Benz OM 441 LA engine, can be used to evaluate the friction reduction from an FE oil for commercial vehicles. A typical measuring result of a comparison between the earlier used 15W-40 oil and a fuel economy oil 5W-30 is shown in Figure 22.72. Both oils also had to fulfill all the other requirements from the other test runs (for example, according to MB 228.3). Multigrade Oil for Commercial Vehicles

Tractive Torque (Nm)

15W-40 (MB 228.3)

285

5W-30 (MB 228.3)

244 (–17%)

Figure 22.72 Tractive torque in the OM 441 LA engine [22-15].

In addition, fuel economy oils also clearly improve pumpability at low temperatures to ensure a much faster supply of oil to the engine after a cold start. This is true for both fresh oil and used oil, as can be seen in Figure 22.73.

Multigrade Oil for Commercial Vehicles

Fresh Oil Oil Pressure 2 bar (s)

Used Oil Oil Pressure 2 bar (s)

5W-30 (MB 228.3)

 7

 9

10W-40 (MB 228.3)

10

14

15W-40 (MB 228.3)

23

35

Figure 22.73 Pumpability in the OM 441 LA engine 0°C [22-15].

Also notable in these results is the low degree of thickening of the used oil as a result of the high quality of these oil types. 22.2.8.5 Break-in Oils At an earlier time, it was necessary to fill new engines with a special break-in oil or initial operation oil that was drained after a relatively short dwell time of 1000–1500 km and contained the metal abrasion from this first operation phase. There were special break-in instructions, and the driver had to drive cautiously to prevent the thin initial operation oils with relatively few additives from being overly stressed. Frequently, they were designed as preservative oil when, for example, the vehicles were to be exported overseas. Because of the improved surface quality of the friction surfaces of the engine and the enormous strides in oil technology, the use of break-in oils in passenger cars and commercial vehicles is no longer necessary. The performance of initial operation oils practically corresponds to off-the-shelf engine oils since they also remain in the engine for a full change interval; however, they generally have somewhat higher corrosion protection. 22.2.8.6 Gas Engine Oils The use of CNG/LNG in automobiles require special engine oils in monovalent operation because the higher combustion

temperatures, and corresponding elimination of the vaporization latent heat, increase the tendency of deposits to form in the combustion chamber and on the pistons. These particularly hard deposits require low-ash additives. In comparison to gasoline, approximately twice the amount of water arises from the combustion of natural gas because of the high HC ratio. Hence, the danger of corrosion is especially greater in shortdistance driving. The absence of high-boiling hydrocarbons in natural gas also generates a greater need for lubrication at the intake valves, which causes higher wear of the valve seat. In bivalent operation, these specific properties are not as predominant, and modern high-performance oils can cover nearly every requirement. 22.2.8.7 Hydrogen Engine Oils In the joint research on alternative fuels, the influence of hydrogen on engine oil in hydrogen-operated spark-ignition engines was also investigated. The different combustion processes in comparison to gasoline clearly affect the engine oil requirements. In the combustion of hydrogen, more than twice the amount of water arises as a combustion product in comparison to the combustion of conventional gasoline. Since additional water is injected into the combustion chamber in some engine designs for regulated combustion, the related increased water vapor in the engine oil under cold-driving conditions requires greater corrosion protection, stronger dispersability, and a greater ability to absorb and release water. On the other hand, there is much less combustion residue and resulting deposits so that the detergent content can be reduced. Recent engine designs for hydrogen operation get by without additional water injection. The risk from additional water to the engine oil has therefore receded. These hydrogen engines can be operated with commercial engine oils. However, experience from operating spark-ignition engines with hydrogen is insufficient to draw any conclusions on the optimum composition of engine oils used in this context. 22.2.8.8 Performance Classes Given the different conditions of use and demands on engine oils, many specifications have arisen over time concerning their composition and performance. These were usually worked out jointly by the engine and mineral oil industries with the collaboration of consumer organizations or military authorities. In addition, the individual automobile manufacturers keep publishing a growing quantity of brand-related specifications. Beyond worldwide approval requirements, there are also regulations limited to the regions of use. These describe both the physical properties and the performance behavior in engine tests. The specifications of, among others, the following associations and institutions describe performances behavior: •• ACEA Association des Constructeurs Européens d’Automobiles •• API

American Petroleum Institute

•• CCMC Comitée des Constructeurs d’Automobiles du Marché Commun

850 | Internal Combustion Engine Handbook

6606_Book.indb 850

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

•• CEC Coordinating European Council for the development of performance tests for transportation fuels, lubricants and other fluids •• MIL-L US MILITARY Lubricants Specification •• DEF British Ministry of Defence •• OEM Original Equipment Manufacturer Þ Automotive manufacturers.

heavy commercial vehicle diesel engines. Since 2004, only combinations of A and B groups have been possible (A1/B1, A3/B3, A3/B4, and A5/B5). In addition to the number for the performance classes, they can also include the year in which the respective specification took effect, for example, ACEA A3-98. Figure 22.74 presents the ACEA specifications for service fill oils for spark-ignition engines with their most important features.

Usually, a distinction is drawn among passenger car spark-ignition engines, passenger car diesel engines, and commercial vehicle diesel engines. Since 1996, the European ACEA specifications have been the successor to the no longer current CCMC specifications that have existed since the 1970s. In addition, API classifications are frequently required. The MIL specifications are now irrelevant in Europe. In recent years, the individual requirements and the related releases of the European automobile manufacturers have become the most important.  CEA specifications A These represent the current standards for engine oils for European automobile engines. In addition to a few selected American test engines, they cover primarily engines of European construction and design. The test conditions correspond to European driving conditions. They define the minimum requirements for physical and chemical laboratory tests and full engine test bench tests. Since a few American engine tests are also prescribed, there is a certain amount of overlap with American API classifications. The abbreviations for the ACEA specifications are no longer oriented around the previously conventional names of the outdated CCMC specifications. Whereas CCMC G (gasoline) was used for spark-ignition engines, CCMC D for commercial vehicle diesel engines and PD for passenger car diesel engines, in the ACEA specifications these categories are now termed ACEA A, ACEA B, and AECA E. The group A concerns spark-ignition engines, B light or passenger car diesel engines, and group E concerns

1

1

ACEA

Oil Type

Important Requirements

A1-08

Fuel economy engine oils

HTHS viscosity 2.6–3,5 mPas, Fuel Economy (FE)1

A2-04 Invalid since 2007

Former standard engine oils

HTHS viscosity >3.5 mPas Requirements increased in respect to: Evaporation loss, wear, cleanness, black sludge, oxidation stability)

A3-08

Premium engine oils

HTHS viscosity >3.5 mPas Requirements increased in respect to: Shearing stability, wear, cleanness, black sludge, oxidation stability

A5-08

Premium fuel economy engine oils

HTHS viscosity min. 2.9, max. 3.5 mPas on performance level ACEA A3, FE1

Evidence of fuel savings in MB M111 test with a SAE 15W-40 reference oil.

Figure 22.74 ACEA specification for spark-ignition engines [22-4], [22-15].

In contrast to the commercial vehicle sector, passenger car diesel engines require special additives that are very similar to those for spark-ignition engines because of the higher speed, higher specific performance, higher valve gear load, and more frequent short-distance use. Figure 22.75 shows the ACEA specifications for service fill oils for passenger car diesel engines. Commercial vehicle diesel engines are used under a wide range of operating conditions such as in government vehicles and municipal buses with a continuously changing load at low speeds and in long-distance traffic with a continuous high load at higher speeds. Engine oils for commercial vehicles are

ACEA

Oil Type

Important Requirements

B1-08

Fuel economy engine oils

HTHS viscosity max. 3.5 mPas, FE1

B2-04 Invalid since 2007

Former standard engine oils

HTHS viscosity >3.5 mPas Requirements increased in respect to: Evaporation loss, wear, cleanness, oil thickening, oil consumption

B3-08 Issue 2

Premium engine oils

HTHS viscosity >3.5 mPas Requirements increased in respect to: Evaporation loss, viscosity layer stability, wear, cleanness, oil thickening, oil consumption

B4-08

Premium engine oil for direct-injecting diesel engines

Compared to B2, higher requirements in respect to piston cleanness in diesel engines with direct injection

B5-08

Premium fuel economy engine oils for direct-injecting diesel engines

HTHS viscosity min. 2.9, max. 3.5 mPa s on higher performance level than ACEA B1

C1-08

Low SAPS engine oils for diesel engines with DPF

Low ash content to 0.5%, HTHS > 2.9, higher FE requirement1

C2-08

Low SAPS engine oils for diesel engines with DPF

Medium ash content to 0.8%, HTHS > 2.9, higher FE requirement1

C3-08

Low SAPS engine oils for diesel engines with DPF

Medium ash content to 0.8%, HTHS > 3.5

C4-08

Low SAPS engine oils for diesel engines with DPF

Low ash content to 0.5%, HTHS > 3.5, higher requirement in respect to evaporation properties

Evidence of fuel savings in MB M111 test with a SAE 15W-40 reference oil.

Figure 22.75 ACEA specification for passenger car diesel engines [22-4], [22-15].

Internal Combustion Engine Handbook | 851

6606_Book.indb 851

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

ACEA

OIL TYPE

Important Requirements

E1-96 Invalid since 1999

Standard engine oils

basic requirements

E2-04 Invalid since 2007

Standard engine oils with increased requirements

Requirements increased in respect to: Bore polishing, piston cleanness, cylinder wear, oil consumption

E3-07 Invalid since 2007

Engine oils for commercial vehicles with ATL engines; corresponds approximately to MB Sheet 228.3/MAN 271

Requirements increased in respect to: Oil consumption, sludge formation, viscosity increase at high soot content in the oil

E4-08

Premium engine oils for commercial vehicles with ATL Higher requirements in respect to: Bore polishing, piston cleanness, engines and air receiver; corresponds approximately to cylinder wear. In addition, restriction of deposits in the turbocharger, MB Sheet 228.5/MAN 3277 extended service intervals, TBN > 12mgKOH/g

E6-08

Premium engine oils for European commercial vehicle engines with exhaust gas aftertreatment, DPF in particular.

Compared to E4, reduced content of substances forming sulfate ash, suitable for diesel particulate filters, for regions with sulfur-free fuels.

E7-08

Premium engine oils for European commercial vehicles

Compared to E4, increases requirements (wear, oxidation, shear stability)

E9-08

Premium engine oils for European commercial vehicle engines with exhaust gas aftertreatment, DPF in particular.

Compared to E6, increased requirements (wear, oxidation); compared to E7, reduced content of substances forming sulfate ash, suitable for diesel particulate filters, for regions with sulfur-free fuels.

Figure 22.76 ACEA specification for heavy commercial vehicle diesel engines [22-4], [22-15].

different in many respects from those for passenger cars. For example, the requirements are particularly high wear protection for the cylinder barrel, very clean pistons over the long term, high dispersability of soot, reserves and high performance for extremely long oil change intervals and low residue formation in turbochargers and intercoolers. Given the present level of technology for European commercial vehicle engines, these requirements can be achieved only by using superior additives. Similar to the requirements in the passenger car sector, the exhaust gas aftertreatment systems pose special demands on the engine oils in commercial vehicles. Here too, the portion of sulfate ash-forming substance is relevant for particulate filters, for example. Figure 22.76 shows the ACEA specifications for heavy commercial vehicle diesel engines. Section 22.2.8.8 provides information on the engine tests stipulated for the individual classes.  PI classifications A The description of performance requirements of engine oils for automobile engines was started much earlier in the United States than in Europe. In its engine oil classifications, API draws a distinction only between passenger car engines and commercial vehicle engines. Because the number of passenger cars with diesel engines in the United States is very low, they have no classification for light diesel engines. The engines used for the tests are made in the United States, and the test conditions favor American driving conditions. All the viscosity grades according to SAE are permissible. While the engine tests up to API SG can be done by the respective manufacturers, they need to be reported and registered since the introduction of API SH (CMA Code) when an API category for the product is claimed. At the same time, it is possible to obtain a license through API that allows an API label to be affixed to the packaging. It can be assumed that, in the course of 2005, the next level of API classification will be introduced: the API SM. Figure 22.77 shows the present API specifications for passenger car engine oils.

These API classes have been on engine oil packaging for decades in addition to the SAE grades and are listed in the operating instructions. The user can then check if the quality corresponds to that prescribed by the automobile manufacturer. The API classification for commercial vehicles is more multifaceted since the design of commercial vehicle diesel engines made in America is sometimes substantially different from that of European models. In earlier classifications, they refer to the MIL specifications (see Section 22.2.8.3). In addition, it should be noted that two-stroke diesel engines can frequently be found in the American market, yet not so much in Europe. Among the test runs stipulated almost exclusively for the Caterpillar, single-cylinder diesel engine have been incorporated into later test runs for modern engines by Caterpillar, Cummins, Mack, and Detroit Diesel. Figure 22.78 summarizes the API classifications for commercial vehicles. Parallel to the API classifications in the United States are the ILSAC certification (International Lubricant Standardization and Approval Committee) that uses the API classifications for passenger car engines in cooperation with the American Automobile Manufacturers Association and Japan Automobile Manufacturers Association to offer a classification of oil quality and usefulness for engine oil packaging that is more consumer-oriented. The outdated ILSAC GF-1 corresponded to API SH, while ILSAC GF-2 corresponds to API SJ. API SL and ILSAC GF-3, as well as API SM and ILSAC GF-4 (with FE test) are valid.  IL specifications M Engine oil specifications have been established for automobiles in the U.S. Army since 1941. The requirements have been continuously adapted as these engines have developed. The term “HD oils” (heavy duty) arose in this context for oils under high stress in diesel engines, and it continues to be used by consumers to this day. This represented the transition from exclusively unalloyed mineral oils to alloyed oils

852 | Internal Combustion Engine Handbook

6606_Book.indb 852

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

1 2

API Class

Year of Introduction

SA1

1925

Important Requirements Unalloyed engine oils. Addition of pour point improvers and foam suppressors possible.

SB

1930

Slightly alloyed engine oils with lower wear, aging and corrosion protection.

SC

1964

Engine oils with increased protection against scoring, oxidation, bearing corrosion, cold sludge and rust.

SD

1968

Improvement of API SC with increased protection against scoring, oxidation, bearing corrosion, cold sludge and rust.

SE

1972

Improvement of API SD with increased protection against oxidation, bearing corrosion, rust and lacquering.

SF

1980

Improvement of API SE with further improved protection against oxidation and wear.

SG

1989

Improvement of API SF with further improved oxidation stability and improved wear protection.

SH

1992

Corresponds to API SG, however, the engine tests for API SH must be registered with a neutral institute, unlike API SG.

SJ2

1997

Corresponds to API SH with additional lab testing against high-temperature deposit formation. Regulated exchangeability of basic liquids, stricter test regulations regarding read-across.

SL2

2001

Corresponds to API SJ with stricter high-temperature resistance, lower evaporation loss and higher wear protection.

SM2

2005

Corresponds to API SL, raised wear protection, corrosion protection and aging behavior, restriction of phosphorous and sulfur contents.

Service Fill. Valid.

Figure 22.77 API classifications for passenger car engine oils [22-4].

1 2

API Class

Year of Introduction

CA1

Mid-1940s

For naturally aspirated diesel engines and occasionally low-load spark-ignition engines. Protection against bearing corrosion and ring-groove deposits.

CB

1949

For naturally aspirated diesel engines using poorer-quality diesel fuel with a high sulfur content. Occasionally also in spark-ignition engines. Protection against bearing corrosion and ring-groove deposits.

CC

1961

For naturally aspirated diesel engines with medium load. Occasionally also in spark-ignition engines with high loads. Protection against high-temperature deposits, bearing corrosion and cold sludge in spark-ignition engines.

CD

1955

For naturally aspirated, supercharged and highly supercharged turbodiesel engines using diesel fuel with very high sulfur content. Increased protection against deposits in the ring--groove area at high temperatures and against bearing corrosion.

Important Requirements

CD-II

1985

For two-stroke diesel engines with increased requirements of wear protection and deposits.

CE

1984

For highly supercharged diesel engines under high loads at low and high speeds. Improved protection against oil thickening, piston deposits, wear and oil consumption, compared to API CD.

CF-42

1990

Improvement over AICE in piston cleanness and oil consumption.

CF2

1994

As CD, but for indirect-injection diesel engines with a wide range of diesel fuels with sulfur content above 0.5% (m/m). Improved control of piston cleanness, wear and bearing wear.

CF-22

1994

For two-stroke diesel engines with increased requirements on cylinder and piston ring wear, also improved deposit control.

CG 42

1994

For high-load, high-speed four-stroke diesel engines in street use, as well as off-road use, with diesel fuel with a sulfur content of 0.5% (m/m). Specially suited for engines meeting the emission standards of 1994. Also covers API CD, CE and CF-4. In addition, increased oxidation stability and protection from foaming.

CH-42

1998

Compared to CG-4, further increased requirements for diesel engines meeting the emission standard of 1998. Sulfur content in diesel fuel up to 0.5% (m/m). During extended oil change intervals, increased protection against noniron corrosion, thickening because of oxidation and oil-insoluble contaminations, foaming and shear loss.

CI-4

2002

For fast-running four-stroke diesel engines meeting the 2004 emission standard. Requirement in the context of exhaust gas recirculation and fuel sulfur content of up to 0.5% m. Higher requirements compared to CH-4.

CJ-4

2006

For fast-running, four-stroke diesel engines meeting the 2007 emission standard. Requirement in the context of exhaust gas recirculation and fuel sulfur content of up to 0.50% m. Exceeds the requirements of CI-4, taking into account the requirements of exhaust gas aftertreatment systems such as oxidation catalytic converters and particulate filters.

Commercial = Large consumer sector. Valid.

Figure 22.78 API classifications for commercial vehicle engine oils [22-4].

Internal Combustion Engine Handbook | 853

6606_Book.indb 853

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

Manufacturer

Specification

Designation

Engine Type

Requirements

BMW

Special oils and long-life oils

Special oil:longlife oil

Passenger car gasoline and diesel

ACEA A3/B3 plus additional BMW engine and foaming test, FE oils0W-X and 5W-X. Long-life oils for flexible oil change intervals, version 04 with Low SAPS requirement for DPF

M 3271-1

Commercial vehicle gasoline

CNG/LPG special oils

M 3275

Commercial vehicle diesel

ACEA E3, more stringent physical requirements, high-performance oils

M 3277

Commercial vehicle Diesel

ACEA E3 plus OM 441 LA according to MB sheet 228.5 plus deposit test, high-performance oil for max. long oil change intervals

M 3477

Commercial vehicle Diesel

Sulfate ash content max. 1.4 m% for engines with exhaust gas aftertreatment

MB Sheet 229.3/229.5

Passenger car SI and ACEA A3, B3 and B4 plus MB engine tests and special requirements, diesel high-performance multigrade oils for significantly extended intervals

MB Sheet 229.31/229.51

Passenger car SI and ACEA A3, B3, B4 plus further engine tests, Low SAPS requirement for diesel DPF

MB Sheet 229.1

Passenger car SI and ACEA A2 or A3 and B2 or B3 plus MB engine tests, multigrade oils for diesel extended interval

MB Sheet 227.0

Commercial vehicle diesel

ACEA E1 plus additional evaluation criteria in OM 602A, single-grade oils for short oil change intervals

MB Sheet 227.1

Commercial vehicle diesel

ACEA E1 plus additional evaluation criteria in OM 602A, multigrade oils for short oil change intervals

MB Sheet 228.0

Commercial vehicle diesel

ACEA E2 plus additional stricter evaluation criteria in OM 602A, singlegrade oils for normal oil change intervals

MB Sheet 228.1

Commercial vehicle diesel

ACEA E2 plus additional stricter evaluation criteria in OM 602A, multigrade oils for normal oil change intervals

MB Sheet 228.21

Commercial vehicle diesel

ACEA E3 plus additional further tightened evaluation criteria in OM 602A, single-grade oils for extended oil change intervals

MB Sheet 228.31

Commercial vehicle diesel

ACEA E3 plus additional further tightened evaluation criteria in OM 602A, multigrade oils for extended oil change intervals, SHPD2 type

MB Sheet 228.51

Commercial vehicle diesel

ACEA E4 plus additional further tightened evaluation criteria in OM 602A, multigrade oils for maximum oil change intervals, USHPD3 type

MB Sheet 228.51

Commercial vehicle diesel

ACEA B3, B4 and E6-06 plus MB in-house test

501 01

Passenger car SI and ACEA A2 plus VW-specific engine and aggregate tests. Standard naturally aspirated multirange oils diesel

505 005

Passenger car naturally aspirated and ATL diesel

ACEA A3 plus VW-specific engine and ACEA B3 plus VW-specific diesel engine and aggregate tests Standard or FE multirange oils for normal oil change intervals up to approx. the end of model year 1999.

502 00

Passenger car SI

ACEA A3 plus VW-specific engine and aggregate tests under special inclusion of long-term stability. Standard FE multirange oils

505 01

Passenger car diesel

SAE 5W-40 special oil for DI diesel engines with pump-nozzle fuel injection systems, normal intervals. VW-specific engine and aggregate tests under special inclusion of long-term stability and fuel economy. HTHS viscosity reduced to ≥2.9 and <3.4 mPas. Comprehensive manufacturer tests. For vehicles from approx. model year 2000 with extended oil change intervals. Not suitable for older vehicles.

503 006

Passenger car SI

Specific in-plant testing in Audi ATL SI engines with high specific output

503 018

Passenger car SI

Successor of 503 00, HTHS viscosity >3.5; backward compatible

504 007

Passenger car SI

HTHS viscosity reduced to ≥2.9 and <3.4 mPas.

506 006

Passenger car diesel

Comprehensive manufacturer tests. For vehicles with DI diesel engines without pump-nozzle injection system, from model year 2000 with extended oil change intervals. Not suitable for older vehicles.

506 01

Passenger car diesel

Comprehensive in-plant testing under particular consideration of longterm stability and fuel economy. HTHS viscosity reduced to ≥2.9 and <3.4 mPas. For vehicles with DI diesel engines with pump-nozzle injection system and extended oil change intervals.

507 00

Passenger car diesel

Successor of 506 00/506 01, HTHS viscosity >3.5; backward compatible with some few exceptions, low SAPS requirement for DPF

MAN

Mercedes

MAN standards

MercedesBenz Operating fluid regulations

4,5

VW/Audi

VW standard

For XW-30 or 0W-40 multirange oils, additional testing in OM 441LA-Test with predimensioned bearings and tappets. 2 Super High Performance Diesel oil. 1

Ultra Super High Performance Diesel oil. No new releases have been granted since 1997. 5 Combinations are possible and common as 502 00 and 505 00. 3

6

4

7 8

Only combined together. Only in combination with 507 00. Obsolete from 04/2009, replaced with VW 504 00.

Figure 22.79 Important engine oil specifications of some automobile manufacturers [22-4], [22-17].

854 | Internal Combustion Engine Handbook

6606_Book.indb 854

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

that received chemical additives for the first time. The only slightly alloyed oils for spark-ignition engines were termed “premium oils” in contrast to the HD oils for diesel engines. Although the MIL specifications were originally intended only for military use, they were used throughout the world for a long time after World War II in the civilian sector in performance recommendations for engine oils. For the military, the specification MIL-PRF-2104G has been valid since 1997. It allows single-grade engine oils SAE 10W, 30, and 40, and multigrade SAE 15W-40 as viscosities. The requirements of these specifications correspond to elements from API CF, CF2, and CG4. In addition to fulfilling chemical–physical requirements, the oils for tactical vehicles in the U.S. Army must also fulfill special friction tests because of the specific construction of tactical military vehicles such as tanks. For a few years, the use of “MIL” to identify oil performance has been allowed only if the corresponding oil is approved by the American military.  utomobile manufacturer specifications A Beyond the API classifications and ACEA specifications, European automobile manufacturers require special performance classes for the release of individual engine oils that must be fulfilled in addition to API and ACEA specifications whose requirements are sometimes clearly exceeded. Given the amazing progress in engine technology, the requirements are subject to a continually accelerating process of change. The fulfillment of the special requirements is confirmed by a written release. Some automobile manufacturers provide lists of approved oils. The most important automobile manufacturer requirements are compiled in Figure 22.79. There are other special requirements, some of which are associated with the formal release, from other automobile and engines manufacturers such as Ford, Peugeot, Porsche, Renault, Rover (passenger cars and commercial vehicles), as well as DAF, Iveco, MTU, Scania, and Volvo (only commercial vehicles). ACEA

A/B/C

E

Engine test methods To fulfill the requirements established in the individual engine oil specifications, binding engine tests are stipulated in addition to the usual physical and chemical requirements. They are updated from time to time as needed, usually every two years in the case of the ACEA. Some automobile manufacturers only recognize tests that are done at neutral, specially approved test institutes. But first, let us take a historical retrospective of the many outdated test methods of the last decades. By the end of the 1950s, spark-ignition engine oils had to undergo the MS test sequences in the framework of API classification for American V8 engines by General Motors (Seq. I/II/III), Chrysler (Seq. IV), and Ford (Seq. V). To be certified according to MIL specifications, diesel engines had to pass the Caterpillar single-cylinder test that runs over 480 h with L IA/E naturally aspirated engines and L-1H, L-1D and L-1G ATL engines.In addition, oils underwent supplementary test runs L-38 and LTD over 40 or 180 h in the smaller CLR (coordinating lubricant research) Labeco single-cylinder engine. In Germany, at the beginning of the 1960s, the suitability of alloyed oils was tested in the MWM KD 12E single-cylinder diesel engine in test methods A and later B over a 50-h transit period in reference to piston cleanness and ring clogging. In England, oils were tested in the Petter AV.1 single-cylinder diesel engine with a 120-h transit period in combination with the Petter W. 1 single-cylinder spark-ignition engine with a 36-h transit period to attain DEF approval. Soon to follow was the certification test demanded by Daimler-Benz in MercedesBenz four-cylinder passenger car diesel engines. Rapid engine development, the demand for further improvements in reliability, longer service life, longer oil dwell times, and falling levels of oil consumption have required ever-new test engines and test methods to fulfill the requirements in the ACEA specifications. Today we have correspondingly suitable, Europe-wide specified motor oil tests that are listed in Figure 22.80.

Method

Test Designation

Type

Main Criteria

CEC L-53-T-95

Mercedes-Benz M111SL

R4-cylinder SI engine

Black sludge, cam wear

CEC L-54-T-96

Mercedes-Benz M111FE

R4-cylinder SI engine

Fuel economy

CEC L-88-T-XX

Peugeot TU-5JP-L4

R4-cylinder SI engine

High-temperature deposits, ring clogging, oil thickening

CEC L-38-A-94

Peugeot TU-3M S

R4-cylinder SI engine

Valve gear wear

ASTM D-659300

Ford Sequence VG

R4-cylinder SI engine

Low-temperature deposits, wear, ring clogging, piston cleanness

CEC L-78-T-99

VW TDI

R4-cylinder DI diesel engine

Oil thickening, piston cleanness

CEC L-093-04

Peugeot DV4TD

R4-cylinder DI diesel engine

Wear (valve gear, cylinder), piston cleanness, sludge

CEC L-099-08

Mercedes-Benz OM646LA

R4-cylinder DI diesel engine

CEC L-101-08

Mercedes-Benz OM501LA

V6-cyl. diesel engine ATL

Piston cleanness, sludge, oil consumption, bore polishing

CEC L-099-08

Mercedes-Benz OM646LA

R4-cyl. DI diesel engine

Camshaft wear

ASTM D 5967

Mack T-8E (or T11)

R6-cyl. ATL diesel engine

Oil thickening from soot

ASTM RR: D-2-1440

Cummins ISM

R6-cyl. ATL diesel engine

Oil filter clogging from soot, valve gear wear, sludge

Mack T12

Mack T12

R6-cyl. ATL diesel engine

Cylinder and ring wear, oil consumption

Figure 22.80 Engine tests for ACEA specifications [22-4], [22-15].

Internal Combustion Engine Handbook | 855

6606_Book.indb 855

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

In addition, engine oil tests for API classification and particularly those of the European automobile manufacturers need to be observed. For API class SJ for spark-ignition engines valid since 1996, the sequence VI A test is also provided for determining the fuel economy of an engine oil. For diesel engines, API CH-4 became effective in 1998 with test runs in the CAT I K and Cummins NTC 400. Test specifications and procedures are voluntarily observed, and the type and composition of lubricants being tested or developed are voluntarily maintained with the assistance of the European Engine Lubricant Quality Management System for engine oils, a joint initiative of ATC (Technical Committee of Petroleum Additive Manufacturers) and ATIEL. The European Technical Association of the European Lubricants Industry ATIEL and the European ATC have developed a fixed set of regulations (ATC Code of Practice and ATIEL Code of Practice) to which the member companies can voluntarily submit by providing an annual written letter of conformance. They thereby certify that the performance classes of the oils that they manufacture or sell are based on exact and controlled tests corresponding to the prescribed conditions of the two Codes of Practice carried out in test facilities certified according to EN 45001. The tests are reported and registered at the European Registration Centre which organization does not, however, publish any specific release lists. The list of the companies voluntarily participating in the quality assurance system is available to consumers and can be requested from ATIEL and ATC or viewed on the Internet. 22.2.8.9 Used Oil Evaluation While oil is in the engine, many foreign materials collect in it, primarily residues from fuel combustion such as soot. In diesel engines these are especially uncombusted hydrocarbons, acidic reaction products, abraded elements from engine wear and water. The load on the oil generated by these foreign materials consisting of liquid (low-molecular) and solid (high-molecular) aging and reaction products naturally changes the physical and chemical states of the used oils. Physically, the viscosity changes (usually by thickening but also by dilution with fuel condensate) particularly in cold seasons. The chemical changes particularly affect the alkalinity reserve as a measure of active ingredient consumption. These changes are evaluated, and the abrasion elements in the used oil are determined in a used oil analysis. This is an important tool for determining the state of the oils while developing the engine and engine oil, and for evaluating the state of the engine and the engine oil in relationship to the dwell time in the engines of large fleets. In evaluating used oil, the effects of different operating conditions are monitored. Passenger cars, in particular, second cars, are generally operated under stop-and-go conditions with many cold starts that are seldom broken up by long drives. On the other hand, approximately 10% of all users primarily utilize their automobiles for long trips with a continuously high load. Hot operation and cold operation, of course, have radically different effects on the engine oil condition. The most conventional points for analyzing used oil are the following:

•• dilution by fuel •• viscosity at 40 and 100°C •• alkalinity ⇒ Active ingredient reserve; base number and acid number ⇒ TBN/TAN •• dispersability •• nitration ⇒ black sludge •• overall soiling ⇒ solid foreign materials, oil-insoluble aging products •• abraded elements and dirt ⇒ iron, copper, chromium silicon content •• water and glycol content ⇒ leaky coolant circuit •• spectrometric infrared analysis according to DIN 51451 ⇒ identity.  hysical changes P The oil can thicken, that is, increase in viscosity during operation because of the evaporation of highly volatile oil components, from the increase in solid foreign materials from combustion, abrasion, and wear, and by oil aging from the oxidation and polymerization of oil components. Longer transit periods under a high load at a high speed promote a rise in viscosity. This makes cold starting and the supply of oil to critical lubrication sites more difficult and thereby increases fuel consumption. Oil thickening is, therefore, one of several important criteria for establishing the oil change interval. Some of the engine oil testing methods cited in Section 22.2.8.7 serve as a standard; however, the tests required by individual automobile manufacturers are the ones that are primarily used. The decrease in viscosity from oil dilution is primarily from fuel and water, in particular, in cold and short-distance driving. Uncombusted fuel components and water vapor from combustion condense in the cold engine, pass by the piston rings and enter the oil pan. In modern, low-polluting engines that use electronically controlled mixture enrichment in cold operation, the condensation tendency is less. Oil dilution also rises when there is incomplete combustion in a cylinder, for example, from spark plug failure or damage to the nozzle. In modern engines, this is the exception because an increased service life, higher component quality and electronically controlled ignition ensure reliable operation. Finally, multigrade oils can experience a permanent viscosity loss from the shearing of V.I. improvers when they are not sufficiently shear resistant (see Section 22.2.5.3). Figure 22.81 shows the substantial amount of oil dilution from extreme short-distance driving measured by the fuel in the used oil in fleet tests of SI engines in typical second car operation. The effect of a drive on the highway measured in a 1.4-L engine is notable. The oil dilution of 2.5% measured after evaporation of the fuel should correspond to the general dilution in alternating operation between short distances and long distances. However, FAME components from biodiesel can, because of their high boiling point of above 200°C, cannot be “evaporated out” by driving on the highway and will accumulate in the engine oil. Diesel vehicles with postinjection for DPF regeneration in particular are affected.

856 | Internal Combustion Engine Handbook

6606_Book.indb 856

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

Distance (km)

2.0-L Engine

1.8-L Engine

1.4-L Engine

Fuel Content in %

Fuel Content in %

Fuel Content in %

  1000

 3.5

 7.5

 5.5

  2000

 7.0

18.0

15.0

  4000

 6.5

20.5

12.0

  6000

12.5

19.5

10.0

  8000

15.2

20.5

11.5

10,000

15.6

27.0

17.5

12,000

18.5

 2.5*

14,000

17.0

 2.5

16,000

17.5

*After highway driving.

Figure 22.81 Oil dilution from fuel in spark-ignition engines after extreme short-distance driving [22-15].

We should not overlook the fact that the oil dilution important for establishing the oil change interval can be masked by the opposite effects of oil thickening. In modern engines, the coolant water circuit that is thermostatically controlled by an oil–water heat exchanger usually quickly heats the cold oil by first heating the coolant water. The oil is then quickly heated to operation temperature, and condensation products can evaporate. If the engine oil temperature rises during operation above that of the coolant water, the oil is cooled via the heat exchanger by the coolant water or the radiator. The rise in the metal content from extreme short-distance driving in three automobiles can be seen in Figure 22.82. Iron is used as an example that can originate from wear of the cylinder wall or the valve gear. The effect on engine wear of such engine oils strongly diluted by fuel is obvious. The investigation of the disassembled engines after traveling 10,000 km under these conditions shows that significant wear arose on the cylinder barrels, piston rings, bearings, and valve gear.

Distance in km

2.0-L Engine

1.8-L Engine

1.4-L Engine

Iron Content in mg/kg

Iron Content in mg/kg

Iron Content in mg/kg

  1000

 10

  7.5

 20

  2000

 15

 10 

 50

  4000

 25

 45 

 75

  6000

 40

 75 

 90

  8000

 80

110 

100 (7500 km)

10,000

100

250 

650

12,000

175

14,000

400

16,000

650

18,000

800

Figure 22.82 Metal abrasion in spark-ignition engines from extreme shortdistance driving [22-15].

Additional steps are required before the oil change interval can be substantially extended by additional bypass filters, both in spark-ignition and diesel engines. Spark-ignition engines have less filterable combustion products in the oil than diesel

engines. Hence, bypass oil filters are not recommended for spark-ignition engines. Diesel engines, especially commercial vehicle diesel engines with a large amount of oil can benefit from reducing the insoluble components in the oil. It has been shown, however, that most impurities remain in the oil. Most of the solid foreign materials in the oil have a particle size of 0.1–0.5 μ m, yet the pore width of the finest oil filter is much greater. The dispersant effect of oil additives is much stronger than the adsorbing force of the filter medium. The same holds true for low-molecular aging products. The filter is totally incapable of slowing the natural breakdown of additive efficiency. Only increasing the oil volume by installing an additional filter can help. A proportional extension of the oil change interval appears possible. An oil additive container that can be used for heavy commercial vehicles would produce a better result and help prevent additional special waste from the required disposal of the bypass filter cartridges that must be regularly changed. The goal of modern flexible maintenance intervals is to cover major differences in the severely varying operating conditions in real life and to ensure maximal oil change intervals according to the actual operating conditions. Using various data, some of them from mixture formation, the system constantly captures the load parameters for the engine oil such as operating temperatures, number of cold starts, used fuel quantity, oil level and temperatures, time and distance values, and more. Using computing models derived from fleet testing, the probable condition of the engine oil is inferred and the maintenance interval is calculated specifically for the load—a so-called indirect quality monitoring of the engine oil in the vehicle. The electronics enables the monitoring of the engine oil state in automobiles with sensors in the oil circuit that also register oil topup. The quality of oil when it is exchanged in a garage can even be electronically specified to yield the fuel consumption for a certain engine oil quality and, hence, determine the length of the change interval.  hemical changes C The alkalinity reserve in fresh oil is defined by the TBN. This is a measure of the ability of the oil to neutralize acidic combustion products to reduce or prevent residue formation, corrosion, and wear. This is contrasted with the total acid number (TAN) that indicates the amount of weak and strong acids in the used oil. Both are used to evaluate used oil. Values beyond pH-9 (highest alloyed diesel engine oils) are termed strong base number (SBN) and those under pH-4 (need oil change) are termed strong acid number (SAN). Figure 22.83 presents the assignment of base numbers and acid numbers. A gradual decrease in the neutralization ability of used oil of up to 50% in comparison to the TBN of fresh oils is generally held to be acceptable. pH value

TBN

SBN

TAN

1–4 4–9 9–11

SAN Decreases

Increases

Decreases Increases

Figure 22.83 TBN and TAN [22-4].

Internal Combustion Engine Handbook | 857

6606_Book.indb 857

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

With the general availability of consumption-optimized SI engines prior to the introduction of engines with regulated catalytic converter running with stoichiometric air–fuel ratio, black sludge formation became a prevalent issue some years ago. The cause was the operation of the engine with a lean air ratio which generated, because of hotter combustion, more nitrogen oxides which entered, with the combustion gases, the crankcase and thus the engine oil via the piston rings. The nitrogen oxides were transformed there into NO2 either in the gas phase or by reacting with oil components. The NO2 then reacts with polar additive components to form organic nitrates and, hence, the problematic black sludge. This process is also termed nitration. The amount of organic nitrates in used oil is an indicator of its continued usefulness. With the development of suitable engine oils and fuel additives, the formation of black sludge could be significantly reduced. The introduction of catalytic converters and the related operation at λ = 1 attenuated the problem. In SI engines with direct injection that operate in the lean range depending on the design, steps must be taken to keep this problem from repeating itself. Finally, we should note that modern passenger car engines generally consume 100 mL of oil every 1000 km; given the large amounts of oil circulating in the engine over conventional oil change intervals of 15,000 km, there is no real need for topping up. The oil may need to be replenished in engines with a long change interval that are increasingly entering the market. This is frequently determined by a sensor in the oil pan and signaled to the driver. In automobiles that are primarily driven long distances, it may be highly advisable to top up the oil given the many cited negative influences on the reserve of agents to prevent the oil volume from dropping too far, and to refresh the reserves of chemically active ingredients. However, we must remember that excessively low oil consumption nearly always indicates harmful oil dilution from fuel. In larger commercial vehicle diesel engines, oil consumption of up to 400 mL/1000 km is common after the breaking-in phase. 22.2.8.10 Racing Engine Oils Oils for engines in competitive vehicles must be optimized for their respective purposes. Let us consider the example of engine oils in modern Formula 1 racing engines in comparison to earlier Grand-Prix formulas in the 1930s. In earlier compressor engines with a very high specific performance (120 kW/L at 7000 rpm) a mixture of castor oil and synthetic esters was used primarily to prevent piston seizing. Castor oil is a vegetable oil from the seeds of the castor oil plant native to Brazil and India. It consists of 80%–85% glycerides of ricinic acid and glycerides of other organic acids. A disadvantage is its insufficient oxidation stability and the formation of resin-like deposits that forced the engine to be disassembled and cleaned almost every time it was used. For today’s 3.0 l naturally aspirated engines that produce more than 300 kW/L at 19,000 rpm, frequently only fully synthetic, highly fluid oils are used that which optimized for lowest-friction resistance with maximum-shear resistance and high-temperature resistance. They must have high oxidation stability, high wear protection, and particularly effective foam suppression because of the

extremely high speeds and oil movement in the dry sump oil tank and engine. Increased dilution of the oil with fuel is a given when there is particularly rich combustion during high performance. High dispersability must be ensured to avoid the ejection of foreign materials and additives. On the other hand, a racing oil of this kind does not have to be tailored for cold starts and must remain intact for only a very short life—just one race or approximately 300 km. Likewise, cost is not a consideration. For long-distance racing such as the 24-h Le Mans, the requirements are naturally more stringent; in addition to increased performance reserves, oil consumption and oil topping up are also factors. 22.2.8.11 Wankel Engine Oils The same engine oils used for reciprocating piston engines are used to lubricate rotary engines. This is understandable for economic reasons because these engines are not very widespread, although the special features of the rotary engine would certainly be better served with tailored engine oils— preferably with low-ash additives. In rotary engines, some of the oil is used to lubricate the apex seal and is continuously combusted in the process. Because of the system-related high oil consumption of approximately 1 1/1000 km and the continuous need to replenish the oil supply, and because of the constructive features of the rotary engine, the otherwise applicable considerations such as low evaporation loss, high oxidation stability, and high wear protection, are not as important. More recent engine designs (Mazda) realize an oil consumption of 0.5–0.6 L/1000 km.

22.2.9 Engine Oils for Two-Stroke Engines

Given their design, two-stroke engines require a different type of lubrication than four-stroke engines because a forced-feed lubrication system cannot be used with crankcase scavenging. A distinction is drawn between conventional mixture lubrication in which a small concentration of special engine oil is premixed with the fuel and increasingly popular fresh oil lubrication that is load and speed-dependent where oil comes from a separate oil tank. Over the course of development of the two-stroke engine and increased environmental awareness, the mixture ratio was reduced from an initial 1:20 to 1:25, 1:50, 1:100, and finally to 1:150 with a concomitantly profound increase in performance. Nevertheless, the oil consumption of the two-stroke engine has always been several times higher than the four-stroke engine. Given the continuous participation of the oil in combustion, the related deposits that arise on the spark plug need to be dealt with in gas exchange openings, and in the exhaust system. Oils for two-stroke engines, therefore, require a clearly different lubricant technology than oils for four-stroke engines. The basic requirements for two-stroke oils are the following: •• high solubility in the fuel •• high corrosion protection because the crankshaft drive and bearing are continually exposed to environmental air •• minimal residue formation during combustion (spark plug/ exhaust outlet)

858 | Internal Combustion Engine Handbook

6606_Book.indb 858

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.2 Lubricants

•• scoring and seizure protection for piston rings, piston skirt and cylinder barrel •• low-smoke and low-noise combustion. The dispersability and the viscosity temperature behavior so vital in four-stroke engine oils are of no importance. Multigrade oils are not used. The required performance is ensured by selecting suitable basic oils and special additives. Primarily SAE-30 basic oils are used. To deal with exhaust smoke, which is particularly critical today, polyisobutylene and synthetic esters have proven to be particularly suitable basic liquids for suppressing smoke in exhaust gas. Detergents, dispersants, and corrosion and rust protection additives are used to attain the previously cited properties. Primarily ash-free substances are used, especially since EP requirements need to be met. They are also advantageous with regard to environmental requirements. 22.2.9.1 Performance Classes The API classes TA to TC were used earlier to evaluate the quality of two-wheel, two-stroke oils; TA was used for mopeds, TB for scooters and motorcycles, and TC for high-performance engines. The required engine test runs are no longer feasible since the stipulated engines are no longer built. However, API TC (CEC TSC-3) still remains in effect. It was replaced by the Japanese Automotive Standard Organization (JASO) and International Standard Organization (ISO) specifications (previously global). Given the dominance of Japanese two-stroke engine manufacturers, JASO specifications are the most referenced. The ISO specifications that are valid worldwide differ only slightly. Figure 22.84 shows the JASO and ISO classes that have been introduced since 1996. They apply to air- and water-cooled two-stroke, two-wheel engines and evaluate the performance of the oils in terms of lubricity, engine cleanness, and freedom of the exhaust system and exhaust smoke. It has become increasingly important to avoid visible and noxious exhaust smoke. Today, practically every powerful two-stroke branded product must fulfill JASO-FC or ISO-L-EGD requirements. The specifications of the latter allow manufacturers to claim superior performance. The classification National Marine Manufacturers Association TC-W3 also covers the biodegradability of two-stroke oils for outboard engines. These oils can also be used in chainsaws. The classification TISI 1040

(Thailand Industrial Standards Institute) is not relevant in Europe; it is applicable only for the Thai market and especially deals with smoke from oil in the exhaust. JASO

ISO

FA

—

Remarks

FB

L-EGB

FC

L-EGC

Low-smoke

—

L-EGD

Low-smoke

Figure 22.84 JASO and ISO classes [22-17].

In the initial developmental stage of two-stroke oil, the oil–fuel mixture had to be made in a “mixer.” “Self-mixing” two-stroke oil with a solubilizer soon became available in small containers and was added to the gasoline in the gas tank. The widespread auto-lube lubrication in modern twowheelers with two-stroke engines makes it unnecessary to premix oil and fuel outside or inside the gas tank. The oil is contained in a separate tank and metered into the flow of the air–fuel mixture depending on the load and speed. Both the life of the oil and environmental requirements can therefore be taken into account by specifically reducing or increasing the amount of oil in the fuel. 22.2.9.2 Test Methods Figure 22.85 provides the physical characteristics for oil specifications for two-wheel, two-stroke engines corresponding to international requirements (ISO) and Japanese requirements (JASO). Test Purpose

Test of

Viscosity at operating temperature

Minimum viscosity at 100°C 6.5 mm2 s–1

Spark plug gap bridging

Limit on sulfate ash content: ISO max. 0.18% (m/m) JASO max. 0.25% (m/m)

Service life of of oxidation catalyst JASO: no phosphorus permitted Safety during storage and transport

Flash point corresponding to national law

Figure 22.85 Characteristics of two-wheel, two-stroke oils [22-15].

Figure 22.86 summarizes the engine tests for two-wheel, two-stroke oils by Japanese manufacturers.

Test Purpose

Engine

Test Conditions

Test Criteria

Safety against piston seizure and scoring

Honda DIO AF 27

Alternating load at 4000 rpm; spark plug seat temperature 160–300°C mixture ratio 50:1

Decrease in torque after cold start and at operating temperature

Engine cleanness at piston rings, piston skirt, combustion residue

Honda DIO AF 27

Full load at 6000 rpm mixture ratio100:1 JASO 1 hour

Evaluation of the engine components after test conclusion

Smoke formation in the exhaust gas

Suzuki SX 800R

Partial load and idling at 3000 rpm mixture ratio 10:1

Evaluation of visible smoke

Cleanness of the exhaust ports

Suzuki SX 800R

Load change at exhaust gas temperatures 330–370°C at 3600 rpm mixture ratio 10:1

Threshold value of the vacuum pressure in the intake system

Figure 22.86 Engine tests for two-wheel, two-stroke oils [22-15].

Internal Combustion Engine Handbook | 859

6606_Book.indb 859

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

22.3 Coolant

Density in kg/m3 at % (V/V) Monoethylene Glycol

The coolant consists of water plus radiator protector. The radiator protector, available as concentrate, provides frost and corrosion protection. The radiator protector and water are usually mixed at a ratio of 1:1, which provides sufficient antifreeze protection in nonarctic areas, and the required corrosion protection. Water alone is insufficient for today’s cooling systems. Water for cooling systems should have the following optimum characteristics:

  50

  40

  30

  20

10°C

1084

1073

1051

1035

30°C

1075

1063

1038

1030

50°C

1064

1049

1031

1022

70°C

1050

1037

1025

1015

90°C

1038

1025

1015

 995

Figure 22.88 Density of the coolant MEG [22-4].

•• Water hardness 5–9 degrees German hardness

Concentration % (V/V)

Boiling Point (°C)

•• pH value at 20°C 7–8

 0

100.0

10

101.5

20

103.0

30

104.5

40

106.5

50

109.0

•• Chlorine ion content

max. 40 mg/L

•• Total > chlorides + sulfates max. 80 mg/L.

22.3.1 Frost Protection

At temperatures below the freezing point, the coolant must be protected against freezing or it will expand and cause unwanted high system pressure and can potentially destroy the cooling system, engine blocks, and cylinder head. Frost protection is achieved by adding glycols—multivalent alcohols—to the coolant water. Figure 22.87 shows the characteristics of the three glycols suitable as radiator frost protectors.

*MEG/water mixtures at normal pressure.

Figure 22.89 Boiling points [22-4].

Monoethylene Glycol % (V/V)

Ice Flaking Points (°C)

Pour Points (°C)

 0

  0

  0

 −2

−2.5

Monoethylene Glycol

Monopropylene Glycol

Diethylene Glycol

 5 10

 −4

 −5

Empirical formula

C2H6O2

C3H8O2

C4H10O3

15

−6.5

−8.5

Density at 20°C (kg/ m3)

1113

1036

1118

20

 9.5

−12

Boiling point (°C)

 198

 189

 245

30

−17

−20.5

Melting point (°C)

 −12

 −60.0

 −11

40

−27

−32

Specific heat at 20°C (kJ/kg K)

2.3

2.5

2.3

50

−37

−47

Figure 22.87 Characteristics of glycols [22-15].

Monoethylene glycol (MEG) is the most used radiator frost protector. Measuring the coolant density provides a way to easily and quickly monitor its concentration. In Figure 22.88, the measured density at the respective measured temperature is portrayed as an indication of the concentration. As expected, the density increases with the concentration and falls as the temperature rises. A coolant concentrate based on MEG has a higher boiling point than water, which promotes engine efficiency. Today, we find coolant temperatures of up to 120°C at 1.4 bar system pressure. The boiling points for the respective MEG concentration are shown in Figure 22.89, whereas Figure 22.90 presents the behavior of water/glycol mixtures under cold conditions. The specific heat of a coolant, that is, its heat-absorbing property or ability to absorb and conduct engine heat, should be very high. It rises with the temperature, but drops with the MEG concentration.

Figure 22.90 Frost protection from monoethylene glycol [22-4].

22.3.2 Corrosion Protection

The coolant concentrate contains carefully harmonized agents (corrosion inhibitors) that prevent corrosion from arising on the different metals that contact the coolant. Figure 22.91 provides information on the arising corrosive substances and the required inhibitors. Individual inhibitors can protect one metal, but may corrode others. The concentration of the individual agents is also a factor. If it is too high, it can be as problematic as if it were too low. The synergies between the individual components also need to be considered. A sufficient reserve alkalinity neutralizes the acidic substances that enter the coolant uncontrolled from the exhaust gas or oxidation products of glycol. The primary corrosion inhibitors are the following: •• benzoate/nitrite •• nitrite amine phosphate-free inhibitors (NAP) •• silicate-free inhibitors (OAT).

860 | Internal Combustion Engine Handbook

6606_Book.indb 860

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

22.3 Coolant

Characteristic

ASTM Test Method

kg/m3

1,110 to 1,145

D 1122

°C

maximal −37

D 1177

Property

Unit

Density at 15.5°C Freezing point 50% (V/V) in distilled water Boiling point (undiluted)

°C

Boiling point 50% (V/V) in °C distilled water

minimum 163

D 1120

minimum 107.8

D 1120

Vehicle paint is attacked

—

No attack

D 1882

Ash content

% (m/m)

maximal 5

D 1119

pH value 50% (V/V) in distilled water

—

7.5 to 11.0

D 1287

Chlorine content

mg/kg

maximal 25

D 3634

Water

% (m/m)

maximal 5

D 1123

Reserve alkalinity

ml

*

D 1121

Gap corrosion In gaps in the cooling system where the coolant does not evenly circulate, corrosive components may increasingly deposit and thus advance corrosion. Cavitation

Fluctuations in the system pressure of the coolant circuit may cause vapor bubbles in the cylinder head and the water pump. These bubbles collapse when the pressure rises again. Because of this pressure surge, material abrasion occurs at the metallic surface which may cause seizing.

Given the anticipated greater use of magnesium as a castalloy component, precise investigations are required to see if the type and composition of presently used radiator protectors can satisfy the new requirements.

Bibliography

*To be agreed between manufacturer and user.

22-1. Reinauer, B. Erdgas im schweren Nutzfahrzeug am Beispiel des ECONIC, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008.

Figure 22.91 ASTM standard D 3306 for coolant based on MEG (physical/ chemical characteristics) [22-15].

22-2. Werner, M. and Wachtmeister, G. Dimethylether—Dieselalternative der Zukunft, MTZ 07.–08.2010, Wiesbaden, Juli–August 2010.

Market-ready coolant concentrates usually contain approximately 93% (V/V) MEG and up to 7% (V/V) corrosion inhibitors. In addition to the corrosion inhibitors, there are small amounts of other additives such as antifoaming agents, sequestering agents for complexing calcium and magnesium ions in hard water, silicate stabilizers, denaturants, and dyes. This yields a complex mixture. To make sure that all requirements are fulfilled, the coolant should not contain less than 40% (V/V) coolant concentrate.

22.3.3 Specifications

Because of the complexity of the coolant concentrate, it must meet the characteristics in the corresponding specifications to be permitted. These describe the quality and performances behavior. Standardized methods are used to determine the described measured values. Figure 22.91 shows the ASTM standard D 3306 for coolant based on MEG and the requirements for coolant performance. There are also many regulations provided by individual automobile manufacturers for radiator protectors. The following additional information regarding coolant requirements should also be noted. Deposits

Deposits in the cooling system must be prevented to ensure heat dissipation. With an excessively high water hardness, calcium and other minerals may precipitate from approximately 60°C which may deposit at critical locations regarding the transfer of heat in particular.

22-3. Esch, T., Funke, H., Roosen, P., and Jarolimek, U. Biogene Automobilkraftstoffe in der allgemeinen Luftfahrt, MTZ 01.2011, Wiesbaden, Januar 2011. 22-4. Worm, J., Szengel, R., and Kirsch, U. TSI und CNG von Volkswagen— eine ideale Kombination, IAV 3. Tagung Gasfahrzeuge, Berlin, 17 Sept. 2008. 22-5. Portmann, D., Keller, K.-H., and Mülbert, K. Die nächste Generation Mercedes Erdgas Sprinter, IAV 3. Tagung Gasfahrzeuge, Berlin, 17 Sept. 2008. 22-6. Thien, U.K.F., Pucher, P., and Weber, G. Analyse eines CNG (Compressed Natural Gas) Fuel System in Real-Life Operation, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008. 22-7. Schüle, H., Treinies, S., Höge, M., and Magori, E. Ein neues Konzept für den zukünftigen Betrieb von DI-Motoren mit Erdgas, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008. 22-8. Berner, H.-J., Bohatsch, S., Ferrari, A., Hoffmann, B., and Bargende, M. Strahlgeführte Erdgas-Direkteinblasung zur Erzielung höchster Prozesswirkungsgrade, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008. 22-9. Thewes, M., et al. Zukünftige Kraftstoffe für moderne DI-Ottomotoren. 19. Aachener Kolloquium, 4–6 Oktober 2010. 22-10. Walther, D., et al. Clean and Protect: Kraftstoffe für heutige und zukünftige Motoren. 6. MTZ-Fachtagung: Der Antrieb von morgen. Hat der Verbrennungsmotor eine Zukunft? 25 und 26 Januar 2011. 22-11. Schult-Bornemann, K.-H. Weltweite Energieprognose bis 2030— Basisdaten von ExxonMobil. 6. MTZ-Fachtagung: Der Antrieb von morgen. Hat der Verbrennungsmotor eine Zukunft? 25 und 26 Januar 2011. 22-12. Hardler J., et al. Der 1.4l 118 kW TSI für E85 Betrieb—Die Erweiterung der verbrauchsgünstigen Ottomotorenlinie von Volkswagen 32. Internationales Wiener Motorensymposium 5 und 6 Mai 2011. 22-13. Aral Forschung Archiv. 22-14. Aral (Hrsg.). Verkehrstaschenbuch 2000/2001. 43. Aufl. Bochum, 2001.

Hot water corrosion

In modern high-performance engines, the temperatures at surfaces in contact with coolant can be extremely high.

Surface corrosion

All metallic surfaces are attached by corroding substances because of their relative roughness.

22-16. Aral (Hrsg.). Fachreihe Forschung und Technik—Dieselkraftstoffe. Bochum: 2001.

Contact corrosion

A number of different metals is present in the cooling system. If, for example, an iron particle deposits at an aluminum surface, a local element forms which may cause holes in its surface.

22-17. Aral (Hrsg.). Fachreihe Forschung und Technik—Ottokraftstoffe. Bochum: 2001.

22-15. Aral (Hrsg.). Fachreihe Forschung und Technik—Kraftstoffe für Straßenfahrzeuge, Grundlagen. Bochum: 1998.

22-18. Aral (Hrsg.). Fachreihe Forschung und Technik—Umweltfreundliche Kraftstoffe. Bochum: 1995.

Internal Combustion Engine Handbook | 861

6606_Book.indb 861

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 22 Operating Fluids

22-19. Aral (Hrsg.). Fachreihe Forschung und Technik—Kraftstoffadditive. Bochum: 1995.

22-31. Walther, D. Entwicklung im Kraftstoffbereich, 5. Emission Control 2010, Dresden, 10 June 2010.

22-20. Aral (Hrsg.). Fachreihe Forschung und Technik—Alternative Kraftstoffe. Bochum: 2001.

22-32. Eichlseder, H., Spuller, C., Heindl, R., Gerbig, F., and Heller, K. Konzept für die Dieselähnliche Wasserstoffverbrennung, MTZ 01.2010, Wiesbaden, Januar 2010.

22-21. Waldmann, H. and Seidel, G.H. Kraft- und Schmierstoffe, Sonderdruck ARAL AG aus Automobiltechnisches Handbuch. 18 Aufl. 1965, Berlin: Walter de Gruyter, Ergänzungsband, 1979.

22-33. N.N. Biokraftstoffe—Die Alternative?—Titelthema, MTZ 12.2010, Wiesbaden, Dezember 2010.

22-22. Aral (Hrsg.). Fachreihe Forschung und Technik—Schmierstoffe Grundlagen/Anwendung. Bochum: 1997/98.

22-34. N.N. Zwischen Acker und Labor—Titelthema, MTZ 12.2010, Wiesbaden, Dezember 2010.

22-23. Aral (Hrsg.). Fachreihe Forschung und Technik—Schmierstoffadditive. Bochum: 1996.

22-35. Schüth F., et al. Zukunft der Energie—Was kommt nach Öl und Gas? 32. Internationales Wiener Motorensymposium 5 und 6 Mai 2011.

22-24. Basshuysen, R.V. and Schäfer, F. (Hrsg.). Lexikon Motorentechnik. Wiesbaden: Vieweg Verlag 2006.

22-36. Stimming U., et al. Wasserstoff—Energieträger der Zukunft? 32. Internationales Wiener Motorensymposium 5 und 6 Mai 2011.

22-25. Menrad, H. (Hrsg.). Alkohol Kraftstoffe. Wien: Springer Verlag, 1982.

22-37. Hardler J. Mobilität im Spannungsfeld globaler Energieketten 32. Internationales Wiener Motorensymposium 5 und 6 Mai 2011.

22-26. DEKRA (Hrsg.). Betriebsstoff-Liste. Stuttgart: MotorPresse-Verlag, 1999. 22-27. Reinauer, B. Erdgas im schweren Nutzfahrzeug am Beispiel des ECONIC, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008. 22-28. Schüle, H., Treinies, S., Höge, M., and Magori, E. Ein neues Konzept für den zukünftigen Betrieb von DI-Motoren mit Erdgas, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008. 22-29. Lenzen, B. and Hohenberg, G. CO2-Potenziale von LPG versus Diesel- und Hybridkonzepten im realen Fahrbetrieb, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008.

22-38. Dinjus, E. and Dahmen, N. Das Bioliq-Verfahren—Konzept, Technologie und Stand der Entwicklung, MTZ 12.2010, Wiesbaden, Dezember 2010. 22-39. Janssen, A., Jakob, M., Müther, M., and Pischinger, S. Maßgeschneiderte Kraftstoffe aus Biomasse—Potenzial Biogener Kraftstoffe zur Emissionsreduktion, MTZ 12.2010, Wiesbaden, Dezember 2010. 22-40. Lumpp, B., et al. Oxymethylenether als Dieselkraftstoffzusätze der Zukunft, MTZ 03/2011, Jahrgang 72.

22-30. Grote, A., Willand, J., Becker, Bernd., and Gerlicher, H. Der neue Wasserstoffmotor von Volkswagen für Flurförderzeuge—aufgeladen, direkteinspritzend, flexible, IAV 3. Tagung Gasfahrzeuge, Berlin, 18 Sept. 2008.

862 | Internal Combustion Engine Handbook

6606_Book.indb 862

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23 Filtration of Operating Fluids 23.1 Air Filter 23.1.1 The Importance of Air Filtration for Internal Combustion Engines

In the early 1920s, automobiles frequently suffered from the very limited durability of the engines used. High levels of dust on the unsurfaced roads of the time caused severe piston-ring wear, leading to declining engine performance and ultimately resulting in the need to replace the old piston rings with new ones to restore the engine’s vitality. The introduction of so-called ‘air-cleaners’ [23-1] made it possible to double repair and servicing intervals to 4000 km. Today too, modern engines also draw in considerable quantities of dust while running, varying depending on the place of operation. Over the course of their service life, modern engines draw significant quantities of dust, depending on the place of operation. These particles ultimately reach the oil circuit, where they are partly responsible for engine wear, and the degree and intensity of component wear (e.g., conrod bearings) depending on the number of particles and their type, geometry, size, and hardness. If these particles are able to enter the bearing clearances, they produce scoring on contact surfaces or generate secondary particles, because they disintegrate between the contact surfaces as a result of the high mechanical loads occurring there. Also relevant in terms of wear are all sizes of particles up to the nominal gap width, because compound friction can occur in the connecting-rod bearings and these particles are able to enter the spaces there. It is the function of filters to remove such particles from the engine’s intake air supply [23-2]. Modern filters achieve particulate-removal efficiencies of up to 99.9%, signifying that, of 1000 particles of the specified diameter, at most one particle is able to pass through the filter element. The nonseparated particles are able to reach

the oil circuit, where they have to be removed by full-flow and bypass oil filters.

23.1.2 Impurities in Engine Intake Air

Atmospheric air contains dust particles, the diameters of which can range from 0.01 to 2000 μ m. Approximately 75% of airborne impurities are of a magnitude of between 0.1 and 100 μ m, this range and its concentration distribution being highly geographically dependent.

23.1.3 Data for Air Filter Medium Assessment

The function of modern air filters is to reduce the pollution in the intake air to acceptable levels under given operating conditions. Specific filtration values are available in various specifications. These data include dust absorption capacity, separation efficiency, surge strength—that is, no passage of dust even if surges occur in the flow of intake air—and stability, signifying retention of the filter element’s pleated structure (see Figure 23.1), even under wet conditions, such as

Figure 23.1 Damaged pleat structure causing reduced filter performance (left) and original condition (right).

Internal Combustion Engine Handbook | 863

6606_Book.indb 863

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

driving in heavy spray and during heavy rain, for example. Adequately stable pleat geometry during operation and, in particular, in the case of wetting is also important for troublefree filter-element function, because maximum dust capacity can be maintained only if the pleated compartments remain uniformly arranged. In addition, the filter medium must also tolerate engine oils, fuels, and crankcase gases, that is, it also needs to have chemical resistance. The materials must remain thermally stable up to engine compartment temperatures of approximately 85 to 90°C. Modern filter fabrics have a significantly lower pressure loss at the filter element despite identical dimensions. The filter media suited for these tasks are manufactured from cellulose or synthetic fibers or mixtures of both. The most recent developments in the field of filter media are orientated around the longer servicing intervals of up to 120,000 km for automobiles. Media with a pronounced ‘gradient structure’ (as far as pore distribution is concerned) are currently used. As seen in Figure 23.2, multilayer media

consist of filter papers in combination with melt-blown layers (layers of synthetic fibers). This increases dust storage capacity and thus service life by up to 30%. Media consisting of pure plastic fibers (synthetic fabrics) (see Figure 23-2) permit the generation of greater gradient structures, and thus a dust capacity is enhanced by as much as 50% [23-3].

23.1.4 Measuring Methods and Analysis

Filter media are tested under standardized conditions defined in DIN ISO 5011 [23-4]. The removal efficiency is determined using one of two methods. According to the standard, there is the gravimetric removal efficiency calculated from the ratio of the increases in mass of the filter medium to be tested to total dust mass introduced during the test (Figure 23.3). To obtain detailed information on the performance range of filter media for the wear-relevant particles, it is necessary to measure removal efficiency as a function of particle size. This is shown as a function of particle diameters between 0.1

Air filter media

Composites (paper + meltblown)

Paper

Fabric

Figure 23.2 Filter media for use with internal combustion engines.

30

Flow resistance rise [mbar]

25 20 15

Paper filter element

10 5 0

0

0,2

0,4

0,6

0,8

1

Addition of water [I]

1,2

1,4

1,6 Fabric filter element

Figure 23.3 Flow resistance when wet.

864 | Internal Combustion Engine Handbook

6606_Book.indb 864

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23.1 Air Filter

and 10 μ m in Figure 23-4. Removal efficiency increases steeply after only a slight pressure increase has occurred, and nearly 100% of the particles are captured. It is necessary to take both removal efficiency and dust absorption capacity into account to determine the efficiency of a filter medium. The latter can be determined by feeding dust onto the filter element under load at a constant volumetric air flow until a pressure drop of 20 mbar occurs.

23.1.5 Demands on Modern Air Filter Systems

Modern air filter systems, that is, a system consisting of a “raw air” pipe with an air inlet, a filter housing, which dampers to reduce intake noise, a filter element, a clean air line with a hot-film air mass sensor, and the engine intake manifold, must be geometrically adapted to the restricted space available under the hood. This results in the development of new filter media that require significantly less space without sacrificing filter performance. The filtration figures required for automobiles (diesel) and trucks are shown in Figure 23.5. The efficiency of the entire air-intake train depends on the performance of the individual components such as the location of the air-intake point on the dust-intake and water-intake minimization criteria. The intake area should be located in sheltered, low-turbulence zones on the vehicle, for instance, under the covering in the wheel arch or in areas with no through flow in the engine compartment. Ideal placement in trucks is above the cab roof (overhead intake) or to the side of the cab, to achieve longer servicing intervals. A flow-optimized filter housing assures complete exploitation of the medium’s potential for dust absorption and removal efficiency, because the medium is not subjected at any point

to excessive approach flow velocities, and thus, high removal efficiency can always be guaranteed. In addition to filtration of the combustion air, a further important function of the air filter element is that of keeping dirt away from the hot-film air mass sensor. Fouling this with particles affects the engine control system, clearly signifying that filtration does have a direct influence on driving comfort and convenience.

23.1.6 Design Criteria for Engine Air Filter Elements

It is necessary for the design of an engine air filter element to differentiate between the filter fineness required for Figure 23.5. The filter medium surface area is calculated on the basis of the engine’s air volume requirement per unit of time (V (m3/ min)) [23-5]. Sufficient surface area is provided for cleaning this volumetric flow of air so that the air’s flow velocity never exceeds a certain critical velocity range of v < vcrit. At excessively high filtration velocities, this results in the particles not being captured on the fibers during passage through the filter. The excessively high pulse results in rebounding of the fibers, and the particles are able to pass through the filter medium. Permissible flow velocities vcrit may differ greatly, depending on the filter materials used, such as filter paper, multilayer media (composites), or synthetic fabrics. The following critical velocities vcrit and dust absorption capacities result for typical filter media when the removal efficiency demanded for gasoline and diesel engines is taken into account (see Figure 23.5). The necessary filter surface area is calculated in the following example.

100

Particle size [µm]

99 98 97

Degree of fraction separation after 1 mbar rise in pressure

96 95

Degree of fraction separation after 20 mbar rise in pressure

94 93 0,1

1

Degree of separation [%]

Figure 23.4 Dust fraction removal efficiencies for a standard filter medium.

10

Filter medium

vcrit (cm/s)

Gravimetric removal efficiency % Passenger car(gasoline) Passenger car (diesel) Truck

Paper

10

>99.5

>99.8

>99.9

190–220

100–120

Multilayer medium

17

>99.5

>99.8

>99.9

230–250

100–120

Fabric

33

>99.8

>99.8

>99.9

900–1100

230–250

Spec. dust capacity g/m2

Weight per area unit g/m2

Figure 23.5 Critical filtration velocity, specific laboratory dust capacities, removal efficiencies of filter media, and weights of filter media per unit of area.

Internal Combustion Engine Handbook | 865

6606_Book.indb 865

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

Surface area is calculated to achieve the required removal efficiency on the basis of volumetric air flow (e.g., 5 m3/min) and vehicle type (automobile/ diesel), taking into account flow velocity. The specific dust capacities of the filter medium are exploited to achieve service intervals of, for example, 60,000 km (= 200 g of laboratory dust). A square meter of standard filter paper is needed to achieve the required dust capacity of 200 g, critical velocity then being exceeded, however, with the result that the higher figure of 1.25 m2 is selected for the ultimate design.

23.1.7 Filter Housings 23.1.7.1 Design of Filter Housings In addition to the ‘available space’ problem, ease of servicing, that is, trouble-free filter-element changing, is of great importance in filter housing design. This functionality is linked directly to the need for housing tightness. Practically, all air filter elements feature a polyurethane (PUR) seal. Filter housings can be classified into flat and round filter element types (see Figure 23.6). In the case of rectangular filters, the PUR section seal is located in the housing groove and clamped axially by the filter housing cover. Round filter elements predominate in the goods and utility vehicle sector, because of their greater simplicity of sealing and higher stability. The design of filter housings is essentially determined by the requirements for homogeneous flow in the housing. This minimizes pressure drop and achieves uniform flow onto the surface of the filter element.

23.2 Fuel Filters Fuel-injection systems on modern gasoline and diesel engines react extremely sensitively to even the very smallest contaminant in the fuel. Damage can be caused, in particular, by particulate erosion and, in diesel engines, also by corrosion resulting from the portion of water in the fuel. Such contamination is composed both of extremely hard mineral particles and of

organic particles such as soot and tar. Diesel fuels sold under DIN EN 590 must achieve particulate contents of lower than 24 mg/kg. International automotive industry associations recommend values of below 24 mg/kg. Diesel fuels sold in Germany generally have particle contents of less than 10 mg/kg, although the above-mentioned limit for particle concentration is, in some cases, greatly exceeded in the fuels available around the world. The fuel filter must be able to reliably capture all larger particles (>15 μ m). Studies into wear on newer injection systems also document the wear relevance of the finer fraction of approximately 5 μ m. For this reason, removal efficiency for the 3- to 5-μ m particle-size fraction has become established as a characterizing dimension for filter fineness in recent years. As a result of their much greater injection pressures, diesel injection systems require greater protection against wear than gasoline injection systems, and thus finer filters have been developed. In addition, the fuel filter must also possess adequate capacity for storage of particles.

23.2.1 Gasoline Filters

Modern gasoline engines are equipped with electromagnetically operated injection valves. It is the fuel filter’s task to prevent erosive wear on the electrical injection valve and the ingress of wear-causing particles into the engine’s combustion chambers. Engines featuring gasoline direct injection require significantly finer fuel filters for protection against wear than systems for injection into the intake manifold. The reasons are that, on the one hand, pressures higher by a factor of 30 occur at the injection valve and, on the other hand, other components, such as the pressure accumulator and the pressure control valve, must also protect the injection system against the entry of particles. 23.2.1.1 Required Filter Finenesses A specified filter fineness (initial removal efficiency in accordance with ISO/TR 13353, Part 1, 1994) is the result of dynamometer and field tests performed by engine and injection system manufacturers jointly with the filter producers. Figure 23.7 shows recommendations for minimum filter fineness for

Figure 23.6 Air filter element types.

866 | Internal Combustion Engine Handbook

6606_Book.indb 866

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23.2 Fuel Filters

100 95

h3–5 µm [%]

85

Initial removal efficiency in interval 3-5 µm ISO/TR 13353: 1994

67

More difficult conditions

50

Minimum requirements

25 Otto engine

Indirect injection up to 4 bar

Direct injection up to 120 bar

gasoline engine injection systems. Even finer fuel filters are necessary to achieve efficient protection against wear if the engine is to be operated with high particulate concentrations in the fuel or the outside air (more difficult conditions). 23.2.1.2 Designs of Gasoline Fuel Filters The main form of gasoline fuel filter is the inline filter. In some types, the pressure control valve is also integrated into the filter head. The filter housings are made of plastic, aluminum, or steel, depending on their position in the engine compartment (crash safety) and the vehicle manufacturer’s specification. Maintenance-free ‘vehicle lifetime’ fuel filters are increasingly being demanded for gasoline engine vehicles. A further trend arises as a result of the lowering of hydrocarbon emissions limits [23-6]. This makes it necessary to integrate all external components of the low-pressure circuit, such as the injection pump, fine filter, and pressure control valve into a single in-tank unit. Further elements such as the tank fillinglevel transducer, the swirl pot, which can be actively filled via an ejector, and the optional prestrainer, which serves to protect the pump, can also be integrated into the in-tank unit. Figure 23.8 shows a modern lifetime filter element. Such filter elements are also currently made in complex, noncylindrical geometries, to achieve maximum exploitation of available space in the in-tank unit.

Figure 23.8 Star-pleated filter element for lifetime filtration in an in-tank unit.

Figure 23.7 Recommended minimum filter fineness for gasoline engine fuel filters.

23.2.1.3 Structure of Filter Element and Filter Medium The requirements currently imposed on filter fineness and dirt storage capacity necessitate innovative filter concepts. With the same filter fineness, a simple strainer (surface-action filter) achieves only approximately a tenth of the filter service life (particle storage capability) of a modern deep-bed filter medium. The use of deep-bed filter media that, in addition, are also installed with extremely high packing density is therefore necessary to meet the performance data specified for gasoline fuel filtration. This is achieved, in particular, through the star pleating (see Figure 23.8). The pleated construction, which possesses a large filter surface area, is supported on a pressure-proof central tube. The fuel flows radially from the outside to the inside of the filter. So-called spiral V filters, which consist of concentric filter–paper compartments, are also in use as an alternative. Deep-bed filters currently consist mainly of ultrafine cellulose fibers. More recently, spiral V elements have been increasingly incorporating multilayer filter media with ultrafine synthetic fibers. The cellulose fibers of the filter medium are sheathed with special fuel-resistant impregnating agents.

23.2.2 Diesel Fuel Filters

The rapid development of diesel engine technology permits extremely high fuel efficiency in both automobile and commercial vehicle engines. In all modern diesel engines, the fuel is injected directly into the cylinder. The function of the fuel filter is to protect all the components of the high-pressure injection system. The filter can, for this purpose, be located in the low-pressure circuit, either on the pressure side in the feed line to the high-pressure pump or on the suction side in the inlet line to the fuel presupply pump. If, on the pressureside, a differential pressure of up to 6 bar, depending on the design of the system, is available for fuel filtration, significantly more than with the suction-side arrangement. Installation on the pressure side, therefore, predominates on commercial vehicle engines and also increasingly on automobile engines. 23.2.2.1 Required Filter Finenesses The introduction of modern solenoid-valve-controlled diesel injection systems made it necessary to significantly increase

Internal Combustion Engine Handbook | 867

6606_Book.indb 867

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

filter fineness [23-6], [23-7], [23-8]. It quickly became apparent, particularly in the case of commercial vehicle pump–nozzle systems, that the use of the up-to-then finest filters (initial removal efficiency η 3–5 μm = 45%, until 1997 the finest filter grade used in Europe for distributor injection pumps) resulted in wear of the solenoid valve’s seats. This damage became apparent in practice in the form of declining engine output, rough running (caused by differing degrees of wear on the individual injectors), and greater generation of soot. The necessary filter fineness is determined primarily by the assessment of wear following field testing of the vehicle. This time-consuming procedure is increasingly being superseded, or at least augmented, by special wear studies performed by the manufacturers of diesel engines and injection systems. A fuel specially blended with ultrafine particles (with ISO 12103 M1 or A2 test particulates, for example) is used on the injection pump test bench. The injection-flow loss that occurs on the equipage with diverse filters is used as an indicator of wear and is then correlated with the filter’s initial removal efficiency. Both in field tests and in the context of test bench trials, indicators relevant for practice correlate extremely well with the initial removal efficiency determined in accordance with ISO/TR 13353 (1994) in the 3 to 5 μ m particle-size range. Tests performed on pump line nozzle engines (commercial vehicle, United States) also confirm the relevance of the ultrafine fraction of approximately 5 μ m particle size for wear [23-9]. Figure 23.9 shows recommendations for minimum filter fineness for gasoline engine injection systems. As in Figure 23.7, differentiation is made between normal and more difficult conditions.

media. Filter media built up of fibers with a hydrophobic surface are particularly suitable for these applications. The water removed collects in a reservoir at the bottom of the filter and is discharged via manual or automatic systems when this water collecting space has filled. Testing for removal of water is performed by adding a 2% emulsion of water in diesel oil and is described in ISO 4020. 23.2.2.3 Designs of Diesel Fuel Filters It is necessary to differentiate between diesel fuel filters that can be opened and filters that are replaced completely with the filter housing during servicing. This category of filters includes sheet steel, aluminum, and plastic inline filters. Greater crash safety requirements have resulted in a significant revival of the sheet steel type. Additional elements such as a water outlet, a water sensor (in the form of a conductivity sensor), a thermostatic valve (for the return of hot fuel), and heating systems can also be integrated into inline filters. Another very widely used type of filter, one that cannot be opened, is the easy-change disposable fuel filter. The disposable filter is screwed onto a filter head featuring an external thread, and is sealed by an external elastomer element. Particularly in the case of commercial vehicles, the demand for maximum filter fineness and good water removal performance combined with longer servicing intervals can be met only by multistage filtration concepts. Here, a prestrainer for removal of water and prefiltration of particles is first installed on the pressure or suction side. The fuel then passes through a pressure-side fine filter in which the ultrafine particles are removed. Figure 23.10 shows an easy-change disposable fuel prestrainer for commercial vehicles into which many extra functions (e.g., a thermostatically controlled fuel return system, an electric heating system, a filling pump, and sensors for water-level and differential pressure) have been integrated. Housing-type fuel filters are filters that can be opened. The housing cover is unscrewed for filter changing, and only the filter element is replaced. For easier servicing, the housing cover is usually located at the top. Housing-type fuel filters may also contain additional elements such as a pressurecontrol valve, a thermostatic valve, an electric heating system, and water-level and differential pressure sensors. In modern types, the filter element itself consists entirely of nonmetal and, therefore, easily thermally recyclable materials (Figure 23.11).

23.2.2.2 Water Removal Damage caused by local impairment of lubrication and, in particular, by corrosion can occur if water reaches the highpressure side of a diesel injection system [23-10]. In many cases, the fuel filter, therefore, also performs the additional task of eliminating this free or emulsified water. Measures to prevent the ingress of impermissibly high quantities of water are generally required for distributor injection pump (volumetric efficiency—VE) and common rail (CR) systems. Because of their short contact times, unit injector systems are relatively insensitive, but also require a system for removal of water if extremely high water levels are anticipated. Water is currently mainly eliminated by coalescence on special filter

More difficult conditions

Common-rail

Pump nozzle system

25

Distributor pump

50

Time-controlled distributor pump

67

Inline pump

h3–5 µm [%]

100 95 85

Minimum requirements

Figure 23.9 Recommended minimum filter fineness for diesel fuel filters.

868 | Internal Combustion Engine Handbook

6606_Book.indb 868

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23.2 Fuel Filters

• Optimal functionality by heating in the inlet system

Filter head with 3-bore flange

• Compact design with high integration of functions • Doubled service life time of the MANN+HUMMEL full-flow filter

Ball valve (option)

• High dynamic rigidity due to three-bore flange • Specific solutions attained by modular concept • Bio-diesel (RME/PME) resistant

Various element sizes Water collecting space Drain plug

Heating (option) Hand pump

Figure 23.10 Diesel fuel prestrainer for commercial vehicle, an easy-change disposable type.

tools have recently been rationally augmented via the use of flow simulation (CFD) for the calculation of particulate removal efficiency in the fibrous deep-bed filter media [23-11]. Discoveries based on the optimization of the fiber structure have been integrated into the modern composite filter media. Figure 23.12 shows the schematic structure of the multigrade filter media. Contaminated fuel

Figure 23.11 Metal-free filter element for a commercial vehicle housingtype fuel filter.

23.2.2.4 Structure of the Filter Element and Filter Medium Filter elements for diesel fuels are usually of the spiral V pleated type. The demand for high filter fineness combined with long servicing intervals necessitates the use of innovative filter media. Multilayer composite filter media (Mann + Hummel multigrade) are capable of achieving ultrahigh filter finenesses with simultaneously high particulate storage capacities. Increased performance can be achieved only through systematic optimization of the composition and structures of the fibrous filter media. Ultramodern test methods based on the measurements of removal efficiency using automatic particle counters are used for this purpose. Empirical development

Clean fuel

Meltblown prestrainer layer with ultra-high contaminant capacity

Very dense cellulose-based ultra-fine filter layer

Figure 23.12 Schematic structure of the multigrade filter media.

Performance data for both particle storage capacity (service life) and filter fineness increases significantly compared with conventional cellulose-based mixed fiber media. Thanks to the hydrophobic properties of the base materials and the ultrafine fiber diameter, the layer of synthetic fibers on the approach side features extremely good water coalescence. Removal of water is, therefore, accomplished on the approach side. Performance data superior to single-layer filter papers can also be achieved using so-called hybrid fiberglass papers. These filter media contain 5% to 20% microglass fibers with fiber diameters of approximately 1 μ m. These filter media,

Internal Combustion Engine Handbook | 869

6606_Book.indb 869

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

which are used outside Europe, are not entirely undisputed, because the ultrafine, brittle glass fibers may detach.

23.3.1 Wear and Filtration

23.2.3 Performance Data of Fuel Filters

Multigrade filter media, developed to achieve maximum particulate storage capacity, are fuel filters suitable for use in in-tank units (lifetime use) for gasoline engines. The use of filter media tailored to the achievement of maximum filter fineness is necessary in the case of high-performance dieselfuel filters. The multigrade concept achieves this without sacrificing service life. Figure 23.13 shows the service life and removal efficiency of a filter element containing MULTIGRADE F_GAS for lifetime gasoline filters and the performance data for diesel filters incorporating the MULTIGRADE F_HC and the MULTIGRADE F_HE ultrafine filter media compared with those of an early standard (cellulose/polyester hybrid fiber paper, until 1997 the finest diesel filtration medium used in Europe).

Degree of separation h [%]

100 90 80 70

Degree of separation EN ISO 4020

60 50 40 30

Cellulose/PES mixed-fiber medium MULTIGRADE F_HC MULTIGRADE F_HE MULTIGRADE F_GAS

20 10 0

0

25

50

75 100 125 Test time [min]

150

175

23.3 Engine Oil Filtration

200

Figure 23.13 Performance data for gasoline and diesel fuel filters incorporating conventional and modern filter media.

23.2.3.1 Outlook In the future, continuing advances in diesel and gasoline injection will also be accompanied by a continuous increase in the performance data of filter media and complete fuel filters. It will be possible to improve on the already extremely high standard only via the use of innovative and ever finer fibers as the filter medium material and via optimization of the filter medium’s microstructure. The focus of future refinements in filter media for diesel fuel will be on the increasing filter finenesses combined simultaneously with longer service intervals and greater compactness. Future diesel fuel filtration systems will include not only the integrated additional functions already used today, but also solutions for maintenance-free disposal of the water removed from the fuel. In the case of gasoline engine fuel filters, the paramount technological challenge in the future will be the achievement in compact lifetime filter elements for in-tank units of the higher filter fineness required for direct injection (DI) engines.

The oil in the engine has primarily the task of lubricating moving components but also to cool hot engine components (such as piston head cooling) and to seal by applying thin oil films in the cylinder running surfaces and the power transmission system. Particulate contamination in wear-relevant sizes cause strongly increased abrasion in the lubricating gaps between moving engine components and damage surfaces, causing further wear. This results in increased oil and fuel consumption, reduced engine performance, and significantly increased environmental impact due to poor exhaust values [23-12]. All the dirt that enters the engine from all the diverse possible sources collects in the engine oil. Dust particles from the atmosphere, allowed through even by good air filters, thus get into the oil. The remaining dirt from production and assembly of the engine enters the oil, as does metallic abrasion (wear particles) and soot from incompletely combusted fuel. Water (condensate) from combustion and uncombusted fuel (oil dilution), which form a complex mixture with the reaction products of the oil, such as oil oxidation products and additive reaction products, for example, must also be added. Because the engine oil is continuously conveyed in the circuit, the contaminants accumulate in the engine oil and, to protect the engine, must be removed as much as possible from the oil or, at least, maintained below a critical threshold value. Soot accumulation in the engine oil is also critical because higher concentrations will cause a higher engine oil viscosity and reduced lubrication capacity. A delayed oil change can cause severe engine damage [23-13]. The wear relevance of particles depends also on the engine itself, its tolerances, and the dimensions of the bearing clearances. In this respect, engine design has advanced largely in the last years. The requirements due to increased combustion pressure (engine downsizing) and higher energy efficiency has resulted in a dramatic reduction in the permissible manufacturing tolerances. Particles of a size of approximately 1 μ m are, therefore, now critical—especially at high concentrations. Particularly dangerous are particle sizes ranging from approximately 8 μ m up to approximately 60 μ m, as is demonstrated by wear measurements. Figure 23-14 shows a typical wear diagram. Tracer-marked metals were used to quantitatively capture metallic abrasion of the engine and to show the different wear relevance of different-size particles having entered the engine. Particles of a similar size as the lubricating gaps are particularly critical. These statements can also be applied to other engines, that is, the smaller the gap widths, ever smaller particles will become wear relevant. Another issue to be noted is that abrasion increases at high concentrations of very small particles in the oil. Coarser particles are also highly dangerous, however, because they become comminuted (i.e., they disintegrate), and then become part of precisely the most dangerous particle-size fractions.

870 | Internal Combustion Engine Handbook

6606_Book.indb 870

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23.3 Engine Oil Filtration

Abrasion [mg/mg]

0,4

0,3

Oil cooler

Bypass valve

0,2

SI engine/ diesel engine

Oil pump

0,1

0

10

20

30

40

Particle diameter [µm]

50

Pressure regulating valve

60

Figure 23.14 Wear diagram illustrating the wear effects of various-sized particles in a combustion engine.

The oil filters (Figure 23-15) act as a sink for the particulate contaminations. These are deep-bed filters, which enable the oil to perform its intended functions during the period stated by the manufacturer. The oil filter element can neither extend oil-changing intervals nor retard degradation processes.

Figure 23.15 A selection of different oil filters [disposable filter (with metal screw-on housing), metal-free elements for installation in oil filter modules, in each case with paper, composite, and all-synthetic filter media]

Oil pan

Fig. 23.16 Oil circuit of a combustion engine with a full-flow oil filter.

filter. A pressure-control valve feeds the surplus mass flow of oil back into the sump. The oil flowing to the engine must completely pass through the full-flow oil filter, generally resulting in a compromise between filter fineness (which generally signifies pressure losses in otherwise comparable filter media) and the size of the filter. The full-flow oil filter also features a bypass valve, which opens as a function of the differential pressure across the filter, and thus assures oil supply to the engine, which has even greater priority than filtration of the oil. This may be necessary, for example, in the case of extremely low temperatures and highly viscous oil. It is, therefore, extremely important that the filter element is not expired, i.e., clogged. Otherwise, this bypass valve remains at least partially open during normal operation and allows continuous passage of an unfiltered side stream of oil, with all the consequences of elevated wear. The maximum differential pressure recommended by automobile manufacturers is between 1.0 and 2.5 bar. Filter media manage to achieve these requirements by using appropriate pleating and embossing methods. Figure 23.17 shows a section through a typical disposable filter. The spiral V filter element and the overflow valve (at bottom) can all be seen, as can the one-piece check valve, which prevents the filter element from running empty when the engine is stopped, and thus assures that the oil’s full lubricating action is immediately available after starting.

Single-part return lock on silicone basis

Long-life elastomer seal

23.3.2 Full-Flow Oil Filters

Fig. 23.16 shows a schematic view of a full-flow circuit and a bypass circuit. This is acceptable provided high ultrafine particle levels are not anticipated between service intervals. In the full-flow oil circuit, used in both gasoline and diesel engines in automobiles, the oil pump draws the oil in from the oil sump (non-pressurized); if necessary, the oil is cooled in the oil cooler and then passes through the full-flow oil

Teflon-coated bypass valve

Long-life filter medium

Fig. 23.17 Sectional view of a screw-in disposable filter.

Internal Combustion Engine Handbook | 871

6606_Book.indb 871

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

Such multifunctional oil modules are increasingly superseding conventional easy-change disposable filters, via the integration of even more functions, such as a heating system, an extra separator for crankcase venting, and mounting functions for other engine components.

A more complex solution is illustrated in Figure 23-18, which shows a complete oil filter module. The oil cooler is also directly integrated here, in addition to the metal-free filter element, permitting environmentally friendly disposal. Also included are sensors for the measurement of oil pressure, temperature, and (in the future) oil quality, permitting precise determination of the current condition of the oil.

23.3.3 Removal Efficiency and Filter Fineness

Unlike the situation with filter finenesses for engine intake air and fuel filtration, there is no specified minimum removal efficiency in the case of engine oil filtration. Various manufacturers define the filter fineness for oil extremely diversely in their specifications; the separation curves specified are correspondingly diverse. Figure 23.19 shows a range of different separation curves as a function of particle size. Filter fineness is defined via the diameter value of 50% of the particles removed from the oil. Full-flow oil filters that remove particles with diameters of 9 and 12 μ m serve as the standard. However, more recently developed filter media that filter engine oils extremely finely, with a filter fineness of 4 μ m, are now available. It must be noted, however, that with increasing filter fineness, the mass of contaminants retained by the filter increases. For this reason, these filters require a larger contaminant storage capacity at the same change intervals. The downside, however, is that contamination capacity generally decreases with increasing filter fineness within a specific medium type. Filters with increased filter fineness must, therefore, have larger dimensions or require higher developed media. So-called lifetime oil filters that assure the reliable capture of relatively large particles but omit the filtration of small particles have not succeeded. These filters must have extremely good chemical resistance and more than 200% more dirt storage capacity than standard oil filter elements of the same size and, therefore, permit longer change intervals. Due to the higher wear within the engine over time, the use of such filters is not recommended. The filter media used in oil filtration are deep-bed media. Cellulose, cellulose and synthetic fiber mixed media, composites

Figure 23-18 Oil filter module for a four-cylinder diesel engine with a metal-free filter element and various integrated additional functions (oil cooler, pressure and temperature sensors, alternator mounting, check and overflow valves, and oil and water flow control).

100 % 90 %

Degree of separation [%]

80 % 70 % 60 %

Synthetic fibers Multigrade OHC Multigrade OHE Long-life Standard cellulose Life-time

50 % 40 % 30 % 20 %

according to ISO 4548-12 Load: 1.2 l/cm 2 h, final delta p 2 bar

10 % 0%

0

5

10

15

20 25 30 Particle diameter [µm]

35

40

45

50

Figure 23.19 Fraction removal efficiency curves of six different oil filter media.

872 | Internal Combustion Engine Handbook

6606_Book.indb 872

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

23.3 Engine Oil Filtration

from cellulose and synthetic fibers, glass fiber and cellulose mixed media, glass fiber media, and fully synthetic media are used. Due to their lower chemical resistance, pure cellulose media are not suitable for use with fully or partially synthetic fuel economy oils or in long-life filters [23-14]. Their resistance can be significantly increased by synthetic fibers (as a mixed medium, in a composite or as a full mat). Media containing glass fibers exhibit excellent filtration properties and attain a very high chemical resistance. However, they are commonly rejected by many automotive manufacturers due to the risk of fiber migration.

Full flow filter

Oil cooler

Bypass valve Throttle

Bypass filter

Oil pump

Diesel engine

Pressure regulating valve Oil pan

23.3.4 Bypass Oil Filtration

The structure of the lubricating system according to Section 23.3.2 is sufficient when a high portion of very small particles is not anticipated between service intervals. If this is the case, as it is, for example, in diesel engines with long change intervals (commercial vehicles) or a high soot accumulation in the engine oil, a bypass oil filter is used in addition to the full-flow filter. A high concentration of ultra-fine particles can have a wear relevance similar to that of larger particles. Also much feared is so-called ‘bore polishing’, an effect in which areas on the cylinder sliding surfaces become so extremely finely polished that the surface quality of the metal itself prevents the adhesion of oil films, with the result that the lubricating film breaks down. Figure 23.20 shows the basic structure of an oil circuit with a full-flow and a bypass filter. Via a throttle valve, a side stream of approximately 5 to 10% of the total volumetric flow of oil is diverted, cleaned in a bypass filter and returned to the oil pan. This bypass filter usually contains significantly finer filter media capable of removing ultrafine particles such as soot from the engine oil. To achieve the corresponding filter effect for soot particles <1 μ m, the filtration velocity must be lower than in the full-flow filter and the filter medium must be correspondingly finer. The flow rate decreases as the deep-bed filter becomes increasingly clogged, but filter fineness increases [23-15], [23-16],[23-17],[23-18],[23-19],[23-20].

Figure 23-20 Oil circuit of a combustion engine with a full-flow and a bypass oil filter.

Another extremely effective method of removing ultrafine soot particles via a bypass flow is achieved via the use of centrifugal filters. A centrifugal filter consisting of metal or plastic (lower weight and more environmentally safe disposal) is installed in place of the larger bypass filter element. Only the oil pressure, which accelerates this centrifuge up to as much as 10,000 revolutions per minute, is needed to drive it. The oil returns, without application of pressure, from the centrifuge housing into the oil pan. Figure 23.21 shows the plastic rotor of a centrifugal filter. The driving nozzles at the lower end of the rotor are easily visible. This figure also shows a top view of the sectioned rotor, in empty condition first, and then after 500 hours of use in a vehicle. The high centrifugal forces achieve not only good removal efficiency for the fine particles that are relevant, but also the filter cake is also extremely compact. These centrifugal filters are a genuine alternative to bypass filter elements. The rotor, filled with compact, ultrafine particles, is simply taken out and replaced with a new one at every service, which can be selected in accordance with the oil-changing and full-flow oil filter-changing intervals. Such compact centrifugal filters may also be of interest for diesel engine automobile applications. The even more stringent

New rotor

after 500 h

cut-away

500 g mass of dirt

Centrifugal filter

Ancillary conditions: Old oil from a commercial vehicle diesel engine; filled at an oil test stand; already filtered through a full-flow filter

Figure 23-21 Structure of a plastic centrifugal filter for bypass oil circuit removal of ultrafine particles. The sectional view of the rotor, new and used.

Internal Combustion Engine Handbook | 873

6606_Book.indb 873

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 23 Filtration of Operating Fluids

legislation with regard to exhaust gas standards (EURO 5 and EURO 6) are in a conflict of aims between reduction of oxide of nitrogen and reduction of soot particle concentrations. In the course of optimizing the combustion process toward lower NOx concentration and avoiding the, otherwise required, selective catalytic reduction technology, soot generation is inevitably increased. This results in an increased engine oil contamination with soot. For service intervals of the same or extended duration, centrifugal filters represent an alternative to oil bypass filter elements with larger dimensions.

23-10. Projahn, U. and Krieger, K. “Diesel-Kraftstoffqualität—Erkenntnisse aus Sicht des Einspritzlieferanten,” Proceedings 9. Aachener Kolloquium Fahrzeug- und Motorentechnik, Pischinger, S. (Hrsg.), Aachen, pp. 929–944, 2000

Bibliography

23-14. Dahm, W. and Daniel, K. 1996. Entwicklung der Ölwechselintervalle und deren Beeinflußbarkeit durch Nebenstromfeinstölfilterung, MTZ Motortechnische Zeitschrift: Wiesbaden. 57 (6).

23-1. Katz, H. 1927. Die Luft-, Brennstoff- und Ölreiniger im Kraftwagen. Autotechnische Bibliothek, Bd. 80, Berlin W62. Richard Carl Schmidt & Co: Berlin. 23-2. Purchase, D.B. 1997. Handbook of Filter Media. Elsevier Science Ltd: Makati.

23-11. Klein, G.-M., Banzhaf, H., and Durst, M. “Fuel Filter Solutions for Future Diesel Injection Systems,” Proceedings World Filtration Congress 8, Brighton, UK, pp. 887–890, 2000 23-12. Mach, W. and Trabandt, T. 1998. Auswirkungen fester Fremdstoffe in Gebrauchtölen auf das Verschleißverhalten von Dieselmotoren. Mineralöltechnik: Hamburg. 10. 23-13. Spanke, J. and Müller, P. 1997. Neue Ölwechselkriterien durch Weiterentwicklung von Motoren und Motorenölen, MTZ Motortechnische Zeitschrift: Wiesbaden. 58 (10).

23-15. Banzhaf, H., Yates, B., Trautmann, P., and Durst, M. “Small elements, big performance, best price—challenges for oil filter elements,” Proceedings International Filtration Conference, San Antonio 2008.

23-4. Lechner F., Dissertation, University of Heidelberg, 2000.

23-16. Banzhaf, H., Yates, B., Trautmann, P., and Durst, M. “Small Elements, Big Performance, Best Price—Challenges for Oil Filter Elements,” Proceedings of the 9th International Filtration Conference: San Antonio, 2008.

23-5. Affenzeller, J. and Gläser, H. 1996. Lagerung und Schmierung von Verbrennungsmotoren. Springer Wien: New York.

23-17. Durst, M. and Klein, G.-M. (Hrsg.). 2006. Filtration in Fahrzeugen. Expert-Verlag: Renningen

23-6. Worldwide Fuel Charter. 2000. Broschüre der Internat. Verbände der Automobilindustrie. ACEA, Alliance, EMA, and JAMA., Berlin.

23-18. Durst, M., Klein, G.-N., Moser, N., and Trautmann, P. “Filtration und Separation in der Automobiltechnik,” CIT 79 (11), pp. 1845–1859, 2007.

23-7. Klein, G.-M. “Kraftstofffilter. Kraftfahrtechnisches Taschenbuch,” Bosch, Bauer, H. (Hrsg.), 23. Aufl. Braunschweig, Vieweg: Wiesbaden, pp. 436–437, 1999.

23-19. Fell, A., Samways, A., and Wächter, H.-G. 2004. Entwicklung einer neuen Freistrahlzentrifuge für Dieselmotoren: Wiesbaden. MTZ 65(9): pp. 664–669.

23-8. Klein, G.-M. “Changes in Diesel Fuel Filtration Concepts.” Proceedings 2nd Int. Conf. Filtration in Transportation, Stuttgart, Bergmann, L. (Hrsg.), Filter Media Consulting: LaGrange; USA, pp. 45–49, 1999

23-20. Reyinger, J., Durst, M., and Klein, G.-M. 2004. Kraftstofffilter für zukünftige Diesel- und Ottomotoren mit Direkteinspritzung: Wiesbaden. MTZ 65(3): pp. 196ff.

23-3. DIN/ISO 5011. International Organization of Standardization 5011:2014 cancels and replaces previous editions.

23-9. Bessee, G., Yost, D., Jones, G., Becktel, D. et al., “High-Pressure Injection Fuel System Wear Study,” SAE Technical Paper 980869, 1998, doi:10.4271/980869. Warrendale USA

874 | Internal Combustion Engine Handbook

6606_Book.indb 874

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24 Calculation and Simulation The use of computer-simulation methods during development has become increasingly established in the industrial environment in recent years because of the significant contribution it makes to increased efficiency. Alongside expansions of the available technical and scientific software, and the rapid advances in computing power achieved, this success can essentially be attributed to the anchoring of the process chain concept in the computer aided engineering (CAE)-assisted development process. This has made it possible to tailor the level of detail and informational output of the computation methods used to the engine’s state of development and, therefore, to the problems requiring solution at that particular point. Also included is the recognition of the fact that the value-creation factor in the computation procedures can be optimized only through a holistic view using design methods shaped by CAD practice. The consequence is that the procedure of •• provision of the necessary geometry, materials property data, characteristics data, and so on in a form suitable for calculation •• definition of the load and boundary conditions for the particular problem •• drafting of a computation model •• evaluation and interpretation of the results must be anchored in the design process, with due account taken of time and cost management factors.

24.1 Strength and Vibration Calculation 24.1.1 Procedures and Methods

Stress and vibration studies play a central role in component design. They are, on the one hand, a precondition for optimum materials efficiency and, thus, have a direct effect on manufacturing costs, and, on the other hand, it is possible in many cases to achieve functional improvements—a reduction in the weight of the crank web, for example, produces a direct reduction in vibration amplitudes of the machine as a whole. The calculation methods used in the development process are orientated around the problem itself, the accuracy levels required, and the time and resources allocated. In many cases, component dimensioning can be accomplished using engineering concepts with simplified relationships. This makes it possible to achieve valuable information for support of the design process extremely quickly and efficiently. The use of more complicated procedures is necessary in the case of complex load states and of components, the geometry of which is defined by open-form surfaces. The finite-element method (FEM) has been proved to be the most effective instrument for this purpose; it permits the simulation of the loads resulting from static and dynamic forces, as well as from temperatures [24-1]. Figure 24.1 shows the finite-element model of an engine. The geometry of the structure is simulated using spatial

Internal Combustion Engine Handbook | 875

6606_Book.indb 875

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

elements. In conjunction with a defined load, the deformations and loads are determined by the computer program within every element and, therefore, at every point in the component. The cost of these results is the input for the generation of this model. Despite the availability of high-performance software systems for support in model generation, a significantly greater time input is required than in the case of the use of the classical engineering formulas. It has only been the consistent use of CAD on a more widespread basis that has made it possible to integrate these simulation methods as a fully accepted and acknowledged aid in the development process. The geometric data processed by CAD are adopted directly from computer systems that generate the required models largely automatically.

Figure 24.1 Finite-element model of an engine.

In addition to questions of component dimensioning on the criteria of strength, vibration problems also play a dominant role in the design of engines. Strength issues resulting from vibration amplitudes dominate in the area of low frequencies, acoustic issues arise from high frequencies. FEMs are used for the treatment of the latter. An important field here is the calculation of acoustic performance deriving from gas forces and the oscillating masses involved. It is necessary to differentiate between structure-borne noise, which is essentially transmitted via the engine bearings and, thus, affects the noise level in the interior of the vehicle, the engine’s acoustic radiation, and the muzzle noise emissions from the intake and exhaust systems, where attention is drawn to [24-2] in the context of the mathematical optimization of these latter. The so-called p-method and the boundary methods have, today, only historical importance for the integration of the calculation processes in the CAD environment. The first allows the use of a mostly unstructured grid of the component to be analyzed; invisible to the user, load factors and their distribution within the spatial elements are determined internally by higher order polynomial formulations in accordance with the specified accuracy levels. The boundary method is reduced to the discretization onto the geometry of the structure’s surface. Because of principal disadvantages (load values only on the surface, high computing times, and limited to linear problems), this method for load analysis is not widely used.

The boundary method is, however, extremely well suited to the calculation of the acoustic field resulting from acoustic radiation from, for example, a complete engine. Today, all questions posed for component dimensioning are answered using the FEM. Strength tasks for the design of components and examinations of their dynamic properties are in the focus. Although the computation methods themselves are reliable, the simulation of the loads involved remains the basic problem, in a number of different forms. During the rotation (crankshaft) or translation (crankshaft drive) of a component, dynamical effects determining the load occur. To analyze these forces, the so-called multibody systems (MBSs) have become established and will be discussed in detail further below. The load data for the calculation of a connecting rod can be determined relatively easily; however, the simulation of the complete engine, considering dynamic effects, is necessary for the determination of forces and the resultant stresses in the crankshaft. Stress analyses of fittings, such as alternator mountings, are similarly complex, because their critical load states are the result of vibrational excitation. At least equally complicated is the investigation of loads caused by thermal exposure. These occur, most particularly, in the cylinder head as a result of the temperature gradients present there. As discussed at a later point, the determination of the participating coefficients of heat transfer from flow calculation is an essential precondition here. In addition to stress calculations, mathematical methods are also used to obtain information on component deformation. This includes oval distortion of cylinder sleeves as a result, for instance, of pretensioning of cylinder-head bolts and exhaust manifold expansion. All these loads occur cyclically and can, therefore, be evaluated only in conjunction with a service life statement. Further procedures allow the use of calculated stresses for an estimate of the component’s service life, as long as the load progression is known [24-3]. These types of analyses require additional information, such as materials properties and machining condition. Special calculation methods are used and will be discussed in detail in Chapter 24.1.3. In the context of dynamic effects (crankshaft and valves), the components are often subject to major movements. Different processes must be used, because the FEM expects minor deformations. For the stress analysis of such phenomena, the MBS method has become established. Unlike the FEM, in which the components to be studied are broken down into individual cells, components are described in the MBS procedure in the form of bodies with inertia and flexibility properties. The solution supplies dynamic variables, such as velocity, acceleration, forces, for everybody in the time range, considering all geometrically nonlinear effects. The model generation is usually complex and time-consuming, because the nonlinearity of system parameters, such as attenuation, must also be stated to obtain dependable results. MBS methods are used primarily for the determination of forces under operating conditions as an input for stress studies using the FEM. MBS methods are capable of simulating flexible component properties only to a limited extent, however. The

876 | Internal Combustion Engine Handbook

6606_Book.indb 876

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

two methods are, therefore, used in combination for special problems. Parts of the structure, whose flexibility properties are significant, are defined using finite elements (FEs), and the modal parameters (frequencies and intrinsic geometries) determined from them become an element in the MBS model. The precondition is that the component described using FE exhibits linear-elastic behavior. A typical example of such a procedure is the calculation of the acoustic radiation of the engine-transmission train. As a result of its elasticity, the cylinder and crankcase unit produces a vibrating surface with low amplitudes that can be determined extremely easily using finite-element procedures, whereas excitation of forces is the result of rotating and oscillating masses that are modeled using MBS methods. The FEM is also used as the basis for mathematical optimization procedures [24-4]. These can be used extremely efficiently to perform special tasks such as minimization of component weight, for example. Specified parameters, such as material thickness or factors determining geometry, are calculated in such a way that the target quantity—weight, for example—is minimized. Computation offers great potential in this area, because the results obtained in this way are generally better, and are available in a significantly shorter time, than those achieved using the classical “trial-and-error” procedure. Computing-time consumption is high, however, and restricts the number of parameters (also referred to “design variables”). These optimization methods start from a given component geometry and vary only its parameters (for example, the coordinates of significant points), whereas topology optimization, as a special variant of these methods, makes it possible to design the geometry of a not yet defined component in such a way that a specific variable (for instance, weight) is minimized. Although design work is also additionally necessary for the ultimate definition of the component geometry, this optimization step generally supplies valuable information for its design. Such methods are used primarily for component dimensioning, that is, for the optimization of material exploitation for a given static load. Even complex problems such as dynamic stresses and the consideration of nonlinear effects can be included [24-5].

Empirical values are used for damping, which mostly affects the vibration amplitudes. Figure 24.2 shows a typical finite-element net of a crankshaft. Cuboid elements or the so-called trias (pyramid on the basis of a triangle) are used.

Figure 24.2 Finite-element mesh of a crankshaft.

The former not only enable higher accuracy but also make higher demands on the mesh generation. Powerful computer systems allow one to attain satisfactory accuracy when a sufficiently high number of tria elements are used. The geometry description of the CAD model is the starting point for mesh generation. Mature software systems enable one to mostly automatically create the FEM mesh. To attain a high level of efficiency, it is essential that the calculation is anchored in the development process. Project planning, ensuring the availability of the geometry data (created in CAD) and the material and machining information required for calculation, is essential. It has been found beneficial to keep the organization of calculation and design as close as possible. The results are deformations, vibration forms and the corresponding frequencies, and the stress progressions on the surface. The variables required for design decisions are graphically processed to display, for example, the size and locations of stress concentrations (Figure 24.3).

24.1.2 Selected Examples of Application 24.1.2.1 Crankshaft Strength Information on stresses and, therefore, on the strength of a component plays an essential role in the component’s dimensioning. Methods of determining loads are, therefore, an important element in the design process. The crankshaft is loaded by the crankshaft drive mass forces which act, with the gas forces, as oscillating forces on the shaft. The dynamic loads from the accelerations of piston and connecting rod are determined in an MBS analysis, and the progression of gas forces is taken from an indicator diagram. The crankshaft is mapped with a finite-element model, enabling a realistic consideration of strength characteristics and mass distribution.

Figure 24.3 Stress distributions because of gas and mass forces.

Internal Combustion Engine Handbook | 877

6606_Book.indb 877

1/19/16 8:55 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

24.1.2.2 Strength of an Exhaust Manifold The design of the exhaust manifold is one of the most difficult problems in component dimensioning for an engine. The thermal load exposure and the high dependence of material properties on temperature necessitate intensive integration of experimental technology in layout and design; in many cases, adequate durability can be achieved only via the selection of a high-cost material. Simulation methods permit the optimization of a cost-intensive process to a particularly great extent. The specified load consists of an accumulation of a number of heating and cooling phases. The loads to be determined occur, on the one hand, as a result of the temperature gradients in the manifold and, on the other hand, as a result of the fact that the manifold’s expansion is hindered, because it is fixed to the cylinder head. The FEM is used for the calculation of the relevant stresses and temperature distributions; the determination of coefficients of heat transfer is accomplished using a 3-D simulation of flow within a manifold, as will be described at a later point. The dependence on temperature of all materials data must, of course, be considered. The coefficients of heat transfer initially established, and the exhaust-gas temperatures determined from 1-D computation are used to calculate the (nonsteady-state) temperature field, from which nonlinear finite-element calculation is used to produce the stress distribution, from which an estimation of the number of endurable load cycles can be made, using suitable materials data. The result is a representation of the fracture-endangered locations (Figure 24.4).

process. Generation of the model is extremely complex, however, because the crank web, an essential factor for the simulation of the relevant excitation forces, must be additionally described using FEs, in addition to a detailed simulation of the structure of the crankcase, gearbox, cylinder head, and so on. The fact that the crankshaft bearings exhibit nonlinear behavior as a result of the hydraulic properties of the lubricating film must also be considered. For these reasons, a simplified method is used in practice, to obtain at least qualitative information for initial drafts. The so-called pulse echo method replaces the excitation forces caused by the crank drive by standard forces that are attached to the main bearings; this avoids the necessity of modeling the lubricating film and the crank drive. The result is an evaluation of the transmission behavior of the structure from which, even at this early point, significant decisions can be derived. Complete modeling of the engine is then used for detailed optimization. In all the cases, acoustic velocities on the surface are used for the identification of those areas that exert the main influence on radiation. A practical and actual application is described in [24-6]. Figure 24.5 shows a calculated velocity distribution from which critical parts of the surface can be identified. The velocities found on the surface permit determination of the acoustic pressure in the immediate vicinity from a subsequent calculation. The values determined are, however, interesting only for specific issues in vehicle engines, such as the radiation behavior of the oil pan.

Figure 24.5 Distribution of acoustic velocities on an engine surface. See color section page 1093.

Figure 24.4 Distribution of breaking load cycles of an exhaust manifold. See color section page 1093.

24.1.2.3 Acoustics The acoustic radiation behavior of the engine is essentially determined by component geometry. Mathematical simulation makes it possible to obtain significant information for design and material selection at an extremely early stage of the design

The behavior of the engine in the vehicle is a decisive factor for the acoustics in the vehicle passenger space. Simulation of this necessitates further “escalation” of the modeling process—a simulation of the vehicle components surrounding the engine, such as the body shell and suspension, is required here. Because of the time and financial input necessary, such a simulation cannot be integrated into the development process and can be used only to obtain information on the engine’s acoustic performance in the vehicle in the context of basic research activities.

878 | Internal Combustion Engine Handbook

6606_Book.indb 878

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

24.1.2.4 Valve Actuation Dynamic Matching of valve-spring stiffness with the cam lift curve is accomplished from kinematic observations that also include the acceleration and deceleration of the valve. Superimposed on these movements are the vibrations of the valve and of the components of the entire camshaft drive and timing system. The resultant forces are definitive for the valve timing system’s service life. MBSs, using nonlinearities such as those of the chain drive system and damping mechanisms, for example, can be registered. Figure 24.6 shows the computation model of a valve timing system, using the intrinsic dynamic of the chain, and the valve springs and the camshaft are simulated. To register all the forces acting in the engine, the chain or belt used to drive the camshaft and the crank drive are depicted additionally in an engine model, from which the dynamic behavior of the entire engine is simulated.

Figure 24.7 Optimization of the geometry of a connecting rod. Left: plot of stress in initial state. Right: plot of stress in optimized state. See color section page 1093.

24.1.3 Piston Calculations

Figure 24.6 MBS model of a valve train.

24.1.2.5 Component Optimization One of the designer’s essential tasks is that of determining that component geometry for which, with a minimum of material input, the permissible loading level is not exceeded. The FEM provides valuable assistance in this process, because its use makes it possible to evaluate all possible variants and, thus, select the best. This target can be achieved much more efficiently via the utilization of optimization methods using— within a limited scope—component geometry that can be determined on the basis of an optimization strategy, considering certain background and boundary conditions. Weight was defined as the quantity to be minimized in the case of the connecting rod shown in Figure 24.7, and permissible stress as the boundary condition. The starting point and the result are shown in Figure 24.7 (left) and (right), respectively. The shaded areas represent the stress distributions; an increase in the weight of only 6 g makes it possible to achieve a reduction of around 40% in maximum stresses.

24.1.3.1 Overview The pistons in modern combustion engines belong to the most thermomechanically stressed components in the engine. Current trends in engine development such as downsizing, direct injection, and turbocharging result in a further increase in the thermomechanical stress on the piston having made the development process even more demanding. The combination of thermal stress because of hot combustion gases and cyclical mechanical loads caused by high peak pressure alternating with high centrifugal forces at simultaneous specification of demanding weight objectives demands a very precise component design to reliably meet the required operational strength. In this context, the calculation has also the task to present or verify further design-related potential for weight reduction, because a weight reduction of the oscillating masses, such as the piston group, may also reduce the masses of the counter-weights, the balancing shafts, and the flywheel. Secondary effects are a higher engine dynamic, smoother running, and less bearing stress. Because of the complex 3-D structure and the manifold loads, it is impossible to use reliable design pistons using analytical formulas. For this reason, numerical calculation methods such as the FEM [24-7] and operational strength calculation [24-8] have become an indispensable component of the entire development process. Because of shortened development times and permanent cost pressure, CAD is used to design a component geometry during a virtual product development phase; this

Internal Combustion Engine Handbook | 879

6606_Book.indb 879

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

geometry is verified for its function by numerical processes, such as FEM, and, subsequently, further developed in several development loops until a virtual functional reliability is achieved. Prototypes will be manufactured only after virtual approval, followed by a verification made in component and engine trials. Because the piston, with the wristpin and the connecting rod, transfers the combustion pressure onto the crankshaft, it is necessary that these components are designed together to achieve optimal system performance. Furthermore, the geometry of the piston boss is largely determined by the length and diameter of the wristpin and the width and angle of the large conrod eye, which means that these values should be defined at the beginning of the development process to avoid principal geometry changes during a later development phase. The predefinition of these basic dimensions can be reliably executed on the basis of analytical formulas for the mean bearing load in the piston boss in the conrod end and the flexure and oval distortion of the wristpin. The process of piston calculation is divided in the following partial steps, which are representative for cyclically thermomechanically loaded components (Figure 24.8): •• creation of the finite-element model on the basis of the CAD geometry •• thermodynamic simulation of the combustion process for determining the thermal conditions •• FE calculation of the temperature map •• FE calculation of tensions and deformations for each load event to be analyzed •• assessment of operational strength considering different load events •• evaluation of the calculation results. The result of the piston calculation is, if the criteria for operation strength have been met, the virtual component approval or, if the criteria have been failed, a list of detailed suggestions for the improvement of the component design.

CAD model

Structure of the FE model

The dynamic performance of the piston in the cylinder is analyzed using secondary piston motion calculations. The hydrodynamic damping of the oil film between piston and cylinder wall, in particular, is here considered. The term “secondary piston motion” refers to the rotatory and translatory movement to demarcate from the primary axial movement. An important result of these calculations is the contact pressure between the piston and the cylinder as a function of crank angle. It permits the derivation of information on the piston skirt wear and, in an extreme case, on the danger of piston seizure in the cylinder. In addition, the frictional power and the secondary kinetic energy of the piston during change of contact are calculated to assess the noise emission (noise, vibration, and harshness) of the piston. The different piston concepts also pose different demands on the virtual component design. In aluminum pistons, for example, the operational strength and, therefore, the thermal and mechanical stress are the most important criterion. The secondary piston motion is more a comfort criterion because of the higher flexibility of the piston skirt. In steel pistons, on the other hand, the operational strength is a lesser criterion because of the high strength of the basic materials. Important design criteria are here the temperature distribution because of the lower thermal conductivity, and the secondary piston motion because of the lower resilience and the component weight optimization because of the higher specific weight. Overall, calculated design is very successful to meet opposing requirements. All such observations must never disregard that the interactions with other components have a decisive influence on the piston’s operating performance. Deformations of the cylinder, for example, influence secondary piston motions and, therefore, the contact pressure between the cylinder and the piston. These secondary piston motions, for their part, have an influence on the dynamic behavior of the piston rings, and, thus, on blow-by figures and oil consumption. The result, ultimately, is that only observation of the overall system (cylinder, piston

Mechani- OperaThermo- Thermal tional FE calcucal FE dynamics lation calculation strength

Report

Figure 24.8 Partial processes of piston calculation. See color section page 1094.

880 | Internal Combustion Engine Handbook

6606_Book.indb 880

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

rings, pistons, wristpin, connecting rod, and bearings) can provide detailed information on operating performance. 24.1.3.2 Demands on Piston Materials and Their Characteristics Because of the high specific demands, new piston materials have been developed on the basis of eutectic aluminumsilicon casting alloys and are successfully established in the marketplace. These alloys differ mainly by their copper and nickel content and the type of their structure modification. The objective of the alloy development is to optimize the mechanical and physical material properties in a targeted manner. These properties mostly depend on the microstructure of the cast pistons. Figure 24.9 shows a typical example of the microstructure of an eutectic piston alloy.

Figure 24.9 Microstructure of an eutectic piston alloy.

This highly complex microstructure with visible primary a aluminum, eutectic silicon, and various intermetallic phases is mostly distinguished by additional, microscopic, finely distributed precipitations, which can only be seen in a transmission electron microscope and providing a crucial contribution to the capacity of a piston alloy. Their type, size, and shape essentially determine the mechanical properties of the alloy in dependence to the temperature. Their occurrence, on the other hand, is determined by the chemical composition of the alloy and the processes during piston manufacturing. The key for the development of modern piston materials is, therefore, a comprehensive understanding of the interconnections among composition, process, and microstructure. For the precise calculation of temperature, thermal and mechanical tension, and the operational strength of a piston, accurate material data for each material are required. The principal data comprise the density, thermal conductivity, specific thermal capacity, Young’s modulus, and Poisson’s ratio. They are expanded by material data for describing the elastic-plastic material characteristics, the thermally caused relaxation, and the operational strength. Because piston alloys are usually special developments of the piston manufacturers, the definition of material data is part of their core expertise.

These data must be experimentally determined for the entire permissible temperature range. For the necessary number of trials, the statistic validation of the measured values must be ensured for every temperature level. More than 500 individual trials are required for a complete characterization of a piston alloy. 24.1.3.3 Creation of the Finite-Element Model on the Basis of the CAD Geometry To create the reference geometry for the finite-element model, certain details in the finished part geometry not required for the calculation (such as groove bottom radii of the ring grooves) are suppressed and exported in a suitable CAD format. This simplification is necessary to keep modeling efforts and computing times within an acceptable range. On the basis of this reference geometry, an FE mesh is created in a preprocessor where extremely high-loaded zones, such as the piston recess of a diesel piston or the shaft connection of an SI piston, are meshed relatively finely to attain high accuracy while less loaded zones are meshed correspondingly more coarsely. Because computing times disproportionately increases with the number of nodes, the optimal mesh size is determined by balancing sufficient accuracy and permissible computing time. Because of the complex piston geometry, mainly tetrahedron elements with parabolic extension functions are used for meshing, because these elements allow an automatic volume meshing. The alternative of a semiautomatic meshing with linear hexahedron elements requires a significantly higher effort in model creation with a nearly equivalent result. Ideally, the same FE model can be used for thermal and mechanical calculations. Depending on the component symmetry and the circumstances, either a full model or a half model with partition in the conrod oscillation plane is built. For the correct representation of the nonlinear contact situation, an assembly consisting of piston (if present, with ring carrier and piston bushing), wristpin, top half of the connecting rod (if present, with connecting rod bushing), and a suitable simplification of the cylinder or the cylinder block are used in structure-mechanical calculation (Figure 24.10). The compression ring set can be ignored in this type of calculation. All contact pairs are coupled using suitable contact conditions depending on the calculation type. The shaped bore of the piston boss and the accurate outer contour of the piston skirt are precisely captured, considering each ovality and the axial profile by defining the gap dimension for each node, for the polygonal discretization of the contact surfaces not resulting in a fault of the contact tensions. 24.1.3.4 Thermodynamic Simulation to Determine the Thermal Marginal Condition The correct temperature map and, therefore, the thermal marginal condition required for the determination are the foundation of an informationally useful piston calculation. The temperature directly affects the thermal expansion, the thermal tensions because of temperature gradients, and the local component strength because of the temperature dependency of the material characteristics. However, during the initial

Internal Combustion Engine Handbook | 881

6606_Book.indb 881

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

with coefficients of heat transfer and limit temperature are used as standard. A decisive advantage of this procedure is that the effects of various engine combustion processes on the piston temperature can be efficiently simulated. Furthermore, alternative fuels can be considered with little effort in the work process calculation.

Figure 24.10 Finite-element models of a piston assembly in explosion representation.

virtual development phase, temperature measurements are not yet available, and the precise calculation of the thermal marginal conditions is indispensable. In addition, a thermodynamic simulation of the work process can be performed on the basis of the operating parameters, which results in the heat input in the piston at the combustion chamber side (Figure 24.11). Using suitable models, the heat flows over the piston rings, and the piston skirt onto the cylinder running surfaces and the heat flow into the oil are also determined.

24.1.3.5 FE Calculation of the Temperature Map Because of the thermal inertia of the piston and the highfrequent temperature load at the combustion chamber end during a work cycle, temperatures cannot be assumed to be time constant over a work cycle in piston areas not close to the surface. These temperatures are solely determined by the operating state over the work process. Thin surface layers at the piston head are subject to cyclical temperature load superimposed of the stationary temperature map within a work cycle because of the change from hot combustion gases and cool fresh air. The resulting thermal tensions represent an additional stress on the material, which can be considered during the subsequent operating strength calculations by usually applying slightly increased safety factors. To calculate the temperature distribution in the piston, the temperatures of the combustion gas, the cylinder, and the oil at injection cooling (or the oil mist) are determined with the corresponding coefficients of heat transfer by executing a thermodynamic simulation (see above). This information is used to calculate the piston’s temperature map by means of FEM. The maximum temperatures generally occur at the design output working point. Typical temperature maps are shown for aluminum pistons in passenger car SI engines (Figure 24.12), aluminum pistons in passenger car diesel engines (Figure 24.13), and steel pistons in diesel trucks (Figure 24.14).

Piston Rings/Liner

Figure 24.11 Piston heat flow balance [24-9].

Piston cooling by the oil jet, in particular, must be considered. Using the determined heat flows, the spatial distribution of the thermal marginal condition is derived and applied to the finite-element model. Convective marginal conditions

Figure 24.12 Temperature of an aluminum piston for an SI engine in a passenger car at rated capacity. See color section page 1094.

882 | Internal Combustion Engine Handbook

6606_Book.indb 882

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

Standard Piston Applications

Extreme Piston Applications

Aluminum piston in SI engines

290/−310°C

350°C

Aluminum piston in diesel passenger cars

360/−390°C

420°C

Aluminum piston in diesel trucks

300/−350°C

360°C

Steel piston in diesel trucks

350/−460°C

500°C

Table 24.1 Maximal temperatures of different piston applications

Figure 24.13 Temperature of an aluminum piston for a diesel engine in a passenger car at rated capacity. See color section page 1094.

Figure 24.14 Temperature of an steel piston for a diesel engine in a truck at rated capacity. See color section page 1095.

The maximum piston temperatures are listed for various applications in Table 24.1. The extreme values represent very hot engine applications or thermally unfavorable combustion recesses. It must be noted, however, that such high temperatures will reduce the mechanical resistance of the piston because of the associated reduction in strength. The permissible limits for piston temperatures predominantly result from the temperature-dependent fatigue strengths under alternating load of each alloy and are, therefore, a

load-depending variable. The melting of individual alloy phases can be observed only under extremely abusive conditions. Furthermore, the maximum temperatures of the ring grooves and the piston boss must be observed, because the functioning of the piston rings and the wristpin are no longer assured at excessively high temperatures. In the course of the development process, temperature measurements are usually taken in the engine as soon as the first prototypes are available. At a significant deviation of the temperature distribution between forecast calculation and measurement, the temperature map is adjusted for simulation, and the subsequent calculation steps are repeated. 24.1.3.6 FE Calculation of Tensions and Deformations for Each Load Event to be Analyzed The kinematics of the crankshaft drive and the resulting forces on the piston are described in depth in Chapter 6.1. In addition to the vertical piston force from the combustion pressure and the acceleration forces, the piston-side force or normal force must be noted, which acts as a reaction force from the cylinder on the piston, independence on the piston force and the applied conrod angle (Figure 24.15 and Figure 24.16). In fast-running SI engines, in particular, the piston-side force represents a significant load on the piston skirt. For efficient modeling, the real dynamic work cycle for the most important operating points—usually the maximum torque and maximum output—are reduced to the smallest possible number of static equivalent load events (for example, centrifugal force at top dead center (TDC) or maximum pressure load shortly after CDC). For each equivalent load event, a separate static calculation is made. In the subsequent operational strength calculations, all equivalent load events are analyzed as a joint load collective that should result in a comparable damage and the complete dynamic cycle. On the basis of the temperature map for the relevant load point, the thermal expansion of the piston in the cylinder is first calculated (Figure 24.17). The bulging of the piston head and the increasing enlargements of diameter from the skirt end to the piston head are clearly apparent. This diameter enlargement under operating conditions is anticipated during the design phase by providing suitable places at the fire land and the ring lands and a corresponding shaft profile. The thermal expansion of the piston must be also considered during the collision control with the valves. Because of the inhomogeneous temperature distribution, significant thermal tensions occur in diesel pistons, in particular.

Internal Combustion Engine Handbook | 883

6606_Book.indb 883

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

Vertical forces SI piston 75 Fz (pressure) 60

Fz (total) Fz (mass)

Power [kN]

45

30

15

0

Intake

Compression

Work

Exhaust

-15 0

60

120

180

240

300

360

420

480

540

600

660

720

Crank angle [°]

Figure 24.15 Vertical piston force of an SI engine piston (maximum output operating point). See color section page 1095.

Lateral forces SI piston 12 Lateral (2000 rpm) 10

Lateral (5000 rpm) Lateral (7000 rpm)

8

Ignition pressure [bar]

Lateral force [kN]

Ignition pressure (5000 rpm) 6 4 2 0 -2 -4

Intake

Compression

Work

Exhaust

-6 0

60

120

180

240

300

360

Crank angle [°]

420

480

540

600

660

720

Figure 24.16 Piston-side force of an SI engine piston, in dependence on the speed. See color section page 1096.

884 | Internal Combustion Engine Handbook

6606_Book.indb 884

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

cylinder pressure progression can be obtained by measurement or simulation. The latter is usually used during the design phase for new engines when the corresponding measurements at the real engine are not yet available. Alternatively, one can use values from comparable applications or target value specifications for the calculation. Table 24.2 provides an overview of the maximum cylinder pressure for different engine concepts. Engine Concept

Maximum Cylinder Pressure

Nonsupercharged SI engines

70–90 bar

Supercharged SI engines

100–120 bar

Supercharged passenger car diesel engines

165–190 bar

Supercharged truck diesel engines 200–240 bar Table 24.2 Maximum cylinder pressure in different engine concepts

Figure 24.17 Thermal deformation of a diesel piston (fifty times enhanced representation). See color section page 1096.

For example, the thermal expansion at the hot recess is impaired by the relatively cooler piston head resulting in the build-up of compression stresses at the recess and tensile stresses at the fire land. Because a large portion of the thermal stresses is decreased by relaxation because of the high component temperatures, a fact must be considered during simulation (Figure 24.18). In the next step, the cylinder pressure and the acceleration of each design point are applied to the component. The

When selecting the design cylinder pressure, it must be ensured that this component is not overdimensioned as this would compromise the goal of weight optimization. Because the maximum cylinder pressure, in SI engines in particular, is subject to relatively high fluctuations because of differences in the mixture formation, a statistic analysis is required. By assuming a Gaussian distribution of the maximum cylinder pressure with mean value p and standard deviation σ s, and a linear damage accumulation of the piston, a sufficient design cylinder pressure of p + 1s is achieved. Sufficient means here that a cyclical load of the component at constant design cylinder pressure causes the same damage effect over all the load cycles as would the total real distribution of the maximum cylinder pressure from p − 6s to p + 6s. During operation, cylinder pressures may actually occur that exceed the design cylinder pressure. These events, however, are considered during design because of the statistical analysis.

 Figure 24.18 Thermal stresses (third principal stress) before and after relaxation. See color section page 1096.

Internal Combustion Engine Handbook | 885

6606_Book.indb 885

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

Gas force

Gas force

Figure 24.19 Mechanical Von Mises stress with twenty-five times scaling of mechanical deformation at maximum ignition pressure (CDC). See color section page 1097.

Cylinder wall pressure-end Section, running plane

The mechanical deformation and the mechanical share of the Von Mises comparative stress of a diesel piston is shown for the maximum pressure load event shortly after the TDC (Figure 24.19). For a clear representation, the shares in thermal deformation and thermal stress have been subtracted from the original thermomechanic state; in addition, the mechanical deformation is shown in a twenty-five times enhanced representation. In the illustration, the force transmission over the wristpin onto the connecting rod and the corresponding oval distortion and flexure of the wristpin can be clearly seen. Because of the shaped bore in the piston boss used in this example, an even force transmission without critical load peaks can be achieved. In high local contact pressures as they are generated at the boss inside without shaped bore, an elastic-plastic calculation is required to correctly map the plastic deformation of the boss and the corresponding shift of the contact pressures. Because of the system behavior, the design of the piston itself and that of the piston pin (internal/ external diameter, length, and so on), supported length (total piston pin length and width of the little end boss), the design of the upper little end boss (parallel or trapezoidal), and the contact pressure between the piston and the cylinder all have a significant influence on piston deformation. The comparative stresses, according to Von Mises, of the wristpin under maximum ignition pressure are shown in Figure 24.20 with twenty-five times enhanced mechanical deformation. An internal cone is usually used for weight reduction, because the ends of the pin exhibit only low stress. Figure 24.21 shows an example of the mechanical stress of the recess of a diesel piston. To illustrate the relevant stress component, the mechanical circumferential stress in a cylindrical coordinate system is shown along the center axis of the engine cylinder with the thermal stress shares subtracted. Because of the combustion pressure, a global flexing of piston occurs over the wristpin causing tensile stresses in the pin plane and the compression stresses in the running plane. With a suitable selection of the ovality of the piston boss and the corresponding change of the support of the piston on the bolt,

Section, pin plane

a partial shift of the stresses between piston boss and recess boundary can be achieved in the pin plane. By modifying the shaped bore profile of the piston boss and, therefore, a shift of the equivalent bearing pressure point, a partial shift of the stresses at the recess boundary can be achieved between the running plane and the pin plane. 24.1.3.7 Assessment of Operational Strength For assessing the operational strength, a number of load changes and the type of stress are used as classification criterion. A very high number of load changes (>105) with predominantly elastic stress is called a long-time fatigue or high-cycle fatigue, while a relative low number of load cycles (≤105) with predominantly elastic-plastic stress is called a short-time fatigue or low-cycle fatigue. The superimposition of a cyclical mechanical stress

Gas force

Figure 24.20 Von Mises stress in the wristpin with twenty-five times scaling of mechanical deformation at maximum ignition pressure (CDC). See color section page 1097.

886 | Internal Combustion Engine Handbook

6606_Book.indb 886

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.1 Strength and Vibration Calculation

difficult to see because of the complexity of the stress, and, for this reason, is considered in the simulation mostly by additional safety factors. To determine the operational strength of the piston, the fatigue strength values of the piston alloy must be known independence of the temperature for different load cycle numbers. For this purpose, temperature-dependent Wöhler curves are determined from stress-controlled continuous vibration tests for 105–108 vibration cycles (Figure 24.22). In a double-algorithmic diagram, this results in a linear relationship between the stress amplitude Sa and the vibration cycle N: log N = −b log Sa + C





At a known point (NA, SA) on the Wöhler curve, the following applies: N ⎛ Sa ⎞ = N A ⎜⎝ SA ⎟⎠



Figure 24.21 Mechanical circumferential stress in the recess of a diesel piston. See color section page 1097.

with a cyclical thermal stress is known as thermomechanical fatigue (TMF). The stress on the piston because of the ignition pressure occurring in every work cycle at constant operating state of the engine is, therefore, an isothermic HFC stress as the piston temperature is constant. Changes in the operating states (for example, from idle to full-load and back) and the corresponding transient changes in the piston temperature and the thermal stresses are a low-cycle TMF stress. The transient thermal stress of thin surface layers at the piston head, caused by hot combustion gases in every work cycle, is a high-cycle TMF stress, which, however, is experimentally

−b



Because of the differently severe stress by compression and centrifugal forces at simultaneously present thermal stresses, the mean stresses in the piston are usually not zero, for which it is necessary to consider the dependence of the permissible amplitude of the mean stress by applying a mean stress correction according to Smith or Haigh. The term damage D of an operating state is defined as the relationship between the number of required cycles n and the number of endurable cycles N: D=



n N

If the damage index exceeds the value 1, a failure of the component before reaching the demanded service life is probable. A more intuitive representation is given with the safety factor SF, defining, for a number of load cycles, the ratio between the

log S

200°C 300°C 400°C

0

1

2

3

4

log N

5

6

7

8

Figure 24.22 Wöhler diagram. See color section page 1098.

Internal Combustion Engine Handbook | 887

6606_Book.indb 887

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

permissible tension S and the existing stress S. If this value is larger than 1, the component is safe:

sF (N) =

Szul (N) S

To calculate the service life, the temperature distributions and the stress states of the individual working points for every surface node are set against temperature-dependent fatigue strengths under alternating load. At a single-axis oscillating stress between two load states, the service life to be excepted can be directly derived from the comparison of the existing stress amplitude and the mean stress with the permissible amplitude for each stress ratio. For multiaxis, nonproportional stress states in the piston, a simple analysis using the primary stresses is not possible, and the “critical plane method” is used. Here, the mean stress and the stress amplitudes of two load events at a time are calculated in an initial step from the two local stress states for a sectional plane selected at will. The amplitudes and mean stresses operating at this sectional plane permit, given the use of monoaxial or multiaxial damage hypotheses, the determination of the damage acting on this sectional plane. The sectional plane is then rotated step by step about all three spatial axes, and the amplitudes and the mean stresses are recalculated with the corresponding damage for each new orientation. The plane with the highest damage is designated the “critical plane.” The damage index of this critical plane is used for the calculation of anticipated service life. The safety factor can also be calculated. The spatial distribution and each critical zone can be displayed and analyzed using common FEM postprocessors (Figure 24.23). Because materials characteristics data determined experimentally are available only in the form of monoaxial tests on sample rods and, in addition, necessary simplifications have been made to the model, the real operating conditions in the

engine are precisely mirrors in the simulation. On the basis of comprehensive engine trials, minimum safety factors are defined for the different stress zones, to be able to meet the real operating conditions with the required safety. The simple load case of an oscillating component loading between only two working points examined above is frequently used, but it does not, of course, represent the actual loads occurring under an extremely large number of differing operating steps. Suitable damage accumulation hypotheses, such as the Miner rule, in which the ratio of the required cycles to the endurable cycles is generated and added up across all the operating states, for example, must be used if the influence of multiple operating states on service life is to be considered:

D=∑ (i)

ni Ni



A failure of the component prior to reach the required overall service life must be anticipated when the accumulated damage exceeds the value 1. 24.1.3.8 Evaluation of the Calculation Results When evaluating the calculation results, it must be ensured that every piston point that is relevant for operational strength is examined. By comparing the results with the results of previous design versions, modification instructions can be derived for further improvement of the product design. To ensure a reliable and robust evaluation process, the process must be either specified and controlled in detail or mostly automated by microprogramming. In this manner, variations in the evaluation by different personnel can be minimized. Because of the similar topology of even very different piston designs, a high degree of automation is possible and makes sense commercially.

 Figure 24.23 Safety factor. See color section page 1098.

888 | Internal Combustion Engine Handbook

6606_Book.indb 888

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

The numerous results from temperatures, stresses, and safety factors for all the evaluation areas are summarized as reports, and result in an overall assessment of the design. With a verified design version, it is possible to approve the virtual component.

consumption, mechanical loading, and acoustics. Although oneor quasi-dimensional computation methods are used here, the concepts used permit a spatially differentiated output of coefficients of heat transfer for finite-element calculations in which the information of component strength can be ascertained. Following the initial phase of approximate component dimensioning, locally and chronologically resolved information on velocities, pressures, temperatures, and mixture compositions, which can be obtained only via 3-D flow calculations, is needed to support the development process.

24.2 Flow Calculation The results of flow simulations provide the development engineer with a detailed insight into engine processes that, on the one hand, disclose potentials for improvement and, on the other hand, serve for virtual verification of function. Nowadays, exhaust system concepts, for example, are evaluated using 3-D flow simulations. The potential of this “virtual product development” becomes rationally exploitable only in interaction with the greater use of optical measuring methods, because the tuning of numerical models on the basis of measured data is necessary in many applications. Because the computing processes have attained a high degree of maturity some years ago, the focus in the past few years has shifted to the development and acceleration of processes to use the developed computing processes end-to-end on a broad scale. The objective of a modern development process is the optimization of all flow-guiding components using the computational fluid dynamics (CFD) process throughout the development. This applies equally to air and exhaust systems, cooling and oil circuit, and the events in the cylinder, including injection and combustion. The starting point for engine development is the simulation of the gas charge exchange processes, which, taken with a working-process calculation, supplies trend information on the engine’s performance data, emissions behavior, energy

Air mass meter

Exhaust turbochargers

Intake silencer

24.2.1.1 Charge Cycle and Work Process Calculation A 1-D calculation of charging is a rational tool not only in the engine’s conceptual phase but also throughout the entire development process, and can be used for the harmonization of components because of the interaction of a range of differing influences (flow, gas dynamics, energy conversion, and so on). Such a 1-D procedure has become established as an optimum compromise between the development input and the accuracy required for the definition of flow phenomena. The advantage of this method over the classical 3-D concepts can be found in its significantly simpler model generation and the much lower computing time requirement. These must be set against the disadvantage of the fact that physical variables, such as flow velocity, are locally averaged, making it impossible to resolve local effects. It is, therefore, necessary, to obtain reliable results, to correlate the properties that cannot be depicted in the computing model empirically, where they are relevant to the result. Figure 24.24 shows a schematic representation of the air and exhaust-gas routing of an eight-cylinder engine with the important units of the computing model for charge change

Preliminary catalytic converter

Preliminary catalytic converter

Air mass meter

24.2.1 One- or Quasi-Dimensional Methods

EGR cooler

EGR-valves

EGR-valves

Intercooling

Air mass meter

EGR cooler Intake silencer

Figure 24.24 Schematic view of air and exhaust-gas routing, showing elements of the charging calculation [24-10].

Internal Combustion Engine Handbook | 889

6606_Book.indb 889

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

calculation emphasized. One important result of simulation is cylinder volumetric efficiency. The results permit the dimensioning of the intake and exhaust systems (pipe lengths and cross sections) and the optimization of the actuating times of the valve timing system. The intake manifold tubes and the exhaust system are represented by pipes, in which the gas flow is simulated one dimensionally with all the dynamic effects occurring. Both the calculated and measured coefficients are used for the inclusion of the pressure drop at pipe bends and branches. The exhaust turbocharger is defined by characteristic maps. A dominant factor in the calculation is the progress of combustion during energy conversion. The 1-D method does not make it possible to resolve the extremely complex combustion processes. The integration of a working-process calculation into the charging program, which calculates using a specified equivalent combustion progression, can be used here. A semiempirical equation, which was proposed by Wiebe [24-11] is frequently used for the determination of the rate of heat liberation. In addition to the empirically established parameters, it is also necessary to introduce a factor from the measured indicator diagram. This method ignores ignition retardation and is, therefore, used mainly for slow-running (low-revolution) gasoline engines. Modifications to an empirical model for the precalculation of combustion progress need to be made here for the study of diesel engines and, in particular, of common rail diesel engines with direct injection to consider the special circumstances involved, such as multiple injection, variable rail pressure, and variable EGR rate (see [24-12], for example). The same also applies to the generation of gasoline engines featuring direct gasoline injection. Rate of heat liberation and, therefore, combustion progress, can be directly used if they are known. The two-zone model permits a more precise simulation of combustion. This assumes a homogeneous mixture and a rotationally symmetrical combustion chamber that is subdivided into a zone of combusted gas and a zone of uncombusted gas. The flame front, the propagation rate of which being determined by the law of combustion, forms the dividing plane. Tuning the model is accomplished via assessment against measurements performed on the high-pressure part of the system. The advantage of this procedure is the fact that it permits at least approximate determination of coefficients of heat transfer on the combustion chamber wall, and these can be further used for the calculation of stresses caused by temperature exposure. The two-zone model is suitable most especially for gasoline engines, because of the assumptions that are used to form the basis for the model. Data from the engine’s predecessor will continue to be needed for tuning mathematical models until it becomes possible to develop generally applicable combustion progress models and to determine the range of parameters taken from existing models. The knock models developed in the past few years can help in answering additional questions about the probability of the occurrence of undesired ignitions. However, because of the

number of input parameters to be selected and missing local resolution of the procedure, the informative value is limited. In addition, they do not resolve the principal problem of the lacking general applicability of the modeling approaches. Detail turbine and compressor characteristics fields will be required in the case of examination of supercharged engines, particularly for the study of their part-load operating ranges. The lack of such data complicates the use of the calculation procedure in the engine’s conceptual phase. An important application case is here the examinations of the interaction between turbocharger and engine to obtain information about capacity, efficiency/consumption, and reaction behavior, the so-called turbocharger mapping. In this case, the stored maps of various turbochargers are used to develop the optimal combination of turbocharger and engine. The literature shows some approaches for the calculated design of turbocharger operational maps, such as [24-13], which can be used in place of the stored maps. 24.2.1.2 Calculation of Vehicle Cooling Circuits Simulation of the entire vehicle cooling circuit on the basis of a 1-D calculation method is a standard application in the design of the engine cooling system. In a procedure similar to the generation of charging models, all the components in the cooling circuit—essentially the water pump, engine, thermostat, coolant and oil cooler, hoses, and header tanks—are connecting to one another in a network and characteristics indices or fields assigned to them. Influencing factors are instantaneous volumetric flow, system pressures, and coolant temperatures. A useful step in the assessment of a range of differing cooling concepts is the simulation of the cooling circuit with 1-D simulation of the radiator/fan group, considering significant heat flows (see [24-14], for example). This makes it possible to calculate the effects of variations in fan and heat exchangers, or the influence of different operating states, on the vehicle cooling data achieved. 24.2.1.3 Calculation of Oil Circuits It is also possible, using similar methods to those described above, to depict the oil circuit of the engine in the form of a 1-D mathematical model. Attention must be devoted here, in particular, to model the oil flow in the main bearings of the crankshaft and camshafts, to obtain realistic results. More complex problems in the design of an oil circuit to assume additional tasks, such as actuating a camshaft shifter, are increasingly answered in the simulation considering the dynamic processes. For example, dynamic pressure peaks or a cavitation risk can be evaluated using the mechanical excitation of various engine components. 24.2.1.4 Simulation of the Fuel Hydraulic Circuit The fuel hydraulic circuit is a further example of an application for this type of calculation method. This involves simulation of highly dynamic processes, in which compressibility factors must also be considered, however. In particular, high-pressure diesel-injection systems with variable rail pressures, it is necessary to calculate the effects

890 | Internal Combustion Engine Handbook

6606_Book.indb 890

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

of fuel-line routings on the individual injectors in terms of the homogeneous distribution of fuel flow to the individual cylinders. Simulation of the effects of rail pressure control on metering of fuel flow is a further application. The behavior of the injector itself need not be resolved in detail but can, instead, be described using characteristics fields. 24.2.1.5 Overall Process Analysis In many cases, knowledge of the steady-state operating behavior of engines is not enough. Instead, it is necessary to simulate transient engine states, particularly in the field of the design and validation of electronic engine management system functions. Specific data on the vehicle itself, the power train, the transmission, engine electronics, ambient conditions, and driver behavior, are, therefore, all included in the computation model in addition to precise thermodynamic modeling of the engine [24-15]. Simulation of vehicle acceleration for the adjustment of the guide-vane mechanism of an exhaust turbocharger with variable turbine-blade geometry, considering various control strategies, may be mentioned here as a specimen application.

24.2.2 3-D Flow Calculation 24.2.2.1 Finite-Volume Method The computation methods mentioned above are not capable of describing flow processes with local resolution. A passage through extremely complex geometries, such as that of a water jacket or the inlet ports of the engine, particularly necessitates the use of 3-D flow simulation. Figure 24.25 shows the mathematical model of the water jacket of a four-cylinder engine.

Figure 24.25 Computing grid for the water jacket of a four-cylinder engine.

Some years ago, the model shown in the illustration consisted of approximately 500,000 volume cells that, taken together, fill the structure through which flow occurs. The same model typically consists, today, of more than 2,000,000 cells. Every geometric detail of the water jacket is resolved here. The finite-volume methods have become established over

finite-difference and finite-element procedures for industrial applications. The conservation equations of mass and impulse are resolved in every individual volume unit. This set of equations is referred to the Navier–Stokes equations (see [24-16], for example). In practice, all technically relevant flow processes have a turbulent character. Further equations, the so-called turbulence models, are needed for the description of flow turbulence. The classical route leads via chronological averaging of the Navier–Stokes equations. The additional terms generated in this averaging operation must also be modeled. Even today, the k-εε model, which necessitates the solution of two additional equations for the characterization of a turbulent longitudinal and time scale, is a standard procedure. Augmentations to this model, the so-called “nonlinear turbulence models,” are, nowadays, increasingly being used in applications in which the interaction of various turbulent eddy structures significantly determines flow topology. The wall-boundary layer must be resolved, to permit adequately accurate treatment of detachments not attributable to discontinuities in the geometry and problems of heat transfer. The so-called “low re-turbulence models” are used then in these fields. The disadvantage of such methods can be found in the large number of computation cells and the resultant elevated level of need for resources such as hard-disk capacity and computing time. The energy conservation equation must also be solved in cases in which additional thermodynamic problems, such as compressible flows, and flows involving energy input or heat transfer, are examined. The system of partial, nonlinear differential equations must be discretized and linearized in an initial operation, to permit the finding of a solution for it. Significant difficulties are created here by the so-called convection term, which describes the transportation of a physical solution variable over the system boundaries of the individual volume units. The commonly used discretization procedures differ in the accuracy with which they define the physical situation. This applies, in particular, to the depiction of steep solution variable gradients. The disadvantage of a complex discretization procedure can generally be found in the associated difficulty of achieving a convergent solution. A solution is generally achieved from an iterative procedure. In an industrial context, the solution of flow problems generally signifies the solution of a boundary value problem. This means that even areas of the solution zone located well downstream are capable of causing flow phenomena such as the occurrence of a detachment throughout the entire solution zone, even well upstream. Decisive importance, therefore, is attached to the definition of boundary conditions. A description of the numerical methods of fluid mechanics can, for example, be found in [24-17]. The three velocity components, pressure, temperature, density, and characteristic turbulence data (degree and length of turbulence) are available as solution variables within each volume element. In addition, the coefficient of heat transfer

Internal Combustion Engine Handbook | 891

6606_Book.indb 891

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

and the heat flux transferred to the wall can be determined in every the so-called “wall cell.” These factors are, in many cases, further processed in finite-element programs, with the aim of obtaining information on component strength. Networking of the flow zone is an essential step on the road to the solution of fluid mechanical problems. It is here that the greatest advances have been achieved in the past five years. The precondition for “automatic” networking is a completely closed surface. The efforts being undertaken, therefore, shift toward CAD systems, to ensure the devotion of attention to the provision of the best-processed surface possible as early as the design stage. There are, essentially, five trends: •• automatic networking on the basis of a Cartesian computation grid, which is not tailored to the body but depicts the precise geometry using additional algorithms •• automatic networking on the basis of tetrahedral grids (for example, the Delauney method), in some cases offering the capability of converting tetrahedrons in the interior of the computation zone back to hexahedral elements •• semiautomatic networking based on block-structured grids •• semiautomatic networking on the basis of hexahedral elements, with toleration of a few tetrahedral or polynomial elements •• mostly automatic networking on the basis of polyhedral cells. Each of the methods mentioned above has both advantages and disadvantages; depending on the required computation grid quality (particularly important in the solution of heat transfer problems), the necessary informational accuracy of the computation results, the available engineering hours, the quality of the CAD surface, the available computer capacity, and the total available time within which reliable results are to be produced. The principle that the method must be holistically evaluated and is only as good as the weakest link in the chain applies here. A missing boundary condition cannot be made good by an extremely finely resolved computation grid. On the contrary, a computation grid containing many distorted volume elements increases computing time and diminishes the accuracy of the results obtained. 24.2.2.2 Simulation Methods Based on the “Lattice Gas Theory” A calculation method that does not solve conservation equations at the macroscopic level (in finite volumes), but instead resolves the processes on a microscopic scale (impact processes at the molecular level), has become established, primarily in the field of simulation of flow on vehicle exteriors. Because of the advantages listed below, this method is widely used in the examination of engine space flows. Applications for engine flows, however, have remained exceptions. This is a particle-based procedure, in which particles exist in time and space, and move in discrete directions at discrete velocities. All the relevant flow variables can be determined using statistics based on discrete kinetic theory.

Turbulence modeling is accomplished here using very large eddy simulation. Further information on this technique can be found, for example, in [24-18]. The benefits of the method are: •• automatic generation of the computation grid •• the absence of numerical errors •• the use of simple solution methods, because there is no need for solution of partial, nonlinear equations •• explicit procedure, that is, nonsteady-state processes are reflected with chronological accuracy •• short turnaround times with low portion of engineering hours. The disadvantages of the method are: •• Absence of an experience base in engine applications. •• Steady-state solutions can be depicted, now, only via transient calculations and in the form of an asymptotic approximation. •• Difficulty of treatment of turbulence. •• High resource requirement. 24.2.2.3 Combination of Calculation Methods of Differing Complexity The increasing complexity of the problems requiring solution results in the combination of calculation methods with one another. This involves both the linking of various 1-D procedures with one another, the coupling of a 3D technique with a 1-D code, and the combination of two 3-D methods. It must be noted here that increasingly coupled problems from different topics are solved with one software package to reduce the number of interfaces and to maintain the efficiency of the computation process. An example for this is the star-ccm+ code [24-19], which was originally only a CFD code. “ccm” is the acronym for “computational continuum mechanics” and refers to the possibilities of solutions for general continuum-mechanical problems. One-dimensional cylinder-charging calculations are, for instance, combined with 3-D CFD simulations to permit registration of gas-dynamic processes beyond the limitations of the 3-D calculation field. For reciprocation, the 3-D calculation supplies significantly better results in terms of the reflection of pressure waves at branch points, and in terms of the pressure loss in the passage of flow through complex components. Simulation of the recycling of exhaust gas into the intake system may be mentioned as an example here. One example of the combination of two 3-D codes is the computation of component temperatures in the engine. It is necessary here to examine heat transfer on both the gas and the waterside. Coupling is accomplished as follows. Heat flows are transferred to the FE method from the CFD code, as the result of a calculation performed in isolation. The FE program, after computation, passes for its part the wall temperatures back to the CFD code. Following the exchange of a number of solutions between the two programs, the heat-flow distributions and wall temperatures will have reached a stable ultimate state.

892 | Internal Combustion Engine Handbook

6606_Book.indb 892

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

24.2.3 Selected Examples of Application 24.2.3.1 Coolant Flow in an Engine’s Water Jacket The limits on performance-enhancing modifications to engines are increasingly defined by component strength. Efficient engine cooling under all engine-operating conditions is, therefore, a precondition for modern engine developments. The available water space is networked on the basis of CAD data, and a fluid mechanical simulation is performed (see Figure 24.25). The result is the ability to perceive zones of extremely low or extremely high velocities. Zones of low velocities are indicative of low convective cooling. Such situations must be avoided in the vicinity of heavily exposed components, such as the exhaust port, for example. High velocities in zones in which the pressure level has already dropped severely and coolant temperatures that are already extremely high result, on the other hand, in the danger of generation of cavitation bubbles as a result of local falls below vapor pressure. An important target factor is the uniform cooling of all cylinders. Variations in the size of the passage apertures in the cylinder-head gasket usually make it possible to achieve this homogeneous distribution. Figure 24.26 shows an example of the distribution of the coefficient of heat transfer around the heavily exposed exhaust ports of a five-valve engine. Systematic variation of the size and location of the passage apertures in the cylinder-head gasket has been applied to achieve the optimization of flow through the zone between the two ports.

necessary for this purpose are either adopted iteratively from an FE simulation or solved directly in the flow solver in the form of a so-called “conjugate heat transfer” problem. The computing grid around the component in which the heat-conduction equation is solved is expanded for this purpose. The convective flow of heat is corrected in both cases by a quantity containing the heat flux resulting from film and bubble boiling. Correction is eliminated as a boiling model is implemented in addition as described in [24-20], for example. 24.2.3.2 Charge Cycle The targets of modern gasoline engine development are •• high torque throughout the entire speed range •• low fuel consumption •• low exhaust emission. Cylinder charging has a significant influence on these targets. The requirements listed above cannot be met with a fixed valve-timing setting. Only variation of the valve timing as a function of speed and load range permits attainment of the target criteria. Many simulations of the influence of timing settings on volumetric efficiency, residual gas content, and charging energy are necessary for this purpose. One possibility of rationally reducing the number of simulations to be performed can be found in “experiment planning” using experiment design. Optimization of target criteria throughout the speed and load ranges can then be achieved with only a few simulation steps [24-21]. 24.2.3.2 Mixture Formation and Energy Conversion

Figure 24.26 Distribution of coefficients of heat transfer in the vicinity of the exhaust ports in a four-cylinder, five-valve engine. See color section page 1099.

Present-day development in computation methods includes solution of the energy equation in the coolant and the coolant’s temperature-dependent physical data, and calculates the heat flow yielded by the component. The component temperatures

24.2.3.2.1 Charge movement in SI and diesel engines The introduction of direct-injection diesel and gasoline engines has heightened the importance of knowledge of the internal engine processes. First to be mentioned is the charge movement, which is optimized in terms of its alignment (swirl or tumble) and its intensity [24-10]. As standard, the inlet ports are, today, assessed on the basis of transient flow simulations, considering each valve stroke curve and using a two-equation turbulence model. In this manner, the charge movement forming in real engine operation can be simulated, including the injection process in some cases. A current trend is to use large eddy simulations (LES) to analyze cyclical fluctuations. Figure 24.27 shows the iterative approximation of the integral values of swirl and flow through the inlet ports of a diesel engine to the target values, which envisaged a simultaneous reduction of swirl intensity and flow resistance in the ports. A subsequent simulation of transient internal cylinder flow additionally makes it possible to evaluate the influence of the integral values of swirl and flow on charge movement at the point of injection or ignition. Such simulations, whether transient or steady state, have become routine. The 3-D flow simulation is dependent on the

Internal Combustion Engine Handbook | 893

6606_Book.indb 893

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

+10

Swirl [%]

0

–10

–20

–30

–40

0

+1

+2

Flow rate

+3

boundary values of a charging calculation, however, because it is not yet possible to observe the entire engine from the intake to the exhaust system within rational turnaround times. The quality of such boundary values is also determined by the quality of the subsequent, locally and chronologically highly resolved, 3-D flow computation. 24.2.3.2.2 Mixture formation in SI engines with direct injection Mixture generation in the engine is subsequently examined on the basis of the simulation of charge movement. In simulation terms, this signifies that it is necessary to calculate a multiphase flow. Numerical treatment of a gas composed of a number of components (for description of an air-fuel mixture in an internal combustion engine, for example) can be performed with one additional equation per component (a type of mass conservation equation), whereas the input for the examination of an additional phase in the same computation zone is considerably more complex. The second phase was long calculated in a LaGrange reference system (co-motive reference system). The term “Euler/LaGrange observation” is used in this context. A motion equation considering mass, energy, and impulse exchange is solved for every particle and/or every droplet. The Euler/LaGrange formulation has the weakness of failing in case of high mass loadings (for example, in the so-called “dense” sprays, such as those generated by highpressure diesel injectors). Euler’s formulation, which also regards the droplets of a spray as a continuum, contrasts with this. An Euler/ Euler formulation necessitates extremely high input and encounters its limitations in extremely dilute sprays, for example. Today, attempts are made to apply both formulations simultaneously in separate zones and, thus, achieve an ability to exploit the advantages of the respective method [24-22]. Simulation requires as its input a so-called “spray model,” which describes the injection process. Injection systems are,

+4

+5

Figure 24.27 Iterative approximation of the integral values of the inlet ports to the target values [24-10].

nowadays, usually first tested under controlled conditions (pressure and temperature) in a pressure chamber. These experimental data are used for the calibration of the simulation. The sprays requiring description can, on the other hand, be measured extremely well and fulfill essential criteria that are a precondition for the Euler/LaGrange observation mode. In addition, other spray models using geometrical parameters of the injectors to model the primary spray decomposition and, therefore, directly returning the input for further mixture formation simulations, are becoming established. Figure 24.28 illustrates, using the chronological development of the average-numbered droplet size and the (also average-numbered) droplet velocity, the quality achievable using a calibrated spray model. Two different radial positions, 15 mm below the injector, were selected for a comparison of the calculation and the phase Doppler anemometer (PDA) measurement. Validation of the calculation methods for specific applications, as performed above for spray calculations, for example, is an important step toward the acceptance of simulation results and toward the integration of calculation methods in the development process. Attention is, for this reason, drawn again and again to the validation experiments in the examples presented below. A spray model calibrated in this way simulates the injection process while the engine is turning. Significant target factors are •• visualization of the fuel-air mixture •• description of the chronological and local spread of the mixture “cloud” •• analysis of the interaction between the injection jet and the charge movement. Figure 24.29 shows an example of the droplets remaining in the combustion chamber at ignition TDC. This distribution shows that large droplets, in particular, accumulate underneath

894 | Internal Combustion Engine Handbook

6606_Book.indb 894

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

Off-axis position

Droplet diameter

Droplet velocity

On-axis position

Simulation PDA measurement

Time

Time

Figure 24.28 Comparative assessment of calculated and measured data, shown in the form of development of droplet size values against time (top series) and velocity (lower series). The left series in the figure applies to the jet axis, while the right series shows comparison with an off-axis position [24-23].

24.2.3.2.3 Mixture formation in SI engines with intake manifold injection The wall film formed in the inlet port and on the valves necessitates complete observation of up to 20 engine cycles. Physical modeling of the droplet and the wall interaction and of the secondary detachment of droplets out of the wall film is extremely difficult. Simulation of the problem necessitates extremely long computing times, with the result that only a few experiments in which the processes of inlet-manifold injection have been simulated with local and chronological resolution are known.

Figure 24.29 View of fuel droplets remaining in the combustion chamber at ignition TDC.

the exhaust valves. This, with the distribution of the evaporated fuel in the combustion chamber, makes it possible to derive valuable information on necessary modifications to the engine concept. Simulation permits the evaluation of the influence of engine speed, load point, combustion chamber geometry, injection characteristics (jet angle), injection timing, and so on. It must also be noted for qualification, however, that only a process averaged for many cycles is depicted in the simulation. Effects caused by the fluctuations in the charge movement or the injection process are not considered here. Thus, the field of engine fine tuning (for the moment) remains closed to simulation.

24.2.3.2.4 Mixture formation in diesel engines with direct injection Diesel injection jets have a significantly higher impulse than direct-injection gasoline engine systems. Injection pressures may rise above 2000 bar. Cavitation phenomena that have effects on the nascent spray can occur within the injector. The occurrence of cavitation bubbles is examined and minimized by simulating the nozzle internal flow. As a result of the high impulse, many extremely small droplets are fed into the combustion chamber in a compact spray-cone zone in an extremely short time. Droplet collisions, thus, occur as a consequence of the high droplet densities. The aerodynamic forces acting on the droplet result, on the other hand, in further disintegration of the droplets. These evaporate within a very short time. An intensive droplet and wall interaction nonetheless occurs as a result of the proximity of the injector and the piston at the point of injection. Whereas the mixture generation and combustion phases in a gasoline engine occur at separate times, in a diesel engine, the liberation of heat influences mixture generation as a result of the short ignition delay.

Internal Combustion Engine Handbook | 895

6606_Book.indb 895

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

The first task in simulation is the calibration of the spray model, a process that can be orientated essentially only around integral measured factors, such as the penetration depth of the jet of liquid. Measured registration of high-pressure diesel injection is extremely difficult, because conventional PDA measuring systems for measuring droplet sizes and velocities are unable to supply any information throughout large zones of the “dense” spray. Quantitative measurements of fuel-vapor concentration using Raman spectroscopy are, for their part, possible only in zones in which no further droplets exist. The consequence is that it is extremely difficult to assess the quality of a spray model intended to determine the spread behavior against time of both the liquid and the gas phases. A further complicating factor is that droplet evaporation and, therefore, also mixture preparation, are decisively determined by the so-called “entrainment” of the hot ambient air drawn into the jet of liquid. It is not possible to resolve such entrainment adequately on the basis of the commercial CFD codes currently used. Spatial resolution of the computer grid with edge lengths of around 0.01 mm are necessary in the vicinity of the injection nozzle in the case of injector hole diameters smaller than 0.2 mm. This results in an enormous number of computer cells and, therefore, in extremely long computing times. This weakness in the simulation of diesel-engine processes, which affects all the downstream and simultaneously occurring phenomena, such as combustion and droplet and wall interaction decisively, is the subject of intensive ongoing research [24-24]. Simulation of mixture generation in diesel engines has, therefore, not yet become a standard application. The existing computation results on diesel combustion were mainly achieved only after the adaptation of model constants. The validity of these model constants for the variation of the injection system or for other engine-relevant parameters is not estimable. Predictive calculations are, therefore, not possible on the basis of the current development status of simulation technology. 24.2.3.3 Simulation of Combustion in an SI Engine Procedures and methods for the experimental study of combustion processes in engines are developing simultaneously, as well as in parallel, with simulation methods. Potentials for the validation and coordination of numerical combustion models are, therefore, emerging. On the other hand, analysis of a combustion simulation permits a deeper insight into the processes occurring in the engine and, thus, augments the knowledge gained from experimental studies. A significant difficulty in the description of combustion phenomena is the fact that the instantaneous conversion rate depends both on local flow state and on the thermodynamic variables of state, that is, pressure, temperature, and composition. The combustion processes occurring in engines can be subdivided into the following categories: •• homogeneously premixed combustion (full-load working point of a gasoline engine) •• diffusion-controlled combustion (full-load working point of a diesel engine with direct injection and without preinjection)

•• partially premixed combustion (part-load working point of a gasoline engine with direct injection) •• combination of diffusion-controlled and partially premixed combustion (part-load working point of a diesel engine with direct injection and preinjection). Physical models have been developed for each of the abovementioned categories of combustion processes. The first models included the so-called “eddy dissipation” models (see [24-28]). These were used, in particular, for the simulation of diesel combustion. It is presupposed here that the chemical processes occur significantly more quickly than the mixture of fuel and air at a microscopic level, for which a turbulent time scale constitutes the limiting factor. These conditions are fulfilled in diffusion flames. This category of models has been and is being developed ever further. Such models are now capable of generating a link between the turbulence-dominated conversion rate and the chemical conversion rate times. The so-called “flame area” models [24-29] were developed specifically for homogeneous combustion processes (complete mixing of air and fuel). The flame front and its rate of propagation are calculated using an advance variable. Expansions of this model with an equation, which describes the mixing state, also permit an application to partially premixed combustion, such as occurs in part-load ranges in gasoline engines featuring direct injection. “Flamelet” models are based on a detailed chemical reaction mechanism. Inclusion of the influence of turbulence is accomplished via the introduction of flame stretching factors, which are determined by the turbulent characteristics data. The solution of complex chemical conversion rates can be accomplished in advance and supplied to the simulation in the form of lookup tables. The representative interactive flamelet model reduces the complexity of including coupling between turbulence and chemistry, as a result of the fact that only between one and a maximum of twenty different turbulent characteristics are included in the calculation of instantaneous conversion rates. The probability-density function (PDF) combustion models, in which the interaction between turbulence and chemistry is described by a multidimensional PDF, constitute a different category of combustion models. The chemical conversion rate is. in many cases, described here using a single-stage global reaction. Detailed information on these combustion models can be found, for example, in [24-30]. Figure 24.30 shows an example of a comparative assessment of a calculated and measured development of combustion with a stratified charge in the AVL-DGI engine, shown in the form of a section through the cylinder and the injector axis. The PDF combustion model was used in this case. Formation of a flame core with a dominant direction of propagation in the zone of stoichiometric mixture composition initially occurs, immediately after the initiation of flame generation. Combustion simulations are currently used for the comprehension of phenomena observed in actual engine operation. Because of the still unsolved problems of modeling, the solution

896 | Internal Combustion Engine Handbook

6606_Book.indb 896

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

Stratified combustion

Measurement/calculation comparison CFD modeling

LIF measurements

24° before TDC

18° before TDC

Concentration distribution, reaction products

20° before TDC

Flame luminosity

Figure 24.30 Combustion with charge stratification—comparative assessment calculation versus measurement [24-31]. See color section page 1099.

of questions concerning ignition stability appears, now, to be premature, and the predictive calculation of absolute pollutant concentrations in new engine concepts unrealistic. A new engine concept that is in development for some years is the homogeneous charge compression ignition engine. In this engine, a homogeneous mixture ignites simultaneously in the entire combustion chamber. The simulation of mixture autoignition is the largest challenge from a modeling point of view [24-32]. 24.2.3.4 Exhaust Gas Treatment The subject of “numerical flow simulation” in the field of exhaust-gas aftertreatment is increasingly gaining in importance as demands for lower exhaust emissions from new vehicles multiply. Originally, the focus was on questions about the operational strength of the catalytic converter in SI engines, in particular. Today, these problems are also defined: •• conversion characteristics •• function test of lambda sensors •• design of diesel particulate filters (DPFs) •• diesel exhaust gas aftertreatment using a selective catalytic reduction (SCR) process. The simulation of the approach flow to a catalytic converter has become a routine application. To be able to classify the

quality of a variant, a uniformity parameter must be defined (see [24-33], for example). The accuracy of the calculation method for a surging flow in the exhaust manifold has been demonstrated in the context of a validation study [24-34]. Figure 24.31 shows a comparative assessment of calculation against measurement for the plot against time of a mass-flow velocity at an engine speed of 2000 1/min. The measurements were made using an optical measuring method in a manifold cross section located immediately downstream from the junction of the exhaust flows from cylinder 3 and cylinder 4. It has, now, been demonstrated that calculations made on the basis of boundary conditions that do not change with time are adequate for the obtainment of trend statements concerning durability and quality even in the context of concepts with closeto-engine catalytic converters. The cylinder under examination in each case is charged with the mass flow that occurs briefly in engine operation during the so-called “preexhaust thrust.” The objective of mathematical optimization is uniform charging of the monolith, to reduce the volume, and, therefore, the space required and costs for the catalytic converter. In addition, a criterion for the durability of the coating of the catalytic converter is derived from the calculated pressure gradients in the first computer cell layer in the monolith. The combination of flow simulation in exhaust systems with a strength calculation has already become established, to the extent that the coefficients of heat transfer from the CFD

Internal Combustion Engine Handbook | 897

6606_Book.indb 897

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 24 Calculation and Simulation

The limitations on correct calculation of heat transfer and the physical period of several seconds requiring simulation do not yet permit the use of 3-D flow simulation for clarification of cold-start performance.

100

Mass-flow velocity

50

0

Measurement Calculation

–50

0

90

180

270

360

450

540

Degree crankshaft angle

Figure 24.31 Comparative assessment of calculated versus measured mass-flow velocity in an exhaust manifold as a function of crank angle.

computation are transferred to the FE program, to calculate there the thermomechanical load exposure of the component. All the statements made above apply primarily to naturally aspired engines. Engines featuring exhaust supercharging require considerably more complex numerical description. Information on conversion performance cannot be derived from the computation results of a 3-D flow simulation. For this purpose, the surface reactions and the heat transport within the monolith are resolved. The decisive boundary conditions for all the calculations of exhaust-gas conversion are statements of the engine’s raw emissions, which must be stated in a chronologically resolved form. This information can, now, be obtained only from measurements. Approach flow onto the lambda probes can also be optimized by computer methods. For this, the correlation between the results about speed, temperature and concentration of a, usually, transient flow calculation and measured lambda sensor signals must be established. The design of the DPF is the focus for the development of the exhaust gas aftertreatment of diesel engines. During filter regeneration, high thermal loads occur in the filter material which may, in extreme cases, cause the destruction of the DPF. The spatial soot distribution in the filter and the physical–chemical process during soot burn-off are the major influences of regeneration in this context, which can be correctly represented only in 3-D simulation processes [24-35]. The discussion about the cleanliness of diesel engines, held mostly in the U.S., has resulted in the development of exhaust gas purification process following the SCR principle in vehicles. This technology was used, for the first time, in the 1980s for purifying the exhaust of power stations. The main focus of the simulation lays on the analysis and optimization of the mixing behavior of the injected urea solution.

24.2.3.5 CFD Applications in the Development of Supercharged Engines Today, there is a trend to supercharged engines even in the SI engine sector, which are fitted with turbocharger, compressor, or a combination of several charging units. This results in a series of additional applications for numerical examination that represent a modification of previously described computing methods. The main new topic in this context is the support of turbocharger design using CFD, including the consideration of the interaction between the charger and the engine.

Bibliography

24-1. Deger, Y. 2008. Die Methode der Finiten Elemente. ExpertVerlag: Renningen. 24-2. Enderich, A. and Handel, R. 1999. “Mündungsschallprognose mit der Finiten Element Methode.” MTZ 60, 1. 24-3. Haibach, E. 1992. Betriebsfeste Bauteile. Springer Verlag: Berlin. 24-4. Vanderplaats, G. N. Numerical Optimization Techniques for Engineering Design—with Applications. McGraw-Hill, Inc. 24-5. Albers, A. et al. 2005. ‘Topology Optimization of Dynamic Loaded Parts Using Multibody Simulation and Durability Analysis.” NAFEMS Seminar Optimization in Structural Mechanics, Wiesbaden, 27/28 April 2005. 24-6. Fischer, P., Nefischer, P., and Kraßnig, G. 2000 “Akustikberechnung lokal gedämpfter Motorstrukturen unter Einbeziehung stochastischer Erregungsprozesse.” Kongress Berechnung und Simulation im Fahrzeugbau, Würzburg, VDI Berichte 1559. 24-7. Bathe, K.-J. 2002. Finite-Elemente-Methoden. Springer Verlag: Berlin. 24-8. Haibach, E. 2002. Betriebsfestigkeit. Springer Verlag: Berlin. 24-9. Merker, G. P. 2006. CTI-Seminar Motorsimulation. 24-10. Nefischer, P. et al. 1999. “Verkürzter Entwicklungsablauf durch Einsatz von CAEMethoden beim neuen Achtzylinder-Dieselmotor von BMW, Teil 2: Thermodynamik und Strömung.” MTZ 60, 11. 24-11. Wiebe, I. 1970. Brennverlauf und Kreisprozess von Verbrennungsmotoren. VEB Verlag Technik: Berlin. 24-12. Barba, C. et al. “Empirisches Modell zur Vorausberechnung des Brennverlaufs bei Common-Rail-Dieselmotoren.” MTZ 60, 4. 24-13. King, A. J. 2002. “A turbocharger unsteady performance model for the GT-Power internal combustion engine simulation.” Dissertation, Purdue Universität. 24-14. Betz, J. et al. 2001. “Gumpoldsberger, T.: Entwicklung von Motorkühlsystemen bei erhöhten Anforderungen an transiente Betriebsbedingungen.” In: 22. Int. Wiener Motorensymposium, Wien. 24-15. Ertl, C., Kronawetter, E., and Stütz, W. 1997. “Simulation des dynamischen Verhaltens von Dieselmotoren mit elektronischem Management.” MTZ 58, 10. 24-16. Zierep, J. 1990. Grundzüge der Strömungslehre, 4. Aufl. G. Braun Verlag. 24-17. Ferziger, J. H. and Peric, M. 1996. Computational Methods for Fluid Dynamics. Springer Verlag. 24-18. Durst, F. (Editor). 2001. “Lattice Boltzmann Methods: Theory and Applications in Fluid Mechanics.” LSTM Erlangen, KONWIHR Workshop, March 2001.

898 | Internal Combustion Engine Handbook

6606_Book.indb 898

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

24.2 Flow Calculation

24-19. http://www.cd-adapco.com/products/STAR-CCM_plus/index.html. 24-20. Kobor, A. 2003. “Entwicklung eines Siedemodells für die Simulation des kühlmittelseitigen Wärmeübergangs bei Verbrennungskraftmaschinen.” Dissertation, TU Graz. 24-21. Knight, J. W. 1989. “Design of Experiments and Simultaneous Engineering.” Society of Automotive Engineers, International Congress, Detroit, February 1989. 24-22. Wan, Y. P. and Peters, N. 1997. “Application of the Cross-Sectional Average Method to Calculations of the Spray Region in a Diesel Engine.” SAE 972866. 24-23. Blümcke, E. 2000. “Strategies for Simulating Transient In-Cylinder Flows Emphasizing Mixture Formation.” In: Proceeding of Workshop on CFD in Automobile Engineering, JSAE 2000 (edited by Kobayashi, T. and Ahmed, S. R.) 24-24. Islam, M. 2002. “Numerical and Experimental Investigations of High-Pressure Diesel Sprays.” Doctoral Thesis, Imperial College London. 24-25. Küntscher, V. (Hrsg.). 1995. Kraftfahrzeugmotoren Auslegung und Konstruktion. Verlag Technik: Berlin. 24-26. Society of Automotive Engineers, Inc. (Hrsg.). 1997. SAE Fatigue Design Handbook (AE-22), 3. Aufl. Society of Automotive Engineers: Warrendale, USA.

24-28. Magnussen, B. and Hjertager, B. 1976. “On mathematical modelling of turbulent combustion.” In: 16th Int. Symp. On Combustion. 24-29. Weller, H. G. et al. 1994. “Prediction of Combustion in Homogeneous-Charge Spark-Ignition Engines.” In: COMODIA 94. 24-30. Peters, N. 1992. “Fifteen Lectures on Laminar and Turbulent Combustion.” ERCOFTAC Summer School Proceedings, Aachen. 24-31. Tatschl, R. et al. 1999. “Mehrdimensionale Simulation der Gemischbildung und Verbrennung in einem Ottomotor mit Benzin-Direkteinspritzung.” In: 7. Tagung “Der Arbeitsprozess des Verbrennungsmotors” (edited by Pischinger, R.). 24-32. Kung, E. H. et al. 2006. “A CFD Investigation of Emissions Formation in HCCI Engines, Including Detailed NOx Chemistry.” 2006 Multidimensional Engine Modeling Users’ Group Meeting, Detroit, MI, 2 April 2006. 24-33. Bressler, H. et al. 1996. “Experimental and Predicitve Investigations of Close Coupled Catalytic Converter with Pulsating Flow.” SAE 960564. 24-34. Bratschitz, B., Blümcke, E., and Fogt, H. 2000. “Numerische und Experimentelle Untersuchungen an der Abgasanlage des Audi 1.8 l 5V-Turbomotors.” In: VDI Berichte Nr. 1559. 24-35. Hinterberger, C., Kaiser, R., and Olesen, M. 2006. “3D-Simulation von Rußbeladung und Dieselpartikelfilter-Regeneration.” MTZ.

24-27. Draper, J. (Hrsg.). 1999. Modern Metal Fatigue Analysis. Safe Technology: Sheffield, UK.

Internal Combustion Engine Handbook | 899

6606_Book.indb 899

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 900

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25 Combustion Diagnostics— Indication and Visualization in Combustion Development 25.1 Discussion Combustion diagnostics is always applied in engine development when unexploited potential compared with thermodynamically possible targets is ascertained during the measurement of consumption, output, and emissions. Given the high targets set for modern engines, thermodynamic combustion analysis using the measurement of cylinder pressure is always a fixed element in the development sequence. Measurements of cylinder pressure are augmented with a whole series of measured variables that define fluid state and component functions. Such “indicated data,” which are generally registered in a form classified by cycle and crank angle, depending on the particular assignment, form the basis for thermodynamic evaluation of combustion and for the optimization of the adjustment parameters for the engine. A comparative assessment of them against the theoretically possible targets available from engine-simulation computations provides guides for an appropriate development activity. Such a development activity is concerned essentially with charging, mixture generation, turbulent charge movement, and, ultimately, flame propagation. Within the scope of the thermodynamic information it supplies, engine indicating provides indications of deficiencies in these processes, but because of the characteristics of the sensor system used, it is unable to provide any information on the local events or on the component-related causes of the deficiencies. The desire then arises in this context to determine through direct insight into engine processes precisely what it is that prevents the achievement of the theoretically possible potentials. This is accomplished using flow and combustion visualization methods.

The potentials for rendering internal engine flow, mixturegeneration, the combustion processes visible, and applying optical measuring methods are as many and diverse as the questions they are intended to answer. Of many methods tested in the laboratory, however, only a very few are actually suitable for the practical use on engines nearing maturity for series production. A number of these procedures exploit flame radiance as a signal source and, therefore, possess the potential of directly indicating the way in which changes in the engine affect local flame propagation processes. Such methods are described here in more detail under the subject of “combustion visualization.” For these task definitions, series-manufactured sensors and measuring instruments are increasingly available, so that these flame diagnostics methods are also used in the routine process of combustion development. A number of other methods that can be used in the subsectors of an operating characteristic field, given corresponding adaptation of the engine, are also examined here, in addition to this central aspect of combustion visualization.

25.2 Indication The term “indicating” is used to designate the measurement and depiction of the plot of cylinder pressure against time or crank-angle position. Pressure indicating occupies a central ranking in combustion development [25-1], [25-2] as a result of the great significance of cylinder pressure for thermodynamic comprehension of

Internal Combustion Engine Handbook | 901

6606_Book.indb 901

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

of the course, duration, and peak intensity of combustion, enabling him to judge the thermodynamic quality of the combustion process. With other measured data, pressure indicating and mass balance also make a considerable contribution to the drafting of an energy balance and loss analysis for the various combustion processes. A comparative assessment of the loss distribution in a direct injection (DI) gasoline engine when the engine is operated at the same working point in various ways is shown in Figure 25.4. The tasks of the presented indicating example mostly concern the work in concept development and the optimization of combustion processes. Beyond this, even for modern engines, it is necessary to indicate in phase of calibration to derive characteristics from the indicating data for a transient assessment of the combustion or the active control of mixture formation and charge change using actuators.

25.2.1 Measuring Systems

The main structure of an indicating measuring system for the measurement of pressure can be described as follows: •• Pressure Transducer: This is installed either directly by a special boring in the combustion chamber or by special adapters in existing borings, such as those for the spark plugs or glow plug. The examples of typical piezoelectric pressure sensors are shown in Figure 25.5. •• Measuring Amplifier: This amplifies the measuring signal received from the pressure transducer to a voltage range of a magnitude suitable for transmission across long cable distances to the data-acquisition unit while at the same time assuring a high signal-to-noise ratio. The length of cable between the pressure sensor and the amplifier is always kept as short as possible, to achieve high signal quality.

20

22

18

20

16

18

14 12 10 8 6

16 14 12 10 8 6

4

4

2 0

Cumulative combustion Combustion curve [%] progress [KJ/°KW]

In addition to the actual signal conditioning, the modern amplifier also assumes the communication with the sensor to provide parameterization information for data recording as required by the particular sensor or measuring task.

Cylinder pressure [bar]

Cylinder pressure [bar]

combustion in engines, and is used for much more than just for the analysis of the plot of pressure [25-3], [25-4], [25-5]. The sensor systems, data-acquisition systems, and result analysis facilities necessary for this purpose have become widely used because of the availability of user-friendly measuring systems. The registration of supplementary characteristic variables, such as measurement of the injection process, ignition current, and thermal variables, proceeds as a natural progression from pressure indicating. Indicating takes a particular rank in the application of engines as here; in direct data exchange with the engine’s control electronics, the engine actuators can be optimized. High-pressure indication in the cylinder is mostly used for combustion analysis. An example of measurement of cylinder pressure in a gasoline engine is shown in Figure 25.1. The pressure signal acquired using the crank angle as the time base is used to generate the pv diagram or, given knowledge of the cylinder filling, the progress of combustion determined using a combustion model. In addition to pressure indicating in the cylinder, low-pressure indicating on the inlet side in the cylinder and on the exhaust side constitutes the precondition for the analysis of charging and for the determination of the masses of the gas available in the cylinder for combustion. Low-pressure indication, with measured data for intake-manifold pressure, cylinder pressure, and the pressure plot determined upstream from the turbine of the exhaust turbocharger, is shown in Figure 25.2. The mass flows calculated from the measured data for lowpressure indicating using a charging model are also plotted. Calculated plots for pressure and mass flow that occur when the chronological outlet-valve lift sequence is changed in a charging simulation to achieve the optimization of volumetric efficiency are shown in addition to these measured data. A comparative analysis of combustion in a number of different combustion processes at a part-load working point is shown in Figure 25.3. The comparison of the plots for pressure and combustion derived for their part using a simple combustion model very quickly provides the observer with an overview

2 0,0

0,2

0,4

0,6

Relative volume

0,8

1,0

–30 –20 –10 0

10 20 30 40 50 60 70 80 90

Crank angle [°CA]

Figure 25.1 Pressure indicating: pv diagram and combustion plot analysis.

902 | Internal Combustion Engine Handbook

6606_Book.indb 902

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.2 Indication

350000

0.2

AÖ basic setting

0.1 Mass flow Off valve

Mass flow On valve

abs. pressure [Pa]

250000

0 AÖ optimized

Cylinder pressure

200000

–0.1

Intake manifold 150000

–0.2

100000

Calculated mass flow [kg/s]

300000

–0.3 Measured data for basic setting

before turbine

Simulation for optimized timing settings

50000 0

90

180

270

360 450 Crank angle [°CA]

–0.4

540

630

720

Figure 25.2 Low-pressure indicating in the inlet, cylinder, and outlet train. Charging calculation for plot of mass flow, simulation of optimized mass flows, and plots of pressure.

40 30 20

1000

10

800

0

600

Combustion progress [KJ/°KW]

400 120

200

homogeneous externally-ignited stratified jet-directed homogeneous auto-ignited HCCI stratified wall-directed

90 60

0

Cumulative combustion curve [kJ/m3 ]

Cylinder pressure [bar]

50

30 0

–30

–15

0

15

30

45

60

Crank angle [°CA]

Figure 25.3 Comparative analysis of combustion in a number of different combustion processes: 2,000 L/min and 2 bar brake mean effective pressure (BMEP).

Internal Combustion Engine Handbook | 903

6606_Book.indb 903

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

100

admitted energy, relative to λ > 1.0 [%]

Exhaust loss 80

60

40

20

Conversion Combustion Wall heat

Leakages Gas change Friction eff. Work

0

homogeneous, I = 1.0

stratified, > 1.0, without EGR

stratified, > 1.0, with EGR

Figure 25.4 DI gasoline engine: loss distribution in stoichiometric and stratified operation: 2,000 L/min and 2 bar BMEP.

For special events (peak pressure or noise analysis, for example), these algorithms may be implemented in the amplifier and can thus be provided without data recording. •• Data-Acquisition System: This is connected to the measuring amplifier and the crank-angle index mark generator, as well as to a PC for the control of the entire system. Its principal function is that of recording the necessary measured data with the required measuring resolution. In addition to this basic function, result computations performed as early as during the measuring sequence, that is, in “real time,” are gaining increasing importance. •• Real-Time Characteristics Calculator: Generates control signals— predominantly for the control of one-cylinder engines—on the basis of measured pressure progressions compared with target values predefined as model characteristics for the specific combustion process. The control signals affect, by the engine control unit (ECU), the actuators influencing the combustion progression (for example, injector, ignition, timing settings, valve stroke, and so on). •• System Operation: This is accomplished by a special PC software that makes it possible to parametrize the entire measuring system and the measurement itself, to obtain characteristics data and calculations, to define algorithms for the determination of characteristics data or the calculation of results from the measured data, and to display theses measured and calculated values. In addition to the actual measurement, this software also controls data storage and the communication to integrable measuring subsystems (such as slower data recording) and to the superordinate automation system. •• Postprocessing: This is used for the presentation and processing of the measured data. More complex computations, comparisons of results, and documentation procedures are performed here using corresponding graphical and computing aids. The scope is adjusted in each case by the user to match the need for the performance of testing.

Application areas: Combustion values, Determination of efficiency, Energy balances, Friction map, Limit value monitoring, Misfire detection, Combustion noise, Knock detection, Vibration excitation, Residual gas determination, EGR adjustment, Automatic characteristic map optimization, Injection process, Mechanical stress Figure 25.5 Design examples for piezoelectric sensors in cylinder pressure measuring.

25.2.2 Quality Criteria

•• Sensors: Sensitivity, signal dynamics, and natural frequency are critical in this context to meet the demands of the particular measuring task. In particular, sensors for practical use on the test bench must be insensitive to varying thermal and mechanical conditions of service and must reliably offer high long-term stability.

•• Measuring Amplifier: In addition to a low-noise amplification, the short-circuit strength and long-term stability are very important for these units. •• Data-Acquisition System: The pure “measured-data acquisition” function occurs here immediately after “real-time” result analysis as early as during the measuring sequence.

904 | Internal Combustion Engine Handbook

6606_Book.indb 904

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.2 Indication

Direct indicated characteristic data for the classification of the pressure plot is determined from the measured data itself by the statement of peak pressure pmax, location of the peak pressure α to pmax pressure gradient dp/dα , location of maximum pressure gradient α to (dp/dα )max, and maximum rate of pressure increase dp/dα 2. Indirect indicated characteristics data are available in this form for indicated average pressures pmi, pmi-HD (high-pressure) and pmi-LW (charge change), friction agent pressure mr, onset of combustion, duration of combustion, and energy conversion points. •• Such real-time analyses are subject to continuous modification depending on needs and appropriateness, and as a function of computing potentials. •• Postprocessing: Measured data and the results from real-time analyses stored in databases are implemented in offline analyses defined by the user. The utilization of open-readable data formats that can be accessed by a whole series of user functions either defined or configured for the particular case is decisive in this context. This makes it possible, for example, from the indicating procedure, model functions, or characteristics variables for quick and effective assessment of combustion on the criteria that relate to the topics as follows: •• maximum component load exposure •• noise generation caused by combustion •• knocking and misfire detection •• engine lean-running limits •• optimum energy conversion.

25.2.3 Indicating—Prospects

Under the clearly formulated preconditions of thermodynamic combustion analysis and as a result of technical advances in the fields of sensor systems and data acquisition, indicating has attained a central ranking in engine development. This established position has resulted in the desire to use cylinderpressure measuring systems not only for the purpose of analysis in the context of combustion development but also for the monitoring of engines during operational service. For this purpose, special modules have been created in the operator software, which reduces the data to be stored to a minimum and, at the same time, document, with sufficient accuracy, combustion phenomena—wherever they occur—and/or trigger safety functions in the monitoring system. In addition, indicating as a functionality becomes more common in the engine operation. The introduction of such a functional diagnosis arrangement will be decisively orientated around the suitability of the sensor systems for mass production and around its direct benefits in use. The use of pressure indicating as a guide factor in combustion processes that are capable of realizing their potential in everyday use by precisely cycle-orientated regulation of combustion has a special technical attraction. Along the suitability of sensor systems for mass production, development in this field will be determined by the availability of actuators relevant to combustion. In addition to this task, engine monitoring and

statements about the unit state must be mentioned as further objectives of such utilization.

25.2.4 Cycle-Precise Signal- and Model-Based Engine Control

Innovative combustion processes use, in addition to cycleprecise effective actuators such as injection and ignition, other fast actuators for the control of the thermodynamic mixture state precisely to the cycle. To be able to profitably use these fast control processes in the engine, the parameters of the actuators must be adjusted to the actual requirements and operating states. The requirements are defined by the driver command and by engine models; the current operating states are taken from engine diagnosis; the thermodynamic mixture state is determined from the characteristics preferably derived from the cylinder pressure signal. In the ideal situation, it is determined, at the beginning of every work cycle and with current driver command, how the actuator parameters must be corrected for the next engine cycle. Such fast engine controls are helpful for the development and the operation of all those combustion processes in which ignition and combustion plot are determined by the charge state and not by the injection progression or externally supplied ignition as is the case in conventional engines. 25.2.4.1 Sensors, Signals, and Signal Processing Conventional pressure sensors are used for the determination of combustion plot characteristics in the practical realization of mostly experimental combustion processes that use autoignition for the combustion of partially or fully homogenized mixtures in, at least, limited load-speed ranges [25-6], [25-7]. From this, the peak intensity of the combustion can be determined in the indicating system or a real-time characteristic calculator. This peak intensity can be regulated for subsequent cycles by the actuators that affect the charge state by changing, for example, valve timing and valve stroke, considering the load demand on the engine. The limit distance to misfiring and overly fast combustion is recorded by the pressure rise speed. The precise control of the filling composition is of crucial importance, in particular, in view of highly dynamic operation as it is required in vehicle applications. As a rule, any sensors that reliably register ignition, peak intensity, and combustion noise in every cycle and cylinder and whose signals can be used in very fast algorithms as measuring variables for combustion regulation are suitable. Pressure sensors are also used in current applications mostly, because the control variables can be directly derived from thermodynamically relevant values. When using alternative sensors, their equivalent functionalities in the generation of the required control variables must be developed and proved. Signal processing and the evaluation algorithms, in particular, must be optimized in view of the limited computing capacity of engine control systems to ensure combustion regulation in all operating states.

Internal Combustion Engine Handbook | 905

6606_Book.indb 905

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

25.3 Visualization 25.3.1 Functions and Discussion

The function of optical diagnosis methods in engine development is to provide insights into those flows, mixture generation, and combustion processes whose behavior cannot be adequately interpreted from the results of conventional indicating methods. In the development of combustion systems, questions concerning the detailed courses of the processes decisive for optimal combustion arise from normal pressure indicating in comparison to thermodynamic calculations and three-dimensional combustion modeling. The main interest focuses here on the following topics in particular: •• the influence of flow within the engine on combustion •• fuel-jet spread and mixture-generation processes •• mixture state: homogeneity–heterogeneity of the cylinder charge and its temperature •• combustion in the case of supplied ignition: flame core formation, flame propagation, burnout of the end gas zones, spontaneous ignition of end gas, combustion anomalies •• combustion in case of autoignition: ignition sites, diffusion combustion, soot generation and burn off, air efficiency, and flame temperature. •• evaluation of operating states causing irregular combustion, recognition of risks, and causes. Such questions are studied at a number of levels: •• In the Context of Fundamental Research: Research and measuring methods in which the automobile engine aspect determines the general orientation but in which the test apparatus may diverge extremely greatly from actual engine operation are used for basic functional analyses. •• Component Tests: In this case, standardized test procedures are applied for the comparative evaluation of component properties. •• In research engines with a combustion chamber made optically accessible by special components and sight glasses. Engine operation is subject to compromises resulting from the use of sight glasses and auxiliary components. •• Actual Engine Operation: Optical sensor systems and measuring methods are orientated here specifically around the needs of engine operation uninfluenced by the measuring activity. Flame-observation methods under such real engine conditions are the central subject of this chapter. Advances in design and application of radiation sensor systems increasingly open the possibility of a touch-less measurement of the temperatures of heavily loaded combustion-chamber components.

25.3.2 Visualization Methods for Real Engine Operation

In the context of fundamental research and in the arrangement of component tests, the study subject is always adapted to the specific question and to the needs of the particular test technology, whereas the emphasis is on undisrupted engine

functioning in the case of visualization of actual engine operation. The limitations that result for the adaptation of the test technology to engine operation are, therefore, correspondingly restrictive in this field. The following questions are presented: What results can be obtained using modern visualization methods under the restrictions imposed by real engine operation? 25.3.2.1 Radiant Properties of Gas, Gasoline, and Diesel Flames The visible radiance that occurs in the combustion of AC flames is the result of the chemiluminescence of the molecules formed during combustion and of the thermal radiation of soot. The spectral composition of the dominant contributory radiant ranges is shown in the emission spectra in Figure 25.6; flame photographs are shown as an example in Figure 25.7. The components fuel (CH), intermediate products (OH and CO), and radiant CO2, H2O, and O2 portions, along with other molecules and radicals, are generally found in the oxidation of CH molecules. The thermal radiation of soot particles also contributes to the flame’s intrinsic luminosity in cases in which low-oxygen combustion results in the generation of soot. This particulate radiation can contribute to flame luminosity with a significant intensity if local rich combustion occurs in a stratified charge; in diffusion combustion in diesel engines, flame radiance is massively dominated by such thermal soot radiance. 25.3.2.2 Flame Spectroscopy The spectral intensity distribution of the emission spectra contains information on the concentration of the radiant molecules and their initial components on their temperature and on the temperature of the radiant soot particles. Since the preconditions for thermal equilibrium are not present in many cases, the context of the transient processes occurring in combustion in engines, given the lifetimes of the radiant molecules, and the fact that severe local gradients occur in the measuring volumes registered, exploitation of spectral radiance properties for quantitative measurements is limited to special cases (lambda [25-9], OH temperature [25-9], and soot temperature [25-10]). Standardized measuring methods are, therefore, restricted to only a few applications. In diffusion flames, for example, they exploit thermal soot radiation, either in spatially integral form [25-11] or by image evaluation methods applied to flame photographs for the determination of soot concentration and the temperature of the diffusion flame (Figure 25.8, [25-12]). Development-relevant results for the clarification of improved soot burn off with divided injection are shown in this context in Figure 25.9 [25-13]. 25.3.2.3 Flame Propagation in Premixed Charges With Supplied Ignition Following ignition and the formation of the flame core, the flame should propagate in such a way that chronologically optimum and locally uniform and complete combustion of the charge occurs. Flame propagation is driven by the advance of the flame front under the influence of turbulent charge motion. A flame image and the flame-front structure generated by charge turbulence are shown in Figure 25.7.

906 | Internal Combustion Engine Handbook

6606_Book.indb 906

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.3 Visualization

Premixed propane flame Crankshaft angle resolution: 1.0 °CA, 750 1/min, l: 1.0, Ignition timing: 20° before TDC

Output of OMA x 1,5

Early Stage of Combustion Crank Angle:

OH

20°A

CO–O

7°A

Output of OMA

Crankshaft angle resolution: 1.0 °CA, 1000 1/min, l: 2.0, Epsilon: 16.0, Injection: 9° before TDC

Crank Angle:

16°A

250 300 350 Crank Angle: 16°A

Diesel flame

Late Stage of Combustion

CH

400

45°A 500

450

250

300 350 Crank Angle: 21°A

400

450

500

6°A

OH

51°A 250

300

350 400 450 Wavelength nm

500

250

300

350 400 450 Wavelength nm

500

Figure 25.6 Plot of the spectral emission from a premixed propane flame and a diesel flame (Kuwahara, Ando [25-8]).

An intensive interaction between the flow and the flame advance can occur across the broad phases of combustion, because the processes of turbulent flame propagation and directional flow within the engine may occur at comparable rates. This is the field for the potential optimization of combustion within the engine. Gasoline flame: central spark plug location, image through the piston of a "glass engine"

Diesel flame: View through a combustion chamber endoscope

14 ° CA after BDC

Base_1602Fhi300_n0060

5.0 ° CA after BDC

14 ° CA after BDC Injection start: 1 °CA after BDC

Figure 25.7 Flame photograph.

The following development-relevant questions primarily arise in this context: •• How far is flame propagation as achieved now removed from the ideal situation described above? •• What changes can be made to improve flame propagation? The presence of a homogeneously premixed cylinder charge is assumed. This prerequisite is mostly met in spark ignition (SI) engines with intake-manifold injection in normal operation but must be ensured by the design of the mixture formation organs. In gasoline engines with DI, the assurance of this

prerequisite is a central component of the development process and an important application of optical flame diagnostics [25-14]. 25.3.2.3.1 Irregular combustion Because of increasing the power density in supercharged engines, the corresponding increased thermal load can cause uncontrolled ignition events. Such “irregular combustions” represent a functional risk for the components of the combustion chamber. For this reason, recognizing risks and causes gains, increasing importance in combustion development. Because those events occur sporadically and are bound to the real engine operation, their analysis during normal engine testing poses very high requirements on combustion measuring 25.3.2.4 Flame Propagation in Diffusion Combustion in a Diesel Engine Ignition and combustion are determined here by gas state, and decisively by the characteristics of the injection process. The goal is an optimal air utilization and efficient soot burn off after the end of injection. The spread of the diffusion flame (Figure 25.7) is determined by the injection and turbulent diffusion of the “cloud” of fuel vapor and its interaction with internal flow. Flame radiance immediately after autoignition is driven by the chemiluminescence of the reactants, but then very quickly becomes dominated by the thermal luminescence of the soot particles. Primary development questions are as follows: •• How can air efficiency be raised by modifying the design of the injection system and charge movement? •• What changes to the injection system and charge movement would affect soot generation and burn off and reduce soot emissions? •• How can excessively high flame-temperature peaks be avoided?

Internal Combustion Engine Handbook | 907

6606_Book.indb 907

1/19/16 8:56 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

25.3.3 Visualization of Combustion in Real Engine Operation by the Flame’s Intrinsic Luminescence

The engine is operated for this purpose on an engine test bench or in the vehicle in a dynamometer test. Analyses that augment standard engine indicating with information on the local and chronological progress of flow, mixture generation, and combustion, permitting the derivation of guidelines for systematic engine improvement, are of interest here for engine development purposes. The flame’s intrinsic radiance is used primarily here as the study subject, since it is accessible with the lowest technical input, thus very largely also avoiding the disruption of combustion by the observation process. In an ideal case, the chronologically resolved, three-dimensional propagation of the flame in the combustion chamber should be registered by a visualization method to permit the ascertainment of deviations in combustion progress from the theoretical optimum. However, for technical reasons, this rigorous requirement can be achieved only with great restrictions. 25.3.3.1 Technical Realization: Flame Propagation Imaging flame-photographic and measuring methods that record the radiant intensity of the flame and derive from it the information on the local and chronological progress of

25.3.3.1.1 Flame photography The advantage of direct flame observation by an endoscope is that picture capture using a camera immediately provides an image that can be interpreted. With the appropriate selection of window position and angle of view into the combustion chamber, flame images that show extremely clearly the spread and turbulent structure of the diffusion flame, for example, in diesel engines, can be obtained (Figure 25.7). The position of the flame in the combustion chamber can then be determined by image superimposition with reference images in a subsequent image-processing step [25-12]. In flame photography, a luminescent, continuously changing cloud of gas with a heavily textured surface is depicted. The surface of the flame itself is registered in this process depending on the flame’s transparency radiation from its interior. Given these properties in the image subject, image quality is influenced by the following factors: •• Motion-Induced Blurring: This can be minimized using correspondingly short camera shutter times.

120 100

•• Variable Subject Distance: A well focused image is achieved because of the endoscope’s high f-stop number and short focal length, provided the flame surface is sufficiently distant from the image-forming lens. The dimensions of the subject are distorted as a result of the short focal length, the expansive flame cloud, and the variable subject distance, however.

80 60 40 20 0

Hole dia - µm Hole dia-mm

NOx - ppm

FSN

NOx ppm

Smoke FSN

257 269

1.83 1.20

0.227 0.195 Area with temperature above 2400 K [mm2]

combustion are available for the technical achievement of flame observation. Optical access in the case of flame photography is achieved by combustion-chamber windows. The combustion chamber is then depicted in the focal plane of a suitable camera by endoscopes. The system element relevant to engine operation in this context is the combustion-chamber window. This is either positioned on the combustion chamber by special borings [25-15] or used in specially adapted engine components [25-16].

Øhole = 0,227 mm Øhole = 0,195 mm

700 600 500 400 300 200

•• Optically Thin (Transparent) Flames, Premixed Flames in Gasoline and Gas Engines: In these cases, the highly textured flame surface and the diffuse underlying layers of combustion charge are imaged. The inner zone and the far boundary zone of the luminescent flame cloud dominate as soon as the flame touches the viewing window [25-16]. Criteria for the evaluation imaging systems include the following:

100 0 –5

•• Optically Dense Diffusion Flames (Diesel): Here, only thin surface layers contribute to the generation of the image. Because of the high level of absorption in the diffusion flame, the flame image is of low informational value only if the diffusion flame touches the viewing window.

0

5

10

15

20

25

30

Crank angle [°ATDC]

Figure 25.8 Combustion analysis, diesel engine: narrowing of the injection hole bore, effects on NOx, and soot emission. Also temperature-zone progress of diesel flames (Larson [25-12]). Surface temperature analysis shows the mechanism of action for enhanced soot burnout.

•• Combustion-Chamber Windows: Their size must not disrupt the operation of the engine. •• Image Transmission: Lens angle of acceptance, f-stop number, and spectral transmission range. •• Camera Characteristics: Local resolution (number of pixels), sensitivity (light yield), spectral sensitivity, signal dynamics, imaging frequency, exposure time, and shutter-induced attenuation.

908 | Internal Combustion Engine Handbook

6606_Book.indb 908

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Spec. soot rate 1/deg CA

Soot mass mg

Needle lift mm

25.3 Visualization

0.3

0 0.4 Speed = 1360 min–1 BMEP = 4.9 bar

0.2 0 0.4 0.2 0 –0.2

Normal injection Split injection (26 mm3/str.)

–0.4

Flame temp.

2400 2200 2000 1800 –10

0

10

20

30

40

50

60

70

Soot-g/kWh

0.14 Speed = 1360 min–1 BMEP = 4.9 bar

0.10 0.06 0.02

BSFC-g/kWh

250 Normal injection Split injection 26 mm3/str.

240 230 220 210 3

4

5

6

7

8

NOx -g/kWh

25.3.3.1.2 Flame radiance The advantage of high local resolution achievable in flame photography because of endoscope and camera characteristics is not exploitable for all the problems relevant to engines and, in many cases, is actually problematical because of the high volumes of data involved. Fixation on a single-window site for the observation of important propagation processes can also be an excessive restriction. A remedy is provided here by observation methods in which flame propagation is reconstructed from the measurement of flame radiance in the limited areas of the combustion-chamber volume. This can be accomplished in a relatively simple arrangement by means of “photoelectric barriers,” which, installed in the shell of a spark plug, detect the propagation of the flame core [25-17] or, in the form of a multichannel arrangement distributed throughout the combustion chamber, track flame advance [25-18]. Combinations of small front-lens elements or “micro-optic” components and individual fiber-optic conductors produce many potentials for the design of directional and locally

9

10

11

Figure 25.9 Analysis of soot radiance in a diesel engine. Greater soot burnout is achieved because of split injection. The result reduced the NOx soot tradeoff and has no effect on consumption [25-13].

delineated registration of flame radiance. The arrangement shown as an example in Figure 25.10, for instance, registers visible radiance from five narrowly defined conical zones in the combustion chamber. A measuring signal typical of a single-channel system is also shown with the pressure curve for combustion in Figure 25.10. In addition to unequivocal localization, high signal quality (sensitivity, signal-to-noise ratio, and signal dynamics), intensity calibration of all measuring channels, and, particularly for the evaluation of knocking combustion, high chronological resolution matching pressurewave spread are definitive for the utilization of such signals. A sensor arrangement installed in the shell of a spark plug is shown in Figure 25.11. As the installation diagram illustrates, growth in the flame core is tracked. The photographic image (obtained in a glass engine) and the signal are shown for comparison purposes. Since the spark-plug sensor records radiant intensities, intensity calibration of all measuring channels is always an element in the measuring procedure. The user must select intensity thresholds for the evaluation

Internal Combustion Engine Handbook | 909

6606_Book.indb 909

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

Viewing cone

80

160

40

80

deg °CACA

0

Combustion chamber

-10

0

10

Rel. intensity

Lens

Pressure - bar[bar] Pressure

Fiber-optics conductor

0 20

30

40

Figure 25.10 Registration of flame radiance using micro-optic components, radiant intensity of the flame in the registration zone of a viewing cone, and, for comparison, the pressure signal for combustion.

of the results obtained. Reliability of results is achieved here by the comparison of graduated threshold values. 25.3.3.1.3 Flame tomography Maximum benefit is obtained from the multichannel measurement of flame radiance, if the geometrical arrangement of the combustion-chamber sections observed can be used for tomographic image reconstruction. This can be achieved using a sensor arrangement that superimposes an optical observation grid on the cross section of the combustion chamber [25-19]. The arrangement of a number of observation cones is shown schematically in Figure 25.12. Local flame intensity can be reconstructed from the measuring signals from all the channels of the observation grid and the knowledge of the individual registration zones. The examples of this from a DI gasoline engine are shown in Figure 25.13. At low swirl, it becomes apparent that intensive luminescent diffusion

combustion occurs in the recessed zone of the piston, resulting in correspondingly high soot emissions. With swirling flow, the central recessed zone is obviously better mixed with air, with the result that no excessive diffusion combustion or soot emission occurs. Local resolution in flame tomography is determined by grid density and is approximately 3–5 mm in technically practicable systems. This falls short by several orders of magnitude of the high local resolutions achievable using imaging cameras. Because of the distribution of the sensors around the entire circumference of the combustion chamber, however, the combustion-chamber cross section is registered uniformly and with no image-field distortion, with the result that flame propagation can be plotted clearly and with a homogeneously distributed resolution throughout the cross section. Along the depiction of intensity, flame propagation is also reproduced extremely conveniently in the form of progressive 14 1/(ms) 7

7

7 70 Int.

1/(ms) 14

7 Flame arrival time, evaluated for intensity threshold

5

Figure 25.11 Flame core generation and observation using a spark-plug sensor. The result illustrates the symmetry or asymmetry of the flame core and its predominant direction of propagation. See color section page 1100.

910 | Internal Combustion Engine Handbook

6606_Book.indb 910

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.3 Visualization

flame-front contours after image reconstruction and with the setting of a threshold value. The propagation forms typical of certain flow conditions in modern four-valve engines are shown in Figure 25.14. Knowledge of these and ascertainment of their dependence on operating conditions and/or the design of engine components can provide decisive indications for improvements. Examples of applications derived from development practice on extremely diverse engines can be found in various publications [25-20], [25-21]. The central advantage of measurement of flame radiance and reconstruction of flame propagation can be found in the arrangement of the sensor system specifically for each particular engine and flexibility in signal recording, the chronological resolution of which can be precisely adjusted to the requirements of the specific measuring task. This is illustrated in Figure 25.15, using the example of a knocking spot distribution. The inadequate progress of the flame into the left side of the combustion chamber results here to a greater extent in end gas autoignition. Flame propagation, its one-sided retardation on the wall of the combustion chamber, and the resultant autoignition can be routinely registered by flame tomography.

Figure 25.12 Arrangement of a micro-optic sensor system in the cylinder-head gasket for tomographic flame reconstruction. See color section page 1100.

without swirl

Rel. intensity

900 100

with swirl

10 0

–20

–10

TDC

10

20

30

40

50

°CA

Figure 25.13 DI gasoline engine: flame tomography shows the local position of bright, soot-producing diffusion flames. Swirling flow produces a significant improvement. See color section page 1100.

Discharge side

Intake side

Lateral extension

Squish area influence

Figure 25.14 Flame propagation: tomography with sensor system installed in the cylinder-head gasket. The isolines indicate the progress of the flame front against time. The influence of internal flow on flame propagation is clearly visible. See color section page 1101.

Internal Combustion Engine Handbook | 911

6606_Book.indb 911

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

5 Out

Out Degree CA

In

In -4

35 Out

Out %

In

In

•• spectral sensitivity, signal sensitivity, and signal-to-noise ratio •• signal dynamics, particularly at high signal amplitudes

0

•• localization of the individual channels and calibration procedures for multichannel systems

Figure 25.15 Flame tomography supplies comprehensive documentation of flame propagation and knocking spot distribution. See color section page 1101.

1 bar –2

20

25

10

Int. var.

Channel nr.

•• signal evaluation and data reduction.

Swirl: low

4000 1/min 15

1 22,4

25.3.3.1.4 Spark-plug sensor systems The results obtained in observation of flames using tomographic sensors and model concepts of flame propagation processes also make it possible to study certain problems using simplified measuring methods. Because of the lower level of local resolution, the signal-acquisition systems of such methods must be precisely tailored to the signal patterns of individual combustion phenomenon. The spark-plug sensors featuring built-in fiber optics already examined above have proved their capabilities for the study of flame core formation. There are spark-plug sensors specifically for the observation of autoignition in the knocking operation of the engine that registers the engine’s compression volume using a rotating fan-type sensor. Signal evaluation is matched to the specific propagation characteristics of the pressure and density wave occurring upon autoignition of end gas [25-22]. An overview of the sensor principle, signal pattern, localization, and result statistics available to the development engineer as an aid to decision making in the context of component modification is shown in Figure 25.16. The criteria for the utilization of microsensor systems for flame radiance are as follows:

medium 35 5

°CA ATD

25

1 40

high 20 % Frequency 0

Figure 25.16a Determination of knocking spots by a fan-type sensor—presentation of results: single-cycle and derived knocking spot statistics. See color section page 1102.

912 | Internal Combustion Engine Handbook

6606_Book.indb 912

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.3 Visualization

500

Diffusion flame

Off –60 °CA 100

Injector FSN = 0.0: Signal pattern of an ideal premixed combustion

Flame brightness [rel. unit]

On

FSN = 0.8: Smoke measurement indicates that improvement is necessary; the flame pattern hints where to start

25.3.3.1.5 Flame pattern evaluation for optimization of mixture formation Sensors registering the major areas of the combustion chamber in separate measuring channels at high temporal resolution are also well suited to indicate the local difference in the flame luminosity. For example, this is beneficial to the evaluation mixture formation. In the development and application of gasoline engines with DI in particular, this enables a cylinderand cycle-precise analysis of the soot-generating diffusion portions of a gasoline flame (Figure 25.16b). The local and temporal analysis of the signal patterns supported so a fast and systematic optimization of the mixture formation for minimal soot generation [25-14]. 25.3.3.1.6 Ignition site registration in irregular combustion An irregular combustion occurs spontaneously in extreme cases and must, therefore, be registered with a measuring system that is event-triggered to record sensor signals occurring before and after the ignition event. The sensor system itself must be capable to capture the ignition site with reliable locus resolution. For this objective, spark-plug sensors registering the flam radiation from the combustion chamber in up to eighty separator channels have proved their worth, which are also capable to determine the ignition sites outside of the electrode space of the spark plug. Sample results are shown in Figure 25.17. Transient recorders record the signal and, triggered by an event trigger, retain the relevant combustion cycles in circular buffer mode [25-23].

Premix flame

0

Figure 25.16b Determination of flame pattern by a fan-type sensor. The polar diagram shows when and in which combustion-chamber sector flam anomalies occur. The flame brightness is evaluated from the unfiltered intensity signal. See color section page 1102.

Scaling of the flame brightness

Visio spark plug sensor channel directions

Visio sensor channel nr

–60 °CA 100

EX

EX

IN

IN

No Flame

irregular ignition location Flame

Deg CA Figure 25.17 Sensor system and sample signal for determining the ignition sites during irregular combustion. See color section page 1102.

Local peak temperature

736°C

746°C

25.3.3.1.7 Touch-less temperature measuring in the combustion chamber Imaging Combustion-chamber optics used for flame registration can also be used for the measurement of the radiation temperature, if the signal registration has been adjusted and infrared (IR)-sensitive signal converters are used [25-24]. An example of temperature measuring at spark plugs in the full-load operation of a freely drawing gasoline engine is shown in Figure 25.18a. The influence of the screw-in depth can be recognized from the intrinsic radiance. Using suitable calibration processes and procedures, it is possible to achieve a measuring accuracy of ±10 K.

5 mm

°C

9 mm

°C

Figure 25.18a Thermal image of a spark plug in full-load operation The screw-in depth of the spark plug exhibits measurable temperature differences. See color section page 1103.

Internal Combustion Engine Handbook | 913

6606_Book.indb 913

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Crank angle resolved dual spot measurement

580 560 540 520

100 °C

Max. valve temperature [°C][degC] Maximal ExValve Temperature

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

500

valve stem

480 460 440 420

valve seat 100 200 300 400 500 600 700 800 900 1000 Cycle [-]

cycle to cycle temperature of two exhaust valves, stationary engine operation Figure 25.18b Measuring the radiation temperature at exhaust valves. Spark-plug sensor system with multichannel optics, use in normal engine operation. See color section page 1103.

25.3.3.1.8 Constant temperature measurement Spark-plug sensors and IR-sensitive radiation converters form the foundation for a crankshaft-resolved temperature measurement in the combustion chamber. The results of the measurement at valves and comparisons of maximum temperatures are shown in Figure 25.18b. With corresponding calibration procedures, it is here possible to achieve a measuring accuracy of ±10 K as well. 25.3.3.2 Measuring Instruments and Systems The listed procedures for evaluating flame characteristics are increasingly used in the routine testing and operation of combustion process development and engine calibration. This development is driven by standardized methods and seriesmanufactured sensors and measuring instruments by some manufacturers and the requirements of engine developers. An indication for the migration of the previously researchoriented measuring methods to routing practice in engine testing is the increasing number of specialist publications discussing these topics (see the bibliography to this chapter).

has become an indispensable aid in the development of DI systems, for instance. The precondition for the use of subject illumination is always the optical access necessary for this purpose, which must be available simultaneously with the optical access for subject imaging. In special cases, a single window into the combustion chamber can be used for both the tasks; the necessary flexibility and quality can in many cases be achieved only through separate access ways, however. In an extreme case, the engine is equipped with large windows for this purpose, or glass components are substituted for the piston and the cylinder sleeve (Figure 25.19, [25-25], [25-26]). Engines like these can be operated under conditions closely approximating reality but within a restricted load and speed range, and provide the precondition for the application of appropriate visualization methods.

25.3.4 Visualization of Illuminated Processes

There is a whole series of questions in combustion development that can be studied only by the active illumination of the processes involved. In the simplest case, the subject—a jet of fuel, for example—is diffusely illuminated and depicted using a suitable camera. Illumination and imaging technology make it possible to exploit the properties of the subject by determining which velocity fields, fuel distribution, or distribution of specific combustion products can be rendered visible. Visualization of the fuel using the laser-induced fluorescence (LIF) method

Figure 25.19 Maximum optical access into the combustion chamber by the use of a glass cylinder and glass windows in the piston. The engine is operated for short periods with ignition for measuring purposes.

914 | Internal Combustion Engine Handbook

6606_Book.indb 914

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

25.3 Visualization

25.3.4.1 Visualization of Mixture Distribution The development of DI gasoline engines, in particular, has accelerated the need for practicable methods for the observation of charge stratification. LIF methods, in which the molecules of fuel or of a tracer substance are made to fluoresce in a planar section of laser light, have proved their value for this. The resultant fluorescence is recorded using suitable cameras and, thus, initially supplies a qualitative image of mixture distribution. Very careful performance of these tests makes it possible to obtain a quantitative assessment of fuel concentration by applying calibration procedures, and the evaluation of pressure-dependent and temperature-dependent fluorescence yields to the intensity distribution of such images [25-27]. Significantly less effort and only relatively simple image analysis methods are needed for the generation from groups of individual images of a probability analysis of a particular distribution recurring reliably at a certain crank angle in every cycle. These distribution statistics on the presence of mixture clouds meet the requirement for practically relevant and informationally useful visualization methods; examples are shown in Figure 25.20. 25.3.4.2 Visualization of Velocity Fields 25.3.4.2.1 Particle image velocimetry Here, the movement of scattered particles either naturally present in the flow field, such as droplets of fuel, or added to the flow as tracer particles is captured by double illumination or double exposure. The velocity field in the flow field observed is determined from the evaluation of particle travel in the interval between making the two images. 25.3.4.2.2 Doppler global velocimetry This is a method equivalent to particle image velocimetry for area visualization of velocity fields. Here, too, tracer particles for scattering of the light beamed in are added to the flow field. The Doppler shift generated in the particles during the

-68° CA

-68° CA

scattering process is evaluated as the velocity signal by an extremely narrow-band illumination and correspondingly modified spectral filters [25-28].

25.3.5 Visualization: The Future

Methods for the visualization of internal engine processes have long been in use in basic research; the results are also used in the verification of computing processes for three-dimensional simulation of flow and of combustion in engines. Visualization methods have come into more widespread use in engine development only since the development needs of modern combustion processes necessitated comprehensive detail understanding and optimization of processes within engines. Unlike indicating methods in which the measurement of cylinder pressure occupies a central position as a result of the thermodynamic significance of the pressure signal, capture or registration of flame propagation has achieved similar significance in the field of visualization. Here, too, theoretical understanding of optimum combustion provides clear guidelines for flame propagation. Its metrological observation can then supply the development process with the necessary systems for component optimization. Unlike the situation in indicating methods, however, visualization methods are only at the very start of their potential applications in engine development. The benefits demonstrated up to now must be consolidated by flexibility in sensor systems and precision in results. The large and diverse range of methods necessitates standardization of central measuring functions and the potential for uncomplicated integration of sensor and measurement technology innovations into open measuring systems. Since the evaluation of combustion is performed ultimately on thermal dynamic criteria and on the basis of the results of emissions measurements, linking the results obtained from visualization, indicating, and waste-gas measurements is a central requirement for the development for combustion diagnosis systems [25-29].

-38° CA, 2000 1/min, 3 bar p mi

-38° CA, 3000 1/min, 1 bar p mi

Figure 25.20 Gasoline DI: fuel distribution in the injection process and after deflection from the pistons. The stability of distribution states is determined from individual images using image statistics. Green/red: fuel vapor with increasing stability. Blue/white: fuel droplet with increasing stability. See color section page 1103.

Internal Combustion Engine Handbook | 915

6606_Book.indb 915

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 25 Combustion Diagnostics—Indication and Visualization in Combustion Development

Bibliography

25-17. Geiser, F., Wytrykus, F., and Spicher, U. 1998. “Combustion Control with the Optical Fibre Fitted Production Spark Plug.” SAE Technical Paper 980139.

25-2. Heywood, J. B. 1988. Internal combustion engine fundamentals. McGraw-Hill: Columbus, OH USA.

25-18. Spicher, U., Schmitz, G., and Kollmeier, H. P. 1988. “Application of a New Optical Fiber Technique for Flame Propagation Diagnostics in IC Engines.” SAE Technical Paper 881647.

25-1. Pischinger, R., Kraßnig, G., Taucar, G., and Sams, Th. 1989. Thermodynamik der Verbrennungskraftmaschine. Springer: New York, USA.

25-3. Witt, A., Siersch, W., and Schwarz, Ch. Oktober 18–19, 1999. “Weiterentwicklung der Druckverlaufsanalyse für moderne Ottomotoren,” in Der Arbeitsprozess des Verbrennungsmotors, 7. Tagung, Graz. 25-4. Binder, S., Zipp, W. Mai 18–19, 2006. “Charakterisierung von Verbrennungsgeräuschen und Strategie zur Optimierung mittels Indizierung,” 7. Int. Symp. für Verbrennungsdiagnostik, AVL Deutschland. 25-5. Leifert, Th., Fairbrother, R., and Moreno Nevado, F. Juni 10–11, 2008. “Durch Messung unterstützte thermodynamische Analyse von Zylinder internen Vorgängen unter transienten Bedingungen,” 8. Int. Symp. für Verbrennungsdiagnostik, AVL Deutschland. 25-6. Agrell, F., Angstrom, H. E., Eriksson, B., Wikander, J., and Linderyd, J. 2003. “Transient Control of HCCI Through Combined Intake and Exhaust Valve Actuation.” SAE Technical Paper 2003-01-3172. 25-7. Fürhapter, A., Piock, W. F., and Fraidl, G. K. Feb. 2004. “Homogene Selbstzündung: Die praktische Umsetzung am transienten Vollmotor,” MTZ, 65(2): pp. 94–101. 25-8. Kuwahara, K., Ando, H. Mai 18–19, 2000. “Time series spectroscopic analysis of the combustion process in a gasoline direct injection engine,” 4. Int. Symp. für Verbrennungsdiagnostik, Baden-Baden, AVL Deutschland. 25-9. Hirsch, A., Philipp, H., Winklhofer, E., and Jaeger, H. July16–19, 1996. “Optical Temperature Measurements in Spark Ignition Engines,” 28th EGAS Conf., Graz. 25-10. Gstrein, W. 1987. “Ein Beitrag zur spektroskopischen Flammentemperaturmessung bei Dieselmotoren,” Dissertation, Techn. Univ. Graz. 25-11. Hötger, M. 1995. “Einsatzgebiete der Integralen LichtleitMesstechnik,” Motortechnische Zeitschrift, 56(5):pp. 278–280. 25-12. Larsson, A. 1999. Optical Studies in a DI Diesel Engine. SAE Technical Paper 1999-01-3650. 25-13. Chmela, F., Riediger, H. Sept. 13–15, 2000. “Analysis methods for the effects of injection rate control in direct injection diesel engines,” Thermofluidynamic processes in diesel engines, CMT Valencia. 25-14. Winklhofer, E., Nohira, H., Beidl, Ch., Hirsch, A., and Piock, W. F. “Combustion Quality Assessment for New Generation Gasoline Engines,” JSAE 20045451. AU: Kindly provide the publication year for 25-14 25-15. Winklhofer, E. Juli 2001. “Optical Access and Diagnostic Techniques for Internal Combustion Engine Development,” J. Electron. Imaging, 10(3). 25-16. Wytrykus, F., Duesterwald, R. 2001. “Improving Combustion Process by Using a High Speed UV-Sensitive Camera.” SAE Technical Paper 2001-01-0917.

25-19. Philipp, H., Plimon, A., Fernitz, G., Hirsch, A., Fraidl, G., and Winklhofer, E. 1995. “A Tomographic Camera System for Combustion Diagnostics in SI Engines.” SAE Technical Paper 950681. 25-20. Liebl, J., Poggel, J., Klüting, M., and Missy, S. June 2001. “Der neue BMW Vierzylinder-Ottomotor mit Valvetronic,” MTZ Motortechnische Zeitschrift, 62(6):pp. 450–463. 25-21. Grebe, U. D., Kapus, P., and Poetscher, P. Nov. 16–18, 1999. “The Three Cylinder Ecotec Compact Engine from Opel with Port Deactivation—A Contribution to Reduce the Fleet Average Fuel Consumption,” 18th Int. VDI/ VW Conf. Braunschweig. 25-22. Philipp, H., Hirsch, A., Baumgartner, M., Fernitz, G., Beidl, Ch., and Piock, W. Winklhofer, E. 2001. “Localisation of Knock Events in Direct Injection Gasoline Engines.” SAE Technical Paper 2001-01-1199. 25-23. Kapus, P., Sauerwein, U., Moik, J., and Winklhofer, E. Sept. 2005. “Ottomotoren im Hochlasttest,” 10. Tagung “Der Arbeitsprozess des Verbrennungsmotors,” Institut für Verbrennungskraftmaschinen und Thermodynamik, TU Graz. 25-24. Hirsch, A., Kapus, P., Philipp, H., and Winklhofer, E. Sept. 2009. “Risikoanalyse und Entwicklungstechniken für DI Otto Brennverfahren hoher Leistungsdichte,” 12. Tagung “Der Arbeitsprozess des Verbrennungsmotors,” Institut für Verbren-nungskraftmaschinen und Thermodynamik, TU Graz, eingereicht. 25-25. Winklhofer, E. Fuchs, H., and Fraidl, G. K. 1995. “Optical Research Engines—Tools in Gasoline Engine Development?” Proc Instn Mech Engrs, 209(4): pp. 281–287. 25-26. Gärtner, U., Oberacker, H., and König, G. Apr. 21–22, 1998. “Analyse der Brennverläufe Moderner NFZ Motoren Durch Hochdruckindizierung und Verbrennungsfilmtechnik,” 3. Internationales Indiziersymposium, AVL Deutschland. 25-27. Ipp, W., Egermann, J., Schmitz, I., Wagner, V., and Leipertz, A. 2001. “Quantitative Bestimmung des Luftverhältnisses in Einem Optisch Zugänglichen Motor mit Benzindirekteinspritzung,” in Motorische Verbrennung, Leipertz, A. (Hrsg.), Erlangen, BEV Heft 2001.1, pp. 157–172. 25-28. Willert, C., Röhle, I., Beversdorff, M., Blümcke, E., and Schodl, R. Sept. 2000. “Flächenhafte Strömungsgeschwindigkeitsmessung in Motorkomponenten mit der Doppler Global Velocimetrie,” Optisches Indizieren, Haus der Technik, Essen, Veranstaltung Nr. H030-09-033-0. 25-29. Winklhofer, E. Beidl, Ch. and Fraidl, G. K. Mai 18–19, 2000. “Prüfstandsystem für Indizieren und Visualisieren—Methodik, Ergebnisbeispiele und Ergebnisnutzen,” 4. Internationales Indiziersymposium, AVL Deutschland.

916 | Internal Combustion Engine Handbook

6606_Book.indb 916

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26 Fuel Consumption Reducing fuel consumption and exhaust emissions has become one of the central tasks of vehicle development in recent years. The reasons for this can be found not only in legal requirements but also in a more conscious approach to fossil fuel reserves and increased environmental awareness, both on the part of customers and vehicle manufacturers. Despite vehicle weights increasing all the time, it has been proven possible to significantly reduce fuel consumption in recent years (see Figure 26.1). Also directly dependent on fuel consumption are CO2 emissions, a field in which the Association of European Automobile

Manufacturers submitted to the European Union (EU) an undertaking to reduce values to 140 g/km (≅5.7 L/100 km fuel consumption) by the year 2008. The commitment was not fully implemented. An agreement between the European Parliament and EU member states at the end of 2008 intends to lower carbon dioxide limits for new vehicles to 120 g/km by the year 2015. This equates to a fuel consumption of 4.9 L/100 km. This limit value shall be achieved with the following measures: An average CO2 output of 130 g will be achieved in the future with more economical engines. A further 10 g of CO2 shall be

Figure 26.1 Fuel consumption and vehicle curb weight trend of passenger cars and station wagons registered in Germany. Data according to [26-1], expanded.

Internal Combustion Engine Handbook | 917

6606_Book.indb 917

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Percentage of fleet that must observe the limit value of 130g CO2/km

saved by using biofuels, rolling-resistance-optimized tires, and more economical air conditioning. According to the EU, the average output for a new car is now 160 g CO2/km. The value limits need to be achieved gradually, as shown in Figure 26.2.

Special rules apply for manufacturers with an annual production of less than 10,000 vehicles. They can apply to the European Commission for derogations to the limit values. There is also financial relief for vehicle manufacturers producing less than 300,000 new cars annually. Setting a limit value of 95 g CO2/km from the year 2020 is now being discussed.

26.1 General Influencing Factors A certain quantity of energy in the form of fuel is required to overcome various driving resistances. Possibilities for reducing fuel consumption include improving the efficiency of the power source and the powertrain, as well as reducing the vehicle’s various driving resistances. The required tractive force for overcoming the driving resistance is calculated as follows:

Z = FL + FR + FSt + FB



where Year

Figure 26.2 A gradual introduction of the CO2 limit value up to 2015.

Whereas 65% of a manufacturer’s new cars are supposed to reach the value of 130 g CO2/km in 2012, this Figure must rise to 100% of vehicles by the year 2015. If the limit values are not adhered to by the manufacturers, then fines will be payable in the period between 2012 and 2018; these are presented in Figure 26.3.

FL = aerodynamic drag FR = rolling resistance FSt = climbing resistance FB = acceleration resistance.

26.1.1 Aerodynamic Drag

Aerodynamic drag increases by the square of the resulting approach flow velocity, that is, with longitudinal approach flow from ahead, at the square of travel speed: Aerodynamic Drag:

FL = cw ⋅ A ⋅ 0.5 ⋅ rL ⋅ v 2



where cw = drag coefficient ρ L = air density v = driving speed A = bulkhead surface area.

Figure 26.3 Fines for exceeding limits in the period between 2012 and 2018 and from 2019.

The cw value, as a shape factor, and bulkhead surface area A, as a size factor, are susceptible to design manipulations. The bulkhead surface area can be reduced only to a limited extent, since a certain size must be achieved for the passenger cell, and all the various equipment assemblies and modules must be accommodated. The trend of the cw value since 1950 is shown in Figure 26.4. The limits are set on the cw value as a result of design trends, vehicle complexity, the necessary flow through the engine compartment and interior, freedom of wheel movement, antilift provisions on both axles, flow through the wheel housings for cooling of the braking systems, underbody flow for cooling of the exhaust system, and necessary attachments (such as mirrors, windshield wipers, antennas, and handles). Studying Figure 26.4, it is possible to detect certain stagnation in the efforts to reduce the drag coefficient. This often leads to an

918 | Internal Combustion Engine Handbook

6606_Book.indb 918

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.1 General Influencing Factors

Figure 26.4 Trend of the cw value since 1950.

increase in the drag coefficient with the respective successor models as the cross-sectional area is generally larger. Even though the New European Driving Cycle (NEDC) only has an average speed of 33.2 km/h, the cw value is still for a fastback vehicle with an exit drag coefficient of 0.32 and consequently achievable consumption improvements in the NEDC [26-32], very important in terms of fuel consumption. Therefore, it can be assumed that a 10% reduction in the cw value effectuates a fuel consumption improvement in the NEDC of 2.5%. The viable improvements of the cw value by introducing individual measures are potentially shown in Figure 26.5 [26-32]. The greatest fuel consumption potential for a fastback vehicle from the Golf class is provided by the rear. A consumption improvement of 4.7% is feasible in the NEDC here. Several measures on the chassis, such as a smooth underside, rear diffuser, and optimizing the airflow in area around the wheels, display similar fuel consumption improvement potential. Minimizing the radiator air flow with a louver or flaps has a consumption benefit of approximately 1.6%. This measures in the A-pillar area by sacrificing the externally fitted mirrors and delivers a fuel consumption reduction of 0.8%.

A Fiat Punto with a “fluid tail” method [26-33] is shown in Figure 26.6. This method demonstrated an 18% improvement in the cw value. The air flow is focused because of the “forward-bent” surfaces on the rear. The rear axle rims are shaped like a radial fan, which can easily be seen in Figure 26.6. The fan fills up the caster offset of the rear wheels with air.

Figure 26.6 Fiat Punto with a “fluid tail” [26-33].

The lower limit for a realistic, future cw value is now 0.2. The influence of the cw value on maximum speed and fuel consumption is shown in Figure 26.7.

26.1.2 Weight

Vehicle weight is a prevailing factor in acceleration and uphill travel. Vehicle weight in this context plays a linear role in resistance. Figure 26.5 Potential measures for reducing the cw value.

Climbing Resistance:

Fa = m ⋅ g ⋅ sin a



Internal Combustion Engine Handbook | 919

6606_Book.indb 919

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.7 The influence of cw value on speed and fuel consumption. Data according to [26-2].

Acceleration Resistance: Fa = ei ⋅ m ⋅ a





Rotating Mass Factor: ei =

qRedi 2 m ⋅ Rdyn

+1

where g

= 9.81 m/s2

m = vehicle mass including load

α

= lane angle of ascent

θ Red = reduced rotating mass moment of inertia Rdyn = dynamic tire radius i

= observed gear step.

The average vehicle weight continues to grow and/or stagnate at the same high level. The reasons for this can be found in increasing demands for convenience in the form of electric actuators for windows, sliding roofs, mirrors, and seats, and in higher equipping specifications, air-conditioning, seat heating, and power steering. In addition, the safety equipment developed in the past 20 years, such as traction and brake control systems, electronic stability systems, active shock absorbers and transverse stabilizers, airbags, and seat-belt tensioners, has also increased vehicle weight. The trend toward larger engines with the associated powertrain components has a weight-increasing effect, as does the increasing selection of diesel engines. Further increases in weight are the result of enhanced crash safety and corresponding additional automobile body structures. The trend in the weight of various vehicle categories in recent years is shown in Figure 26.8. The increase in weight means that more powerful (and heavier) engines are necessary to achieve the same performance;

weight then begins to escalate. A number of concepts exist to combat this weight spiral. A large portion of total body shell and suspension system weight can be reduced by intelligent lightweight structures and by replacing steel for lighter metals, mainly aluminum alloys. Logical vehicle concepts should also use lower power, supercharged engines to reduce the weight of the entire powertrain and result in working point shifts to the higher mean-pressure ranges. Such concepts are increasingly being offered in series production (see Chapter 26.2.1 Downsizing). Every 100 kg of weight reduction reduces fuel consumption by 0.15–0.2 L/100 km (NEDC). In the HYZEM driving cycle, in which the vehicle weight plays an important role because of the high driving dynamics, a fuel consumption reduction of approximately 5% can be reasonably expected with a 10% weight reduction.

26.1.3 Wheel Resistance

Wheel resistance is made up of components resulting from the deformation of the tires upon contact with the road surface from bearing friction losses, water-displacement resistances caused by the displacement of water on wet road surfaces, and toe-in resistances and lateral force resistances. The largest portion is made up of rolling resistance. This can be derived approximately from vehicle weight and the coefficient of rolling resistance, which summarizes the resistances generated at the tire. Rolling Resistance:

FR = fR ⋅ FG



where f R = coefficient of rolling resistance FG = vehicle weight.

920 | Internal Combustion Engine Handbook

6606_Book.indb 920

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.1 General Influencing Factors

Figure 26.8 Weight trend since 1981 for different vehicle classes. Data according to [26-3].

The coefficient of rolling resistance varies between the orders of magnitude of 0.01 and 0.04, depending on tire type, road surface, and vehicle speed. Approaches for reducing the rolling resistance can be found in the vehicle models that have a particularly optimized driving resistance, which virtually every large manufacturer offers. It is not possible to completely exploit the consumption-reducing potential of low-rolling resistance tires, since a drastic reduction in rolling resistance also incurs losses of drive smoothness, adhesion, and tire grip in wet conditions. It is necessary to not only select a different tire material mixture but also reduce the tire width and increase the inflation pressure. An increase in the tire pressure also reduces the tire’s flexing energy. The coefficient of rolling resistance of a passenger car tire, for example, can be reduced by approximately 25% if the tire pressure is increased by 0.5 bar. The coefficient of rolling resistance increases with vehicle speed.

The fuel consumption advantage of the commercially available energy-saving tires of the size, 195/65R15, was determined against the traditional tires at three constant travel speeds (50, 100, and 130 km/h). On average, depending on the brand, there was a fuel consumption advantage of between 0 and 0.29 L/100 km [26-36]. With everyday use, General German Automobile Association estimates a fuel consumption advantage for tires with an optimized rolling resistance of 0.15 L/100 km, with a vehicle mileage of 15,000 km a year, which equates to 22.5 L of fuel.

Coefficient of Rolling Resistance:





fR = C0 + C1 ⋅ v + C2 ⋅ v

4



where C0, C1, C2 = tire-specific constants v = driving speed. Tire rolling resistance and, consequently, consumption can be influenced by the particular rubber mixture selected. The greatest influence can be achieved by varying the material of the running surface. Low absorption mixtures make it possible to reduce rolling resistance by up to 35%. The improvements achievable by means of variation of materials are only in the 1%–5% range in the case of other tire elements, such as the sidewalls and bead.

26.1.4 Fuel Consumption

The various factors produce in total the following influence on distance-related fuel consumption (mpg):

∫ be ⋅ Be =

r ⎡⎛ 2⎞ ⎤ 1 ⎢⎝⎜ m ⋅ fR ⋅ g ⋅ cosa + 2 ⋅ cw ⋅ A ⋅ v ⎠⎟ ⎥ ⋅⎢ ⎥ ⋅ v ⋅dt hü + m ⋅ ( ei ⋅ a + g ⋅ sin a ) ⎢⎣ ⎥⎦ v ⋅dt ∫



When the vehicle is stationary, some other variables are idling fuel consumption (approximately 0.5 to 1 L/h) and, either continuously or intermittently, the power consumption of electrical consumers that extend into the kilowatt range. Examples are found in Figure 26.10. The installed electrical output for medium class vehicles is now approximately 3.5 kW, and for premium class vehicles is 5 kW. Approximately 560-W electrical output (voltage = 14 V and current = 40 A) results in the overall efficiency levels in the combustion engine map for providing electrical energy for the on-board supply system of maximum 26% (see Figure 26.11). In broad map areas, the efficiency is considerably lower. The crankshaft starter generator, the consumption potential of which is shown in Figure 26.12, is an alternative to the three-phase generators now widely used.

Internal Combustion Engine Handbook | 921

6606_Book.indb 921

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Parameter

Unit

Parameter

Unit

Be—distance-related fuel consumption

g/km

m—vehicle mass

kg

be—specific consumption

g/kWh

fR—rolling resistance coefficient

—

η ü—powertrain efficiency

—

g—gravitational acceleration

m/s2

α —lane angle of ascent

°

v—vehicle speed

m/s

ρ —air density

kg/m3

ei—rotating mass additional factor in gear i

—

cW—drag coefficient

—

a—longitudinal acceleration

m/s2

A—bulkhead surface area

m

t—time

s

2

Consumer

Power Consumption (kW)

Consumer

Power Consumption (kW)

Rear window heater

0.12

Onboard display

0.15

Windshield heating

0.3

Audio system

0.2

Wiper motor

0.1

On-board computer

0.15

Exterior lighting

0.16

Ventilation fan

0.1

Control unit power supply 0.2

ABS/FDR pumps

0.6

Fuel pump

0.06

Radiator fan

0.2

Gasoline injection

0.06

Total

2.4

Efficiency of generator at I= 40A

Efficiency of combustion engine

Torque [%]

Torque, Efficiency [%]

Efficiency

Overall efficiency of electrical energy

Figure 26.11 Efficiency map of a combustion engine (gasoline engine), efficiency characteristic curve of a generator, and overall efficiency map for the provision of electrical energy in the lower image (without belt losses). See color section page 1104. Function/Property

Overall Fuel Saving Potential

Start/stop (ECE cycle)

approximately 15%

Efficiency increase/42-V on-board power supply Braking energy recuperation Boost mode Figure 26.12 Consumption reduction potential achieved by crankshaft starter generator [26-5], extended.

Figure 26.9 Parameters and units.

Figure 26.10 Power consumption of electrical consumers in a passenger car.

26.2 Engine Modifications Characteristic data for modern gasoline and diesel engines are shown in Figure 26.13. Diesel engines now enjoy significantly better fuel consumption than gasoline engines. Diesel engines generally have a slightly higher power-toweight ratio than gasoline engines. Because of their various advantages and disadvantages, both the methods will continue to be used in future for those applications to which they are most suited. The modern diesel engine is superior in terms of fuel consumption to the gasoline engine, but the gasoline engine possesses a greater potential than the diesel engine. This is exploited by fully variable valve trains and direct injection (DI), downsizing, crankshaft starter generators, friction reduction, variable compression ratios, variable swept volumes including cylinder shutoff, and supercharging, and is estimated at 30%–40% consumption reductions. Therefore, lower specific fuel consumption could be achieved than with current supercharged DI diesel engines. The consumption potential of the diesel engine is estimated at 15%–25%. The successful provisions here also include the crankshaft starter generator, minimization of friction losses, improvement of mixture formation (see Chapter 26.2.2), and supercharging.

26.2.1 Downsizing

Increasing the mean pressure makes it possible to increase the effective output with the same displacement. Sensible approaches include supercharging, since modern naturally aspirated engines already exhibit relatively high mean pressures, which are difficult to increase without supercharging. With a smaller displacement, the same performance data are attained as with a larger engine. Smaller engines have lower absolute friction power and heat up quicker after a cold start. Because of the shift in the operating point and higher

922 | Internal Combustion Engine Handbook

6606_Book.indb 922

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Maximum Speed (r/min)

Maximum Compression Ratio: ε

Maximum Mean Pressure (bar)

Specific Power Output (kW/L)

Fuel Consumption Optimum Point (g/kWh)

Gasoline engine Naturally aspirated for passenger engine cars With supercharging

5,500 … 8,500

up to approximately 12

up to approximately 13

90

minimum 225

up to 6,800

up to approximately 10.5

up to approximately 19

up to approximately 125

minimum 225

Gasoline engine for motorcycles

17,000

up to approximately 13

up to approximately 12

50 ¼ 150

n/a

Diesel engine for passenger cars (DI)

Naturally aspirated engine

up to 5,000

up to approximately 20

up to 9

up to approximately 30

approximately 210

With supercharging

3,500 … 4,500

16 … 21

up to approximately 25

up to approximately 105

approximately 205

Engine Type

Figure 26.13 Comparison of characteristic data from gasoline and diesel engines.

mean pressure, the same performance profile is achieved in the sectors of better thermal efficiency, that is, usually at higher loads and lower speeds. In diesel engines, for instance, increasing injection pressure increases the mean pressure. According to [26-6], an increase in the injection pressure from 600 to 1,000 bar produces an increase in the mean pressure of 17% with the same specific consumption. Modern injection systems already reach a pressure of more than 2,000 bar. In the past, the mean-pressure increase thus possible was used to obtain a greater power with the same displacement. However, genuine downsizing concepts have been on the market for some years now, which is largely attributable to changing customer preferences and political impetus. In gasoline engines, thermal efficiency can be improved by raising the compression ratio. The resultant improvement in fuel consumption is shown in Figure 26.14. To avoid knocking, the compression ratio at full load is restricted to ε ≈ 12 (intake manifold injection). Considerably higher ratios of up to ε = 15 are possible under partial-load operation. A variation of compression ratio is desirable for

Figure 26.14 Impact of compression ratio on specific fuel consumption [26-12].

the purpose of optimization. This even makes it possible to avoid the efficiency loss in supercharged engines in the partial-load range, if the compression ratio is adjusted here. The high supercharging that is now possible produces a further improvement in thermal efficiency. Implemented concepts use the transfer of the crankshaft by means of eccentric adjustment or the tilting of a so-called monohead, which comprises the cylinder and the cylinder head. In summary, downsizing signifies the relocation of frequently traversed working points to ranges with lower specific fuel consumption. Since this range is in the high load region, the engine must be designed in such a way that it is operated under this high load for the major part of the load block used by the customer. The lower maximum power is gaining a growing amount of customer acceptance. The operating point shift is clearly discernible in Figure 26.15, when considering the points of equal speed/power. Combustion engines with twincharging or sequential supercharging constitute a promising downsizing concept. Here, the engine is supercharged by a compressor and an exhaust-driven turbocharger (TC) or by two geometrically different TCs depending on the concept. With twincharging, a compressor assumes the charge cycle work in the lower speed range, joint operation between the TC and the compressor ensues in the medium-speed range, and the TC assumes sole responsibility for the gas exchange in the upper speed range. Sequential supercharging uses two geometrically different exhaust-driven TCs. One involves a TC with a small turbine and compressor diameter, and the other involves a TC with significantly larger internal diameters. Sequential supercharging is regulated in a similar way to twincharging. The small exhaust-driven TC assumes the lower speed range. There then follows a short period of cooperation between both the TCs before the larger TC works independently again in the upper speed range. The advantage of a concept with two exhaust-driven TCs is primarily in the transient operation. By dividing into two supercharging units, charge pressure builds up much quicker. Both the concepts have been implemented in series development. This concept achieves an effective reduction of the engine displacement and, thus, the overall geometry of the engine (friction loss advantages), yet with unchanged effective output.

Internal Combustion Engine Handbook | 923

6606_Book.indb 923

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.15 Shifting the operating point with different engine concepts [26-30].

Furthermore, because of the operating point shift mentioned previously, a sizable part of the specific fuel consumption can be saved. Turbo engines with DI offer another advantage in terms of efficiency. Thanks to improved internal cooling with DI, it is possible to operate the supercharged engine at high compression ratios. In addition, the engine can be largely driven with a stoichiometric or leaner than stoichiometric mixture [26-20]. Because of the tolerance to exhaust temperatures at 1,050°C, enrichment is mostly no longer necessary. This leads to a significant reduction in fuel consumption. Other requirements for implementing this concepts include the use of thermally heavy-duty catalytic converters that are able to resist the increased exhaust temperatures on a permanent basis. Significant improvements have also been made recently to the TC response in transient operation. This was achieved with the so-called twin-scroll chargers, which allow separate exhaust gas routing from the cylinders up to the turbine housing. Two cylinders are centralized with four-cylinder engines in an exhaust gas line. This principle utilizes the kinetic energy from the exhaust gas stream, significantly better than the systems with only one exhaust gas feed to the turbine. The fuel consumption of a current downsizing concept is shown in Figure 26.18.

The disadvantages and risks of downsizing primarily include higher component load as a result of higher mean pressures (impact on service life), more elaborate mixture formation systems to get large charge masses ready for combustion in a targeted and efficient way, greater degree of complexity, complex coordination of the load cycle response, and the customer acceptance, particularly for midclass vehicles and higher segments.

26.2.2 Downspeeding

The adjustment of the overall transmission ratio (see also Section 26.3) is provided as a logical consequence of downsizing. A longer overall transmission ratio is selected with the result that the working point of the engine shifts to a lower speed at higher mean pressure for the same driving power, consequently, approaching map areas of better specific fuel consumption in theory. Concepts such as these were already in series production in the 1980s (e.g., mean pressure engine from Audi, BMW’s eta engines, and overdrive characteristic of the highest gear from several manufacturers). Yet, because the raw material situation (mineral oil reserves) was not so critical at that time and because of costs, they only achieved relatively low numbers.

924 | Internal Combustion Engine Handbook

6606_Book.indb 924

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.16 Twincharging Volkswagen AG, VW Golf GT, 2.0-L four-cylinder inline gasoline engine DI [26-20].

Lowering engine speeds to reduce fuel consumption in the absence of compensatory measures, such as supercharging or increasing engine displacement, is also designated as dieselization. In these instances, the vehicle’s acceleration capability is diminished. Strategies such as these are an option for vehicles with automatic transmissions and are represented in series production. The transmission control unit undertakes very early gear shifts, if the driver is rated as economically orientated. If the driver requests more dynamics, then upshifts are performed later and the low-speed concept is temporarily abandoned. A compromise between customer acceptance in terms of acceleration capability and fuel consumption effect is necessary for manual transmission vehicles.

26.2.3 Diesel Engine

The three diesel combustion systems, the prechamber, swirl chamber, and DI processes, were used along one another up to the late 1980s. The introduction of DI in passenger car diesel engines resulted in the increased replacement of prechamber and swirl chamber engines. Most of the new passenger car diesel engines use DI. The up to 15% lower specific fuel consumption of the DI diesel engine compared with indirect injection engines is primarily the result of the lower heat losses resulting from the undivided combustion chamber that is located in the piston head, and is also the result of the lower losses because of the elimination of the flow between the subsidiary and main combustion chambers. Indirect injection engines achieve an effective efficiency of approximately 36% compared with 43%

in the case of DI. Disadvantages are the steeper pressure rise (noise) and higher NOx emissions compared with indirect injection engines. Softer, more homogeneous combustion is achieved by preinjection of small quantities of fuel and by pulsed injection. In future, this will be achieved with working-cycle-selective and cylinder-selective injection quantity modulation. The precondition is an injection system incorporating high-speed, largely free actuation of the injection nozzles, and complex tuning of the high-pressure fuel supply system up to the nozzles. This is achieved with a common rail, pump-nozzle systems, and solenoid-valve actuated distributor pumps. The first two systems make use of the pressurized storage system for the fuel. The quantity required for injection is obtained from the pressure accumulator. In the case of the pump-nozzle system, storage is restricted to a crankshaft angle window generated by the cam contour; the common rail system provides a continuous high system pressure. Both the systems operate with a peak pressure of between 1,000 and 2,500 bar. The high injection pressure compared with earlier systems has the following effects: •• higher jet velocity and greater penetration depth •• earlier start of mixture formation because of greater reactive surface area and improved distribution •• smaller average droplet diameters and more surface area for reaction •• more intensive mixture formation •• more rapid evaporation

Internal Combustion Engine Handbook | 925

6606_Book.indb 925

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

BMW X1 28i 4 Cylinder Turbo

BMW X1 28i 6 Cylinder Naturally Aspirated Engine

Combustion process

Gasoline engine—DI turbo

Gasoline engine

Engine construction type

Four-cylinder inline

Four-cylinder inline

Displacement

19.956 cm3

2,996 cm2

Compression

11:1

11:1

Output

180 kW/5,000 r/min

190 kW/6,600 r/min

Torque

350 Nm/1,250–4,800 r/min

310 Nm/2,600–5,000 r/min

Consumption NEDC vmax Acceleration 0–100 km/h

7.9 l/100 km 240 km/h 6.5 s

9.4 l/100 km 230 km/h 6.4 s

•• more rapid mixture distribution •• higher conversion rates •• higher degree of homogenization of the mixture “cloud” •• shorter combustion duration •• improved (internal) oxidation of soot and smaller particles. As a result of these advantages, the trend is now in favor of higher injection pressures. A compromise is needed between the consumption advantages resulting from high injection pressures and the increased consumption deriving from the power necessary to drive a high-pressure pump or the camshaft and pump-nozzle elements. The common rail system possesses the lowest maximum drive power requirement. It needs only 40%–50% of the drive power of a distributor pump and only 20% of the drive power of a pump-nozzle system. Looking at current and future emissions requirements, one can reasonably assume the exclusive use of common rail systems. The improved facilities for intervention to control mixture formation and reaction rate in the combustion chamber make it possible to achieve, for the same specific consumption using variable high-pressure injection systems, an increase in the torque that more than balances out the losses resulting from the additional drive torque required. Pilot injection systems and injection rate adjustments can also be achieved with variable nozzle geometries. So-called piezo injectors are also being used more often. They allow the quick and accurate fuel metering necessary for up to five pilot injections, one main injection, and one secondary injection. An electronic engine management system (digital diesel electronics), which controls the injection rate and time on the basis of the working point, is necessary to permit the exploitation of the potential of high-pressure injection systems. Inlet duct geometry must also be tailored to the individual injection system and nozzle geometry in terms of defined flow, and, therefore, mixture-formation processes in the combustion chamber.

26.2.4 Gasoline Engine 26.2.4.1 The Lean-Burn Concept and Direct Injection The significantly higher specific fuel consumption of the gasoline engine in partial load compared with full load can be reduced by operating with excess air, that is, lean-burn operation or stratified charge. The reasons for this are the following:

Figure 26.17 Multistage charging principle, BMW Group, BMW 535d, 3.0-L six-cylinder inline engine diesel, multistage charging principle [26-24].

•• partial dethrottling as a result of higher air demand at the same mean pressure and reduction of charge-cycle work •• enlargement of thermal efficiency as a result of increasing the isentropic exponent •• reduction of wall heat losses as a result of lower mixture densities in the wall zone. Limits are imposed on mixture leanness by the following: •• ignition limit (lean-burn limit) •• incomplete combustion as a result of locally differing mixture compositions •• cycle fluctuations as a result of “wandering” of the combustion center and misfiring •• slowing of combustion rate. In the lean-burn operation, the overall charge mass increases as a result of the air surplus, with rising final compression pressures and temperatures as the consequence. The quantity of heat liberated is transferred into a greater charge, however, thus causing the average process temperature to fall. Both the effects, that is, the larger charge mass and the higher temperature spread, result in an increase in the isentropic exponent κ , causing thermal efficiency to rise. This is defined by the following:

hth = 1 − e 1−k

where

ε = compression ratio κ = isentropic exponent. The reduced charge cycle work makes it possible to increase the effective overall efficiency by up to 4%, depending on the working point and degree of leanness [26-7]. The fuel-saving potential in the part-load range of the experimental engines constructed in the 1980s (normal FTP and ECE cycles) was up to 15%. These lean-burn concepts were not pursued further, however, because of the more stringent exhaust emission legislation. Catalytic aftertreatment causes no problems for lean-burn operation in the cases of hydrocarbon and carbon monoxide emissions. However, high combustion temperatures produce more oxides of nitrogen than the stoichiometric operation, and these temperatures cannot be completely reduced because of the high oxygen content in the exhaust gas. This phenomena

926 | Internal Combustion Engine Handbook

6606_Book.indb 926

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.18 Comparison of fuel consumption conventional and downsizing concept [26-37].

means that other conversion methods, such as NOx storage catalytic converters, are necessary. The potential of lean-burn operation can be better exploited in combination with gasoline DI. The advantages are as follows:

•• internal cooling of the inducted air by the evaporation of fuel in the cylinder, thus making higher compression ratios (supercharging) and, as a consequence, higher thermal efficiencies possible (shifting of the knocking limit)

•• replacement of quantity regulation with quality regulation, reduction, or elimination of throttling losses

•• reduction in full-load enrichment

•• charge stratification by the corresponding injection-jet location and injection rate combined with reduced air flow •• improved dynamics at load changes, delays caused by intake manifold filling, and creation of the wall fuel film are eliminated

•• adjustment of the air ratio to various working points (see Figure 26.19). Global statements concerning the reduction of fuel consumption as a result of gasoline DI have little rationale, since they

Figure 26.19 Gasoline DI strategies.

Internal Combustion Engine Handbook | 927

6606_Book.indb 927

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

depend on the differing variants and working points. Significant consumption potential exists in partial-load operation, since engines can be run with high average air ratios in stratified charge operation. A 10%–15% reduction in fuel consumption can be assumed in the test cycles now widely used in Europe, North America, and Japan. Only slight consumption advantages are achieved in the full-load operation with the stoichiometric operation as a result of improved internal cooling. In doing so, full-load enrichment can be reduced or the compression ratio increased. A certain portion of the consumption reduction is negated again by the regeneration phases necessary in the partialload operation for the reduction of oxides of nitrogen. For this purpose, the engine is operated for a short time (several seconds) at substoichiometric air ratios, to desorb the oxides of nitrogen fixed in the storage catalytic converter during the lean-burn operation. During this phase, the catalytic converter operates like a conventional three-way catalytic converter. Great advantages can be attained with the gasoline DI of supercharged engines. Thanks to the direct delivery of fuel into the combustion chamber and significantly increased internal cooling, it is possible to operate the supercharged engine with compression ratios similar to those of a naturally aspirated engine. This leads to a significant improvement in thermal efficiency. In addition, this also largely cancels out the need for enrichment in the full-load range by using exhaust gas turbines that are insensitive to high temperatures. A huge fuel saving can be achieved in this area simply by relinquishing the substoichiometric engine operation during full load. Modern engines (intake manifold injection) are operated during full power output with mixtures of up to λ = 0.7 to combat the accumulated heat flows generated by internal cooling. However, the principle of stoichiometric full-load operation requires

highly heat-resistant materials throughout the combustion chamber and on the exhaust gas side. Furthermore, there are differences in the specific fuel consumption between the individual combustion processes of gasoline DI. The lowest specific consumption can only be achieved with the jet-directed method (see Figure 26.20). Potential fuel consumption savings with the transition to gasoline engines with DI are detailed in Figure 26.21. Consumption advantages are up to 35% in the partial-load range and still approximately 5% in the full-load range. The consumption saving potential shown in brackets in Figure 26.21 applies to the wall-directed method; the values that are not in brackets apply to the jet-directed method, which is in a better position to implement the theoretical consumption advantages of direction injection.

26.2.5 Combustion Process HCCI

The acronym HCCI stands for homogenous charge compressed ignition . This type of combustion compresses the fuel in the combustion chamber and forms a homogeneous mixture prior to injection, which is then self-ignited by the compression and a considerable externally led residual gas quantity. Consequently, this procedure represents a combination of diesel engine and gasoline engine combustion. Temperatures of approximately 1,000 K are required to achieve self-ignition with traditional gasoline fuels. These high temperatures cannot be achieved with the modern conventional compression ratios (ε = 9 … 13) by compression alone. Therefore, a process temperature increase is adjusted by the exhaust gas of the previous cycle. Self-ignition monitoring can take place for upstream pilot injections, similar to the common rail system in a diesel engine. During this process, the combustion center and indicated mean pressure can be adjusted on

Figure 26.20 Specific fuel consumption of different DI combustion methods, data according to [26-29], extended.

928 | Internal Combustion Engine Handbook

6606_Book.indb 928

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.21 Fuel consumption saving with DI [26-34].

a cylinder-by-cylinder basis. In the NEDC, a fuel saving of approximately 3% [26-18] is possible with this combustion process when compared against the traditional methods.

26.2.6 Variable Valve Train

Variable valve trains are a further method of influencing charge cycle and exhaust gas recirculation, and thus cut fuel consumption and pollutant emissions. Matching of inlet valve stroke and opening duration with the necessary flow of fresh gas makes it possible to achieve a partial dethrottling in the partial-load range. Running is improved as a result of lower valve strokes in the idling and near-idling ranges. The reasons for this can be found in the improved mixing resulting from higher gas velocities (up to the

speed of sound) in the narrow valve gap, and resultant more homogeneous and faster combustion. The reduction in idling speed thus possible offers an additional consumption-saving potential, because of the lower frictional losses. There are two possibilities for the intake opening duration: early closure of the intake when the necessary fresh gas charge is reached and delayed closure of the intake. The intake valve closes only during the compression stroke, when the unneeded charge mass is transferred back into the intake system. Here, losses occur, compared with early closing of the intake, as a result of the double movement of part of the charge mass (see Figure 26.22). In addition, variable valve actuation times and valve lifts make internal exhaust gas recirculation via shifting and

Figure 26.22 Charge cycle loop: conventional, early closure of the intake, and delayed closure of the intake.

Internal Combustion Engine Handbook | 929

6606_Book.indb 929

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

prolongation of the valve overlap possible. The advantages are found in dilution and improved mixing of the fresh charge. The improved mixing brings with it the following effects: dilution significantly reduces the emissions of oxides of nitrogen, since the overall charge mass consists partially of inert exhaust gas. Combustion gas temperatures are thus lowered; less energy is available for the formation of oxides of nitrogen. The consumption benefits are to some extent negated by the lower combustion gas temperatures, resulting in slower and less homogeneous combustion. Cylinder shutoff can also be achieved by the complete closure of the intake valves of individual cylinders for a number of cycles (see corresponding Chapter 26.2.7). The greatest opportunities for the future lie with gasoline engines, where the advantageous effects of variable valve actuation and DI with stratified-charge concept and supercharging are combined. The consumption reduction potential of a number of different concepts is shown compared with the present series production situation with camshaft spread in Figure 26.23. Soon, concepts from the fully variable valve train will also become increasingly important. These primarily include electromechanical and electrohydraulic valve trains. A charge cycle actuation such as this in gasoline engines could save approximately 15% of fuel [26-21]. 26.2.6.1 Ignition System Rapid and homogeneous ignition and high reaction rates are necessary to achieve a high mean pressure and good thermal efficiency. Ignition depends, among other things, on spark plug position and on the quality of spark transmission. A central spark plug location is generally desirable to achieve uniform combustion and short flame paths (see Figure 26.24). Central spark plug location is relatively easily achieved with

the use of four-valve technology cylinder heads. Improved charge gas exchange conditions can thus be combined with the optimum spark plug position. The combustion process in a actual gasoline engine differs from the constant volume combustion assumed in the ideal engine. Observation of energy conversion against the crankshaft angle during the compression and power strokes generates a surface that expresses the progress of combustion. The concentration point of this area should be located approximately 8–10° crankshaft to top dead center to achieve a good internal efficiency. Combustion progress and, therefore, the location of the combustion center, depend on the ignition angle α Z and the air-fuel ratio λ (see Figure 26.24). Energyoptimized early ignition produces low specific consumption; compromises involving a later ignition angle are frequently necessary, however, to avoid knocking and to reduce exhaust emissions. With homogeneous operating strategies, minimum consumption occurs at slightly lean mixture ratios. Ignition of the mixture can be further improved by the use of two spark plugs for two- and three-valve engines. More ignition energy is supplied, with the result that more homogeneous combustion and, therefore, lower cycle fluctuations are achieved even at critical working points, for instance, when idling and during exhaust gas recirculation. More rapid combustion is also achieved, improving thermal efficiency. Optimum combustion center positioning can be achieved with greater certainty. Series operation achieves reductions of approximately 2% in consumption compared with single plug ignition systems [26-9]. In addition, phase-displaced double ignition makes it possible to control the pressure rise and the plot of pressure in such a way that combustion noise can be reduced without delayed ignition and corresponding loss of efficiency. A reduction of 3 dB (A) has been achieved in this field in series operation [26-9].

Figure 26.23 Comparison of consumption potential of different gasoline concepts and diesel DI, data according to [26-8], expanded.

930 | Internal Combustion Engine Handbook

6606_Book.indb 930

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.24 Influence of the spark plug (left), preignition and air/fuel ratio on fuel consumption [26-10].

26.2.7 Cylinder Shutoff

Examination of the fuel-consumption map of a gasoline engine (Figure 26.33) indicates that specific fuel consumption at low engine torques and low mean pressures, in particular, may be more than twice as high as at and around the optimum, high mean pressure point. This consumption disadvantage at low mean pressures is the result of many factors as follows: •• a compression ratio selected for full-load operation •• low flow velocities at the intake valve •• high throttling losses in gasoline engines, as a result of the throttle valve being almost closed •• relatively high friction compared with the engine output •• high wall heat losses. In large passenger car engines with a high power and torque output, only a small fraction of the engine’s available power is required in urban traffic or while driving on the highway. The more power and torque an engine offers, the lower its engine working point drops into the partial-load range. The result is high fuel consumption.

26.2.7.1 Concept for Reduction of Fuel Consumption The basic concept of cylinder shutoff is that of increasing the torque of the individual cylinders in the partial-load range to achieve a working point with a better fuel consumption for these cylinders. Other cylinders are shut off for compensation. Eight- and twelve-cylinder engines are particularly suitable for the cylinder shutoff concept. But there are also six- and four-cylinder engines already in series production. In these engines, half the cylinders can be shut off in each case at low load and speed levels. The cylinders to be shut off are indicated by the engine’s firing sequence on the basis of the rule of maintaining a constant firing sequence even in shutoff operation. In eight-cylinder V engines, an obvious tactic is to shut off two cylinders in each bank, whereas a complete bank can be shut down in a twelve-cylinder engine. Comprehensive control algorithms and pilot control systems for throttle valve, injection, and ignition are needed to eliminate ride smoothness losses during gearshifts. 26.2.7.2 Consumption Benefits in the Partial-Load Range The mean pressure curve for the eight-cylinder engine (operating on all cylinders) and for shutoff operation, as well as the cylinder shutoff range, is shown in Figure 26.25. Depending

Internal Combustion Engine Handbook | 931

6606_Book.indb 931

1/19/16 8:57 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.25 Consumption benefit achieved from cylinder shutoff of an eight-cylinder engine. According to [26-13].

on the working point, the consumption benefits of cylinder shutoff are between 5% and 20%. The consumption benefit for a twelve-cylinder engine at a partial-load point with a cylinder shutoff is shown in Figure 26.26. Improvements in the consumption of 15% and 13% were determined for a vehicle with an eight-cylinder engine (W220 5.0 L) for constant travel at 90 and 120 km/h, respectively; fuel consumption in the NEDC was reduced by 6.5% as a result of cylinder shutoff [26-13], [26-14].

26.2.8 Auxiliary Units

The energy requirement and associated fuel consumption of the auxiliary units should not be underestimated. These include the generator for supplying the on-board supply system, the air conditioning compressor, the power steering

pump, a vacuum pump for the braking booster for engine that are driven dethrottled, the mechanical coolant pump, the oil pump(s), as well as pumps for driving dynamics control systems. A great deal of energy is required to drive these components. By way of example, we will assess the air conditioning compressor at this point. The fuel consumption of an air conditioning system (caused by the air conditioning compressor) can vary greatly. The most important factor is the ambient temperature. The higher the temperature of the outside air, the more the energy is required for cooling. Distance-related fuel consumption in l/100 km stills depends on how long a vehicle requires for a journey of 100 km. If 20 km is covered in an hour in stop and go traffic on a hot day, then approximately 0.4 L of fuel is required for the air conditioning. Projected to 100 km, this would represent a fuel consumption

Figure 26.26 Consumption benefit achieved from cylinder shutoff of a twelve-cylinder engine. According to [26-12].

932 | Internal Combustion Engine Handbook

6606_Book.indb 932

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.27 Influence of oil and coolant temperature upon fuel consumption [26-35].

bar is required for operationally reliable oil supply in this operating range. The oil pressure requirement of the engine increases almost linearly with the engine speed. An oil pump delivery volume increases with its speed, whereas the oil flow through the engine is essentially determined by flow resistance, which is independent of the speed. The engine flow rate, however, is proportional to the oil pressure and, in turn, dependent on the operating temperature. It is particularly at low temperatures that marked discrepancies arise between oil pump delivery volumes and engine oil flow rates. In this range, the oil pump delivery volume is significantly higher than the oil flow through the engine. The “superfluous” oil is, therefore, discharged via a bypass. The mechanical work needed to deliver oil for the engine requirements is lost. This

Friction [kW]

of 2 L/100 km. However, fuel consumption on the interstate can reduce significantly per 100 km, depending on the speed driven. The fuel consumption per time unit remains constant at approximately 0.4 L an hour. If an average distance with an average speed and a typical climate for Germany is taken as a basis, then a fuel consumption of approximately 0.62 L/100 km for the vehicle air conditioning can be expected. With an average fuel consumption of 8–12 L for 100 km, this equates to between 5% and 8% [26-23]. The oil pump is another energy-intensive auxiliary unit. It is usually driven directly by the crankshaft and provides a volumetric oil flow that is dependent on the engine speed. Hot idle conditions usually act as a design point with oil temperatures of 140°C. An oil pressure of approximately 1.0

Tempera

ture cool

ant [°C]

Te

C]

il [°

re o

atu

er mp

Figure 26.28 Friction map with different fluid temperatures [26-35].

Internal Combustion Engine Handbook | 933

6606_Book.indb 933

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

situation can be remedied with demand-controlled oil pumps, which are able to adjust the volumetric oil flow to the oil flow through the engine. This enables an effective improvement in oil pump efficiency. With demand control of the volumetric oil flow, which a few vehicle manufacturer have since brought into series production, a 2% fuel saving can be achieved on average [26-28].

26.2.9 Heat Management Methods for Achieving Fuel Consumption Reductions

Heat management denotes the purposeful influence and utilization of heat flows generated in the vehicle. These objectives include ensuring thermal operational reliability, increasing driving comfort, and reducing fuel consumption and its associated exhaust emissions. The priority of all these efforts to integrate an intelligent heat management system into the vehicle is the attempt to shorten the particularly fuel-intensive warm-up phase. Optimizing heat flows while the engine is at operating temperature is another heat management task. The map-controlled thermostat has become widely accepted in series production as a simple basic heat management system. This system allows the combustion engine to be operated in the partial-load area with significantly higher coolant temperatures. The increased temperature level has a positive effect on fuel consumption. A higher temperature level can be achieved by lowering the coolant and oil heat flows. The impact of the coolant temperature and oil temperature on fuel consumption is shown in Figure 26.27. Higher temperatures of the components surrounding the combustion chamber (primarily achieved by higher water temperatures) provide better combustion conditions in the combustion chamber. Friction power inside the engine is influenced by both the coolant temperature as well as the lubricating oil temperature (see Figure 26.27 and Figure 26.28). The electrical coolant pump is another heat management system with huge potential in terms of fuel consumption reductions. An additional degree of freedom has been achieved by electrically actuating the coolant pump. This freedom becomes obvious in both the warm-up phase and in hot operating state. Forced convection of the coolant can be suppressed in the combustion engine’s warm-up phase. The standing coolant has an isolating effect and significantly reduces the engine’s warm-up phase. The electric coolant pump starts up automatically when it reaches the predetermined limit temperature and is then able, as previously, to discharge the heat by forced convection. Heat removal dependent on heat introduction is possible in the hot operating state. This system does not just provide benefits in the warm-up phase, the coolant circuit can be maintained even after switching off the engine. This is a thoroughly interesting issue in light of the exhaust gas turbocharged engines, which are now coming back onto the market again. Furthermore, the electrical coolant pump provides benefits in terms of power loss, which are considerably less in the broad engine map ranges than with its mechanical counterpart. This means that fuel consumption savings can be realized even in

the stationary operation. With the aid of an electrical rather than a mechanical coolant pump, a fuel consumption saving of approximately 2% [26-20] is possible in the NEDC depending on the vehicle and motorization. The manufacturing costs of the electrical coolant pump are greater though.

26.2.10 Hybrid Concepts

Hybrid vehicles are becoming ever more important. There are very high levels of demand for hybrid vehicles in Asian and North American markets. During the course of the CO2 debate, and in light of rising fuel prices, growing registration numbers and an increased range of vehicles have also been recorded in Europe. It is now possible to choose among four different hybrid concepts: micro, mild, basic, and full hybrid. The classification of the individual classes is shown in Figure 26.29 [26-22]. It is particularly in the partial-load area that hybrid concepts offer advantages in terms of fuel consumption and emissions. The potential CO2 reduction stated in Figure 26.28 can be assigned to the fuel consumption, according to the carbon footprint. One disadvantage of these vehicle concepts is the large weight of the energy storage system, which has a particularly negative effect on fuel consumption during interurban and interstate journeys. Only the combustion engine is active during these operating states and the mass of the electrical energy source being carried along represents an additional driving resistance. The extra weight to be anticipated for different power hybridizations is displayed in Figure 26.30. The additional costs also have to be assessed. The disproportionally increasing percentage of the storage system (battery pack) makes it clear that from a fuelconsumption perspective, hybridizations are only logical up to approximately 20 kW. To ensure the desired battery service life, with current systems, a charging cycle of only 10% of the nominal capacity should not or should only ever rarely be exceeded. The largest part of the additional weight is installed in the form of unusable battery capacity and, therefore, substantially lessens the consumption advantage. In [26-26], a modern hybrid vehicle (Lexus RX400h) is compared against vehicles with conventional drives now found on the market in terms of fuel consumption. In the hybrid vehicle, the combustion engine is connected to a generator by an electronic continuously variable transmission (E-CVT). The generator transfers its energy to a power control unit (PCU), which, in turn, feeds the electric motor or, in appropriate cases, the battery. The engine torque is constantly divided to a ratio of 28%–72% (generator/vehicle). Consequently, 28% of the power produced by the combustion engine is supplied with three levels of efficiency (η generator, η PCU, and η electric motor). This multiplication of three individual efficiency levels inevitably leads to a poor overall efficiency. According to [26-26], the actual driving mode of hybrid vehicles can be subdivided into three phases. In the initial range up to approximately 60 km/h, the combustion engine is cyclically operated, the generator speed is relatively low, and E-CVT transmission efficiency is high. This represents a

934 | Internal Combustion Engine Handbook

6606_Book.indb 934

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.2 Engine Modifications

Figure 26.29 Distinctions between hybrid drive systems [26-22].

significant consumption advantage. However, in the highly dynamic driving style with higher mean pressure requirements, this very quickly leads to an operating point shift in

the areas with poor efficiency. In driving mode on country and highways and interstates, that is, in the speed range from 60 to 150 km/h, there is a minor consumption advantage

Internal Combustion Engine Handbook | 935

6606_Book.indb 935

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.30 Weight prognosis and level of hybridization [26-31].

thanks to the optimization of the combustion engine as well as the special hybrid system operating strategies. At this range, the system was designed so that the generator stops. As a consequence the flow through the electric path is virtually nil, a constant driving mode is required. A soft driving style is also considered optimal to fuel consumption in this phase. However, in the highly dynamic driving style, fuel consumption increases disproportionately compared with conventional drive systems. The hybrid system does not provide any advantages at constant speeds above 150 km. The losses rise markedly and push fuel consumption to ranges that lie well above the conventional drive systems. The fuel saving potential of the individual hybrid drive subsystems can be observed in Figure 26.31.

It is clear from the diagram that the biggest saving is delivered by shifting the combustion engine’s operating point and not through energy recovery. Current trends indicate that hybrid systems with small, highly supercharged combustion engines display the greatest fuel consumption potential. With systems such as these, the combustion engine can be operated for large journey times at its optimum point in terms of specific fuel consumption and, in particular, more than compensate for the disadvantage resulting from the additional weight. These systems are known as range extenders. The combustion engine thereby has the task of charging the battery and is not used to provide driving power.

26.3 Transmission Ratios Transmission ratios of the powertrain are determined in the manual transmission, where the individual gear ratios are selected either manually by the driver, or by an automatic system. In other systems, the transmission ratio refers to the axle drive (differential gear), which features a fixed transmission ratio. The following describes the gear ratio from the transmission (from the engine) to the wheel: •• transmission ratio in the selected gear iGi •• differential gear ratio iD

Figure 26.31 Fuel saving potential hybrid drive system Lexus RX 400h according to [26-26].

•• power splitting and its impact on the engine working point in hybrid powertrains. Overall powertrain transmission in selected gear is therefore i:

iA = iGi ⋅ iD

936 | Internal Combustion Engine Handbook

6606_Book.indb 936

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.3 Transmission Ratios

26.3.1 Selection of Direct Drive

The combustion engine’s torque is converted through the multistep transmission with its individual gear ratios and the transmission ratio of the downstream differential gear according to the demand for tractive effort at the drive wheels. In engines with spur gear drives, it is possible to assign a direct ratio to one of the gears (transmission ratio = 1:1) (coaxial gear). This achieves higher transmission efficiency, since no pair of gearwheels are engaged under load in such “direct drive” dimensions. The gear most frequently used on the road should be selected for the “direct transmission” stage to permit maximum exploitation of this consumption benefit. In passenger cars, this is generally the highest gear, which may reach utilization rates of above 80%. In the lower gears, transmission efficiency is 95%–96%, whereas 98% is achieved in direct drive. The necessary overall transmission ratio is then a function, in this low-friction gear, of the differential gear.

26.3.2 Selection of Overall Transmission Ratio in the Highest Gear

The powertrain’s smallest transmission ratio (iGimax ⋅ iD) has an influence on not only maximum speed, surplus force, and, consequently, its agility, but also fuel consumption, noise emission, and engine wear. Selection of overall transmission ratio is heavily dependent on the vehicle manufacturer’s philosophy, and it is, therefore, not possible to provide any generally valid recommendations.

There are, in principle, three design options; these are shown in Figure 26.32. 26.3.2.1 Design for Maximum Speed Here, the overall transmission ratio is selected in such a way that the driving resistance curve on a level road (wheel resistance plus aerodynamic drag) intersects with maximum tractive effort at the wheel. Only this design achieves the highest possible maximum speed for the vehicle. At this working point, the engine is running at its design speed. 26.3.2.2 Overspeed Design In this case, the powertrain’s overall transmission ratio is higher in the design for highest maximum speed. The intersection of tractive effort at the wheel with the driving resistance curve on a level road is located beyond maximum power, that is, at a correspondingly higher engine speed. This high engine speed level results in higher fuel consumption (see Figure 26.32). The highest possible maximum speed is not achieved using this “short” transmission ratio, but high surplus force that can be used to overcome additional driving resistances is available at the wheel below maximum speed. This high surplus force produces an extremely agile vehicle. 26.3.2.3 Underspeed Design Here, the overall transmission ratio is lower in the design for highest maximum speed. The intersection of tractive effort at the wheel with the driving resistance curve occurs below the maximum tractive effort engine speed. Here, engine speed is lower than that in the other two design modes and,

Figure 26.32 Different design of the overall transmission ratio in the highest gear.

Internal Combustion Engine Handbook | 937

6606_Book.indb 937

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

thus, provides engine working points with favorable fuel consumption. The highest possible maximum speed is again not achieved in this layout, and surplus force at low speeds is slight, with the result that the vehicle does not react so readily to load shifts. If an extremely pronounced underspeed arrangement is selected, it will be necessary to change down in gear as driving resistance increases, since the surplus power available is not sufficient. This driving style sacrifices the consumption benefits of the underspeed design as a result of changing down a gear. Of the extremely short and extremely long ratios in the highest gear shown in Figure 26.32, a specific fuel consumption advantage of 16% results at the restrictive maximum speed for the highly underspeed design. 26.3.2.3.1 Selection of the optimum-consumption gear An unrestricted selection of transmission ratios is, of course, not available if the engine power is to be fully exploited. This is apparent in the case of a journey at maximum speed or full acceleration going up to top revs in all the individual gears. The situation is different when traveling in the partial-load range. The required tractive effort can then be provided in any one of the several gears available. Then, the driver, or alternatively the automatic transmission management, is able to select the optimum-consumption gear without having to accept any disadvantages, at least at constant travel speed. Examination of the consumption map of a combustion engine (Figure 26.33) demonstrates that there is only one working point at which the engine achieves its best specific consumption. The optimum point is in all cases located at high-load and medium-to-low speeds. Specific consumption becomes

greater the further one departs from this point in this map. If unrestricted gear selection is available, the highest possible gear should always be selected, in order that engine load is high and that revs do not rise excessively. This is illustrated in an example in Figure 26.33. A constant power curve, such as is required for a travel speed of 100 km/h by the vehicle, has been plotted in the consumption map. The engine is able to provide this driving performance in gears two to five. It is apparent that the engine reaches zones of curves with lower specific consumptions the higher the gear selected. Consumption at a constant 100 km/h is poorer by 60% if the vehicle is driven in the second gear instead of the fifth gear. The best consumption case would occur if it were possible to select the overall transmission ratio at this vehicle speed in such a way that the engine was operated at a speed of 1,100–1,200 r/min. Consumption savings of 15% would be possible compared with driving in the fifth gear if this were the case. The more gears that are available, the better one can target the engine’s operating optimum for the particular power demand. This is extremely well implemented in the case of a fully variable transmission, selection of the transmission ratio being performed by an electronic system. In this case, it would be possible to achieve the optimum working point for the above 100-km/h journey in Figure 26.33 and exploit the consumption advantage stated. Such transmissions are offered by some manufacturers in series production as so-called CVT with friction-contact transmission. However, part of the fuel saving achieved by the consumption-optimized working point adjustment of the combustion engine is negated by the greater friction present in

Figure 26.33 Consumption map and influence of selected gear at constant travel speed.

938 | Internal Combustion Engine Handbook

6606_Book.indb 938

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.4 Driver Behavior

the CVT compared with form-fit transmissions. The growing number of travel gears in manual transmissions (six gears) and traditional torque-converter transmissions (up to nine gears) in recent years also enable a convergence to the respective consumption-optimal working point.

26.4 Driver Behavior It is known from the maps for specific consumption against torque and speed that combustion engines achieve their best efficiency only within a narrow range of the overall map. This range is at low-to-average speed and high load, depending on the engine design. The driver should remain within this range. This means the following: •• shifting up at the lowest possible speed •• accelerating at low speeds with high loads (no full load with supercharged engines) •• driving smoothly in the highest possible gear, thinking ahead, and avoiding braking •• using only 70%–80% of the vehicle’s maximum speed •• switching the engine off if stationary for longer periods (sensible after 20 s). The potential savings achievable with low gear shift speeds in daily commuter traffic are shown in Figure 26.34. Consumption can be cut even further with an automatic stop/ start system active when the engine is at operating temperature.

Series-produced systems have been available for a number of years and are being used in an ever-broader range of vehicles. Thereby, as per [26-32], in the NEDC, a consumption benefit of 8% is achieved in the urban phase and 4% in the interurban phase. The use of the integrated crankshaft starter generator will permit the more widespread use of automatic stop/start systems in the future. During the warm-up phase, fuel consumption is particularly high as a result of the high friction caused by the cold lubricating oil and the low component temperatures. In addition, fuel consumption is also increased with the enrichment to improve ride comfort (jolting and poor throttle response) and by heating strategies for the catalytic converter, such as delayed ignition with the gasoline engine. Depending on the initial temperature, the increase in fuel consumption during the warming-up phase may be up to 40%–50% compared with the operating-temperature phase in the NEDC. For driver behavior, this signifies making as few cold starts followed by a short journey as possible, and/or the use of engine preheating systems. Studying Figure 26.35 reveals that the driver of a midclass passenger car stops most often during low-to-medium engine speeds and low loads. Depending on temperament, these values shift to higher engine speeds and high loads up to full load. It is clear that the NEDC only describes a fraction of the true driving behavior and, therefore, only possesses limited predictability with regard to fuel consumption.

Figure 26.34 Consumption saving achieved by reducing the gear shift speeds. According to [26-11].

Internal Combustion Engine Handbook | 939

6606_Book.indb 939

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.35 Frequency of gas pedal position for different drivers and routes.

26.5 CO2 Emissions Emissions are classified by their origin. Flora and fauna, volcanoes, oceans, and lightning are all natural emission sources. The emissions caused by humans (“anthropogenic” emissions) are the result of energy conversion processes, industry, transport, domestic heating, slash-and-burn agriculture, and waste incineration. Locally acting emissions include pollutants governed for motor vehicles, such as carbon monoxide (CO), hydrocarbons (HC), oxides of nitrogen (NOx), and particulates.

Globally active are, above all, carbon dioxide emissions (CO2), which are considered to be responsible along other greenhouse gases for global warming. Total CO2 emissions amount to some 800 Gt/year and originate from causes, as shown in Figure 26.36. Approximately 3.5% of total CO2 emissions are of anthropogenic origin (approximately 28 Gt/a). Road traffic accounts for approximately 11.5% of anthropogenic CO2 emissions.

940 | Internal Combustion Engine Handbook

6606_Book.indb 940

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.5 CO2 Emissions

Figure 26.36 Global CO2 emissions per year and their origins [26-15].

Figure 26.37 CO2 concentration, engine torque, and specific fuel consumption over the air/fuel ratio.

26.5.1 CO2 Emissions and Fuel Consumption

The mass CO2 emissions of a vehicle directly depend on its fuel consumption and can, as stated in [26-16], be calculated using the following formula:

mCO2 =

0.85mKr ⋅ 0.429 CO − 0.866 HC 0.273

where mKr = fuel mass CO and HC = emissions factors for carbon monoxide and unburnt hydrocarbons Consequently, every measure designed to reduce fuel consumption leads to a reduction in CO2 emissions. Thus, a reduction in vehicle weight from 1,500 to 1,300 kg brings about a reduction in CO2 emissions of approximately 20 g/km.

It is apparent from the above equation that fuel composition also has an influence on CO2 emissions. The higher carbon content and greater density of diesel fuel lead to, despite the somewhat higher calorific value, higher volume-based CO2 emissions than gasoline at the same level of consumption. A diesel-engined vehicle emits 26.5 g/km CO2 for every 1 L/100 km of fuel consumption, but a gasoline-engined vehicle emits only 24 g/km CO2. The realized energy as the driving power for a midclass vehicle is equal to 3,000 kJ (measured values from the NEDC rolling tests and pure drive power) for the NEDC registration cycle, which is mandatory in Europe. This would equate to a fuel consumption of approximately 73 g of gasoline for a theoretically approved efficiency of 100% for the entire NEDC or alternatively for a distance-related fuel consumption of

Internal Combustion Engine Handbook | 941

6606_Book.indb 941

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

Figure 26.38 Distances traveled and fuel consumption of passenger and commercial vehicles in Germany from 1990–2009 [26-38].

Figure 26.39 Global CO2 emissions as a result of burning fossil fuels [26-15].

942 | Internal Combustion Engine Handbook

6606_Book.indb 942

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

26.5 CO2 Emissions

0.9 L/100 km and CO2 emissions of 22 g/km. These values cannot be undershot in the NEDC with combustion engines, which do not have energy recovery with the driving resistance laid down. W = 3.000 kJ,

s = 10.891 km,

41.100 kJ 741 g Hu = , rKrst.,Gasoline = , kg l 2.400 g CO 2 mCO2 ,Vol.,Gasoline = l Gasoline



mspec. = =

26.5.2 The Influence of Engine Use on CO2 Emissions

W ⋅ mCO2 ,Vol.,Gasoline H u ⋅rKrst.,Gasoline ⋅ s

3.000 kJ ⋅ 2.400 g CO 2 ⋅ l ⋅ kg 41.100 kJ ⋅741 g ⋅ l ⋅10.891 km

= 21.71 g CO 2/km





Thus, diesel results in 20.45 g CO2/km. If an electric drive vehicle with the same driving resistance and similarly with no energy recovery, but with a 100% approved efficiency is assessed, then, with the current German power station mix (2011: 620 g CO2/kWh) in terms of electricity generation, this results in CO2 emissions of 47.44 g/km. W = 3.000 kJ,

s = 10, 891 km,

mCO2 /kWh = 620 g/kWh



W ⋅ mCO2 /kWh

3.000 kJ ⋅620 g CO 2 = s 10, 891 km ⋅ 3.600 s ⋅ kWh = 47.44 g CO 2 /km

mspec. =

20%. These values (gasoline: 17.37 g CO2/km; diesel: 16.36 g CO2/km; electric drive: 38 g CO2/km) represent the absolute minimum for the respective drive modes within the given boundary conditions.



where W = work s = path Hu = lower mixture calorific value

ρ Krst. = density mCO2,Vol., gasoline = volumetrically based on carbon dioxide mass per liter of completely burnt fuel (gasoline) mCO2/kWh = carbon dioxide mass for each kWh of electricity delivered from the German power station mix, and mspec. is from the determined driving performances calculated from carbon dioxide emissions (see the legal principles on determining fuel consumption and emissions). With the vehicle assessed for driving resistances, the current (2011) series-production values are now 146 g CO2/km for the most economical gasoline engine and 109 g CO2/km for the most economical diesel engine. Electric vehicles in the same segment now emit something like similar quantities of CO2 as diesel drive systems. If 100% recovery of the energy used for acceleration is possible with the assessed vehicle, then the overall energy requirement for the NEDC would be 2,400 kJ. This equates to 80% of the energy used without energy recovery. Likewise with 100% efficiency for drive system and energy recovery, the above-stated CO2 emissions would be reduced by a further

The legally regulated exhaust emissions from a vehicle engine depend inter alia on application factors, such as valve and ignition timing, ignition point, and air/fuel ratio λ . Nonregulated CO2 emissions also manifest dependence on the air/fuel ratio, as shown in Figure 26.37. Maximum CO2 emissions are reached at the stoichiometric air/fuel ratio (λ = 1). Gasoline engines equipped with a regulated three-way catalytic converter operate at this air/ fuel ratio. In addition, CO2 is also produced retrospectively from the CO emissions contained in the exhaust gas after a corresponding period of residence in the atmosphere by the reaction of a portion of the carbon monoxide with oxygen in the air.

26.5.3 The Trend in Global CO2 Emissions

Germany had approximately 50.2 million registered motor vehicles in the year 2010. The combined distance traveled by all these vehicles amounts to approximately 682 billion kilometers each year. This distance has approximately doubled in the last 20 years. This continuous rise has also been accompanied by a corresponding increase in the demand for fuel up to the year 2000 as well as in the amount of CO2 released into the atmosphere (see Figure 26.38). Between 2000 and 2008, the total distances traveled on German roads rose by 30.4%, while at the same time, the resulting CO2 emissions fell by 15.8%, which can be directly equated to a reduction in fuel consumption. The use of hydrogen as an automotive fuel would eliminate CO2 emissions. Hydrogen can be produced from solar or nuclear power or also from biomass. In so-called closed circuits, alcohols are obtained from biomass as the fuel for the vehicle engine. The CO2 emissions caused during the combustion of the alcohols are degraded again by photosynthesis during the biomass’s growth phase. However, the considerable energy input and soil occupation necessary for the generation of the biomass should not be forgotten in the context of these so-called closed circuits. The continued expansion of the global population and increasing industrialization in countries such as China and India are also causing an overall increase in world CO2 emissions that cannot be balanced out by the decreases in the CO2 emissions of existing industrialized nations (see Figure 26.39).

Bibliography

26-1. 1999. Federal Ministry of Transport, Building and Urban Development. Verkehr in Zahlen. Deutscher Verkehrs Verlag Hamburg. 26-2. Breass/Seiffert. 2000. Vieweg Handbuch Kraftfahrzeugtechnik, Friedr. Vieweg & Sohn Verlagsgesellschaft mbH: Braunschweig/Wiesbaden.

Internal Combustion Engine Handbook | 943

6606_Book.indb 943

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 26 Fuel Consumption

26-3. 1999. 20th International Vienna Motor Symposium, VDI-Verlag: Düsseldorf. 26-4. Hucho, W. H. (Ed.). 1994. Aerodynamik des Automobils, 3. Auflage, VDI-Verlag GmbH: Düsseldorf.

26-28. Körfer T. et al., May 5–6, 2011. “Fuel Consumption Potential of the Passenger Car Diesel Engine after EURO 6,” 32nd International Vienna Motor Symposium. 26-29. www.borg-warner.com. Accessed Aug. 29, 06.

26-5. Bargende, M. 2001. “Kraftfahrwesen und Verbrennungsmotoren,” 4. Internationales Stuttgarter Symposium, Expert Verlag: Renningen.

26-30. Dec. 2003, “Friedrich Vieweg & Sohn Verlagsgesellschaft mbH,” in MTZ Motortechnische Zeitschrift, Wiesbaden.

26-6. Essers, U. Ed., 1998. Dieselmotorentechnik 98, Expert Verlag: Renningen.

26-31. Borrmann, D., Davies, M., Friedfeldt, R., Philips, P., Pingen, B., Wirth, M., and Zimmermann, D. Oct. 2005. “Antriebe mit Zukunft,” “Downsizing—Konzepte auf der Basis strahlgeführter DI—Brennverfahren,” MTZ Motortechnische Zeitschrift, Sonderheft. Friedrich Vieweg & Sohn Verlagsgesellschaft mbH: Wiesbaden, pp. 20–25.

26-7. Carstensen, H. 2000. Systematische Untersuchung der Konstruktions- und Betriebsparameter eines Zweiventilmagermotors auf Kraftstoffverbrauch, Schadstoffemission und Maximalleistung. 26-8. 2000. MTZ Motortechnische Zeitschrift 3/2000, Friedrich Vieweg & Sohn Verlagsgesellschaft mbH: Wiesbaden. 26-9. 1997. 18th International Vienna Motor Symposium, VDI-Verlag GmbH: Düsseldorf. 26-10. Robert B. G., Ed. 1998. Gasoline Engine Management, Vieweg: Braunschweig/Wiesbaden. 26-11. Aral A. G., Ed. 1998. Kraftstoffe für Straßenfahrzeuge, Fachreihe Forschung und Technik. 26-12. 1997, “Kurbeltrieb für variable Verdichtung,” in MTZ, Verlag Vieweg: Wiesbaden. 26-13. Fortnagel, M., Doll, G., Kollmann, K., and Weining, H. K. 1998. “Aus Acht macht Vier Die neuen V8-Motoren mit 4,3 und 5 l Hubraum,” in ATZ/ MTZ Jahresband, Verlag Vieweg: Wiesbaden. 26-14. Fortnagel, M., Schommers, J., Clauß, R., Glück, R., Nöll, R., Reckzügel, CH., and Treyz, W. Mai 6, 2000. “Der neue Mercedes-Benz-Zwölfzylindermotor mit Zylinderabschaltung,” MTZ, Verlag Vieweg: Wiesbaden, 61(5):pp. 280–291. 26-15. “VDI Report: Das Auto und die Umwelt,” http://www. ivk.tu.wien.ac.at 26-16. Abthoff, J., Noller, C., and Schuster, H. 1983. Möglichkeiten zur Reduzierung der Schadstoffe von Ottomotoren, Fachbibliothek Daimler-Benz. 26-17. 2001. “Sonderheft 25 Jahre Dieselmotoren von Volkswagen,” in MTZ Motortechnische Zeitschrift, Friedrich Vieweg and Sohn Verlagsgesellschaft mbH: Wiesbaden.

26-32. Golloch, Rainer. 2005. Downsizing bei Verbrennungsmotoren, Springer-Verlag: Berlin, Heidelberg. 26-33. 2007. 28th International Vienna Motor Symposium, VDI-Verlag: Düsseldorf. 26-34. Jan. 30, 2008. Wolf-Heinrich Hucho: Luftwiderstand kostet Treibstoff, VDI Nachrichten. 26-35. Morelli, A., Di Giusto, N. 1999. “A New Step in Automobile Aerodynamics—Performance Improvements and Design Implications,” Int. Conf. Vehicles System Progress, Ed. The Ministry of General and Professional Education, Volgograd, Russia. 26-36. Reissing, J. “Spektroskopische Untersuchung an einem Ottomotor mit Benzin-Direkteinspritzung,” Dissertation, Universität Karlsruhe (TH). 26-37. Briesemann, Sven. 2003. Diploma thesis at the Lehrstuhl Fahrzeugtechnik & -antriebe, TU Cottbus. 26-38. www.autobild.de/artikel/oeco-reifen.html. 26-39. Steinparzer, Unger, Brüner, Kannenberg. 2011. “The New BMW 2.0l 4-Cylinder Engine with TwinPower Turbo Technology,” 32nd International Vienna Motor Symposium. 1, VDI-Verlag GmbH: Düsseldorf. 26-40. 2011. Energy Efficiency—Made in Germany, Federal Ministry for Economics and Technology (BMWi). 26-41. Brünglinghaus, C., Winterhagen, J. Feb. 2011. CO2-Grenzwerte stellen die Weichen, ATZ. 26-42. Schwarzer, C. Nov. 2010. Verbrauchsangeben von Hybridfahrzeugen führen in die Irre, Zeit online.

26-18. 14. Aachener Colloquium. 2005. “Fahrzeug- u. Motorentechnik”, IKA Aachen.

26-43. Wetter, H., Schünemann, R., Schombert, K., Mögn, R., and Oröll, M. Juni 2010. Der Audi A1 Fahrspass, Komfort, Effizienz und CO2-Reduzierung, ATZ extra.

26-19. Wallentowitz, H., Neunzig, D. 2001. Reduzierung der Schadstoffbelastung. IKA Aachen.

26-44. Ammon, D. 2010. “CO2-mindernde Fahrwerk- und Fahrdynamiksysteme,” ATZ, 112(10):pp. 770–775.

26-20. Krebs, R., Szengel, R., Middendorf, H., Fleiß, M., Laumann, A., and Voeltz, S. 2005. “Neuer Ottomotor mit Direkteinspritzung und Doppelaufladung von Volkswagen Teil 1: Konstruktive Gestaltung,” MTZ Motortechnische Zeitschrift, Friedrich Vieweg & Sohn Verlagsgesellschaft mbH: Wiesbaden, 66(11):pp. 844–856.

26-45. Weigand, A., Kastner, O., Atzler, F., Rotondi, R., Juni 10, 2010. “Wenzlawski, K. Effiziente Kombination von AGR, Ladedruck und Einspritzdruck zur Reduzierung von CO2 im Spannungsfeld niedrigster Emissionen,” 5. Emission Control, Dresden.

26-21. Aug. 1999, FEV Spectrum, 12th Ed., FEV Motortechnik GmbH Aachen. 26-22. Blumenröder, K., Buschmann, G., von Essen, C., C. Mehler, O., and Voß, B. 2005. “Antriebe mit Zukunft,” “Der Hybridantrieb—eine Technikoffensive,” MTZ Motortechnische Zeitschrift, Sonderheft, Oktober 2005, S. 27–29, Friedrich Vieweg & Sohn Verlagsgesellschaft mbH: Wiesbaden. 26-23. www.behr.de/produkte/fahrzeug/klimatipps. Accessed Aug. 4, 06. 26-24. www.borg-warner.com. Accessed Aug. 22, 06. 26-25. 2006. VDIK eV; Geschäftsbericht Umwelt. 26-26. Hohenberg, G., Indra, F. Apr. 27–28, 2006. “Theorie und Praxis des Hybridantriebs am Beispiel des Lexus RX 400h,” VDI-Bericht: 27th International Vienna Motor Symposium, Band 1, VDI—Verlag: Düsseldorf. 26-27. Kapus P. et al., May 5–6, 2011. “Cylinder Deactivation with 4 Cylinder Engines—An Alternative to 2 Cylinders?” 32nd International Vienna MotorSymposium.

26-46. Zeiser, M., Düsterdiek, T., and Dorenkamp, R. June 10, 2010. “Strategien zur CO2- Emissionsoptimierung von Diesel-Abgasnachbehandlungssystemen,” 5th Emission Control, Dresden. 26-47. Vogt, C. D., Kattouah, P., Kuki, T., and Furuta, Y. June 10, 2010. “Advanced Particulate Filter for Lower CO2 and Future Emission Standards,” 5th Emission Control, Dresden. 26-48. Adam, F., Eckhoff, S., Richter, J. M., Schmidt, M., and Spurk, P. June 11, 2010. “Katalysatortechnologien für Stöchiometrisch und Mager Betriebene Ottomotoren mit dem Ziel der CO2-Reduktion,” 5th Emission Control, Dresden. 26-49. Reulein, C., Schwarz, C. 2010 “Senkung der CO2-Emissionen durch Ladungswechsel,” MTZ, Wiesbaden, 71(11): pp. 760–765. 26-50. Lauxmann, R. June 2011. “Durch Bremsen weiterkommen und CO2 reduzieren—Entwicklung neuer Bremssysteme bei der Robert Bosch GmbH,” 2nd International Munich Chassis Symposium, Munich.

944 | Internal Combustion Engine Handbook

6606_Book.indb 944

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27 Noise Emissions Anybody who has ever experienced a motor vehicle with a rigidly mounted engine or intake noise with no muffler, not to mention “naked” exhaust noise, will be in no doubt that of the many subdivisions of vehicle acoustics, and engine acoustics was the first, and for a long time remained the most important. Passengers’ demands for greater comfort and the concerns of the general public, represented by the law, and drove developments in the engine acoustics field, which has now achieved an extremely high level of sophistication despite the enormous increases in engine output. Scarcely anybody is now surprised at having to glance at the revolution counter when the car is stationary to check that the engine is really still running. On the road, engine noise has been suppressed to such an extent that other sources such as tire and wind noise achieve equal and even dominant levels. A further indication of thorough mastery of engine acoustics can be found in the fact that it has now been possible for many years to conceive of “sound design.” Whereas the engine acoustics specialists’ early tasks concentrated on the suppression of elemental exterior noise and vibration problems, aiming to find solutions in improvements in mufflers, internal engine mass balancing, and flexible mounting, the present-day field of engine acoustics has become significantly more diverse. This is true of both the nature of the noise sources involved, and we may use the classifications: secondary noise radiation, generator noises, and timing mechanism noises, as well and the working methodology employed by engineers, which we can classify into transfer path analysis, noise intensity measurement, holography, vibrometry, dummy-head technology, finite element method (FEM), boundary element method (BEM), and statistical energy analysis (SEA). Taken in conjunction with the enormous range of nonacoustic requirements (fuel consumption, emissions, heat balance, costs, package, etc.), the result is a complexity expressed within the manufacturing companies in the form of correspondingly large work groups and within supplier companies in the form of a high level of specialization. It

is only necessary in this context to think, for example, of components such as exhaust systems, enclosure components, engine bearings, and multiple-mass flywheels. All these branches, however, work on the basis of the same physical principles and methods, and utilize the same basic concepts of physical acoustics and psychoacoustics. We, therefore, summarize the essential concepts and terms in the following section, before going on to examine individual topics. Chapter 27.7 contains a short discussion of widely used analytical methods.

27.1 Basic Physical Principles and Terms Although the term “engine acoustics” does not really make it clear, it is not only “audible” phenomena that play a role here but also the vibrations, which can be felt by the occupants of vehicles and which, as in the case of so-called idling shudder, can reach extremely low frequencies. So-called solid-borne noise is also important because only a very small range of noises actually originate as airborne noise (such as exhaust muzzle noise), most are instead generated in the form of vibration in solid bodies and then radiated by oscillating surfaces (such as inertial forces, gas forces, and toothing forces) and/or need to pass through the body shell in the form of airborne noise on their way into the interior of the vehicle. If we think of piston canting noise, for example, this too in many cases has to traverse at least a short distance in the form of liquid-borne noise (in liquid-cooled engines) on its transmission route. This stage presents a considerable barrier to noise propagation, because neither liquids nor gases are capable of absorbing shear stresses. A significant difference compared with airborne noise can, however, be found in the considerably higher characteristic acoustic impedance (wave impedance), which signifies considerably greater coupling with the solid-borne sound perpendicular to the surface and results,

Internal Combustion Engine Handbook | 945

6606_Book.indb 945

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

inter alia, in it not being possible, because of the significant interactions that occur, to examine solid-borne noise and liquid-borne noise in isolation from one another. The most widely used parameter for describing solid-borne noise is acceleration, which is “popular” as a result of the relative simplicity of measuring it. It must, however, be noted that, unlike airborne sound pressure, acceleration is a directional variable requiring measurements in three dimensions from a single point. In general, solid-borne noise is significantly more complex and diverse to measure than airborne sound because, as a result of the ability of solid bodies to also absorb shearing stresses, they generate a large range of different sound wave propagation forms (many of them simultaneously). The following can be mentioned as examples: dilatational waves (valve stem), flexural waves (oil pan), and torsional waves (crankshaft and camshaft). Because of the need not to influence the vibration system, as a consequence of constricted space or other restricting boundary conditions (temperature, pressure, tightness, etc.), and in particular, because of the desire to acquire information on the prevailing forces, a range of other measured variables such as noncontact path measurement (rotating and thin-shelled components) and measurement of strains (crankcase), pressure distributions (bearing seatings), and forces (engine bearings) are also used alongside acceleration. Another physical phenomenon, rotational oscillation, is also of great importance, particularly in the field of engine and powertrain acoustics. The basic measured variable in this context is generally angular velocity, which can be determined from discrete angle impulses (gearwheels and incremental encoders). Rotational acceleration can, if required, then be determined by differentiation. Compared with solid-borne noise, the recording and quantification of airborne noise, both in the vehicle interior and in the case of exterior noise, is initially relatively easy because space and temperature problems are encountered only rarely and the variable relevant to the human ear (i.e., sound pressure) can be measured directly using microphones. Sound pressure p is the term used to describe the amplitude of pressure fluctuation around static air pressure, with pressure amplitude cycles of 3 Pa, for example, being perceived as extremely loud (for comparison: 1 bar = 105 Pa). Although sound pressure is generally adequate for a description of airborne sound output, the most suitable parameter for quantification of airborne sound radiation (emission) is acoustic power P. This is the total sound wave power penetrating through an imaginary envelope and is calculated from the integral of acoustic intensity I across the envelope s:

P = ∫ I ⋅d s

(27.1)

S

Acoustic intensity represents mean energy transport per unit of area. It is a vector factor aligned parallel to the vector of acoustic velocity v and is calculated from

I = p(t)⋅ v(t)

(time range mean) or from

(27.2)

I=



1 ⋅ Re { p! v! ∗ } 2

(27.3)

(frequency range) analog to mechanical power P = Fv. In this context, acoustic velocity is the velocity of the local oscillating motion of the air particles. The vector character of acoustic intensity can be exploited with measurement technology by locating sources in complicated sound fields and in the measurement of acoustic power, which also becomes possible in a reflecting environment. Evaluation of noise signals is generally accomplished with two criteria. These are the most precise possible evaluations in terms of subjective human perception (we shall restrict our attention here to airborne noise) and in terms of the most efficient extraction of information possible about noise generation and the transmission route. Concerning subjective perception, it is necessary first to note that the human ear is capable of registering acoustic pressures across a range of orders of magnitude of several decimal powers. For this reason, depiction of sound levels on a logarithmic dB scale has become established in acoustics, not only for sound pressure (sound pressure level) but also, for example, for solid-borne noise variables, and a variable proportional to energy is defined in every case ⎛ x2 ⎞ ⎛ x⎞ Lx = 10 ⋅ log 10 ⋅ ⎜ 2 ⎟ dB = 20 ⋅ log 10 ⋅ ⎜ ⎟ dB ⎝ x0 ⎠ ⎝ x0 ⎠ or





(27.4)

⎛ X⎞ LX = 10 ⋅ log 10 ⋅ ⎜ ⎟ dB ⎝ X0 ⎠

in which x is a field variable (e.g., sound pressure and sound acceleration), X is an energy variable (e.g., acoustic intensity and acoustic power), and x0 and X0 are their reference variables. In the case of sound pressure, p0 = 2 ⋅ 10 –5 Pa (root mean square value), in the case of acoustic power, P0 = 10 –12 W, and in the case of acoustic intensity, I0 = 10 –12 W m–2. Human hearing is not only nonlinear but also frequency dependent. Its sensitivity declines significantly as low and extremely high frequencies are approached. The fluctuation range becomes greater as the absolute sound pressure decreases. The various frequency components are therefore simplified with defined weighting curves (DIN IEC 651, Curve A for low, Curve B for medium, and Curve C for high sound volumes), before being consolidated to an overall level, which is then correspondingly indexed [e.g., dB (A)]. Weighted levels are used as the basis for a whole series of legal regulations because of their simplicity, but are suitable neither for diagnosis purposes nor for obtainment of information concerning the “quality” of a noise (see Chapters 27.9 and 27.10), because they do not contain any information on the spectral or temporal structure of the noise. Readily understandable designations that are associated with certain mechanisms, while at the same time provide implicit information on the frequency range involved or on the chronological structure of noise, have become established in the automotive field as the conceptual basis for recurring

946 | Internal Combustion Engine Handbook

6606_Book.indb 946

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.1 Basic Physical Principles and Terms

acoustic phenomena. Figure 27.1 shows a number of typical examples concerning the engine. An essential basis for the analysis of noises is provided, initially, by their spectrum, that is, the classification of a signal into its frequency components. In practice, spectra are determined from digital signals using the so-called fast Fourier transformation, a highly efficient variant of the digital Fourier transformation. This generally supplies narrow-band spectra, that is, a relatively high-frequency resolution, making it possible to obtain an approximate calculation of the classical “coarser” subdivisions into octave (doubling of frequency) and third-octave spectra if required. Only high-frequency resolution permits detailed analysis of the noise. Sinusoidal noise components attributable to inertial forces in the crankshaft drive and therefore possessing defined frequencies appear in the form of narrow peaks in the spectrum (linear spectrum). Such deterministic frequency components, which “wander” proportionally to speed, are referred to as “orders”. A typical example is the second order of the crankshaft in an inline four-cylinder engine that dominates as a result of the unbalanced inertial forces. The fundamental frequency, or first order, can be completely balanced out in a four-cylinder engine. Where the dominance of an order is known from the outset, a so-called “order filter” is used in many cases to examine only the level of this order. The higher orders, which significantly influence sound, also appear in the spectrum, however. Modulations, that is, fluctuations in amplitude (or beat) or frequency significant enough for perception as noise or nuisance, also become visible in the

Noise sources Inertial forces Crankshaft drive, valve train

Booming Humming Humming

Combustion forces

(rot. oscillation)

Control/ Auxiliary units drive (chain, synchronous belt, V-belt ribs)

Roughness

Roughness

Combustion noise (load), knocking

Drumming (belt), screeching

Impact of the valves

"Mechanical" engine noise, clanging

Hydraulic pressure pulsation

Humming

Pulsed-air noises, intake/exhaust system, fan

Booming Humming

Sägen

Heulen

Flow noises, intake/exhaust system, fan

Trickling, quavering. Whirring, hissing 0 0 20

Transmission route into the passenger cell

form of sidebands. These are peaks that occur at the interval of the modulation frequency next to the midband frequencies. An internal combustion engine is a diverse source of noise. If we leave aside the auxiliary units for the time being, not only the so-called aeropulsive and aeroacoustic sources, such as exhaust muzzle noise, intake noise, and fan noise, but also the noise radiation of the oscillating surface of the engine and attachment components are responsible for exterior noise. Their effectiveness can be characterized in the so-called degree of radiation, which is the ratio of actually radiated acoustic power P of a surface S to that of a large (significantly larger than the sound wave length) concurrent-phase oscillating plate with the same mean quadratic speed. If the interior zones oscillating in counterphase are significantly larger than the airwave length, the degree of radiation will be close to 1. The converse case is considerably more difficult to handle. Simplified, the radiated acoustic power is smaller as the number of counterphase zones that are present become larger and more densely located relative to one another (hydrodynamic short circuit). Once the airborne noise has been generated, it can be countered with insulation (energy reflection) and/or attenuation (energy dissipation), with at least a small amount of attenuation always being necessary. Insulation provisions (such as encapsulation) and combined provisions, such as mufflers, are evaluated specifically by degree of transmission or, in absolute terms, by the insertion insulation increment, and the difference in level prior to and following the interposition of a provision

500

Solid-borne noise

1000

Frequency [Hz]

10000

Airborne noise

Figure 27.1 Examples of noise sources.

Internal Combustion Engine Handbook | 947

6606_Book.indb 947

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions



De = Lo.D. − Lm.D

(27.5)

while the characterizing variable for purely airborne noise attenuating provisions, such as noise-absorbing linings and claddings, is their degree of absorption, that is, the ratio of absorbed intensity to incident intensity

a=

Iabsorb . I incident

(27.6)

When interior noise is examined, a further component, actually dominant in the frequency range up to about 500 Hz, is added to the sources already mentioned, namely, transmission of solid-borne engine noise via the body shell into the vehicle interior. In addition to the drive shafts relevant, in particular, in front-wheel drive vehicles, the essential transmission points are in this case the engine mountings. One development target of engine acoustics is, therefore, that of minimizing engine-side solid-borne noise amplitudes at the support points, because the degree of isolation of vibration achievable via the rubber bearings is limited. On the body shell side, the corresponding development goal is minimization of “sensitivity” at the coupling points, which is quantified by the so-called acoustic transmission function [27-3]. This is the frequency-dependent ratio of sound pressure on an interior microphone and dynamic force at the point of excitation:



p!j H! ij = ! Fi

(27.7)

It includes the complete transmission path, including radiation into the interior, absorption there, and the influence of resonance effects resulting from cavities in the passenger cell. So-called input inertance, on the other hand, discloses local weak points in the body shell in the form of the ratio of vibrational acceleration at the point of action of a force in the direction of the force and the force acting [27-12], [27-13], [27-16].

a! I! = ! F

(27.8)

27.2 Legal Provisions Concerning Emitted Noise 27.2.1 Emitted Noise Measuring Procedure

While the level of noise and comfort in the vehicle interior remains a matter of market forces, the operating noise emitted into the environment (“emitted noise”) became subject to official regulation at a very early stage. The measuring procedure applied is internationally standardized in DIN ISO 362. The measuring apparatus used is the product of measuring technology at the time of its origin. The simple measuring apparatus has the advantage that this measurement can be performed identically and with little cost and complexity in all countries of the world. The procedure, in principle,

is shown in Figure 27.2: At a constant 50 km/h, the vehicle approaches a 20-m long measuring section at the start of which the throttle valve is opened suddenly. The noise level during the subsequent, approximately 1.5 s in duration, fullthrottle acceleration process (hence, the designation “accelerated drive past”) is measured in each case by one microphone at a distance of 7.5 m from the center of the track. The maximum dB (A) value achieved during this procedure is the result of the measurement for this gear. Because engine noise increases disproportionately with speed, the result is primarily dependent on the speed level achieved in this short period of time. The gears to be driven are therefore precisely determined: vehicles with manual transmissions are usually measured in the second or the third gear and the result is averaged; vehicles with automatic transmission travel in drive position “D” without using the kickdown.

27.2.2 Critical Evaluation of the Meaningfulness of the Existing Emission Noise Measuring Procedure

The disadvantages of the method now used have been known for some time: a segment of only approximately 1.5 s at a certain vehicle operating state can describe its contribution to traffic noise made under practical road conditions only imperfectly. Over the course of the last few decades, attempts have therefore been made to depict the changes in the actual vehicle collective and its driving conditions using special rules: special rules for recreation vehicles, passenger car diesel engines with direct injection, and sports utility vehicles were integrated into the basic procedure. The high speed dependency and “full-throttle” driving condition together describe, however, a completely unrealistic misuse test that deviates greatly from today’s standard average driving behavior in urban traffic. Because the torque characteristics of engines (particularly diesel engines) have changed considerably in the past few years and as a result, vehicles are driven at significantly lower revolutions, it was time to develop a more practice-orientated measuring procedure, which reflects technical vehicle improvements not only in the drivetrain and tires but also in actual reductions to the traffic noise level. Because of a measuring procedure lacking in practical relevance, the reduction in the nominal type-test values performed by the automobile industry at great expense has led only to a minimal relief of the public in terms of traffic noise. Besides nonrepresentative engine speeds, the reasons for this lies in the fact that the tire/road surface noise plays only a minor role in the currently valid noise measuring procedure.

27.2.3 Future Emission Noise Measuring Procedure and Limit Values

Efforts running since the 1990s to develop a “new emission noise measuring procedure” were derived not only from both the considerations of real-life driving styles but also from the technical necessity to give greater consideration to the

948 | Internal Combustion Engine Handbook

6606_Book.indb 948

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.3 Sources of Emitted Noise

A¢

10 m

10 m

B¢

C¢

C

A

7,5 m Microphone

tire/road surface noise. The objective was to depict “actual driving behavior in urban traffic” as realistically as possible. In addition to the actual measuring procedure itself, global discussions are now focusing on “vehicle classification” and associated “limit values.” The establishment of a new measuring procedure as well as the definition of new vehicle classes and limit values is highly anticipated sometime after 2010. The new measuring procedure is drafted on the established measuring section layout. A realistic “partial load” shall be simulated and is therefore composed of “accelerated journeys” (full load, 50 km/h at the height of the microphone) and “constant journeys” (50 km/h). The accelerations and transmission gears to be driven are determined by the vehicle’s “power-to-weight ratio,” all predetermined values represent the result of comprehensive series of measurements and statistical evaluation of the data contained. With trucks, the driving conditions (exit speed, engine speed) are defined at the end of the measuring section instead of at the beginning as previously. Furthermore, realistic loads are also intended. To test the capacity of the new emission noise measuring procedure, a two-year “monitoring phase” was decided by both the Economic Commission for Europe (ECE) and the European Commission (EC) of the European Union Commission: all typetested new vehicle types released between Jul 01, 2007 and Jun 30, 2009 (ECE) and between Jul 01, 2008 and Jun 30, 2010 (EC) must be measured according to both the previous procedure and the new one. The new limit values and vehicle classes will be derived from the data pool available after the tests. Initial measurements reveal that most vehicles exhibit values that are 2–3 [dB(A)] lower than the previous measuring procedure, which is down to the lower engine speed level and addition of constant journeys. For the vehicle manufacturer, however, this makes it significantly harder to still achieve substantial reductions. This is most likely regarding the reduction of the tire/road surface noise, the “emission noise” property of a tire will have an ever diminishing priority compared with the safety aspects (wet grip, braking, and fast mode capability). Because the overall noise of a vehicle is made up of the drivetrain noise and tires/road surface noise, the total from pure rolling noise of the tire [approximately 65–69 dB(A)] as well as the additional load proportion of the tire from slipping while accelerating (plus approximately1–3 dB) represents the “lower limit for optimizations” of the overall vehicle and subsequently the acoustic power.

B

Figure 27.2 Measurement of emitted noise [27-14].

Furthermore, within the framework of a “General Safety Directive” from the European Union Commission, significantly stronger limit values will be specified for the rolling noise of the tires during their component type-test (rolling at 80 km/h for passenger cars or alternatively 70 km/h for trucks). The values which are then available for the vehicletype test for the tire/road surface noise can be derived from these. The development of quieter tires was previously almost exclusively driven by vehicle manufacturers (particularly of premium vehicles) and their in-house specifications for stricter noise limit values compelling tire manufacturers to make improvements. In addition, at the starting point of the new measuring procedure, a newly defined test track ISO 10844 “new” shall also be reckoned with, which represents the road surfacing used on European roads more than previously. Moreover, the additional provisions (additional sound emission provisions) are contained in the new regulation ECE R51, which prevent the circumvention of the sensible design of the provisions by means of active systems or electronic methods by the vehicle manufacturer. This additional measuring procedure was still in discussion at the time that this book went to print. It remains to be seen whether countries such as the USA, Japan, India, China, or South Korea will join the new ECE regulation. Preparation of a Global Technical Regulation within the framework of the UNECE is planned.

27.3 Sources of Emitted Noise Depending on vehicle speed, the components making up traffic noise can, above all, be traced to two sources, the engine and the tires. Engine noise dominates at low speeds and road-surface/tire noise at interstate speeds. Wind noises, on the other hand, can be ignored. A range of physically diverse effects contribute to engine emitted noise: •• Intake and exhaust muzzle noise. The pressure pulsation excited in the intake and exhaust systems by the gas-charging sequences in the engine result in direct emission of airborne sound waves at the open end of the pipe system (the “muzzle”). Within the spectrum, ignition frequency for the 2nd order in the case of four-cylinder engines and the 3rd order in the case of six-cylinder engines dominates. High-frequency,

Internal Combustion Engine Handbook | 949

6606_Book.indb 949

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

broadband flow noises resulting from high flow velocity at the end tube are also achieved at high engine speeds and full-load operation. •• Secondary radiation from intake and exhaust systems. The pressure pulsations in the interior of the system also cause the pipe and wall surfaces to oscillate, with the result that they also emit airborne noise to the exterior; this then is referred to as secondary airborne noise. •• Noise radiation from the engine structure, that is, the external surfaces of the engine and the transmission unit. Combustion pressures and all irregularly moving components in the engine, transmission, and auxiliary units result in dynamic forces acting on the housing structure and, thus, in motions and deformations of the outer walls, which then become the originating point for emission of pressure waves into the ambient air, that is, radiation of noise. The spectrum of this airborne noise is dominated by high-frequency components above 500 Hz, which are subjectively perceived as mechanical engine noise. The components of mechanical engine noise are as follows: •• pressure increases during compression and combustion, particularly with diesel engines •• the impact of the valve disks on the valve seat •• piston slapping •• valve train gearwheel oscillations •• chatter of the floating wheels in manual transmissions (“gear rattle”) •• pressure surges from the hydraulic pumps.

27.4 Emitted NoiseReduction Provisions 27.4.1 Engine Approaches

Action taken at source, that is, an engine design that produces the least possible noise, is always the most consistent and the most efficient provision [27-2]. The design target must be: •• Maximum stiffness in the force-transmitting housing structure and crankshaft drive, in order that excitation of housing-wall-surface oscillation is minimized. •• Oscillation-reducing design of engine exterior walls by •• ribs or fins (highly dynamic stiffness, for example, engine block and transmission housing) •• isolation, for example, flexible fixing of the cylinder-head cover or intake manifold •• damping, for example, a sheet-metal oil pan. All the listed alternatives conflict with the aims of cost optimization and functional requirements. Ribs and fins: extra weight and space requirement; isolation: Oil tightness, fixing of attachments, and auxiliary units; and damping: heat removal and extra weight. In the case of the oil pan, the ribbed aluminum oil pan has, in particular, become

established over the damped sheet-metal oil pan in aluminum engines, because it is necessary as a load-bearing structural element to increase the stiffness of the engine plus transmission unit (see also Chapter 27.6). •• Noise-absorbing attachment shells as components of a “tight fitting encapsulation” are considered secondary provisions that prevent the radiation of noise. The cylinder-head claddings, which are installed primarily for engine-compartment styling reasons (enclosure of cables and injection lines), also perform an acoustic function as in-mold skinned plastic shells [27-1], [27-8]. •• Softer combustion is the name assigned to the task in the field of mixture formation, particularly in the case of diesel engines, to significantly reduce the subjectively perceived nuisance quality of combustion noise and improve both noise emissions and interior noise levels. •• High-volume intake and exhaust mufflers primarily constitute demands made in terms of space and cost on the overall vehicle conception. Muzzle noise must not supply any measurable contribution to total emitted noise if presentday limits are to be met. This provision is deemed to be achieved if an additional muffler, designated an absolute muffler, achieves no further reduction in the dB (A) level in measurement of emitted noise. In the case of interior noise, however, an audible exhaust system contribution is targeted for the most pleasant engine noise possible, particularly during full-throttle acceleration. This is achieved, inter alia, with the support of analytical and simulation sound design methods during the development phase. Complex solutions involving exhaust throttle blades that prevent low-frequency booming at low engine speeds and reduce flow noises (and exhaust gas back pressure simultaneously) by opening an additional cross section at high engine speeds and exhaust flows are the result of such development activities. •• Reduction of engine speed, one of the most effective provisions for noise reduction, is also a matter of overall vehicle conception. In practical on-road operation, it achieves low noise emissions only if a longer transmission ratio is combined with high engine torque in the low rev range and, therefore, also offers acceptable driving characteristics. •• Noise-optimized auxiliary units, such as radiator fans and generators, already play a role in achieving the low limits imposed for passenger cars, despite their relatively small contribution to overall noise.

27.4.2 Vehicle Approaches

Secondary provisions implemented on the body shell to reduce engine noise emissions are referred to as “spaced” or “offengine enclosures”. The objective is that of using additional components installed in or on the body shell to make the engine compartment a largely sealed space from which only a small amount of engine noise can escape to the exterior. In addition, the peripheries are lined with noise-absorbing materials, such as foam or nonwoven cotton fabric (which, particularly in the lower engine compartment zones, require

950 | Internal Combustion Engine Handbook

6606_Book.indb 950

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.4 Emitted Noise-Reduction Provisions

protection against absorption of oil and moisture, and must be of noncombustible type) to reduce the high noise level in the engine compartment. Practically all series-produced automobiles and those with diesel engines are equipped with a combination of the following enclosure elements (see Figure 27.3): •• Absorbent engine hood lining consisting of foam, nonwoven fiber fabric, of a cassette-module type in some cases, spaced from the sheet metal, in order that panel resonator effects achieve greater absorption at low frequencies, or of honeycomb type, with the effect of Helmholtz resonators. •• Undershield: This term identifies a plastic or metal shell that encloses the engine compartment at the bottom and is required even for reasons of aerodynamic drag alone. Frequently featuring a cutout for the oil pan, to maintain ground clearance, it normally ends shortly before the firewall. In special cases, the front section of the transmission tunnel in acoustically critical vehicles is also closed on its underside,

provided a solution can be found for the cooling problems that then arise. The undershield too features noise-absorbing lining on its inner side with safety against absorption of oil, fuel, and so on, having priority over the best possible noise-absorbing properties. For this reason, skinned foam or cassette-type absorbers are used here instead of open-cell foam materials, which would be desirable from an acoustic viewpoint. •• Closure of the sides of penetrations to the wheel housing by rubber bellows for the track rods and, if necessary, provision of tunnel-like absorption sections lined with foam for the front-wheel drive shafts. •• Closure of cooling-air inlets at the front by, for example, thermally controlled slats; this is an extremely complex provision that is nonetheless practiced in top end diesel engine vehicles. At cold starting, the slats are closed, reducing the diesel engine’s cold “knocking” noise. For functional reliability reasons (freezing in winter), such systems can

Engine hood absorber Material structure – Covering non-woven fabric – Non-woven cotton fabric, pressed

Undershield Material structure Support – PP-injection molding with 40 % glass (short fibers) Absorber – Cassettes made from PU foam, pounded

Figure 27.3 Engine enclosure [HP-Chemie Pelzer].

Internal Combustion Engine Handbook | 951

6606_Book.indb 951

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

be used only behind the radiator; that means there must be sufficient space between the radiator and the engine. The reduction in emitted noise achievable with an engine enclosure is restricted by the size and number of the penetrations necessary for cooling the vehicle. The development of a complete engine enclosure, therefore, is more of a cooling problem than an acoustics one. Vehicle studies incorporating completely enclosed engine compartments achieve spectacular publicity effects, but are generally a long way off surviving a journey through Alpine passes with a trailer or hot weather tests. It is necessary to differentiate between the following in the context of numerical data in dB (A) quantifying the measured noise reduction achieved with an enclosure: •• dB (A) value in reduction of radiated engine noise •• Reduction of emitted noise emission factor, that is, total vehicle noise in the legally prescribed drive past test. For example, in the drive-past test, 74.8 dB (A) is measured for a passenger car, which is the result of the sum of the energy contents of 70 dB (A) tire noise and 73 dB (A) engine noise. A sophisticated engine enclosure would permit a reduction in acoustic energy radiated by the engine to a half, that is, to an acoustic pressure level of 70 dB (A). Tire noise and engine noise of 70 dB (A) each then produce a consolidated level of 73 dB (A). The reduction in emitted noise emission value achieved by the encapsulation is therefore 1.8 dB (A).

27.5 Engine Noise in the Vehicle Interior Although the provisions described in Chapter 27.4.1 also serve to reduce interior noise, the solid-borne noise paths, which dominate interior noise in the lower frequency range, must be added here, which is not the case with emitted noise. All mechanical links between the engine and transmission unit and the body shell constitute potential transmission routes for solid-borne noise, particularly the engine mountings and the drive shafts, which in the case of front-wheel drive vehicles are connected with no intervening insulating elements to the suspension system from where the solid-borne noise can be transmitted via relatively stiff suspension attachment members into the body shell structure. Only the higher-frequency noise components above 500 Hz, so-called mechanical engine noise and combustion noises (Figure 27.1), are transmitted via the airborne noise route from the engine compartment via the firewall and floor plates into the passenger cell. The origins of solid-borne noise oscillations can be found primarily in oscillating inertial forces, because the lowfrequency buzzing and booming noises are present in three to five-cylinder engines (even when the engine is idling). Gas forces, the main causes of rotational irregularity of the crankshaft drive and the counterphase external reaction of the engine and transmission unit in the form of a rotational oscillation about the principal axis of inertia in the crankshaft direction, are the second source of solid-borne noise. Their

contribution to overall noise can be easily differentiated as a result of its nondependence on engine load and the increase in the degree of irregularity at low engine revolutions. Excitation caused by free inertial forces and torques can be controlled only by mass balancing, in terms of number and arrangement of cylinders or additional balance shafts, rotating at crankshaft speed (e.g., balancing of 1st order moments of inertial torque in three-cylinder, five-cylinder, and V6 engines) or at twice crankshaft speed (e.g., “Lancaster balancing” of 2nd order forces in the four-cylinder inline engine). Acoustic improvement is confronted with higher costs and friction losses. Reduction of inertial forces by lower piston and connecting rod masses is possible in theory, but these potential measures have in most cases already been exhausted. The higher orders (>2nd order) are generally no longer the subject of mass balancing studies, because their excitation is significantly slighter and, in particular, because the precondition for mass balancing rigid body behavior in the crankshaft drive and crankshaft housing is no longer fulfilled at frequencies of above 250 Hz. In interior noise, the contribution from the higher orders is noticeable in the form of “roughness. XE ”” in the engine noise and is a problem in the field of passenger car engines of long-stroke (empirical value H > approximately 80 mm) and high-torque crank gears with a greater conrod ratio λ , because the higher orders increase disproportionately with λ . If two or more orders of approximately the same size are adjacent in the spectrum, for instance, 4th, 4.5th, and 5th, their superimposition results in a modulation, a surging fluctuation in noise level, which is perceived as an unpleasant noise characteristic. This “hammering” or “crankshaft rumbling” arises during full throttle operation because the “half orders” originate from combustion (ignition frequency of the individual cylinder in a four-stroke engine). The second source of solid-borne noise, cyclic irregularity of torque delivery, causes buzzing at ignition frequency (2nd order in four-cylinder engines, 3rd order in six-cylinder engines, etc.), particularly in the lower engine speed range, associated with occasionally severe vibration. Because of physical properties, and therefore not subject to influence by the designer, this problem is linked to torque at low speeds; that is, the better the engine’s torque characteristic, the greater the problem of vibration and buzzing. This was for a long time the reason why direct injection diesel engines were only used in utility vehicles. Only years of development effort, including activity devoted to insulation of vibration, resulted in the breakthrough in passenger cars. The rotational oscillations of the crankshaft itself can be largely isolated from the drive shaft by installing a heavy or dual-mass flywheel (Figure 27.4). The reaction forces acting on the crankcase remain unaffected; however, it is the rotational oscillations of the housing that make this problem so difficult to solve. The inertial forces of the valve train and intrinsic timing system noise tend to play more of a role in four-cylinder engines in top-end vehicles, because they impair the engine’s acoustic pattern in a way similar to the noises from auxiliary units (generator “whistling”, “sawing” from hydraulic pumps,

952 | Internal Combustion Engine Handbook

6606_Book.indb 952

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.6 Acoustic Guidelines for the Engine Designer

Figure 27.4 Two-mass flywheel [LuK].

etc.). The inertial forces of the valve train occur as a reaction to the acceleration of the valves, tappets, rocker arms, and so on, and the results are, in principle, the same as for the inertial forces resulting from piston movement. Although they are one order of magnitude lower, they are perfectly capable of codetermining the hum noise level in engines with good mass balancing, that is, six and eight-cylinder engines. The noise of the timing system, referred to as “timing chain screaming” or “synchronous belt screaming,” is located in the mediumfrequency range corresponding to the frequency of tooth engagement of the chain sprockets and belt pulleys. Similar to the situation in the case of gearwheels, it is the result of the periodic subjection of the teeth or the chain links to load and subsequent relief, and in the case of synchronous belts and of expulsion of air upon tooth engagement (air pumping). These screaming noises are audible particularly during idling phases and at lower engine speeds, where they are not yet screened by the increasing level of the other mechanical and combustion noises produced by the engine. In addition, lowfrequency noises may also be generated by string oscillations from chains or synchronous belts [27-11].

27.6 Acoustic Guidelines for the Engine Designer The question of how the objective of achieving low noise generation can be incorporated at the engine design phase has already been the subject of many studies. As far as it is possible without knowledge of a particular design to provide any universally applicable guidelines, these tend toward maximum stiffness. This can be justified in physical terms by

the fact that the deformations responsible for noise radiation and solid-borne noise transmission are, given identical forces, reduced and by the fact that the structural resonances are shifted toward higher frequencies where the amplitudes of the dynamic forces causing excitation become smaller. The latter starts with the first bending mode of the engine and transmission unit, which in the case of a four-cylinder engine must be shifted under all circumstances into the frequency range significantly above the strongest oscillation excitation of the second engine order, that is, ³ approximately 250 Hz in the case of a gasoline engine. The stiffness “weak point” is generally the bolted engine housing flange/clutch bell housing or torque converter bell housing joint, particularly if a nonload-bearing sheet-metal oil pan or a “short skirt” (engine side walls not extended down beyond the main bearing pedestals) prevents transmission of forces below the axis of the crankshaft. This can be remedied by using a die-cast aluminum oil pan with corresponding ribbing and ribs or fins on the bell housing. The general aim is to achieve the greatest level of “straight-line flow of forces,” that is, any indentation or projection of the housing reduces the achievable stiffness. In the case of six or eight-cylinder engines with no 2nd engine order excitation forces, a natural bending frequency above the 1st order, ³ that is, approximately 120 Hz is sufficient. The greater masses of such generally high-capacity engines are lower the natural frequency to such an extent that it is, nonetheless, necessary to design for high bending stiffness, particularly with the engine arranged longitudinally and with long all-wheel-drive transmission systems. Static stiffness analyses or calculations are still sufficient for these ultralow oscillation modes, because the masses of the housing walls have only a small influence on the oscillation mode. The first natural torsion frequency of a standard passenger car engine generally lies above the most powerful rotational oscillation excitation frequency. In special cases, however, this natural oscillation mode can also cause noise problems and necessitate structural stiffening provisions. Here too, projections in the bell housing are typical points of low stiffness. Housing wall natural oscillations play an increasing role in the frequency range above around 500 Hz and can be reflected in the form of “hot spots” in the engine or transmission noise radiation. Such noise problems are so particular to the individual design that no precise recommendation, for ribs/ fins or damping, can be made without measuring-technology analysis or simulation. Larger flat, thin-walled zones should be avoided from the outset, however, by cambering, ribbing, beading, divider walls, and so on. Maximum dynamic stiffness of the engine mountings, to which the flexible engine bearings are fixed, is a further categorical design aim. They should be regarded as cantilever beams with additional masses at the free end (covibrational engine bearing mass), the resonances of which increase the solid-borne engine noise transmitted via them into the body shell by as much as a power of ten. The designer should, therefore, always attempt to achieve more than 1000 Hz for the first natural support frequency, because the engine’s excitation of

Internal Combustion Engine Handbook | 953

6606_Book.indb 953

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

solid-borne noise is no longer dominant. This can be achieved in practice only if the mountings 1. are short; they do not project more than 100–150 mm beyond the engine wall 2. possess an adequate bolt-connecting base on the engine wall (square with approximately the mounting length as the side dimension), which must be correspondingly stiff 3. and take the form of closed tapering hollow-section beams. For this reason, sheet metal supports with an open sectional cross section are scarcely used as engine mountings in modern passenger cars; they are being replaced by die-cast aluminum mountings, in which Figure 27.5 shows an example. Engineend mounting length is a matter of the arrangement of the engine mounting in the overall vehicle concept, in which a whole range of functional requirements must be fulfilled. If, then, it is necessary to select between longer body-shell-side engine mounting brackets and longer engine-side mountings, where the shorter engine-side mounting generally offers advantages from an acoustic viewpoint.

Here again, the aim under normal circumstances must be that of locating this resonance frequency as high as possible, that is, designing the stiffest mounting possible. The apparently elegant and lower-cost arrangement of two or more auxiliary units on one and the same mounting makes achievement of a high resonance frequency difficult because of the accumulation of masses, and cannot be recommended for yet another reason: the solid-borne noise of an auxiliary unit, such as the power steering pump, is transmitted directly into the neighboring auxiliary unit, the air conditioning compressor, and into the body shell via its connecting elements (the air conditioning hoses in this case). It is also generally recommendable not to select a whole number (“integer”) transmission ratio for the auxiliary unit, but 1.1 or 0.9, for example, instead, in order that the vibration excitation generated by the auxiliary units cannot coincide with the frequency of one of the engine orders. Unavoidable belt slippage will otherwise result in the superimposition of two oscillations of almost identical frequency, causing acoustically extremely unpleasant periodic fluctuations in level, so-called beat frequencies, [27-15], [27-17], [27-18].

27.7 Measuring and Analytical Methods

Figure 27.5 Engine mounting.

Similar stiffness criteria must also be taken into account when arranging and fixing the auxiliary units. Their engine arrangement derives essentially from engine compartment packaging and the design of the belt drive. For vibration engineering purposes, they can be regarded as masses connected with certain spring rigidity to the engine mass, vibrating in resonance at the corresponding frequency.

The complexity of the internal combustion engine and its peripherals as a noise source has resulted over the course of time in the development of a large range of experimental methods that would, applied together, provide an extremely detailed picture of the overall system. Because of the sometime considerable effort involved in measuring the pressure distribution in a crankshaft bearing shell, for example, a standard range of diagnostic methods, using which most problems can be solved, or at least identified and estimated, is initially applied in practice. It is necessary to differentiate from these methods for the testing of typical specification data or benchmarks, such as the acceleration level on the engine mountings or the engine’s radiated acoustic power, for which simple prescribed procedures generally exist. A universal tool that is generally utilized at the start of a diagnosis is the application of so-called signature analysis to an airborne noise signal, obtained either from individual microphones (close around the vehicle or in the interior) or from a dummy-head image (Figure 27.6) of the noise. Many spectra are plotted in 3D form as a so-called waterfall diagram (Figure 27.7, left side) or in the form of a (generally color) coded 2D Figure, a Campbell diagram across an engine-speed ramp (Figure 27.7, right side). Such a depiction is extremely useful in particular, because it is possible, on the one hand, to read off the excitation dimension in the form of its relevant orders as obliquely running curves and because, on the other hand, resonances in the transmission route are visible as a result of peaks with a fixed frequency. If additional individual orders of frequency ranges can be filtered out or elevated, it becomes possible to identify the components critical for the problem under study in an acoustic assessment.

954 | Internal Combustion Engine Handbook

6606_Book.indb 954

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.7 Measuring and Analytical Methods

Amplitude [dB]

speed. An increase in these levels may, for instance, indicate deficiencies in the acoustic insulation between the engine compartment and the vehicle interior.

Speed

Figure 27.8 Order level over engine speed, section from Figure 27.7 (image courtesy of AFT Atlas Fahrzeugtechnik). Figure 27.6 Plastic head (photo courtesy of Head acoustics).

If only a few orders are relevant or precise quantitative information is required, one restricts oneself to the depiction of order curves against engine speed (Figure 27.8). In this context, the use of variable frequency filters makes it possible to achieve data reduction even during the measurement. Typical order curves for a combustion engine are those of the largest unbalanced inertial forces and torques and those of the ignition frequency, that is, the 2nd engine order in a four-cylinder inline engine and the 1st, 2nd, and 2.5th engine order in a five-cylinder engine. It is normal practice in the higher-frequency airborne noise range (mechanical engine noise) to subdivide the noise into frequency bands (generally thirds or octaves) and to plot their level against engine

If a detailed analysis of noise radiation is required, the classic yet complex window method, in which small “windows” are opened point by point from a complete, tightly fitting insulating enclosure (consisting, for example, of mineral fiber and lead sheet) and their influence on radiation measured has been superseded or at least augmented by more modern procedures involving less feedback effect on the test object and the acoustic field. In the case of the intensity method, the emitter is scanned point by point at a relatively short distance from the surface. The plotting of acoustic intensity above the projected surface provides a good picture of the distribution of heavily and less heavily radiating zones, making it possible to determine the total radiated acoustic power simultaneously. One disadvantage is the long measuring time, throughout which a stable operating state must be maintained. In addition, an automatic system for Resonance

Frequency

Speed

Amplitude [dB]

Amplitude [dB]

Spee d

Order

Frequency

Figure 27.7 Signature analysis: Campbell diagram (on right), waterfall diagram (on left) (image courtesy of AFT Atlas Fahrzeugtechnik).

Internal Combustion Engine Handbook | 955

6606_Book.indb 955

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

movement of the intensity probe is necessary in many cases for safety and repeatability reasons. Short measuring times and the observation of transient processes permit the array/ beamforming technique, during which the sound pressure is recorded parallel at a certain distance from the acoustic source by a series of specially arranged microphones (an array). Special algorithms permit the assessment of anisotropy of the sound pressure origin, so that a “source dissemination” can be represented for the level of the emitter surface. The complete acoustic field, that is, also directly on the emitter surface, can be calculated with the using mathematical-physical models from measured sound pressure signals during the so-called spatial acoustic field transformation Spatial Transformation of Sound Fields (STSF)/acoustic holography. The sound pressure signals are recorded at a microphone grid located at a distance from the emitter; additional reference sensors are also used to separate independent sources if necessary. Efficient systems also allow for short measuring times. Transient sound events can also be analyzed thanks to newer variations of the procedure [27-21]. So-called operational vibration analysis, which visualizes the motion modes occurring under actual operating conditions, is used where necessary to determine and influence where necessary, the oscillation modes relevant to radiation and/or solid-borne noise transmission. Experimental modal analysis on the other hand utilizes defined artificial excitation (e.g., an impact hammer) and is generally used for calibration of

Transmission cover

computation models or for checking whether certain natural frequencies and modes are within specified limits (e.g., the first natural bending frequency of the engine plus transmission unit)[27-10]. The method most frequently selected for modal analysis and operational vibration analysis is pointwise measurement of accelerations in three directions in each case. With the assistance of a computer, these can then be assigned to the nodal points of a wire-mesh model and a frequency selectively animated in slow motion. Other solid-borne noise signals, such as inductively measured paths, are also suitable in principle for operational vibration analysis. Optical methods are frequently used where high temperatures, rotating and/ or thin-walled components are present. In laser vibrometry surface velocity is measured point by point in one direction as a time signal. This does permit subdivision into frequency components, but necessitates relatively long measuring times as a result of the point-by-point scanning method required and, consequently, stable, steady-state operating conditions. Laser double-pulse holography on the other hand supplies instantaneous images of the deformation of large areas (Figure 27.9), which, however, represents the total deformation in the time between the two laser pulses and permits “natural-modeselective” evaluation only to a limited extent and only by skilled selection of the trigger point and time interval (typical: 0.8 ms).Transfer path analysis is an extremely efficient tool for precise analysis of the solid-borne noise contributions to

Interference stripes (hologram)

Ymm

448 398 347 296 245 194 143 92 41

80 Xmm 129

178

227

276

326

374

423

Zµm 1, 1,140 0, 5 0, 90 0,465 0, 0 0, 15 0,310 0,6 5 472 0

Figure 27.9 Laser double-pulse holography, instantaneous images of the deformation of large areas [Laser lab FHT-Esslingen].

956 | Internal Combustion Engine Handbook

6606_Book.indb 956

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.9 Sound Engineering

interior noise and therefore frequently forms the basis of an acoustic vehicle study [27-4], [27-5]. It is essentially performed in three stages.

5. Roughness: Assesses modulations in the 20–300 Hz frequency range that make a noise seem “rough,” which is not always a negative characteristic (sporty sound).

1. Direct (load cells, strain gages) or indirect (bearing deformations, input impedances) determination of the cutting forces at the points considered relevant

6. Tonality: Used for classification of noises in terms of their pure tonal content in proportion to their noise content [27-20].

2. Determination, with a decoupled noise source, of the acoustic transmission functions from the transmission points to the point of reception (e.g., the driver’s ear) using artificial excitation 3. Determination of the individual contributors to the noise by multiplication of the forces by the corresponding transmission functions. For instance, it is possible to determine the engine bearing and direction contributing most significantly to the buzzing noise perceived at the driver’s ear.

27.8 Psychoacoustics The weighting curves (e.g., A weighting) specified in DIN IEC 651 for acoustic signals are an initial, approximate approach toward taking into account the nonlinear behavior of human hearing. A simple frequency assessment does not, however, allow one to draw a conclusion on the subjectively perceived nuisance quality of a noise. Noises containing pulses, such as diesel knocking, are perceived as particularly unpleasant, whereas uniform sound at a constant level causes a low perception of unpleasantness, which is discernible less in the spectrum than in the time plot of such signals. The objective recording of these differences is a goal of so-called psychoacoustics, which defines for this purpose a range of psychoacoustic parameters on the basis of models and detailed audibility tests. The basis for many psychoacoustic parameters is provided by a modified frequency scale, the tonality scale (0–24 Bark), which is based on the nonlinear frequency/site transformation of the basilar membrane and thus simulates the natural frequency discrimination faculty of the human ear. Using complicated algorithms in some cases, the psychoacoustic parameters can be calculated from measured signals, taking into account spectral and temporal concealment effects, as well as the sensitivities of the ear to fluctuations in amplitude and frequency. The following are common psychoacoustic parameters: 1. Loudness: Linear variable for evaluation of perception of sound volume, using the “sone” (Reference: 1 kHz sinusoidal tone, 40 dB is equal to 1 sone). A computation method (according to Zwicker) is standardized in ISO 532. 2. Sound volume: Level variable for assessment of perception of sound volume, using the “phon” unit. It can be approximately calculated from loudness. 3. Severity: Weighting that emphasizes the high frequencies, which give a sound its severity (Unit: 1 acum). 4. Fluctuation amplitude: Assesses extremely low-frequency (< 20 Hz) modulations in signal level, which are normally considered unpleasant.

These psychoacoustic parameters, taken together, characterize a noise significantly better than a weighted level, with the result that attempts are often made to derive by suitable combinations of the parameters characteristics data for the quality of a noise; however, these characteristics are restricted to certain contexts. We may mention, to illustrate the basic difficulty here, the exhaust noise from a Ferrari, which is considered “good” by a young man, but not by his grandmother. Many carefully tailored hearing studies with assessors drawn from vehicle purchasers and/or experts have been and still are performed to define the correct targets. The basis for this is provided, because of the high level of reproduction quality and manipulation potential of the associated analysis systems, generally by dummy-head recordings (Figure 27.6), actually reproduced in some cases in an original environment and including perception of low-frequency solid-borne noise (hand, foot, and seat vibration). This is also the interface between psychoacoustics and sound engineering.

27.9 Sound Engineering The accepted tenets of vehicle acoustics have for many years now included recognition of the fact that simple “quietness” is in many cases no longer an appropriate target, because a certain level of acoustic feedback from the vehicle is in fact necessary and expected. The function of sound engineering is now that of providing the desired acoustic information with a sound that is as pleasant and, where appropriate, also as typical of the brand as possible, and to strike certain characteristics, such as “sporty,” “powerful,” “dynamic,” or “classy” depending on the vehicle type. Naturally, the sound of the engine is given special significance in this context. At the same time, the engine also boasts the greatest potential for modification of noise, as a result of the large range of transmission routes and the composition of the noise in the form of a mixture of orders. If one sets aside for the time being purely electronic manipulations, with which spectacular but “artificial” results can be achieved, depending on the engine type, it is possible to emphasize individual orders, for example, with modifications to the intake and exhaust systems and thus to generate the required sound, in many cases with a “sporty” or “dynamic” note. Essential fundamentals for the quality of engine noise are set down, however, at the concept phase, where decisions are made concerning number and arrangement of cylinders, firing sequence, mass balancing, housing and crankshaft stiffness, air routing, and so on. Provided the main criteria are taken into account at this stage, a good basic engine sound results almost automatically, signifying that “quieting” of undesirable, higher-frequency components resulting from

Internal Combustion Engine Handbook | 957

6606_Book.indb 957

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

gas forces, timing gear, and auxiliary units, which without doubt increase as engines and vehicles become more complex, is again needed to achieve a “perfect” acoustic pattern. It is not always easy to decide what components must change and to what extent because people’s subjective perception may, on the one hand, be extremely diverse, and on the other hand, the costs involved in the development and implementation of remedies are not insignificant. The models, parameters, and methods of psychoacoustics are, therefore, increasingly used to ease this decision [27-19].

27.10 Simulation Tools The predictive calculation of the vibrations and radiated airborne noise from an engine that is still at the design stage even today remains a demanding target (Figure 27.10). In addition to the large number of degrees of freedom necessary, it is particularly the modeling of the charge cycle and combustion processes, and the nonlinear contact processes (impacts, lubricating films, friction, etc.) between the many moving and fixed components that make the achievement of a meaningful model extremely difficult. The consequences have been a series of specialized models and methods, the results of which generally must first be compiled to obtain an overall statement. The starting point on the structure dynamics side is an finite element (FE) model of the engine plus transmission

unit and an FE or multibody model of the rotating crankshaft drive and, where required, of the timing gear, which are sufficiently fine to record the complex oscillation modes in the acoustically relevant frequency range. Verification is normally performed in subsystems by calibration against experimental modal analyses. One of the main difficulties in this field is the correct recording of attenuation. The most complex and difficult element, however, is calculating the forces acting on the block structure during operation. It is necessary for this purpose to link the individual models, the linking procedure conditions through the hydrodynamic oil film of the bearings and the cylinder sliding surfaces being highly nonlinear. The gas forces acting on the crankshaft drive can, for their part, either be taken from an indication measurement or be calculated using a complex charging and thermodynamic model. The overall result of the simulation is a velocity distribution on the surface of the engine, including the engine mountings. Calculation of the radiated airborne noise, for the purpose of quantification of emitted noise, for example, can be accomplished relatively easily using FEM or BEM, because at this stage only a single, homogeneous medium is involved. When calculating the vehicle interior noise, it is necessary to differentiate between the airborne noise passing through the body shell wall and the solid-borne noise transmitted into the body shell via the engine mountings, drive shafts, and various other connecting elements. The solid-borne noise transmission routes generally dominate at frequencies below approximately 500–1000 Hz.

Charge cycle program or measured indicator diagram

FE model engine transmission block

Non-linear coupling thanks to a hydrodynamic oil film

FE model crankshaft drive

Forced vibrations from the engine surface

High-frequency: Airborne noise

Low-frequency: Solid-borne noise

BE model of airborne noise emission

Transmission function of the engine mounting and all connecting elements conveying solid-borne noise to the body

SEA model of the body

Body noise transfer function Interior noise as the sum of all solid-borne noise parts

Interior noise as the sum of all solidborne noise parts

SEA model of the engine compartment (capsule, or free airborne noise discharge through holes) Exterior noise (engine contribution)

Figure 27.10 Example for the sequence of an engine noise simulation.

958 | Internal Combustion Engine Handbook

6606_Book.indb 958

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

27.11 Antinoise Systems: Noise Reduction using Antinoise

Here, the engine mounting oscillation amplitudes can be regarded as “path excitation” of the “Silentbloc” rubber mountings of the engine mounting system, because they are frequently used as reference factors that must not exceed certain limits. The solid-borne noise component of total interior noise can be calculated for every bearing using the dynamic stiffness of the engine mounting and the acoustic transmission function of the body shell at the bearing points, and these components are then added with correct phasing for all bearings and directions of oscillation. The acoustic transmission functions of the body shell can, for their part, be provided in the form of measured data or in the form of the results of a dynamic FE computation of body shell structure and cavity. Mathematical determination of body shell transmission functions is extremely difficult, because of the complex structure of a lined body shell that incorporates many difficult to define joints. In the higher frequency range, which is generally dominated by airborne noise excitation from the engine compartment, conditions are in some cases further simplified by the fact that it is possible as a result of a broadband noise character and the high density of body shell natural frequencies to observe energy flows for whole frequency ranges using statistical observation methods and ignoring phase relationships, and thus obtain simple algebraic equations (SEA ). Because of the extremely high complexity of a meaningful acoustic model for a complete vehicle—even the development-phase updating of submodels constitutes a not insignificant organizational task—“hybrid models” are increasingly used in practice, that is, the combination of assemblies and modules or input variables registered using measuring technology with computer models of new components. This route is open, in particular, in the case of new developments based on an existing vehicle floor group or an existing engine family [27-9].

27.11 Antinoise Systems: Noise Reduction using Antinoise The elimination of unpleasant noises by artificially generated antinoise is an option that became technically possible with the availability of high-speed digital control systems. The signal of the noise is recorded as close as possible to its source, the opposite phase signal is generated in a computer and is then emitted via an amplification system. The cancellation of harmonic signal components works best of all, for example, one or more engine orders. The analogous mechanical principle is the familiar cancellation of inertial forces from the crankshaft drive by opposite-phase forces using balance shafts. Antinoise systems for four-cylinder vehicles, which capture engine noise in the vehicle interior using a number of microphones and reduce the buzzing of the 2nd engine order by more than 10 dB using systematically located loudspeakers, became available in the 1980s in the form of mature prototypes. Practically all the major automobile manufacturers also had their own development programs in their research

departments, which were unveiled to great publicity in the technical press in the form of experimental vehicles in which the buzzing noises could be switched off during travel at the push of a button. These initially impressive presentations at the same time brought to light a shortcoming: the vibrations generated along the same route by the engine were not susceptible to influence with antinoise by the loudspeakers, but do play a role in determining the subjective impression of comfort. Further developments were, therefore, logically enough, aimed at eliminating the solid-borne noise and, therefore, also the vibration at the points of transmission into the body shell by counterphase controlled vibration generators, in the form of piezo actuators, which further heightened the technical and cost input. The complexity or cost-benefit ratio was in the past also the main reason for antinoise not becoming established in mass produced passenger cars. Put quite simply: “It is too expensive for the four-cylinder vehicles which need it, and the more expensive six- and eight-cylinder engines do not need it.” The frequently expressed vision of “electronics instead of weight” (i.e., of balancing out the costs for the control system by saving on noise insulation materials), was doomed to failure based sheerly on physical principles: the sound package in the passenger car body shell is installed almost entirely for reduction of the high-frequency noise components, the signals of which have a stochastic character and, therefore, no definable phase reference. Signal coherence is, however, the precondition for all stable interference phenomena and, therefore, also for noise cancellation. Recent developments in higher-value four-cylinder engines indicate a pronounced trend toward the classic “balance shaft” solution (Lancaster balancing), through which the inertial forces of the 2nd engine order and the vibrations and buzzing noises caused by it can be completely eliminated. As a result of the entirely different cost framework, antinoise systems are in use in a series of applications in aerospace and plant engineering. Antinoise can ideally be emitted into a helicopter pilot’s helmet close to his ear to cancel low-frequency rotor noise and, thus, improve ease of radio communication and general comfort. Antinoise systems are employed in wind tunnel facilities to cancel out low-frequency fan noises using a loudspeaker system installed directly on the fan outlet. The main field of application is that of low-frequency noise problems, where the standard noise absorption and insulating materials offer little effective help.

Bibliography

27-1. Albenberger, J., Steinmayer, T., and Wichtl, R. “Die temperaturgesteuerte Vollkapsel des BMW 525 tds,” ATZ 94(5):244–247, 1992. 27-2. Basshuysen, R. v. “Motor und Umwelt,” ATZ 93(1):36–39, 1991. 27-3. Bathelt, H. and Bösenberg, D. “Neue Untersuchungsmethoden in der Karosserieakustik,” ATZ 78(5): 211–218, 1976. 27-4. Bathelt, H. “Analyse der Körperschallwege in Kraftfahrzeugen,” Automobil-Industrie 27–33, 1981. 27-5. Bathelt, H. “Innengeräuschreduzierung durch rechnergestützte Analyseverfahren,” ATZ 83(4):163–168, 1981.

Internal Combustion Engine Handbook | 959

6606_Book.indb 959

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 27 Noise Emissions

27-6. Betzel, W. “Einfluss der Fahrbahnoberfläche von Geräuschmessstrecken auf das Fahr- und Reifen-Fahrbahn-Geräusch,” ATZ 92(7/8): 411–416, 1990. 27-7. Ehinger, P., Großmann, H., and Pilgrim, R. “Fahrzeug-Verkehrsgeräusche. Messanalyse- und Prognose-Verfahren bei Porsche,” ATZ 92(7/8:398–409, 1990. 27-8. Eikelberg, W. and Schlienz, G. “Akustik am Volkswagen Transporter der 4. Generation,” ATZ 93(2):56–66, 1991. 27-9. Estorff, O. v., Brügmann, G., Irrgang, A., and Belke, L. “Berechnung der Schallabstrahlung von Fahrzeugkomponenten bei BMW,” ATZ 96(5):316–320, 1994. 27-10. Ewins, D. J. 1984. Modal Testing, Theory and Practice. Research Studies Press Ltd.:Letchworth.

27-17. 1993. “Kraftfahrzeugmotoren,” 3. Aufl, edited by Küntscher, V. Verlag Technik: Berlin. 27-18. 1997. “Handbuch Dieselmotoren,” edited by Mollenhauer, K. Springer Verlag: Berlin Heidelberg. 27-19. 1994. “Soundengineering,” edited by Quang-Hue, V. Expert Verlag: Renningen; Malmsheim. 27-20. Zwicker, E. and Fastl, H. 1990. Psychoacoustics, Facts and Models. Springer-Verlag: Berlin; New York. 27-21. Quickert, M. and Andres, O. 2003. “Moderne Verfahren zur Ortung von Schallquellen am Beispiel schwerer Nutzfahrzeugdieselmotoren,” edited by Tschöke, H. and Henze, W. Motor- und Aggregate-Akustik, Haus der Technik Fachbuch Band 25. Expert Verlag: Renningen.

27-11. 1998. “Geräuschminderung bei Kraftfahrzeugen,” edited by Geib, W. Friedr: Braunschweig and Vieweg and Sohn: Braunschweig.

27-22. Hohnstein, T., Gleiter, U., Glaser, S., and Fritz, T. “Erste Serienanwendung von Steinschlagoptimierten Kunststoff-Ölwannen,” MTZ, Wiesbaden, 2010.

27-12. Heckl, M. and Müller, H. A. 1994. Taschenbuch der Technischen Akustik. Springer Verlag: Berlin; Heidelberg.

27-23. N.N. “Akustikentwicklungen für künftige Motorkonzepte—Titelthema,” MTZ, 2010.

27-13. Henn, H., Sinambari, G. R., Fallen, M., and Erhard, Ch. 2008. Ingenieurakustik, 4. Aufl. Wiesbaden: Vieweg+Teubner.

27-24. Krüger, J., Pommerer, M., and Jebasinski, R. “Aktive Abgasschalldämpfer,” MTZ, Wiesbaden.

27-14. Klingenberg, H. 1991. Band A: Akustik. Automobil-Messtechnik, 2. Aufl. Springer Verlag: Berlin.

27-25. Heuer, S. et al. “Neue NVH-Herausforderungen infolge der dramatischen CO2-Reduktionsvorgaben.” Eleventh Stuttgart International Symposium, March 16, 2010 and March 17, 2010.

27-15. Kollmann, F. G. 1993. Maschinenakustik. Springer Verlag: Berlin; Heidelberg. 27-16. Kremer, L. and Heckl, M. 1996. Structure-Borne Sound, second edition. Springer Verlag: Berlin.

27-26. Meier, C. et al. “Electric drive—a challenge for interior and exterior sound.” Eleventh Stuttgart International Symposium, Automotive and Engine Technology, February 22, 2011 and February 23, 2011.

960 | Internal Combustion Engine Handbook

6606_Book.indb 960

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28 Engine Measuring Technology The testing phase represents a crucial stage in the development of a combustion engine. Basic tasks include validating the engine design, verifying safe adherence to the limit values, as well as optimizing and calibrating entire powertrain. To satisfy these tasks, all the engine specific values need to be clearly and reproducibly recorded. This requires a logical modular design for the test bench overall system, a realistic representation of the load conditions of the engine, as well as a defined level of precision for all measurands. The engine internal characteristic values are particularly important. Determining these values is an indispensable part of test bench measuring technology and is dealt with specifically in Chapter 25, Combustion Diagnostics. Please refer to Chapter 21, Exhaust Emissions for specific aspects of exhaust gas measuring technology and statutory legal provisions. The increasing complexity of modern engines coupled with the simultaneous foreshortening of development times places great demands on measuring technology and this results in significant changes to the test methods integrated into the overall development process. Statistical test planning [design of experiments (DoE)] [28-1], model-supported testing [28-2], as well as the representation of dynamic drive circumstances at the test bench has become standard. A powerful test bench system [28-3] is intended to allow highly automated processes with the shortest possible measuring times as well as proven reliability of results. The required data security and reproducibility of the measured results even for different test benches also signifies an increase in the performance of the methods and tools used. We will provide a brief description of this connection using fuel consumption measuring technology as an example. Specific fuel consumption [see Chapter 3.5] is a key objective in the development of any engine. Computersupported optimization and DoE methods [28-4] are used

during optimization work to keep necessary inspection costs low. With the effect that only relatively few measuring points are available for evaluating the engine behavior in the overall map. Nevertheless, to be able to identify the progress safely depending on the set parameters and the desired optimum, measuring uncertainty must be extremely low in these points. Repeatability within a measuring process is crucial. Absolute accuracy [28-5] is also vital for the exchangeability of results between different test benches. An important factor with the accuracy requirements is the interaction of various measuring technologies. This always involves practice-orientated values from the entire inspection system [28-6]. For example, when determining the specific fuel consumption, fluctuating torque development or insufficient conditioning of the engine, even at high precision levels of the consumption measuring technology, can prevent automated optimization and consequently efficient testing. The following sections shall provide a concise overview of the required measuring technology on practically every engine test bed. Please refer to the bibliography for further measuring technology themes, for instance, oil consumption measurements [28-7], injection system measuring technology [28-8], and optical combustion diagnostics [28-9].

28.1 Measuring Technology in the Test Bench Overall System Modern engine development test beds use a network of measuring devices with up to several hundred measuring channels to record physical parameters from the temperature and pressure measurement through to measuring emitted pollutant masses.

Internal Combustion Engine Handbook | 961

6606_Book.indb 961

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

The size of the test bench depends on the size/power of the engine being tested. Depending on the type test specimen, the following basic test benches are used:

•• Dynamic test benches: They also permit the representation of dynamic operating conditions. Power pack, powertrain, and vehicle test benches are in most cases classified as dynamic test benches.

•• Engine test bed: On the test bench, it is just the engine without any other components from the powertrain. The engine’s flywheel forms the interface between the test specimen and the test bench.

A multitude of tasks are assigned to the test bench automation system [28-10] (Figure 28.1). Depending on the inspection program to be carried out, the engine shifting point must be adjusted and the measurement data need to be recorded and evaluated. Furthermore, modern test bench systems can even be operated without operating personnel. This allows costs and the time taken for development and calibration work to be reduced, especially if modern test scheduling tools and computer-supported optimization are used.

•• Power pack test bench: The engine is tested with the vehicle transmission. The transmission output shaft is connected to the test bench dynamometer. •• Powertrain test bench: The test specimen is the engine with the entire vehicle powertrain. The two dynamometers (there are four for all-wheel drive) are connected to the wheel-side half-axle ends. In some applications, the combustion engine is substituted with an electrical drive machine (prime mover), requiring very special powertrains for the representation of the high-frequency components in the torque development of the combustion engine. •• Vehicle test bench (also called roller test bench): The whole vehicle along with the tires is set onto the roller test bench. The tire contact patches act as the interface between the vehicle and the test bench. The rollers, at least from the vehicle drive wheels, are directly connected to a dynamometer or to several dynamometers. Depending on the inspection operation, a distinction is made between the following main categories of engine test bed: •• Stationary test benches: They allow inspection of the test specimen during stationary conditions.

28.1.1 Dynamometers

The type of dynamometer (Figure 28.2), (dyno for short) is an integral feature of a test bench. It is necessary to distinguish between active and passive dynamometers. Active dynamometers can brake and drive the combustion engine, while passive dynamometers are only able to brake. Active dynamometers therefore also allow towing operation of the combustion engine. They are generally suitable for the four-quadrant operation, where all four operational sign combinations of positive and negative rotation speed and positive and negative torque may occur. This is one of the reasons why they are used for dynamic test benches. However, their great flexibility, higher dynamics, and low maintenance and operating costs (because of frequently active braking) are also utilized, when it is not entirely necessary to drive the combustion engine.

Test bench automation system

Data recording, data reduction

Calculation (online) Spec. fuel consumption. Spec. mass emission.

Volumetric concentrations Gas preparation Analyzers (CO, CO2, O2, NOx, THC)

Forming the mean value Integration

Masses and volumetric flows Gas volume meter Air mass flow sensor Fuel consumption

Internal engine control parameters

Temperatures Low temperatures High temperatures

Ignition angle Start of injection Injection mass...

Pressures

Mech. parameters Speed Force Torque

Pressure (absolute) Pressure (relative) Cycle-resolved

Media

Data storage

Measurement data processing

Engine data recording

Air Coolant Oil Fuel

Control lever adjuster mechanical electronic

passive active

Dynamometer

Engine working point, control/adjustment

Figure 28.1 System design and tasks of a test bench system [28-11].

962 | Internal Combustion Engine Handbook

6606_Book.indb 962

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.2 Mechanical Measurands

Figure 28.2 Examples dynamometers. A pendulum machine on the left and a foot machine on the right.

Hydraulic dynamometer brakes and electrodynamic retarders are mainly used as passive dynamometers for engine test beds. Both models are predominantly constructed mechanically as so-called pendulum machines. Because of their better regulating properties, electrodynamic retarders are preferred to hydraulic dynamometer brakes for engine powers up to approximately 500 kW. For large and stationary braking power greater than 500 kW, hydraulic dynamometer brakes have the advantage because of their lower purchase price and greater robustness. Typical applications include endurance test benches, quality test benches, as well as test benches for simple research and development tasks. Synchronous, asynchronous, and direct current machines are used as active dynamometer brakes. They are constructed both as foot machines and pendulum machines for highly dynamic applications, for example, for the passenger car exhaust measurement with vehicle and driver simulation, for driveability evaluations and automated optimization, as well as for motor racing applications. Extremely low moments or inertia for the highest possible dynamics, for example, for representing the torque shock loads of a combustion engine, can be achieved with permanent magnet machines. Pendulum machines (Figure 28.2) allow a high torque measurement accuracy as well as very dynamic and exact torque regulation. They have a pivoting stator, which means that the supporting force of the stator can be measured at the fixed housing and can be used to determine the torque. They also have the advantage of a particularly simple and accurate calibration option, specifically, of loading a defined stator lever with calibration weights. With the foot machines (Figure 28.2), the stator is permanently connected to the housing. The torque is measured using a torque measuring shaft or preferably with a torque measurement flange. Because of the emerging torque shock load, care must be taken that the sensors have a sufficient overload capacity. In conjunction with a highly developed test bench regulation system, the latest generation of permanent magnet synchronous machines (PMSM) is able to recreate or apply a large proportion of the real-life driving conditions, from both conventional and hybrid powertrains at the engine test bed.

In addition to the simulation of drive train oscillations up to 40 Hz, this includes the vehicle-realistic (realistic) engine start with different starters (ISG, RSG, conventional starters) and the “theta-zero simulation,” during which misfiring can be identified by the on-board diagnostic. Because of the positioning of the particularly small constructed PM machine in the portal framework, it opens up the possibility of designing the exhaust system corresponding to the arrangement in the vehicle and, consequently because of the realistic flow, of performing the calibration of the exhaust gas after-treatment systems (NOx storage catalytic converter, diesel particulate filter—DPF, etc.) in a vehicle-realistic manner [28-12], [28-13].

28.2 Mechanical Measurands Torque and speed are among the most important mechanical measurands at the test bench. For quick so-called crank-anglebased measurements performed, for example, for combustion diagnostics, the running determination of the current angular position of the crankshaft [°crankshaft angle (CS)] is important during each individual rotation. Strain gages are generally used for torque measurements, which are arranged in the force flow of the measuring shaft and measuring flange or, in the case of the pendulum machine, on the bending bar of the load measuring cell. The straindependent change of the electrical resistance of the strain gages is electronically evaluated, taking the stiffness of the strain gage carrier into account. Furthermore, the so-called air gap torque of the dynamometer can be calculated from machine-internal measurands, for example, from the stator current, with the aid of mechanical machine models and can be used especially for highly dynamic applications. The measurement of speed and CS is usually performed once the impulse counting procedure has been carried out, for instance, with hall sensors and a toothed pulley, or with an optical detector for the division markings of a glass plate. For practical reasons, the crank angle index mark generator is often mounted onto the freely accessible end of the engine (Figure 28.4).

Internal Combustion Engine Handbook | 963

6606_Book.indb 963

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

Figure 28.3 Test bench construction with a highly dynamic dynamometer.

Figure 28.4 Crank angle index mark generator for determining speed and current angular position of the crankshaft.

With accurate and chronologically highly resolved measurements, care should be taken that non-negligible torsion of the crankshaft and superimposed rotational oscillations cannot occur while the engine is being operated. The momentary torque values, angular velocity, and rotational angular position of the crankshaft (CS) should therefore by measured as close as possible to the same location or should be analytically referenced to the same point. In addition to the current angle sensors on the market, systems also exist, which are able to detect the direction of rotation or even the absolute angle. With a system based on optical measuring principles, for example, the signals of a marker plate are determined using two identical sensors, which are displaced (usually at 90° to each other) in regard to the phase. The particular direction of rotation is the advantage/disadvantage of a signal. Other

systems with multiple traces on a marker plate allow the output of the absolute angle of rotation if the traces are coded differently. The resolution depends on the number of traces. Measurements from an idle state can be performed with an inductive measuring system, as a signal is present even when silent. The angular position can be determined at any time by recording the spatial distribution of the magnetic field of a permanent magnet rotating alongside the crankshaft. A particular advantage of such inductive systems is the small amount of space needed by the sensor, which can thus be simply combined with an optical angle sensor. If this angle sensor is already calibrated to an indicating system, then two measuring tasks are thus being fulfilled with little additional installation space. Specific sensors are used for special measuring tasks. For instance, optical sensors in accord with the reflection principle for the speed measurement at the turbocharger; piezoelectrical sensors as acceleration and knock detection sensors; the capacitive piston stroke sensor for precise determination of the upper dead point; and the inductive travel sensor for measuring needle stroke for injection nozzles or valve strokes [28-14].

28.3 Thermodynamic:Measurands Combustion engine thermodynamics are primarily determined from the temperature and pressure measurements in the surrounding atmosphere, in the flowing liquids, combustion air, and fuel, as well as in the combustion chamber itself and at the exhaust. A greater degree of accuracy for measuring low temperatures, for instance, the air at different parts of the intake section or of the lubricating oil, is possible thanks to mostly standardized resistance thermometers, such as Pt100 or Pt1000.

964 | Internal Combustion Engine Handbook

6606_Book.indb 964

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.6 Fuel Consumption Measurement

Thermocouples are generally used to measure high temperatures, for example, at the combustion engine exhaust. The thermal inertia of these sensors differs depending on the mechanical design and effectuates a significant limitation of the chronological resolution. To perform pressure measurements, sensors with very specific characteristics are used depending on the measurement task. An inherent distinction is made between absolute and relative pressure sensors. One is used for dynamic pressures and the other for dynamic and static pressures. The most essential characteristics are often the measuring accuracy and chronological resolution adapted to the measuring task. Accuracy is paramount when measuring static pressures, such as barometric absolute pressure, and relative pressure at various points in the intake section of the exhaust system or lubrication circuit. Appropriate industrially available sensors are used. Piezoelectrical sensors are normally used for dynamic pressure measurements, for example, at the injection system or in the engine’s combustion chamber during a power cycle. These sensors are optimized both in terms of the measuring range (combustion chamber pressures of up to approximately 250 bar and injection pressures of up to approximately 2500 bar) and the temporal response (natural frequencies at approximately 100 kHz). Quartz pressure sensors generally require cooling to satisfy cylinder head temperatures and accuracy requirements. Pressure sensors that use GaPO4 measuring elements do not require external cooling and reduce the amount of installation space required in the cylinder head. These sensors are also used in the lower pressure range, for example, for determining the gas dynamics in the intake and exhaust channel.

28.4 Flow Measuring Technology Mass flows of drawn combustion air and the fuel from the combustion engine used to generate power are given pivotal significance at the engine test bed for several reasons. The ratio of air mass flow to fuel mass flow represents a central specific value for the type of combustion (see Chapter 3). The ratio of fuel consumed to work performed (specific fuel consumption) needs to be optimized; this is generally considered one of the most important development goals for modern engines. Another flow measure, the volume of blow-by gas that escapes from the combustion chamber into the crankcase, is often called on as a monitoring variable and used to evaluate the engine condition, but is also a decisive factor in the optimization of the piston, piston ring, and cylinder system. However, the flows or consumption of lubricating oil and coolant are also important. In the following, we will discuss the measurement of the intake air volume, the fuel consumption measurement, and the significance of associated fuel conditioning, as well as the oil consumption measurement, blow-by gas measuring technology, and urea consumption measurement technology. The direct determination of the exhaust gas mass flow is dealt with in the context of the exhaust gas measuring technology.

28.5 Intake Air Volumetric Flow Measurement The systems for determining the quantity of air drawn in by the engine can be divided into measuring procedures based on either volume or mass. Both measuring procedures should be considered equal, provided the density of the intake air is taken into account. Key features of the measuring procedures used at the engine test bed include the dynamic behavior, the pressure loss brought about by the measurement setup, as well as possible distortion of the measuring results because of pulsation of the air column. All systems have one thing in common: the engine air inlet must be tightly connected to the measuring system. The principal volume-based measuring techniques include: •• Flow measurement according to the displacement principle: This category includes, for example, rotary piston gas meters and positive displacement gas meters frequently used in plant engineering. Because of their distinct inertia, they are largely suitable only for stationary engine operation. As a closed system, the pulsation influence is circumstantial. •• Volumetric flow measurement calculated from a differential pressure measurement at an orifice (laminar flow element): Dependent on the volumetric flow, a quantifiable differential pressure forms at a orifice of a certain geometry, which can be used to calculate the volumetric flow by applying the laws of fluid dynamics. The procedure is suited to both stationary and dynamic measurements. In essence, an undisturbed flow profile is present at the measuring point. The orifice should preferably be installed somewhere around the center of a sufficiently long (>20 diameter) section of pipe. The most significant example of the mass-based measuring process is the hot-film anemometer. The functional principle— determining the mass-flow contingent heat transport between, for example, an electrically heated platinum thin-film resistor and another one used as a temperature sensor—is also applied to modern engine controls to determine dynamic air mass flow (see Chapter 18). At the test bench, with its significantly higher accuracy requirements, the hot-film anemometer is also installed in long, straight pipe sections with undisturbed flow. These are, however, sensitive to pulsation, which is why it is recommended to use attenuation containers or flexible connector hoses between the measuring pipe and engine.

28.6 Fuel Consumption Measurement When it comes to measuring the fuel consumption of internalcombustion engines, continuous and discontinuous volumetric as well as gravimetric measuring techniques are known. With volumetric measuring principles, the volume of fuel consumed by the engine is calculated. When determining the mass of fuel consumed, for example, to be able to give the specific fuel consumption (g/kWh), temperature-dependent fuel density must be taken into account. In contrast, the gravimetric measuring procedure directly calculates the mass of

Internal Combustion Engine Handbook | 965

6606_Book.indb 965

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

fuel consumed, with the result that the additional uncertainty of determining the density is unnecessary. Volumetric measuring techniques include measuring procedures touching on the container and displacement principle. •• The vessel principle is applied to the traditional Seppeler vessel and type-related devices. The draining time of a defined volume is measured and converted into fuel consumption. The procedure conducts measurements intermittently and is limited to strictly stationary operating periods. •• With the displacement principle (Figure 28.5), a displacement body from the fuel flow is set in motion and consequently the volumetric fuel consumption is continuously and, depending on the inertia of the sensor, even dynamically calculated. Pressure and gap losses occur with passive displacement

meters, which have a negative impact on engine operation and mass accuracy. Active displacement meters overcome this disadvantage by employing an external drive [28-15]. Common to all measuring techniques mentioned previously is the fact that during the measuring phase, gas bubble separation from the fuel, as it automatically occurs in the fuel tank during normal operation, must be sacrificed. A further development of the servo-controlled displacement meter enables the chronologically highly resolved measurement of intermittent flows from individual injection processes. Figure 28.6 shows the application of this quick volumetric flow measurement in the injector intake of a direct injection gasoline engine, where it is possible to analyze multiple injections with separations times as low as 150 ms.

Interface Engine/Pulse meter combination Electronics Gear meters with different volume contents for measurement range agreement Inlet

Exhaust Light source

Bypass

Piston Light sensor

Figure 28.5 Measuring principle of servo-controlled displacement meters (Δp = 0) [28-16].

∆p-Sensors

Signal-Processing

Pressure Sensor

Piston Sensor

Figure 28.6 Shot-to-shot PLU measuring principle in upstream configuration (measuring device on the high-pressure side of the injector).

966 | Internal Combustion Engine Handbook

6606_Book.indb 966

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.6 Fuel Consumption Measurement

•• The mass flowmeter according to the principle of a Wheatstone bridge contains a total of four orifices, which are used as hydraulic resistors. A pump generates a constant flow via a bridge diagonal plane and therefore has the same pressure losses at the orifices in each branch. If a flow occurs at this moment through the diagonal plane, then a pressure loss that is proportional to the mass flow is measured via the bridge.

Gravimetric measuring techniques include the Coriolis and Wheatstone bridge mass flow devices according to the continuously measured flow principle and the procedures that work according to the vessel principle, such as the drainage weight measurement and the weight measurement using burets and the weight principle. The continuously working mass flowmeters require bubble separators to deaerate the approach flow and return flow. Gas bubbles in the fuel that have not been separated lower the quality of the measuring signal and diminish the accuracy and dynamics of the system.

Mass flowmeters according to the gravimetric container principle are open systems: they are used for gas bubble separation in the measuring vessel: •• The drainage weight measurement is similar to the volumetric container principle (Seppeler vessel). A second drainage vessel is located parallel to the fuel drainage vessel. The second vessel is fitted with a membrane on the basis of pressure compensation and is filled with a fluid of a known density. The drainage of a specified volume of reference fluid is used as the dimension of the consumed fuel mass. This principle, on the other hand, is only suitable for strictly stationary engine operating states.

•• The Coriolis mass flow device is the most important exponent of the flow principle (Figure 28.7). The fuel flows through the pipe sections that are electromagnetically stimulated to oscillate, with the result that the pipes are twisted because of the Coriolis effect. The twist angle is calculated proportionally to the mass flow [28-17].

Flow

•• A buret weight measurement involves using a buret to remove the consumed fuel from the engine and measure the lowering with the aid of a differential pressure sensor. Using the familiar buret cross profile, the consumed fuel mass can even be calculated in the dynamic engine operation.

Fluid Force

Flow

•• With the weighing principle (Figure 28.8), the fuel consumed by the engine is removed from a storage vessel, which has all the properties of a fuel tank—even the incorporation of fuel return and gas bubble separation. Both the accumulated fuel consumption of the engine and the momentary dynamic fuel flow are gravimetrically determined from the continuously defined weight and the ongoing weight reduction of the weighing vessel. Variously sized weighing vessels are used depending on the method and resolution of the weighing procedure. Standard vessels sizes usually suffice, even during whole passenger car exhaust emission test cycles

Fluid Force

F Twist Angle Twist Angle

F

Driving Force

Figure 28.7 Coriolis measuring principle for determining mass flow.

Tare weight

Balance beam

Spiral spring

Capacitive sensor

Measuring vessel

Hydraulic damping equipment

Stable measurement carrier

Flexible spring-pipe element

Fill solenoid valve 12

3

4

1 2 3 4

... Fuel supply ... to the engine ... from the engine ... Vent pipe

Figure 28.8 Functional diagram of a pair of fuel scales.

Internal Combustion Engine Handbook | 967

6606_Book.indb 967

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

and are able to measure the dynamic fuel consumption without interruption and at high levels of accuracy. For high fuel consumptions, for instance, with utility vehicle engines, two pairs of fuel scales connected via a hitchless switch unit are often used. This add-on also allows genuine continuous operation.

system. The noticeably lower distribution of the measured values with suitable conditioning of the fuel temperature (Figure 28.9) is impressive in documenting the impact of conditioning technology on the quality of the measuring results [28-18]. Figure 28.10 demonstrates a favorable arrangement of an overall system for measuring fuel consumption and for conditioning the temperature and pressure of the fuel fed to the engine.

28.7 Fuel Conditioning

specific consumption [g/kWh]

The connection between consumption measuring devices and the engine’s injection system is crucial for achieving repeatability and comparability in the practical measuring operation. It is necessary to use sophisticated conditioning technology to avoid effects such as obvious overconsumption/ underconsumption in the measuring circuit resulting from temperature changes, or to be able to maintain the pressure levels required for the functioning of the injection system for approach and return flows. Stability and precision requirements for temperature regulation in the fuel measurement circuit are approximately two orders of magnitude higher than remaining media conditioning at the test bench. Fuel conditioning is therefore an integral part of modern standard measuring systems and is critical for the quality of practically attainable measuring results. Figure 28.9 shows a comparison of two measuring series with varied exhaust gas recirculation rates (EGR), one with unregulated fuel cooling and the other with fuel temperature conditioning using a highly accurate temperature regulation

28.8 Oil Consumption Measurement Technology Strict legislation and stringent engine specification requirements in terms of oil consumption require high-quality measuring technology in the relevant sections of the development process. Accurately determining oil consumption values is imperative because oil consumption has a direct bearing on particulate emissions. The measured values influence the further development strategy and many decisions in the following application areas: •• optimizing head gaskets, piston rings, or valve guides, as well as investigating obvious signs of wear •• monitoring endurance test runs •• investigation and comparisons of sources of oil consumption •• measurements at various engine map load points •• investigations when oil dilution appears •• production monitoring.

430 428

unregulated fuel cooling

426 424 422 420 418 416 414 412 410

0

5

10 EGR rate [%]

15

20

specific consumption [g/kWh]

395 393

regulated fuel temperature (AVL Fuel Temperature Control)

391 389 387 385 383 381 379 377 375

0

5

10 EGR rate [%]

15

20

Figure 28.9 Effects of fuel conditioning on measuring results [28-19].

968 | Internal Combustion Engine Handbook

6606_Book.indb 968

1/19/16 8:58 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.9 Blow-By Measuring Technology

Ventilation

Buffer volume

Measurement in both flow directions Calibration sensor (optional)

Pressure sensor

Fuel supply

AVL Fuel Temperature Control

Calibration, ventilation

Druckregelung (patented)

Pressure regulating

Gas bubble separation

Fuel cooling and heating

Sensor for mass flow and density

AVL Fuel Mass Flow Meter Coolant

Various options are available for measuring oil consumption. These methods are outlined and compared in Figure 28.11. Typical oil consumption values for modern engines are approximately 0.20 ± 0.05 g/kWh in the utility vehicle sector or 0.50 ± 0.1 g/kWh in the diesel passenger car sector, and can even be a little higher for petrol engines, where it must be noted that consumption increases as the engine gets smaller.

28.9 Blow-By Measuring Technology Measuring the crankcase gases, also known as blow-by, is standard practice these days for engine test beds. On the one hand, there are areas of application in engine research and development for optimizing the cylinder–piston pairing and the piston micrograph, in the development of a favorable piston ring geometry or with the design of crankcase ventilation systems. On the other hand, blow-by measurements are performed to monitor production and for acceptance

Figure 28.10 Diagram of a measuring system for fuel conditioning and consumption measurements.

inspection when monitoring the endurance run and run in at the test bench for quality assurance. Various measurement principles are applied: •• With the float volumetric flow measurement, a resistor body from the blow-by is immersed vertically from bottom to top against gravity. The power exerted by the blow-by flow causes the float to rise or fall, and this information can be used to draw conclusions about the volumetric flow. •• Hot-film anemometers are used in a way similar to that of the intake air quantity measurement. •• With impeller and turbine meter principles, an impeller studded with blades is made to rotate in a measurable way by the blow-by flow. •• The gas meter principle uses two measuring chambers and each chamber has one bellow, which is alternately filled and drained. The movement of the membrane is transferred to a crankshaft and the number of revolutions is counted.

AVL 403S

AVL 406

Draining and weighing

Filling method

Principles

Interacting vessels

Intake and weighing

Draining and weighing

Interacting vessels

Application

For all research and development work concerning the optimization of oil consumption. Monitoring the endurance test

For all research and development work concerning the optimization of oil consumption. Production monitoring

One-time measurement during development or in the field

Monitoring the endurance test

Online measurement of a speed load point

Yes

No

No

No

Measurement while the engine is running

Yes

Yes

No

Yes

Setup duration

4–5 hours (non-recurrent)

15 min

2–3 hours for each test

15 min per test

Minimal measuring duration

5 hours

3 hours per test, which must be repeated three times

10 hours per test, which must be repeated three times

5 hours per test, which must be repeated three times

Measurement sensitive to

Crankcase dynamics, oil aging and ventilation, oil temperature fluctuations, altering the installation height of interacting vessels, and dilution by fuel

Crankcase dynamics, oil aging and ventilation, oil temperature fluctuations, instability of the oil level because of return flows, and dilution by fuel

Oil return flow and hang-up characteristic of the engine, engine cooling, engine shutdown sequence errors, and drainage interval errors

Oil ventilation and oil temperature differences in the engine and inspection glass

Figure 28.11 Comparison: different oil consumption methods.

Internal Combustion Engine Handbook | 969

6606_Book.indb 969

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

The vortex frequency or vortex flow principle is based on the implication that an immersed body can cause an eddy flow and that a constant ratio of vortex distance and vortex effect is generated behind a cylindrical impact body depending on the approach flow velocity. The vortex frequency is proportional to the flow velocity when the flow is equal. An ultrasonic signal that is send vertically to the flow direction and modulated by the vortex is used to collect data. •• With the orifice measuring principle, as already mentioned with the intake air quantity measurement, the differential pressure is measured at a flow obstruction using a differential pressure sensor. The principle enables a large measuring range, return flows can also be recorded through the symmetrical arrangement of the measuring orifice, and practically no deposits occur at the sharp edge of the pinhole aperture that would alter the cross section. The characteristics of various blow-by measuring procedures are summarized comparatively in Figure 28.12 [28-20]. The accuracy statement alludes to measurement uncertainty and refers to the measuring area (full scale). Because the blow-by is generally heavily polluted with oil and the flow can also be strongly pulsating or even flowing backward, it is advantageous to use measuring procedures that are not sensitive to pollutants with the relevant structure and that are able to detect the direction of flow, such as the orifice measuring principle.

28.10 Urea Consumption Measuring Technology Modern exhaust gas after-treatment concepts for diesel engines use selective catalytic reduction (SCR) systems because of their high efficiency at reducing emissions of oxides of nitrogen

Float volumetric flow measurement Accuracy

Hot film anemometer

(NOx) by 60–80% and their associated potential of lowering fuel consumption with engine-internal measures by up to 5%. An SCR system consists of a low-pressure injection system, which injects a watery urea solution (also known by its brand name AdBlue) as a reduction agent into the exhaust flow. It also consists of a catalytic converter, where the ammonia that resulted from the hot exhaust gas stream reacts with NOx to form nitrogen and water. The necessary calibration of urea dosing quantities for the overall engine, injector, and catalytic converter system via the overall engine map requires a precise urea consumption measurement at the engine test bed and stationary conditions, in the same way as during the highly dynamic test cycles (emission test cycles—ETC and federal test procedure—FTP). Unlike the fuel consumption measurement, extremely small flow rates of lower than 10 g/h (<5% of fuel consumption) should be taken in account here, therefore directly resulting in three essential physical requirements for accurate urea consumption measurements: 1. The distance should be as small as possible from the measuring device to the injector. Minimizing the dead space between the sensor and the injector reduces apparent flows caused by temperature changes (see the section on fuel conditioning). 2. Pressure pulsation caused by the low frequency urea injection is present near the injector. This pressure pulsation may not disturb the measuring device. On the other hand, the measuring device may not alter the conditions in the injection system (∆p = 0). 3. The complete ventilation of the SCR injection system. The transportation of air bubbles in the piping system may not be sufficiently secured at potential flow rates of <10 kg/h to guarantee timely expulsion and separation.

Impeller, turbine wheel, impeller

Gas meter

Vortex frequency flow measurement

Orifice measuring principle

~ 5% FS

2%

2% FS

1%

1 - 2% FS

1% FS

0.5%

0.2 - 0.5%

n/a

0.3%

0.5%

< 0.1%

average

large

average

large

large

very small

approx. 8

approx. 28

approx. 6

approx. 0.5

approx. 7

0.2

approx. 1

approx. 0.1 s

approx. 1

approx. 10

approx. 0.002

approx. 0.01

Detection during flow reversal

no

no

no

no

no

yes

Counterpressure at half volumetric flow [Pa]

80

100–800

200

50

400

60

Counterpressure at full volumetric flow [Pa]

500

600–3000

1200

300

1000

300

Reproducibility Sensitivity to pollutants Smallest recordable measurement value [l/min] Response behavior t 90 [sec]

positive

average

bad

Figure 28.12 Comparison of various blow-by measurement measuring principles.

970 | Internal Combustion Engine Handbook

6606_Book.indb 970

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.10 Urea Consumption Measuring Technology

The PLU measurement principle (see fuel consumption measurement section) is very well suited to the smallest quantity measurement under these conditions. Figure 28.13 shows a diagram of a system solution with automatic ventilation system

and protection against drainage with automatic retractive suction when switching off the SCR system. Figure 28.14 shows the dynamic consumption measurement of the urea injection during a transient ETC test cycle.

Figure 28.13 Schematic design of the AVL PLUrea urea consumption measurement system.

Figure 28.14 AVL PLUrea urea consumption measurement: AdBlue volumetric flow rates, accumulated flow volume (above), and pressure and temperature (below), section from a European test cycle (ETC cycle) at 4-Hz timing frequency.

Internal Combustion Engine Handbook | 971

6606_Book.indb 971

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

A dynamic online correlation to the target quantity required by the control device enables deviations or error functions in the dosing system to be immediately detected and to assign specific operating states or causes [28-21].

determined from these travel times, and an additional pressure measurement allows for the mathematical determination of the mass flow and standard volumetric flow [28-24]. Difficulties primarily include the high temperatures required and high sampling rates. Traditional piezoelectrical ultrasonic transducers now allow maximum temperatures of up to 250°C and sampling rates of maximum 20–50 Hz with operation in hot gas. Individual applications can already be covered and very good results can be achieved in some cases using this concept and alternative ultrasound transducer concepts [28-25]. The high temperature area and quick sampling rates for avoiding aliasing effects during exhaust gas pulsation have yet to be achieved. Directly measuring the volume of exhaust gas is possible using the Kármán vortex street method (vortex principle). The proportional frequencies of the flow velocity of the gas and the flow vortex caused by a flow obstacle are detected using a pressure measurement [28-26]. Measuring the pressure loss at a defined flow obstruction, for instance, at an orifice (contraction) or at a simple pipe, is also recommended as a direct measurement of the volume of the exhaust gas. The pressure loss at the flow obstacle, which is both flow dependent and not to be neglected has a negative impact. Moreover, thermal effects and resonance phenomena lead to return flows in the exhaust gas system, which generally speaking cannot be detected using current measuring procedures.

28.11 Direct Exhaust Quantity Measurement Limits are set on the smallest quantities of pollutants that can be detected when measuring emissions using constant flow dilution sampling [constant volume sampling (CVS) principle]. Even highly accurate detectors reveal measurement uncertainties which can no longer be neglected for emissions in the stationary operation of modern vehicles with additional dilution. Therefore, it would be desirable to forgo dilution of these already low emissions. In addition to the high costs of a CVS system, the large amount of space needed has also led to the search for alternatives concepts. The requirements imposed on measuring technology such as this are, however, extraordinarily high [28-22], [28-23]: •• Temperatures of up to 450°C even at the end of the exhaust gas system •• Virtually no pressure drop, so as not to influence the engine operation •• Pulsation rates of between 5 and 50% depending on where the sensor is fitted •• At high chronological resolutions of the measurement data (~20–50 Hz), internal sample rates are required for the sensors from 400 Hz to 5 kHz to avoid aliasing effects.

28.12 Exhaust Gas Measuring Technology

•• Measurement uncertainty of <1% of the measurement value

Besides lowering fuel consumption, minimizing motor vehicle pollutant emissions is regarded as the second essential development task for modern engines. Figure 28.15 provides a rough overview of the most important exhaust gas components during complete and incomplete combustion as well as their concentrations. Various legal limits apply to the permissible mass emission, depending on the pollutant components and engine/vehicle category, which apply to the route (g/km) for motorbikes and passenger cars and to work performed (g/ kWh) for utility vehicles.

•• The lowest possible dead times, to deliver a timely multiplication of the measured concentrations with the exhaust mass flow even with the dynamic operation of the engine. Ultrasonic runtime methods are overwhelmingly used as the measuring technique because of positive experiences in industrial plant engineering. This involves sending an ultrasonic signal once in the direction of the flow and once in the opposite direction and then measuring the travel time. The current flow velocity as well as the speed of sound is

Fuel (CnHmOi) + Air (O2 + N2) lete n mp Co ustio b m o C

CO2

< 15 %

+

H2O

< 10 %

+

N2

< 72 %

+

Inc Co omp mb lete us tio n

O2

CO

< 20 %

<1%

+

HC

< 0,5 %

+

NOx

< 0,5 %

+

PM

< 0,5 %

Figure 28.15 Reference values for exhaust gas composition.

972 | Internal Combustion Engine Handbook

6606_Book.indb 972

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.14 Pollutant Mass Determination at Undiluted Exhaust Gas

28.13 Measurement of the Volumetric Concentration of Gaseous Exhaust Gas Components The concentration of the individual pollutant components is determined using gas analyzers. The measurement principles employed (Figure 28.16), even with research and development test beds, mostly relate to the exhaust gas legislation specifications.

Gas components

Detector measuring principle

CO (carbon monoxide) CO2 (carbon dioxide)

Nondispersive infrared detector

NDIR

NOx (oxides of nitrogen)

Chemical luminescence detector or ultraviolet resonance absorption detector

CLD or NDUV

HC (hydrocarbons)

Flame ionization detector

FID

O2 (oxygen)

Paramagnetic detector

PMD

Abbreviation

Figure 28.16 Standard measurement procedures for determining the concentration of important gaseous exhaust gas components

All common types of detector ascertain the volumetric concentration of the individual components. Please refer to the appropriate literature regarding their functioning [28-27]. The sometimes non-negligible cross sensitivities of many of the instruments should be taken into account. For instance, CO and CO2 are practically only determined in dried exhaust gas for undiluted exhaust gas because of their extreme cross sensitivity to water vapor. The above listed measuring principles, also often referred to as conventional measuring techniques, are the most widespread and are prescribed by most exhaust gas legislation. In addition, other methods are required and employed in the research and development area to be also able to measure gas components, which are not/not yet limited. This is especially imperative with new diesel engine concepts with exhaust gas after-treatment systems, such as NOx storage catalytic converters, SCR catalytic converters, and diesel particulate filters. Multiple-component measuring systems are used in

these cases. The most widespread system is Fourier transform infrared spectroscopy (FTIR). FTIR is an optical infrared absorption measuring method, which is able to measure a multitude of exhaust gas components at the same time. Using a Michelson interferometer, the intensity of individual infrared wavelengths is constantly changing. This leads to absorption of the individual wave lengths in the measurement cell by different gas components in the exhaust gas sample. The concentration of the individual gas components can be determined using complex mathematical formulas from these infrared spectra [28-28]. The engine’s mass emissions are determined by mathematically linking the measured volumetric concentrations with the relevant masses or volumetric flow of the exhaust gas considering the physical properties of the exhaust gas and the components analyzed. For instance, a mathematical adjustment of the volumetric concentrations must be performed even for the above-mentioned measurement in the dried exhaust gas.

28.14 Pollutant Mass Determination at Undiluted Exhaust Gas An undiluted measurement is when a partial flow taken directly from the exhaust flow is led to the various detectors in the analytical system without having been diluted at all. To determine the mass emissions, we can either directly measure the exhaust gas flow (see above). Alternatively, it is possible to analyze the entire engine as an enclosed system, for which the mass balance requires that the air and fuel flows must be the same as the exhaust gas flowing out of the system. These principles, which also form the basis of many other guidelines, allow the simple determination of the exhaust mass flow, however, only when the engine is in stationary operation. Standard ISO 16183 [28-29] and the Euro IV utility vehicle exhaust gas legislation [28-30], which builds on this also allows this method in the transient engine operation. Undiluted measurement in the dynamic engine operation (“modal analysis”) is widespread in research and development, because it can be carried out on any test bed with the appropriate measuring technology (Figure 28.17). However, for

1: CEB II 2. SPC

GFuel = Fuel Flow Rate GIntake = Intake Air Flow Rate

3: Opazimeter

GTp = Tail Pipe Exhaust Flow Rate

Figure 28.17 Engine test bed with exhaust gas measuring system for undiluted measurements.

Internal Combustion Engine Handbook | 973

6606_Book.indb 973

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

the moment, it is not yet permitted for evaluating dynamic test runs that conform to the guidelines (see below). A key factor during the evaluation of such measurements in the dynamic operation is that all the signals needed for the calculation must be corrected by their dead times. Also bear in mind that the undiluted measurement is not suitable for particulate measurements (see below).

Dilution tunnel Dilution air

Full Flow Dilution Tunnel

Secondary tunnel

Dilution tunnel Dilution air

28.15 Pollutant Mass Determination at Diluted Exhaust Gas In the transient engine operation, the dilution method, preferably the full flow dilution, had to, until recently, be applied for the pollutant mass determination conforming to the guidelines. With ISO standard 16183 [28-31], which on the one hand defines the correct modal analysis and on the other hand the correct partial flow dilution for the transient engine operation, this has been incorporated into the European legislation governing commercial vehicles [28-32]. This is likewise already envisaged for pending U.S. legislation [28-33]. The dilution method can generally be divided into full flow and partial flow dilution: •• With a full flow dilution system, [28-34], [28-35] (Figure 28.18), the entire exhaust mass flow from the engine is ideally mixed, that is, diluted, in the dilution tunnel with filtered air. The overall flow at the dilution tunnel discharge is kept roughly constant, hence the designation “constant volume sampler.” The use of a CVS system is practically mandatory for all exhaust emission measurements in the dynamic engine operation that conform to guidelines. The general objective is to determine the total pollutant mass emitted during a test cycle. This can involve determining the accumulated gaseous mass emission from the engine either by a mathematical integration of the chronologically resolved mass flows measured in the dilution tunnel or with a pneumatic integration, where a partial flow, which is always proportional to the main flow of the diluted exhaust gas, is collected in bags and then the volume concentration of the components collected in the bag is determined. The latter method is not suitable for all pollutant components because of the various absorption effects. Specific to particulate measurements, direct sampling of a proportional partial flow is possible only if the maximum mixing temperature does not exceed 52°C (see Exhaust Gas Particulate Measuring Technology section below). If the temperature limit is exceeded during the test, then a second dilution level can be used according to the type of partial flow dilution systems mentioned hereinafter. •• With a partial flow dilution system (Figure 28.18), only a constant small percentage of the exhaust mass flow is fed into an appropriately smaller “tunnel” and is diluted there with clean air [28-36]. Because the costs and space requirements of full flow dilution systems are very high, particularly for utility vehicle test benches, the use of partial flow systems for measuring particulates has been permissible in the

Partial Flow Dilution Tunnel

Total sampling type

Figure 28.18 Schematic representation of the full flow and partial flow dilution systems.

stationary engine operation for some time already [28-37], [28-38], [28-39]. Systems have recently been implemented, which see the entire partial-flow-diluted exhaust flow being guided via the measuring filter (a so-called “total-samplingtype partial flow dilution tunnel”). Even though it is smaller physically, a partial flow dilution system is more complex than a CVS system, because the sample flow must be regulated at a constant ratio to the exhaust flow. Improvements have been made to the controls of modern systems to the extent that the proportionality of the mass flows is even possible for transient driving cycles. Such partial flow dilution systems are only used in exceptional circumstances to determine the mass emission of gaseous pollutants. They are, however, often used as a second dilution level in CVS systems. Please refer to Chapter 21 for further information regarding guidelines and compliant driving cycles, as well as limit values.

28.16 Exhaust Gas Particulate Measuring Technology Unlike gases, which are generally only subject to slight alterations in the exhaust gas system and in the measuring chain, the so-called “volatile” share of the particulates are formed only during cooling and dilution. “Diesel particulates” in the sense of the legislator are therefore defined by the CVS method (see above) [28-40], [28-41]. The most important requirement of particulate measuring technology is that the temperature of the diluted exhaust gas is less than 52°C at the end of the measuring section.

28.16.1 Gravimetric Particulate Measurement

The gravimetric particulate measurement is performed to ascertain the integral measurement of the particulate emission during a certain measuring phase, for instance, during an entire dynamic driving cycle. A constant partial flow sampled from the diluted exhaust gas is drawn over a Teflon-coated glass fiber filter or via a Teflon filter. The mass of the absorbed

974 | Internal Combustion Engine Handbook

6606_Book.indb 974

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.16 Exhaust Gas Particulate Measuring Technology

then displayed as a filter smoke number or as a pollution level (in %). Within a 20% confidence band, the soot content in the exhaust gas can be determined from the paper blackening and effective intake length and can be displayed as a soot concentration (in mg/m3).

28.16.2 Dynamic Particulate Measurement

•• Continual mass determination by determining the sizes and thicknesses of the particulate via electrical mobility analysis combined with an electrical impactor. This device has only recently come onto the market under the designations, “Dekati mass monitor” and “microparticle monitor.”

Dead volume

Effective length

The most important processes not just for determining the emission integrally from the filter weight but also for being able to observe the time sequence of the emission during the measurement of the exhaust gas from the dilution tunnel include:

Suction volume

particulate is determined from the weight increase of the filter. Because the particulate mass at the filter falls as emissions drop, the ambient parameters, such as dilution air temperature and moisture, and filter temperature, are tolerated for longer in the more recent legal regulations. Consequently, with an optimized setting of dilution air flow and sample quantity via the filter, a sufficiently reproducible measurement is even possible for low emissions required according to Euro 4 or US2007 [28-42], [28-43].

Filter paper

•• tapered element oscillating microbalance (TEOM): Mass determination based on the frequency change of an oscillating glass cannula, through which the exhaust gas is guided and onto which a filter is placed, the weight increase of which causes the frequency change.

Figure 28.19 Measuring principle—smoke meter.

28.16.3 Measurement of the Soot or Smoke Emission

28.16.5 Opacity Measurement Using “Opacimeters”

“Soot” or “smoke” is the most noticeable emission from diesel engines and a dominant element of particulate emissions. It mainly consists of “black” elemental carbon particles. The measurement and minimization of soot emissions is a standard task in the development of diesel engines. Of the measuring methods described in the following, the first two have been used for some time, while the newer methods of photoacoustics and laser-induced incandescence (LII) were specially developed for modern low-emitting engines and are associated with increased measurement technology expense.

28.16.4 Smoke Emission Value Measurement with “Smoke Meter”

The smoke emission value measurement has been successfully used for a number of decades for stationary emission optimization at the test bench. Thanks to its easy handling, it is the most widespread test bench measuring method for quantifying soot at diesel engines [28-44]. During the smoke emission value measurement, a freely selectable volume of exhaust gas is sucked in through a clean filter paper (Figure 28.19). The drawn volume of exhaust gas is measured using an orifice measuring section and the effective intake length is calculated using this information (405 mm, pursuant to ISO 10054). The blackening (paper blackening) of the filter paper, caused by the soot in the exhaust gas, is determined using an optical measuring head. The result is

Effective length =

Suction volume - Dead volume - Leakage volume Filter paper

An opacimeter measures the light attenuation (turbidity, opacity) through the particles, especially the soot particles, in the diesel exhaust gases. A measuring chamber with defined measuring length and a nonreflective surface is uniformly filled with exhaust gas. The light attenuation between a light source and a receiver is measured and the result is used to calculate the opacity of the exhaust gas. The basis for this calculation is the Beer–Lambert law (Figure 28.20). This method is mandatory for some legal inspections (e.g., “ECE R24 and ELR tests”). It permits the chronologically resolved measurement of short soot impacts and is until now the standard method for dynamically optimizing the particulate emissions of diesel engines [28-45].

28.16.6 Photoacoustics (PASS)

In the photoacoustics principle (Figure 28.21), intensity-modulated light is radiated through the measuring gas with the “black,” that is, highly absorbent, soot particles. The periodic warming and cooling and resultant expansion and contraction of the carrier gas can be regarded as a sound wave and detected using microphones. Clean air does not produce a signal. With soot-charged air or exhaust gas, the signal increases in proportion to the concentration of the soot in the measuring volume. This constitutes a major advantage over opacimetric methods, where the sample from the “zero signal” of 100% light intensity, the opacity, forms the measured value.

Internal Combustion Engine Handbook | 975

6606_Book.indb 975

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

Lamp

I0

T, p

Beer-Lambert

I

Detector

L I/I0 = Extinction = Absorption + Radiation I0 I I N – · k L =e = 1– [%] N I0 100 [m–1] k [m] L – · k L ) N = 100 · (1 – e T0 [K] [K] T T · p0 k0 = R · konz0 = k · T0 · p p0 [Pa] [Pa] p [m2/g] R 2 konz [g/m ] Index0

... Light intensity without smoke ... Light intensity with smoke ... Opacity ... Absorption coefficient ... Length ... 273 K, Temperature ... Temperature in the measuring chamber ... 1.013 · 105 Pa, Standard pressure ... Pressure in the measurement ... Mass extinction coefficient ... "Smoke" coefficient ... Values for standard requirements

Figure 28.20 Measuring principle of an opacimeter.

28.16.9 Particle Group and Particle Total Microphone as a detector

Modulated expansion

Soot particle

Modulated laser beam

See Chapter 26.6.3.12.

Sound wave

Modulated heating

Figure 28.21 Measuring principle of photoacoustics.

Thus, significantly increasing the sensitivity, the detection limit is less than 10 μ g/m3 [28-32], [28-33].

28.16.7 Laser-Induced Incandescence

The particles are heated by a high-intensity laser to approximately 4000 K. The maximum intensity of the light radiated from the particles can be used to calculate back the mass concentration in the sample volume, while the decay behavior allows conclusions to be drawn about the average primary particle diameter. This method is predominantly used in the measurement of carbon particles [28-46].

28.16.8 Scattered Light Measurement

Methods which simply evaluate the scattered light signal, such as those in nephelometers and tyndalometers, are primarily suitable for larger particles. Devices such as these are also available in a wide variety of handsets. A proportionality to mass charging of the exhaust gas with soot or particulates is, however, possible only to a very limited extent. Consequently, this method is hardly ever used for test bench measurements.

28.16.10 Measurement Data Processing and Evaluation

The test bench automation system with the entirety of the various subsystems is responsible for the essential tasks of collecting and processing measurement data. To do this, first the data recording and data reduction of the measuring and processing values from a few hundred measuring channels need to be counted simultaneously. When performing measurements in the stationary engine operation, the measurement results from the measuring channels calculated with a specific sampling rate are acquired by forming the mean value over a specific measurement time. Indicating signals such as the combustion chamber pressure sequence are generally not chronologically averaged; rather an average combustion cycle is calculated from a certain number of combustion cycles. With dynamic tests, the sequence of the measured values is recorded with a specified sampling frequency. In many cases, preprocessing by integration may be expedient, for instance, when determining the fuel mass consumed during a driving cycle. The quality of the measurement data obtained in this way is an essential requirement for gathering reliable statements about engine behavior. Evaluating the collected data before it is reused is therefore of vital importance. This involves discovering possible anomalies in the measuring system, not in the engine. Errors should be understood as deviations in the measurement values that depart from the normal results scatter. The high rate of measurement and processing data is a serious challenge for the necessary speedy evaluation of the data, which should, if possible, be performed during the test run. An evaluation of all the measurement channels by the measurement engineer or engine developer is rarely possible without automated assistance. Simple data evaluation methods such as limit value monitoring are usually already part of the

976 | Internal Combustion Engine Handbook

6606_Book.indb 976

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

28.16 Exhaust Gas Particulate Measuring Technology

Physical Models & Rules Library Online Data Interface

Online Data

Online Data Access

Online Data

Data-based Models Library Diagnosis Rules & Support Library

Physics-based Check Modules Data-driven Check Modules

Result Display

Result Unifier Module

Faulty Channels Isolation Module

Result Store

Diagnosis & Support Module

Figure 28.22 Plausibility evaluation of the measurement and processing data.

measurement or automation system. Signal processing methods such as recording interference peaks and parasitic frequencies are available for evaluating high-frequency collected data. However, it seems that an additional evaluation of the plausibility of the measurement data is required to be able to ensure measuring quality. This evaluation comes from an awareness of the connections between the measurement values from different data channels. Examples of such validation rules include mass and volume balances or known size relationships in the form of inequalities for temperature and pressure measuring chains. In addition to valuation rules such as these (which can be accurately predicted physically), there is also a multitude of practical knowledge that describes “broad” connections between measuring channels in various engine states. Automated evaluation systems can help here to evaluate the plausibility of the measurement data even in complex and quick measuring systems, (Figure 28.22). Certain configurable modules can perform the evaluation according to physical validation rules that can be parametrized by the user. Other self-learning modules can be implemented alongside, in the first instance to extract “practical knowledge” from the data material already available and then to use this knowledge to evaluate new data.

28-4. Fortuna, T., Mayer, M., Pflügl, H., and Gschweitl, K. “Optimierung von Verbrennungsmotoren mit DoE und CAMEO,” HDT—Second Tagung DoE in der Motorenentwicklung, Berlin, 2003.

Bibliography

28-14. Schmitt, A. et al. “Methodik und Sensorik zur Applikation von Harnstoff-SCR-Systemen am dynamischen Motorprüfstand.” VDI— Conference: NOx—Control, Nuremberg, 2009.

28-1. Kleppmann, W. 2008. Taschenbuch Versuchsplanung—Produkte und Prozesse optimieren, fifth edition. Munich. ISBN 3-446-41595-9. Carl Hanser Verlag München Wien 28-2. Riel, A., Hasewend, W., Bogner, E., and Fischer, R. “Vehicle and Powertrain Modelling in the hole Development Process,” Automobiltechnische Zeitschrift 106(6): 522–531, 2004. 28-3. Apschner, M. and Hauser, G. “PUMA Open: Prüfsystem für Motor und Antriebstrang,” Motortechnische Zeitschrift MTZ 103(3): 228–234, 2001.

28-5. ISO/BIPM guideline. 2008. “Guide to the Expression of Uncertainty in Measurement,” revised eighth edition, Geneva. International Organization for Standardization 28-6. Bier, M. et.al. “Hybridentwicklung auf dem Motorenprüfstand—Ein wichtiger Schritt zu mehr Effizienz im Entwicklungsprozess,” Fourteenth VDI Symposium—Erprobung und Simulation in der Fahrzeugtechnik, 2009. 28-7. Denger, D. et.al. “Effizientere Konzeptevaluierung und Kalibration von Fahrzeugflotten am Motorenprüfstand,” Fourteenth VDI Symposium— Erprobung und Simulation in der Fahrzeugtechnik, 2009. 28-8. Schyr, C. AVL-Indizier-TechDay, Darmstadt, 2008. 28-9. Weidinger, C. R. “Messgüte am Prüfstand für die Motorenentwicklung.” Doctoal Thesis, TU Wien, 2002. 28-10. Gohl, M. “Massenspektrometrisches Verfahren zur dynamischen Online-Messung der Ölemission von Verbrennungsmotoren.” Doctoral Thesis, Hamburg University of Technology (TUHH), 2003. 28-11. Kammerstetter, H. and Werner, M. 2004. Vorbedatung von Steuergeräten am Komponentenprüfstand unter Einbezug kompletter Diesel-/Otto-Einspritzsysteme. Expert-Verlag: Renningen. Messtechnik und Simulation in der Motorentwicklung. Expert-Verlag:Essen. 28-12. Winklhofer, E., Beidl, C., Hirsch, A., and Piock, W. “Flame Diagnostics for Performance and Emissions Development,” Motortechnische Zeitschrift 65(5): 362–369, 2004. 28-13. Werner, M. “Kraftstoffverbrauchsmessung: Kürzere Entwicklungszeiten durch kontinuierliche Messtechnik,” Motortechnische Zeitschrift 99(11), 1997.

28-15. Ebner, H. W. and Köck, K. “Coriolis Fuel Meter—A Modern and Reliable Approach to Continuous and Accurate Fuel Consumption Measurement.” SAE2000 World Congress, Detroit, 2000. 28-16. Graf, F., Hofmann, P., Köck, K., and List, R. “Highly Efficient Test Beds with State-of-the-Art Fuel Consumption Instrumentation,” Motortechnische Zeitschrift 65(7–8), 2004.

Internal Combustion Engine Handbook | 977

6606_Book.indb 977

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 28 Engine Measuring Technology

28-17. Ebner, H. W. and Jaschek, A. O. “Die Blow-By-Messung—Anforderungen und Messprinzip,” Motortechnische Zeitschrift 59, 1998.

28-33. Directive 1991/542/EC of the European Parliament and of the Council, 1991.

28-18. Wiesinger, M. “Entwicklung eines Abgas-Massenflusssensors.” Doctoral Thesis, TU Wien, 1999.

28-34. Directive 1999/96/EC of the European Parliament and of the Council, 1999. http://eur-lex.europa.eu/legal-content/EN/ TXT/?uri=CELEX:31999L0096.

28-19. Lohde, S. “Direkte Abgasdurchflußmessung mittels Ultraschalllaufzeitverfahren.” Doctoral thesis, Naturwissenschaftliche Universität Lübeck, 2001. 28-20. Yassine, M. K. “Parameters Affecting Direct Vehicle Exhaust Flow Measurement.” Society of Automotive Engineers, SAE Paper 2003-01-0781, 2003. 28-21. Gätke, J. 1991. Akustische Strömungs und Durchflussmessung. Akademie-Verlag-Berlin:Berlin. 28-22. Beck, M. and Hinterhofer, K. “Direct dynamic Flow Measurement in the Exhaust of Combustion Engines.” Society of Automotive Engineers, SAE Paper 980880, 1998, pp. 95–104. 28-23. Hardy, J. E., McKnight, T. E., Hylton, J. O., and Joy, R. D. “Real Time Exhaust Gas Flow Measurement System.” Society of Automotive Engineers, SAE Paper 982105, 1998. 28-24. Klingenberg, H. 1995. Automobile Exhaust Emission Testing. Volume C: Exhaust Emissions Tests. Springer-Verlag: Berlin; Heidelberg; New York. 28-25. Engeljehringer, K. “Multi Component Emission Measurement System for Engine Development,” Motortechnische Zeitschrift (7–8), 2006. 28-26. Directive 2005/78/EC of the European Parliament and of the Council. http://eur-lex.europa.eu/legal-content/EN/ TXT/?uri=CELEX%3A32005L0078. 28-27. Code of Federal Regulations. “Title 40 Part 1065—Engine Testing Procedures,” https://www.law.cornell.edu/cfr/text/40/part-1065. 28-28. “ISO 16183—Heavy duty engines—Measurement of gaseous emissions from raw exhaust gas and of particulate emissions using partial flow dilution systems under transient test conditions,” http://www.iso.org/ iso/home/store/catalogue_tc/catalogue_detail.htm?csnumber=68439. 28-29. Directive 2001/1/EC of the European Parliament and of the Council, 2001. http://eur-lex.europa.eu/legal-content/CS/ TXT/?uri=celex:32001L0097. Accessed May 15, 2001.

28-35. ISO 8178-1: 2006—Reciprocating internal combustion engines— Exhaust emission measurement—Part 1: Test bed measurement of gaseous and particulate exhaust emissions. http://www.iso.org/iso/iso_catalogue/ catalogue_tc/catalogue_detail.htm?csnumber=42714. 28-36. Schweinle, G. and Graf, A. “Interaktion von Testverfahren und Fahrzyklen bei PKW Abgasmessungen.” Third International Exhaust Gas and Particulate Emissions Forum, 2004. 28-37. Vouitsis, E., Ntziarchristos, L., and Samaras, Z. “Particulate Matter Mass Measurements for low Emitting Diesel Powered Engines: What’s Next?” Progress in Energy and Combustion Science 29:635–672, 2003. 28-38. Khalek, I. A. “2007 Diesel particulate measurement research, Final Report, Project E-66 Phase 1.” Southwest Research Institute Report 03.10415, 2005. 28-39. Stein, H. J. “Weiterentwicklung der Partikelmessung für Nutzfahrzeugmotoren nach 2005—ISO 16183 und US 2007.” Second International Exhaust Gas and Particulate Emissions Forum, 2002. 28-40. Silvis, W. “Standardized dilution conditions for gravimetric PM sampling—measures to assure results that correlate.” Fifth International Exhaust Gas and Particulate Emissions Forum, Ludwigsburg, 2008. 28-41. Schindler, W., Nöst, M., Thaller, W., and Luxbacher, Th. “Stationäre und transiente messtechnische Erfassung niedriger Rauchwerte,” Motortechnische Zeitschrift 62(10): 808, 2001. 28-42. Pongratz, H., Schindler, W., Singer, W., Striock, St., and Thaller, W. “Transient Measurement of Diesel Engine Emissions,” Motortechnische Zeitschrift 64(10): 824, 2003. 28-43. Beck, H. A., Niessner, R., and Haisch, Ch. “Development and Characterization of a Mobile Photoacoustic Sensor for On-Line Soot Emission Monitoring in Diesel Exhaust Gas,” Analytical and Bioanalytical Chemistry 375:1136, 2003.

28-30. CFR. 66(12), Part 86, pp. 86007–11, 2001. www.epa.gov/omswww/ regs/hd2007/frm/frdslreg.pdf. Accessed January 18, 2001.

28-44. Haisch, C. “Simultaneous Detection of Gaseous and Particulate Exhaust Components by Photoacoustic Spectroscopy.” Fifth International Exhaust Gas and Particulate Emissions Forum, Ludwigsburg, 2008.

28-31. Silvis, W. M., Marek, G., Kreft, N., and Schindler, W. “Diesel particulate measurement with partial flow sampling systems: A new probe and tunnel design that correlates with full flow tunnels.” SAE Technical Paper Series No. 2002-01-0054.

28-45. Schindler, W., Haisch, Ch., Beck, H. A., Niessner, R., Jacob, E, and Rothe, D. “A Photoacoustic Sensor System for Time-Resolved Quantification of Diesel Soot Emissions.” SAE Technical Paper Series No. 2004-01-0968.

28-32. Steigerwald, K. “Weiterentwicklung eines Teilstromverdünnungssystems für die Partikelmessung am instationär betriebenen Dieselmotor.” Doctoral thesis, Technische Universität Darmstadt, 2008.

28-46. Schraml, S., Heimgärtner, C., Will, S., Leipertz, A., and Hemm, A. “Applications of a New Soot Sensor for Exhaust Emission Control Based on Time Resolved Laser Induced Incandescence.” SAE Technical Paper Series No 2000-01-2864.

978 | Internal Combustion Engine Handbook

6606_Book.indb 978

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29 Hybrid Drive Systems 29.1 History In the early stages of automobile development various drive concepts competed with one another. In addition to gasoline and diesel engines, steam engines and electric motors were also used as vehicle drive systems. Ferdinand Porsche, who in 1900 was working for “K.u.K. Motorenwagen- und AutomobilFabrik Jacob Lohner&Co” is considered to be one of the first to have developed a vehicle with a hybrid drive system. The “Lohner-Porsche Mixte” involved a series hybrid drive system with wheel hub electric motor and a four-cylinder combustion engine manufactured by Daimler. Other engineers such as Henri Pieper und Louis Antoine Kriéger [29-1] were also experimenting with hybrid drive systems at this time. In 1904, Henri Pieper among others registered a patent for a “regulating device for explosive force engines coupled with dynamos” at Austrian patent office

[29-2]. From the patent specification, it can be deduced that he had already worked on hybrid drive systems previously. In stark contrast to present day engineering goals, emissions and fuel consumption of the vehicle drive systems were at that time presumably not the motivating factor behind the hybrid drive systems. Rather it was a case of circumventing the insufficient durability of mechanical components, such as the transmission. In the years that followed, the combustion engine prevailed as the sole drive system, particularly for passenger cars, and the electric, hybrid, and steam drive systems disappeared from the market. One of the last manufacturers of hybrid vehicles from this period was the Owen Magnetic brand, which produced hybrid vehicles up to 1921. Aside from isolated vehicle studies and temporary solutions as a result of fuel scarcity in the aftermath of the Second World War, vehicle development concentrated almost exclusively

Figure 29.1 1902, Ferdinand Porsche at the wheel of a LohnerPorsche gasoline-electric-powered automobile in front of his birthplace in Maffersdorf (present day Vratislavice nad Nisou, Czech Republic) [29-3].

Internal Combustion Engine Handbook | 979

6606_Book.indb 979

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

DC to AC Modulating Inverter Lead acid Batteries

Commutating Capacitors

Battery Charging Control Helium Reservoir Sterling Engine Radiator and Fan

Logic & Inverter Controls Induktion Motor (3-Phase)

Starter Motor Alternator

Combustion Air Blower

Figure 29.2 GM Stir-Lec I based on an Opel Kadett [29-5].

on combustion engines. It was not until the end of the 1960s when serious problems associated with air pollution arose for the first time in some major cities and the first exhaust emission standards were subsequently imposed that research again increasingly started to focus on alternatives to gasoline and diesel engines. In addition to gas turbine drive systems and electric vehicles, new hybrid drive systems were also developed, but which never made it to volume production. In 1968, GM [29-4] presented its Stir-Lec I study. A hybrid drive system was constructed based on an Opel Kadett, consisting of a Stirling engine and an electric motor (see Figure 29.2).

Just a year later another concept vehicle was presented by GM in North America: the XP-883 was based on a Vauxhall Chevette (similar to a Chevrolet Chevette or an Opel Kadett City) (Figure 29-3). Unlike the Stir-Lec I, both the combustion engine and the electric motor were now located in the front and the combustion engine was a smaller two-cylinder gasoline engine with 0.573 l (35 in.3) displacement [29-6]. The Japanese company Toyo Kogyo presented a prototype at an automobile exhibition in Tokyo of a future city car, the Mazda EX 005 (see Figure 29.4). Its drive system was a combination of a two-disk rotary piston engine and an electric

Control system Gasoline engine

Transmission Electric motor Battery

Figure 29.3 GM XP-883 [29-6].

980 | Internal Combustion Engine Handbook

6606_Book.indb 980

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.1 History

Figure 29.4 Mazda EX 005 [29-7].

motor with an output of 3 kW [29-7]. As can be seen in Figure 29.4, this vehicle was more of a concept study rather than a close-to-production prototype. In North America in 1973, Victor Wouk and Charlie Rosen combined a Mazda Wankel engine and an electric motor in a 1972 Buick Skylark. The vehicle surpassed the exhaust emission standards valid at that time, consumed 7.84 l/100 km (30 mpg) and reached a high speed of 137 km/h (85 mph). The experimental vehicle was promoted as part of the “Federal Clean Car Incentive Program.” On top of the problems associated with air pollution caused by automobile emissions, in the 1970s, there was also the oil crisis, which meant that lower consumption drive systems received a further push. Research was being conducted into alternative drive systems at the Technical University of Berlin, as well as a number of similar institutions. Work was being carried out on many projects including a Fiat with an electric and hydro drive system as well as a BMW 2000 with a combustion engine and hydraulic motor [29-8] (Figure 29.5). Another interesting drive system combination was presented by Toyota 1977: a Sports 800 with a gas turbine and an electric motor (see Figure 29.6). Back then gas turbines were certainly considered as an alternative to gasoline and diesel engines. In that same year, Volkswagen revealed a VW camper van taxi with a gasoline hybrid drive system for an exhibition at the Museum of Modern Art in New York (Figure 29.7). In 1989, Audi presented the first of three generations of the Audi duo. It was an Audi 100 Avant quattro equipped with a 12.6 PS (8.8 kW) electric motor, which supplied the input at the rear axle. The electrical energy was provided by a nickel–cadmium battery. The front axle was driven by a 2.3-L five cylinder with 136 PS (100 kW; Figure 29.8). The second generation of the Audi duo followed two years later. The output of the direct current electric motor had risen to 28.6 PS (20.6 kW). The 2.0-L four-cylinder diesel engine was now also connected to the rear axle via a Torsen differential. Between 1991 and 1993, Volkswagen carried out a fleet test with 20 Golf II parallel hybrid vehicles in Zürich (Figure 29.9). The project was supervised on the ground by ETH Zürich. A 1.6-L four-cylinder diesel engine with 44 kW and an electric motor with 7 kW were used. Eight vehicles were fitted with a lead-gel battery, six vehicles had a nickel–cadmium battery, and six more were equipped with a sodium–sulfur battery. The consumption of the vehicles in a predefined driving cycle

(Zürich urban driving cycle, forty-one sets of traffic lights over 10.3 km) was 3.8 l/100 km and the electrical output was 21.7 kW [29-12], a series production Golf diesel consumed 8.6 L of diesel. Poor reliability, service life, and storage capacity were cited as problems for the batteries tested.

View A

Engine

TRANS.

Engine

Figure 29.5 Draft of a BMW 2000 with combustion engine and hydraulic motor [29-8].

Figure 29.6 Toyota Sports 800 [29-9].

Internal Combustion Engine Handbook | 981

6606_Book.indb 981

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Figure 29.7 VW camper prototype taxi with hybrid drive system [29-10].

Figure 29.8 Audi duo [29-11]. Figure 29.10 Audi duo III [29-12].

Figure 29.9 Electro hybrid Golf II, fleet test in Zürich, 1991–1993.

In 1997, Audi presented the third and final generation of the Audi duo (Figure 29.10 and Figure 29.11). It had a 1.9-L TDI engine with 66 kW and an electric motor with 21 kW. This time both drive systems operated the front axle. A lead-gel battery was used to store the electrical energy. Based on the A4 Avant, the vehicle could be leased for 60,000 DM (approximately $35,000). Strictly speaking Audi was therefore the first European carmaker to manufacture a volume-production hybrid vehicle. The extremely high purchase price kept demand in check, with the result that production was discontinued barely a few months later. Still in the same year (1997), Toyota commenced sales of the first Prius in Japan. With only 323 customers, the Prius too was hardly a commercial success at first. However, by the following year sales had jumped to 17,653 units.

Figure 29.11 Audi duo III [29-11].

Toward the end of 1999, Honda also brought a hybrid vehicle onto the market. At first, the Insight was only available to buy in Japan and the United States. It was driven by a three-cylinder gasoline engine and an electric motor on the crankshaft. The first Prius model change occurred at the end of 2000: the Prius I went into volume production [29-9]. Yet it was not until the introduction of the Prius II in 2003 that public interest in hybrid vehicles was awoken. Sales success in the USA and Japan, as well as the increased sensitivity regarding CO2 emissions ultimately led all the famous vehicle manufacturers to commence volume production of hybrid drive systems for passenger cars.

982 | Internal Combustion Engine Handbook

6606_Book.indb 982

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.2 Fundamentals of Hybrid Drive Systems (General Overview)

29.2 Fundamentals of Hybrid Drive Systems (General Overview) Hybrid drive systems consist of a combination of at least two different drive systems. Thus hybrid drive systems have at least two different energy converters (in terms of motors) and two different energy storage devices. A similar definition was adopted, among others, from the UNECE in 2003 [29-13]: “Hybrid vehicle (HV) means a vehicle with at least two different energy converters and two different energy storage systems (on vehicle) for the purpose of vehicle propulsion.” In spite of this, a hybrid drive system is often wrongly reduced to the presence of two different engines. The combination of a gasoline or diesel engine with an electric motor has prevailed for vehicle drive systems. However, the definition does not prescribe this classification. In the past, there were also combinations of diesel/flywheel engine and gas turbine/electric motor with corresponding energy storage devices.

29.2.1 Principle

In theory, a hybrid drive system should combine the advantages of two drive systems. The objective these days is lowering emissions and/or the vehicle’s energy consumption. A comparison of a vehicle’s requirement identifier with the supply identifier of a diesel engine, gasoline engine (see Figure 29.12), and an electric motor reveals that combustion engines show serious shortcomings in the low speed range. By contrast electric motors offer an almost ideal supply identifier, however, the energy storage devices represent a problem in this aspect. In the short and medium term, sufficiently efficient and long-lasting energy storage devices are not within sight for passenger vehicles with electric-only

drive systems, with the result that only niche applications are now available. If we look at the interaction between combustion engines and electric drive systems, the potential of one combination becomes clear. The electric motor supplements the supply identifier weaknesses of the combustion engine, whereas the efficient energy storage device, that is, the fuel tank, compensates for the shortcomings of the batteries. It is also possible to convert some of the energy forms, which are not useful to combustion engines, into electrical energy. Furthermore, it is possible to avoid combustion engine operating conditions where increased emissions occur. These include highly dynamic states and low-load operating points. In extreme cases with a hybrid drive system, the combustion engine is only operated in one speed-load point.

29.2.2 Components

Compared with a diesel or gasoline engine, a few adaptions and additional components are required for a hybrid drive system. In addition to the electric motor and electrical storage device, a hybrid drive system also requires an energy management system, an electrical machine connection to the drivetrain and power electronics. 29.2.2.1 Combustion Engine From a technical perspective, only a few modifications are absolutely necessary to the combustion engine. One consideration is that because of the modified operating strategy, for example, frequent engine start/stop, the required catalytic converter temperature must continue to be maintained. Modifications to the engine bearings are also necessary as the additional electrical machines alter the drivetrain’s vibration behavior. If the combustion engine is rather operated at high loads because of the operating strategy selected, then

Output

Torque

Vehicle Electric motor Diesel engine Gasoline engine

Vehicle Electric motor Diesel engine Gasoline engine

Speed

Speed

Figure 29.12 Supply identifier of combustion engines and electric motors. See color section page 1104.

Internal Combustion Engine Handbook | 983

6606_Book.indb 983

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

the corresponding components also need to be reconstrued. Despite this, an adjustment to the new requirements, such as lower dynamics, is also advisable for financial reasons. 29.2.2.2 Electric Motor In principle, all common types of electrical machines are possible. However, direct current machines are less suitable because of their comparatively poor efficiency (just 80%–85%), especially as they are more maintenance intensive than other brushless types of motor because of the brushes required. They are also sensitive to vibrations and shocks. Asynchronous machines (ASMs) are sturdy electric motors, which do not have any wear parts with the correct mechanical and thermal design. These are simple and cost-efficient motors [29-14]. Synchronous machines (SMs), when compared with ASMs, have a greater efficiency and a higher power-to-weight ratio, although system costs are higher because of the magnets, which are required. 29.2.2.3 Generator The generator function can be implemented in various ways. A classic alternator can be used for low outputs (approximately 5 kW) or simple concepts. Because of the presence of electric motors, there is also the option, however, to use these as generators also. If this is not possible for technical reasons, then an electric motor is employed specially to operate the generator when the power requirement is greater. 29.2.2.4 Electrical Energy Storage Device The energy storage device or battery is considered as the electrical machine’s fuel tank. Energy density, power-toweight ratio, service life, and costs are the decisive evaluation criteria. Lead batteries, nickel–metal hydride (NiMH) batteries, and even occasionally lithium-ion batteries are used. While costly lead batteries were frequently drawn on in earlier test vehicles, NiMH batteries are mainly used nowadays. Vehicle manufacturers will switch to the more powerful lithium-ion batteries as soon as costs and functional safety allow. Ultra Caps are not in direct competition with NiMH and Li-ion batteries. Instead their strength lies in the short-term uptake of very high outputs. 29.2.2.5 Transmission If the outputs of the combustion engine and electric motor are supposed to be mechanically consolidated with a hybrid concept, then modifications are required vis-à-vis standard vehicle transmissions. Load type and thermal loading differ greatly from traditional vehicle transmissions. 29.2.2.6 Energy Management System The energy management system has a pivotal significance for hybrid drive systems; however, now there is the driver option of operating with two drive units. Furthermore, it must be ensured that the electrical storage devices contain a minimum of energy to straddle vehicle stationary times or cold weather for instance. A sophisticated energy management system can lead to smaller components, less weight, and ultimately lower costs.

29.2.2.7 Power Electronics The power electronics converts the accumulated electrical energy in terms of voltage, frequency, or polarity into what is required in any particular case. The costs are directly associated with the capacity and current rating required. Because of its size and cooling requirements, the power electronics has a profound influence on the package of a hybrid vehicle.

29.2.3 Functions

Depending on the level of hybridization, additional functions such as electric driving or regenerative braking can be realized with a hybrid drive system when compared with a vehicle drive system with a combustion engine. A distinction can be made between functions, which are only possible with a hybrid drive and those that would theoretically be possible with a gasoline or diesel drive system as well, but which up till now have only been implemented in some vehicles for financial reasons. 29.2.3.1 Start/Stop (Stop/Start) The start/stop or stop/start designation relates solely to the combustion engine in the hybrid drive system. It is shut off when a vehicle is stationary and subsequently consumes no fuel and produces no emissions. The most common example of this is stopping at traffic lights or the stop-and-go operation in traffic. The engine is started and/or setting off is facilitated by the hybrid drive system’s electric motor. Because the electric motor is powerful in relation to the starter, alternative start concepts can be put into practice for high-performance hybrid drive systems, such as a pulse start of high engine-speed start. The start/stop function has since been implemented in pure gasoline and diesel drive systems. They require a more powerful dimensioned starter and battery management for this purpose. 29.2.3.2 Electric Driving If the electric motor and the energy storage device are sufficiently powerfully dimensioned, then electric-only driving is possible. The combustion engine is switched off in this driving condition. A distance of a few hundred meters to a few kilometers (Prius II: approximately 6 km) can be bridged depending on the storage capacity with this concept. This function will grow in significance as more powerful and affordable storage devices are introduced. 29.2.3.3 Shifting the Operating Point Because an additional drive source is available in a hybrid drive system besides the combustion engine, the load requirement of the combustion engine is decoupled from the driver option. The electrical machine can operate in a supporting role according to requirements. Thus, it is able to lower the operating point of the combustion engine or increase the load in the generator mode. The objective behind shifting the operating point is to reduce consumption and/or exhaust emissions. If the dynamic requirements of the combustion engine are reduced, this can also be described as the phlegmatization of the combustion engine. In extreme cases, the combustion engine is only operated at one operating point.

984 | Internal Combustion Engine Handbook

6606_Book.indb 984

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.2 Fundamentals of Hybrid Drive Systems (General Overview)

29.2.3.4 Boost Mode Boost mode is the name given to the short-term support provided to the combustion engine by the electric motor. The duration of the boost mode is generally restricted by the size and current charge state of the electrical energy storage device and may be in the range of several seconds.

29.2.3.6 Regenerative Braking Regenerative braking is understood as the conversion of a vehicle’s kinetic energy into usable energy forms while braking. This kinetic energy is usually converted into electrical energy. The power required to recover the braking energy must be managed by the electrical machine as well as the power electronics and storage devices. Because a huge amount of power is sometimes required during braking, only part of the braking energy is recovered and a mechanical brake continues to be used with greater amounts of energy for technical and financial reasons. Also for legal reasons it is not yet permitted to sacrifice the mechanical brake entirely.

Torque, combustion engine

29.2.3.5 Compensation of Rotational Oscillations If the combustion engine and electric motor are operated simultaneously, the electric motor can be used to compensate rotational oscillations. The fluctuations caused by individual combustion processes can be reduced as a first approximation with a sinus-shaped actuation of the electric motor (see Figure 29.13).

29.2.3.7 Electrical Auxiliary Units In combustion engines, the auxiliary units are usually driven by a V-belt or a chain. In principle, electrical auxiliary units can also basically employed in conventional drive systems; however, their use is benefited by the high-voltage network present in hybrid drive systems. A few electrical auxiliary units are even mandatory because of the start/stop function of the combustion engine. An advantage of electrical auxiliary units is that their output can be adapted to requirements, and therefore it is no longer necessary for instance to operate the coolant pump in the first few seconds after starting the engine. Likewise the flow properties of an oil pump over the engine speed do not equate to demand. Figure 29.14 plots a comparison of the delivery pressure and power requirement of a conventional oil pump against a variable oil pump.

Combustion engine

Drive axle final torque

Rotation in °CA Electric motor

Combustion engine (Transmission to the drive axle)

29.2.3.8 Automatic Parking Besides the sensors that detect the parking space, a controllable vehicle steering mechanism is also needed for automatic parking. If the vehicle is designed to have the ability to maneuver independently, then the drive system and direction of travel must also be automatically controllable. In principle, this is possible with a diesel or gasoline engine drive system, yet hybrid vehicles have the advantage of possessing an easily controllable drive system with their electrical machines and often already have an electrical steering mechanism.

Resulting

Rotation in °CA

Pressure [bar]

Power consumption W]

Figure 29.13 Compensation of rotational oscillations.

Speed [rpm]

conventional oil pump

variable oil pump

Speed [rpm]

Figure 29.14 Comparison of a conventional oil pump and variable oil pump [29-15].

Internal Combustion Engine Handbook | 985

6606_Book.indb 985

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

29.3 Classification of Hybrid Drive Systems Hybrid drive systems can be classified into various types and performance categories depending on their design and the electric power installed respectively.

29.3.1 Types

Hybrid drive systems are broken down into three types according to the combination of the combustion engine and electrical machine. These types are known as series, parallel, and mixed hybrids (see Figure 29.15). The basic principle of these three types is displayed in Figure 29.16. With the series hybrid drive system, the combustion engine and the electric motor are located in a row. The combustion engine drives a generator, and the accumulated electrical energy is then either stored in a battery or diverted

directly to the electric motor. The electric motor drives the drive axle via a mechanical connection. One advantage of the serial hybrid drive system is that the combustion engine can constantly be operated at the optimal operating point, with the result that both consumption and emissions are very small. However, disadvantages include the required mechanical–electrical–mechanical energy conversion, the high installation power, weight, and costs. Both drive machines and the generator must be able to deliver the required output; the drive rating is subsequently available three times. Furthermore, the power electronics must be designed for the maximum vehicle performance. With the parallel hybrid, the combustion engine and the electric motor are arranged parallel in the powertrain. The combustion engine is therefore mechanically connected to the axle. The power from both the combustion engine and the electric motor can affect the drive axle either singly or

Hybrid drive systems

Series hybrid drive systems

Parallel hybrid drive systems

Mixed hybrid drive systems

Combined hybrid drive systems

Power-split hybrid drive systems

Torque-accumulating hybrid drive systems

Single-shaft hybrid drive systems

Speed-accumulating hybrid drive systems

Tractive-force-accumulating hybrid drive systems

Dual-/Multiple-shaft hybrid drive systems

Figure 29.15 Classification of hybrid drive systems [29-17].

Key

Combustion engine Clutch Transmission Generator Electric motor Electrical energy storage device Fuel tank

Figure 29.16 Principle of a series, parallel, and power-split hybrid drive system (from left to right). See color section page 1105.

986 | Internal Combustion Engine Handbook

6606_Book.indb 986

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.3 Classification of Hybrid Drive Systems

simultaneously. Depending on how the interaction between combustion engine and electric motor is realized, a distinction is made between speed, torque, and force addition. The combustion engine and the electric motor do not have to be connected to the same shaft, which means that single-shaft and double-shaft solutions are possible. In addition to the series hybrid and parallel hybrid, mixed forms of these two variations also exist. A mixed form is the power-split hybrid drive system, the use of which separates the power from the combustion engine into an electrical and a mechanical branch. It provides a broad spectrum of operating strategies for the individual motors/ engines. A planetary gear set is often used as a connecting transmission. Another mixed form of series and parallel hybrid can be achieved by arranging a combustion engine and two electrical machines in a row, and a switchable coupling is located between the electrical machines. This concept allows switching between a series hybrid drive system and parallel hybrid drive system.

then micro hybrids operate in the range of H = 0.05, the mild hybrids at H = 0.1, and from H = 0.25 can be described as full hybrids [29-18]. The lines are shown in Figure 29.17 for the values H = 0.25, 0.10, and 0.05.

Figure 29.17 Level of hybridization.

29.3.2 Power Classification

Regardless of their principle, hybrid drive systems are classified as micro, mild, and full hybrid according to their electrical power and the associated functions. They occasionally have even more grades of classification, such as basic hybrid or analogous terms, such as strong hybrid for the full hybrid application. One option for qualitatively expressing this distinction is to reference the power of the electric motor (Pe) to the total power of electric motor and combustion engine (PVKM) [29-18]:

H = Pe/(Pe + P VKM)

As this approach neglects the electrical storage devices, it must be assumed that their capacity is the same as the potential of the overall drive system. If the individual designations, micro, mild, and full hybrid are quantitatively arranged,

The boundaries of the individual power variants are fluid, which means that in some cases a clear allocation is not given. Typical characteristics of the individual hybrid drive systems are shown in Figure 29.18. The functions that can be implemented with them are described in Figure 29.19. Micro hybrids only have a start/stop function and can recover energy in thrust phases to a limited extent. As they are not making a noteworthy contribution to the drive torque, it may be controversial to discuss whether they belong to the hybrid drive systems at all. The output of the electrical machine is somewhere in the region of 5 kW. The most common configuration is a belt starter/generator. In spite of the low output, fuel savings can be as much as 8% compared to a gasoline or diesel drive system.

max. electrical output Voltage, traction network Battery type

Lead-acid type (AGM)

Battery output max. fuel saving Cost increment (market price)

Figure 29.18 Characteristic of various levels of hybridization.

Internal Combustion Engine Handbook | 987

6606_Book.indb 987

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Stop/start High speed start Energy recovery

minimal

Shifting the operating point

high

maximum

reduced

Coasting Boost

weak Figure 29.19 Potential functions of various hybrid types.

Electric driving

term. The following properties are important especially for use in the vehicle: •• low weight •• little space requirement •• low costs •• long service life •• high efficiency.

29.4.1 Electrical Machine

Depending on the effective direction, electrical machines either convert electrical energy into mechanical energy or the reverse. Conversion into mechanical energy is referred to as motor operation, and those into electrical energy are referred to as generating operations. Because the direction of rotation can be either to the right or to the left, there are consequently four quadrants in the speed torque diagram (Figure 29.20).

Generator

29.4 Electrical Hybrid Drive Systems Electrical motors have become established alongside the combustion engine as the second motor in the hybrid drive system. With their supply identifier they compensate for the shortcomings of the combustion engine in the lower speed/ load area, they convert energy with a high level of efficiency (sometimes >90) and they are able to convert mechanical energy back into electrical energy. They can also be significantly overloaded, that is, in the region of a few seconds, in the short

Torque

The mild hybrid, because of its greater electrical output, is in a position to support the combustion engine to a certain extent in addition to the stop/start function. In addition, braking energy can be recuperated to a noticeable extent. Moreover, a slight shift of the combustion engine’s operating point is possible. The output of the electrical machine is somewhere in the region of 15 kW. For the configuration, for example, a crankshaft starter generator (e.g., Honda Integrated Motor Assist (IMA), GM with Continental ISAD) comes into consideration. The fuel savings associated with such a system are somewhere in the region of 10%–20%. Full hybrid allows for the entire spectrum of hybrid drive functions, particularly electric driving. Owing to the available electrical power, the working point of the combustion engine can be almost entirely decoupled at the driver’s request. This measure delivers advantages both in terms of consumption and emissions. More so than with gasoline or diesel drive systems, the fuel consumption of a hybrid drive system is dependent on the driving profile. The mentioned consumption advantages should therefore be seen as the maximum possible potential. The smoother and faster a vehicle moves in road traffic, the lower the consumption advantage compared with the classic drive systems.

Engine Speed Generator

Engine

Field-weakening range

Base speed range

Field-weakening range

Figure 29.20 Functional areas of an electrical machine.

988 | Internal Combustion Engine Handbook

6606_Book.indb 988

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.4 Electrical Hybrid Drive Systems

Electrical machines

Direct current machines

Shunt

Separately excited

Alternating current machines (1-Phasig)

Series circuit

Three-phase machines

Synchronous

Compound

Asynchronous

Short circuit rotor

Slip ring rotor

Figure 29.21 Classification of the electrical machines.

through the machine and the iron losses represent the losses because of constant magnetization of individual components. Bearing friction, rubbing of commutators, and, where appropriate, also fan power are counted as mechanical losses. The individual loss types occur with varying degrees of strength depending on the type and construction of the electrical machines.  irect current motor (DC motor) D Individual electrical machine displays different speed–torque behaviors under load. Three characteristic processes are shown in Figure 29.23. These are called the shunt, series circuit, and synchronous behavior.

Synchronous characteristic

Shunt characteristic

Speed

The speed range in which an electrical machine reaches its maximum torque is known as the base speed range. It is limited, among others things, by the maximum permissible current in the electrical machine/actuation unit. The power must be limited above this speed for thermal reasons, with the result that the torque drops over the speed. This area is known as the field-weakening range. The electrical machine function is based on the Lorentz force, that is, the force effect of a magnetic field on a moved charge and on Faraday’s law. According to the model type, a distinction is made among direct current machines, alternating current machines, and three-phase machines with respective different subclasses (Figure 29.21). Three-phase machines are state of the art in today’s hybrid drive systems. Despite high levels of efficiency, losses from electrical machines should not be disregarded. The power loss is the combination of copper, iron, and friction loss (Figure 29.22); copper losses refer to the heating of the conductor as a result of the current flow

Input power Series circuit characteristic

Copper losses Iron losses Friction losses

Figure 29-22 Electrical machine losses.

Torque

Friction losses

Figure 29.23 Characteristic curve of various electrical machines

Direct current machines are electrical machines, which are operated with direct current. They consist of a stator (column), to which all fixed components are arranged, and a pivoting rotor (anchor). Either an electromagnet or a permanent magnet is installed in the stator. The rotor consists of conductor windings, which are supplied with current via commutators.

Internal Combustion Engine Handbook | 989

6606_Book.indb 989

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Control voltage

Torque

The speed of the motor can be modified with a series resistor in the armature circuit, by changing the magnetic field, or adjusting the armature voltage (Figure 29.24). If the excitation field consists of one magnet, then the speed of the electric motor can only be adjusted with the armature voltage. Direct current motors are divided into shunt motors, series circuit motors, and compound motors according to the circuitry of the exciter and armature winding. Another distinction is made between direct current motors that are separately excited.

Series resistor

Torque

Speed

Armature voltage

Figure 29.26 Characteristic curve of a series circuit motor.

S  hunt motor With the shunt motor, the armature winding and excitation winding are shunted (Figure 29.27).

Torque Figure 29.24 Direct current control.

 eries circuit motors S With series circuit motors, the exciter and armature windings are wired in series (Figure 29.25); consequently, both are loaded with the same current.

Figure 29.27 Shunt wiring.

Separately excited direct current motor With separately excited direct current motors, the excitation winding and armature winding are supplied by two separate voltage sources (Figure 29.28).

Figure 29.25 Series circuit wiring.

Series circuit motors may not be operated without load; otherwise, the speed rises too much and may consequently destroy the motor in some circumstances. One advantage of this circuit, however, is the high starting torque of the motor (Figure 29.26). The speed can be influenced either directly via the voltage or with the aid of an adjustable series resistor.

Figure 29.28 Separately excited direct current motor.

Brushless direct current motor With brushless direct current motors, the function of the mechanical commutator is replaced by an inverter. The rotor carries a permanent magnet, and the stator carries the inductors actuated by the converter. The motors are the same as a synchronous motor in terms of construction.

990 | Internal Combustion Engine Handbook

6606_Book.indb 990

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.4 Electrical Hybrid Drive Systems

Consequently, this motor is very low-maintenance and low-wear, but is, however, more expensive because of the necessary electronics. Electronic disruptions, which can arise with the commutator are gone. However, other interference variables do occur in some circumstances with the power electronics. A  synchronous machines ASMs belong to the three-phase machine family. Three inductors are arranged in their stator such that a magnetic rotating field is formed with appropriate actuation. The rotor consists of individual conductors, which run parallel to the rotational axis and are either shorted to each other at their ends (short circuit rotors) or deflect the emanating current via loop rings (Figure 29.29). When using a short-circuit rotor, wear only occurs on its bearings, meaning that the rotor is therefore long lasting.

nField =



fControl p

The rotor cannot reach the speed of the field in principle, but differs from it according to the load. The deviation of the rotor speed (nRotor) from the speed of the field (nField) is designated as slip (s) and is calculated according to s=

nField − nRotor ⋅100 in % nField



If the rotor is stationary and the rotating field is rotating, then s = 100%; when the speed is the same, then theoretically s = 0%. In the idle mode, that is, during load free operation, the speed of the rotor is almost that of the rotating field (nRotor):

n Rotating field

n Rotor

In the engine operation, the magnetic rotating field induces a voltage in the rotor wire, which results in a current flow. The torque is formed from the interaction of the current flow with the magnetic field, which makes the rotor rotate. The speed of the rotating field (nField) is the result of the control frequency (fControl) and the number of the pole pairs in the stator (p):

Motor operation:

nRotor
nRotor, Idle mode ≈ nField Figure 29.30 shows the characteristic curve of an ASM with the characteristic points: torque when stationary (MA), tipping torque (M K) and speed (nK), as well as synchronous speed (nsyn). The highest torque is designated as tipping torque M K. It decreases with constant stator voltage and stator current quadratically with the control frequency: M~

Figure 29.29 ASM.

1 2 fControl



where

Torque

UStator = constant and IStator = constant.

Speed

Figure 29.30 Characteristic curve of an ASM.

Internal Combustion Engine Handbook | 991

6606_Book.indb 991

1/19/16 8:59 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Torque

In principle, the speed of an ASM can be varied by changing the number of pole pairs, by varying the control frequency, varying the voltage, and with slip ring rotors, also by positioning resistors in the rotor circuit. Controlling the speed with the voltage, however, is only possible to a limited extent, since the motor torque changes quadratically to the voltage, which is why control takes place in the vehicle via the frequency. Because the magnetic flow reduces with the frequency, however, the voltage has to be increased at the same time. Figure 29.31 shows the efficiency map (lines of equal efficiency) of an ASM.

Speed

Figure 29.31 Efficiency map of an ASM.

In addition to its simple and inexpensive construction, the ASM, when compared against a direct current machine, has the advantage of a lower torque ripple and smaller noise emission. S  ynchronous machine SMs (Figure 29.32) like ASMs also belong to the three-phase machine family. In terms of design, the stator is identical to that in the ASM, but only the rotor differs: it consists of a pole wheel.

The individual pole pairs can either be generated by a permanent magnet or by an electromagnet. The permanently excited SM (PSM) has the advantage of no loss in the rotor and is also low wear, because of its high efficiency and high power density. However, the necessary magnetic materials are very expensive. In operation, the rotor rotates with the same speed as the stator rotating field, that is, synchronous to it. This synchronous number (nsynchronous) is calculated with nsyn =



fControl p

where the control frequency and p is the number of the pole pairs. Depending on the load, the rotor, however, lags behind the stator rotating field around the rotor displacement angle α pole (Figure 29.33). If the maximum torque is exceeded, the so-called tipping torque MTorque, then the effective system collapses and the machine continues uncontrollably at a high current consumption. SMs lean toward so-called rotor displacement angle oscillations during operation, that is, the rotor displacement angle fluctuates periodically. These oscillations can be attenuated if necessary with damping bars (damper cage) distributed along the rotor circumference. SMs are unable start up independently, which also becomes clear from its characteristic curve in Figure 29.34. Three options are available for starting up; these are the asynchronous startup, startup with an auxiliary motor, or startup with a frequency converter. The latter of these options applies to the automotive application. Compared to the ASM, higher peak efficiencies and a higher power density [29-22] can be achieved with the PSM, because there is no rotor loss. The characteristic efficiency maps of both machine types are shown in Figure 29.35. The SM reaches its peak efficiency in the medium speed range, close to or at full load. Compared to the ASM, there are advantages in the lower to medium speed range during partial load and full load [29-23], as well as disadvantages at high speeds.

n Rotating field nRotating field n Rotor

n Rotor

Figure 29.32 SM (left: nonsalient pole machine, right: revolvingfield machine).

992 | Internal Combustion Engine Handbook

6606_Book.indb 992

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.4 Electrical Hybrid Drive Systems

nRotating field

nRotating field

nRotating field n Rotor

n Rotor

n Rotor

Idling

n Rotor = n Rotating field

Motor operation M>0 n Rotor = n Rotating field

Motor operation M=Mtilt

n Rotor = n Rotating field

Figure 29.33 Rotor displacement angle during idle operation, engine operation, and maximum torque.

Torque

Switched reluctance machine Reluctance machines (Figure 29.36) use the principle of the lowest magnetic resistance. The stator is made up of various inductors, which are purposefully switched on and off via power electronics and thus generate a switchable magnetic field. The rotor consists of either a soft-magnetized material or of material alternately magnetized around its circumference. If a magnetic field is generated via the inductors in the stator, then the rotor is arranged such that the magnetic resistance is minimal. The switched reluctance machine is a type of reluctance machine. The power-to-weight ratio is less than that of an ASM [29-25]; noise also arises as a matter of principle. One interesting approach is to combine the principle of the reluctance machine with that of an SM to utilize the advantages of both machines [29-26].

Speed Figure 29.34 Characteristic curve of an SM.

Torque

Torque

Particularly because of their high levels of efficiency and small size, PSMs are the preferred concept for use in hybrid vehicles [29-24]. However, it should be observed that permanently excited synchronous motors can build up very high torques if there is a short in the stator winding. This characteristic must be taken into account with the vehicle’s safety concept.

Speed

Speed

Figure 29.35 Characteristic efficiency maps for an SM (left) and an ASM (right).

Internal Combustion Engine Handbook | 993

6606_Book.indb 993

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

nRotating field

Rotor

nRotor

Coil Stator

Figure 29.37 Transverse flux machine.

Figure 29.36 Reluctance machine.

 ransverse flux machine T The transverse flux machine is a special version, with which the magnetic flow is led vertical to the rotation plane and therefore transversally as in Figure 29.37. This type of machine displays a high torque at low speeds; the space requirements of this type of machine are also low. Downsides include a powerful torque ripple and the emergence of noises and vibrations [29-27].  heel hub motors W Wheel hub motors are motors that are incorporated into the vehicle wheel rims and used to drive these (examples are shown in Figure 29.38). One major advantage is the freedom, which is gained in the vehicle package, as the space taken up by the electrical machine, not to mention the transmission,

shafts, and differentials is freed up. Small cars, in particular, benefit from the extra space that can be utilized for batteries, fuel cells, or fuel tanks. Furthermore, the power losses for the above-mentioned components are also eradicated, which benefits overall drivetrain efficiency. Driving dynamic advantages can also be achieved if wheel hub motors are used on all four of the vehicle’s wheels. On the other hand, there are the disadvantages of the unsprung masses of the individual electric motors, the higher costs for many motors and their power electronics, as well as the unfavorable environmental influences at the point of installation such as grime, splash water, and heat. With a wheel hub motor, the speed range in particular must be adjusted. It should lie between 0 and approximately 1500 rpm. Because the most commonly occurring wheel speeds do not lie in a high efficiency range, compromises must be made with the overall level of efficiency.

Figure 29.38 Wheel hub motor (left: Michelin concept [29-28], right: Magnet motor concept).

994 | Internal Combustion Engine Handbook

6606_Book.indb 994

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.4 Electrical Hybrid Drive Systems

29.4.2 Power Range

Electrical machines can be manufactured for a wide power spectrum, from a few milliwatt right up to a few megawatt. Powers of between 5 and 100 kW are interesting for hybrid drive systems, and preferably a high maximum torque or a high maximum speed is required depending on the hybrid concept and installation selected (Figure 29.39). If the electrical machine is designed to be able to start the combustion engine, then it also requires a high start torque. If the electrical machine is used as both a motor and a generator, then, in addition to the drive output, the maximum amount of braking power that can be recuperated is also crucial to the design.

29.4.4 Power Electronics

The power electronics implements the specifications from the control electronics. It switches the electrical energy and converts it, according to requirements, in terms of polarity, voltage, frequency, and phasing. According to DIN 41750 T1/2.85 conversion of alternating current (AC) to direct current (DC) is referred to as rectification, from AC to AC as AC conversion, from DC to AC as inversion, and from DC to DC as DC conversion (Figure 29.41).

High torque Figure 29.41 Transformer tasks.

In a hybrid drive system, the electrical energy flows both from the battery, which delivers a direct current, to the electrical machine, which requires an alternating current, as well as back, meaning that all types of conversion occur. Current converters take care of the individual conversions. High, maximum speeds

29.4.5 Current Converter

High power-to-weight ratio Figure 29.39 Different designs for an electrical machine [29-29].

29.4.3 Control

The electrical machine in the hybrid drive system is controlled by the control and power electronics (Figure 29.40). The control electronics receives orders via a data bus, for instance, a CAN; in addition, it also receives measured values about the electrical machine from various sensors. The actuation of the power electronics is calculated using this information.

A current converter is an electronic circuit constructed from various semiconductor components and performs the respective type of conversion. Diodes, transistors, thyristors, triacs, insulated-gate bipolar transistors (IGBT), and metal–oxide– semiconductor field-effect transistors (MOSFET) are used particularly in this instance. The diodes only conduct current in one direction; if they are set in the opposite direction, they block the current flow. Transistors, IGBTs, MOSFETs, and thyristors can be switched using an electrical signal, so that, depending on the signal, they block/conduct the current in one direction, and block the current, like a diode, in the

Housing Power electronics

Current measurement

Temperature sensor Control system Cooling system Power supply CAN

Speed/ Position sensor

Actuation Locking mechanism

Microcontroller

Figure 29.40 Electrical machine control.

Internal Combustion Engine Handbook | 995

6606_Book.indb 995

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

other direction. A triac basically behaves like two opposing thyristors, that is, it can be used intentionally to switch the current in both directions. The individual electronic switches primarily differ in their ability to switch high currents and in the resulting switching losses. Analogous to their function, current converters are referred to as rectifiers, DC converters, inverters, and AC converters [29-30]. Hence, an inverter converts AC voltage into DC voltage. A DC converter converts DC voltage into DC voltage with a different polarity or voltage. They are also referred to as DC/DC converters and can be used, for example, as voltage stabilizers and charging converters. An inverter converts DC voltage into AC voltage. The AC converter or frequency converter converts AC current into another AC current. It consists of a rectifier, intermediate DC link, and an inverter. If it combines the desired AC voltage from various DC voltage pulses (Figure 29.42), then it is also referred to as a pulse converter (Figure 29.42). Losses occur in the converter when switching the high currents, with the result that it is necessary to cool the components because of the high switching power. For cost reasons, engineers fall back on the combustion engine’s coolant circuit, where the coolant can, however, reach temperatures of up to 115°C. The highest possible voltages and lowest possible switching frequencies are needed to keep losses low. While on the other hand, there is the desire to keep battery voltages as low as possible and switching frequencies high for the electrical machines. Moreover, the selected semiconductor also caps the voltage.

29.5 Energy Storage Devices Energy is required to drive motors and this energy is carried along in corresponding storage devices in the vehicle. Besides

Housing

Filter

the well-known fuel tank, principally used for chemical fuels such as gasoline, diesel, CNG, LPG, and H2, flywheel generators, pressure accumulators, and (rechargeable) batteries are also conceivable alternatives. Hybrid drive systems have at least two different energy storage devices. With today’s hybrid vehicles, this is usually a fuel tank and an electric energy storage device, and therefore we will focus solely on the electric storage device in this chapter. Storage devices for electrical energy are characterized by their ability to store energy in the form of chemical energy or electrostatically and convert this into electrical energy without an intermediate stage via mechanical or heat energy. Two of the most important properties of a storage device are energy density and power density. As you can see in Figure 29.43, there are significant differences here between the various types of storage. While a similarly high specific energy density can be achieved with rechargeable batteries, the strength of the electrostatic storage device is represented by the specific power density. However, with approximately 150 Wh/kg, these systems do not come close to the energy density of gasoline or diesel fuels. In the case of a gasoline engine, its lower calorific value at 40.1–41.8 MJ/kg equates to 11.14–11.61 kWh/kg [29-32]. Even when the weight of the fuel tank has to be factored in, the differences are blatantly obvious. Likewise if we take the required volume of the storage device into account, traditional fuels are vastly superior to batteries and capacitors (see Figure 29.44). The difference between the individual storage types is, however, curtailed somewhat when the subsequent energy converter is also taken into account. The energy from liquid fuels such as gasoline and diesel is converted into mechanical energy with relatively low efficiency by the combustion engine, whereas the energy stored in batteries and converted by the electric motors has an efficiency of over 95% in part. Various elements from the periodic table are available for the battery concept (Figure 29.45).

Inverter

Voltage

U Pulse U Resulting Time

Figure 29.42 Principle of the pulse converter.

996 | Internal Combustion Engine Handbook

6606_Book.indb 996

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Specific energy density [W/kg]

29.5 Energy Storage Devices

Electrolyte capacitors

Figure 29.43 Power density and energy density of various storage devices[29-31].

Weight-related energy density [kWh/kg]

Specific power-to-weigh ratio [Wh/kg]

Gasoline/ Diesel

In future: Liquid hydrogen At present: Gaseous hydrogen At present: Liquid hydrogen

however, in practical implementation, which, for example, can be an obstacle to long service lives and low costs. Therefore, the Li/F battery is still in the foundational research phase. In addition to the energy and power density, other properties are important depending on the application. Therefore battery systems being used in vehicles must meet the following requirements: •• high specific energy •• high specific power •• good charging and discharging efficiency

NiMH

Metal hydride

Volume-related energy density [kWh/l]

Figure 29.44 Energy content of different storage devices in relation to their weight and volume [29-33].

The highest theoretical energy density can be achieved with the pairing of lithium (Li) and fluorine (F). A battery with this material pairing could achieve a theoretical energy density of 6100 Wh/kg [29-34], not counting the weight of the components. Additional properties of the elements play a role,

Reduction

•• high number of charging and discharging cycles •• high safety •• temperature resistant in the automobile range of application •• low costs. The required characteristic of these properties is in each case dependent on the practical use of the battery system. Technically mature and inexpensive lead batteries have become prominent for the application in gasoline and diesel drive systems. Their energy and power density is, however, insufficient for hybrid drive systems, with the result that NiMH and Li-ion rechargeable batteries will be used in future (Figure 29.46).

Oxidation

Figure 29.45 Potential material pairings for batteries with high energy densities.

Internal Combustion Engine Handbook | 997

6606_Book.indb 997

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Storage type

Power-toEnergy weight ratio density [W/kg]

Operating temperature

[Wh/kg]

[°C]

Lead-acid (AGM)

-20 to +70

Nickel-metal hydride - Energy-optimized - Performance-optimized

-10 to +45 -10 to +45

Lithium-ions - Energy-optimized - Performance-optimized

-30 to +60 -30 to +60

Ultra Caps (Performance-optimized)

-40 to +65

Lead-acid batteries essentially consist of two lead electrodes with sulfuric acid as an electrolyte. If the battery is charging, the following chemical reaction takes place:

Pb + PbO2 + 2 H2SO4 ⇒ 2 PbSO4 + 2 H2O

In its charged state, the electrolyte contains approximately 70% acid and 30% water. If the battery is discharging, then the reaction is reversed:

2 PbSO4 + 2 H2O ⇒ Pb + PbO2 + 2 H2SO4

The following partial reactions occur at the individual electrodes: Positive electrodes: PbO2 + HSO – 4+ 3 H+ + 2 e – ⇒ PbSO4 + 2 H2O Negative electrodes:

Figure 29.46 Energy and power densities.



Rechargeable batteries requirements are specially determined for hybrid drive systems by the United States Council for Automotive Research (Figure 29.47). A few of the systems are introduced briefly hereinafter.

Parameter (Units) of fully burdened system

Minimum Goals for Long Term Commercialization

Long Term Goal

Power Density (W/L)

460

600

Specific Power—Discharge, 80% 300 DOD/30 sec (W/kg)

400

Specific Power—Regen, 20% DOD/10 sec (W/kg)

150

200

Energy Density—C/3 Discharge 230 Rate (Wh/L)

300

Specific Energy—C/3 Discharge Rate (Wh/kg)

150

200

Specific Power/Specific Energy 2:1 Ratio

2:1

Total Pack Size (kWh)

40

40

Life (Years)

10

10

Cycle Life—80% DOD (Cycles)

1,000

1,000

Power and Capacity Degradation (% of rated spec)

20

20

Selling Price—25,000 units @ 40 kWh ($/kWh)

<150

100

Operating Environment (°C)

–40 to +50 20% Performance Loss (10% Desired)

–40 to +85

Normal Recharge Time

6 hours (4 hours desired)

3 to 6 hours

High Rate Charge

20–70% SOC in <30 minutes @ 150 W/ kg (<20 min @ 270 W/kg Desired)

40–80% SOC in 15 minutes

75

75

Continuous discharge in 1 hour—No Failure (% of rated energy capacity)

29.5.1 Lead-Acid Battery

Figure 29.47 Requirements of a battery [29-35].

Pb + HSO – 4⇒ PbSO4 + H+ + 2 e –

Various models have come to prominence. A general distinction can be made between open and closed lead-acid batteries. The electrolyte is present in liquid form in the open battery. Besides the main reaction described above, water electrolysis also takes place at the electrodes while charging. This means that water needs to be regularly refilled with this type of construction. The electrolyte is present in either a solid or gelatinous form with this closed construction. The gel is used to create sulfuric acid with the addition of silica; with solid electrolytes, the liquid is bound in a glass mat. Both closed model types have a better, that is, a lower self-discharge rate than the open system. Like other battery systems, lead-acid batteries also display a considerable dependence between the discharge current and usable energy (Figure 29.48). The advantages of the lead-acid battery are its sturdiness and low price. These are two of the reasons why it has established itself in vehicles with combustion engines. Disadvantages include the low energy density, because of the specific weight of lead. Because they are relatively heavy, lead-acid batteries are therefore unsuited to any hybrid drive systems, which go beyond a start/stop function.

29.5.2 Nickel–Metal Hydride Battery

With the nickel metal hydride battery, the negative electrodes are composed of hydrogen, which is stored in a metal and is therefore present in metal form. The positive electrode is nickel(II) hydroxide (Ni(OH)2) and the electrolyte used is caustic potash (KOH). The following chemical reaction takes place in the individual cells: x Ni(OH)2 + M ⇒x NiOOH + MH x, where M stands for a hydrogen-storing metal. The discharging equation is as follows: x NiOOH + MH x ⇒ x Ni(OH)2 + M. The cell voltage of nickel-metal hydride (Ni–MH) batteries is 1.25 V and the energy density of optimized Ni–MH batteries can reach as high as 90 Wh/kg. Power-optimized values of

998 | Internal Combustion Engine Handbook

6606_Book.indb 998

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Voltage in V

29.5 Energy Storage Devices

Figure 29.48 Discharge curve of a lead battery (gel) at various discharge currents [29-36].

extracted charge in Ah

up to 1300 W/kg are possible. As a result, they boast better energy and power densities than lead-acid batteries. However, their self-discharge rate is very high at approximately 40%/ month in the first 15 days [29-36]. Ni–MH cells also display a dependency on the usable energy amount and cell voltage from discharge current (Figure 29.49). Because of their properties, Ni–MH batteries are suitable for use in hybrid drive systems. It is merely for full and plug-in hybrid systems that their energy density is too low.

29.5.3 Sodium–Nickel-Chloride Battery



2 Na + NiCl ⇒ 2 NaCl + Ni2 Discharging occurs in the reverse sequence:



2 NaCl + Ni2 ⇒ 2 Na + NiCl.

The nominal voltage is U = 2.6 V. Because a considerable amount of energy is required to maintain the heat, even when the battery system is well insulated, its use in hybrid vehicles is contentious. The use of sodium–nickel-chloride batteries in fleet vehicles, such as local public transport buses, is perfectly feasible because of their short service lives.

Voltage in V

The sodium–nickel-chloride battery belongs to the family of high-temperature batteries. This means that it only works beyond a certain system temperature (in this instance: 300–400°C). The electrode material is present in solid form

at room temperature; a chemical reaction is prevented with this battery [29-37]. The negative electrode consists of liquid sodium, and the positive electrode consists of nickel chloride. If the battery is being charged, the following chemical reaction takes place:

extracted charge in Ah

Figure 29.49 Discharge curve of a NiMH cell for different discharge currents [29-36].

Internal Combustion Engine Handbook | 999

6606_Book.indb 999

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

29.5.4 Lithium-Ion Battery

Li-ion batteries contain a lithium metal oxide cathode and a graphite anode. A lithium salt or polymer is often used as the electrolyte. Li-ions move back and forth between the electrodes when charging and discharging. The following reaction takes place during charging: LiMeO2 + x C6 ⇒ Li1–xMeO2 + x LiC6 (Me = Ni, Mn, Co) The reaction is reversed during discharging: Li1–x MeO2 + x LiC6 ⇒ LiMeO2 + x C6

Voltage in V

The foremost advantage of the lithium-ion battery is its significantly higher specific energy density compared to leadacid and Ni–MH batteries. The cell voltage is 3.6 V or higher for lithium-ion batteries, depending on the material, and is consequently higher than other battery systems being used for hybrid drive systems. A high cell voltage has the advantage

that fewer cells and subsequently less installation space is required for the individual voltage level needed. As with other battery types, the cell voltage of the Li-ion battery depends on the state of charge (SOC) (Figure 29.50). The cell voltage falls as the SOC falls. But the voltage level is also dependent on the discharge current. Li-ion batteries can be manufactured from various different materials (Figure 29.51). They differ from each other not just in terms of cell voltage, but also in terms of other properties. The various properties of a few materials are evaluated in Figure 29.52. There are differences here, particularly in service life and safety. Compared to the nickel–metal hydride batteries previously used in hybrid drive systems, Li-ion batteries exhibit better behavior at low temperatures. In addition, there is no voltage hysteresis while charging/ discharging, which means that an efficiency of 100% is theoretically possible.

Figure 29.50 Voltage curve for different discharge currents [29-38].

extracted charge in Ah

Potential Li/Li+ in V

Cathode material

Anode material Sn-M-C composite Carbon Graphite

Si/C composite

Capacity in Ah/kg

Li-metal

Figure 29.51 Different materials used to construct lithium-ion batteries [29-39].

1000 | Internal Combustion Engine Handbook

6606_Book.indb 1000

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.5 Energy Storage Devices

Output

Service life (calendrical)

Safety

Costs

Figure 29.52 Comparison of different materials [29-40].

Service life in years

Above average Average Characteristic is in need of improvement Problematic

Process capability

Temperature in °C

Li-ion batteries are still in an early stage of development when it comes to their use in automobiles. Sensitivity to short circuits and overcharging is a serious downside of current Li-ion batteries. For this reason and because very large amounts of energy are stored compared to other applications in the vehicle, safety is now a research priority. Therefore, unlike other battery systems, Li-ion batteries require cell monitoring.

29.5.5 SuperCaps

SuperCaps (also known as UltraCaps or PowerCaps) are electric double-layer capacitors and work like capacitors in principle. Unlike batteries they do not use a chemical concept, but store the energy electrostatically. Figure 29.55 shows the basic construction of an electric double-layer capacitor. The electrodes consist of activated carbon. Aluminum is used as the separator [29-42]. They have a much smaller energy density than batteries (approximately 100 times less than a Li-ion battery); instead,

Figure 29.53 Service life of a Li-ion battery depending on the temperature [29-41].

they have a power density that is around ten times greater. Another advantage over batteries is their large calendrical and cyclical service life of >10 years and >1.5 million cycles. Electric double-layer capacitors exhibit very good charging/ discharging efficiency; observing the frequency dependence of the capacitance during quick charging/discharging cycles is critical. The capacitance of the capacitors falls as the frequency increases. A major distinction between batteries and capacitors is the voltage curve. With capacitors, the voltage is proportional to the charging gradient and, as result, is more dependent on it than battery systems are (Figure 29.56). Because of their properties, they are little suited to a role as sole energy storage device for hybrid drive systems with a high level of hybridization; however, because of their very high energy density, they could be used for short-term energy storage/transfer, for example, regenerative braking, shortduration boost modes, or smoothing peak loads.

Internal Combustion Engine Handbook | 1001

6606_Book.indb 1001

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Service life in cycles

Chapter 29 Hybrid Drive Systems

Figure 29.54 Service life of a Li-ion battery depending on the discharge stroke [29-41].

Negative Electrolyte electrode

Separator Positive electrode

Separator

Activated carbon material

Activated carbon material Aluminum foil

Figure 29.55 Design of a SuperCap manufactured by Epcos.

Charging

Battery

Voltage

Discharging Charging

Discharging l = const.

Time

Figure 29.56 Charging/discharging of UltraCaps and batteries.

1002 | Internal Combustion Engine Handbook

6606_Book.indb 1002

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.5 Energy Storage Devices

Battery housing High-voltage interface Onboard supply

Module Module electronics

Commands, information Power supply

Cells

Commands, information

Cooling system

Battery management system

CAN

Figure 29.57 Battery system.

29.5.6 Battery Management

Battery management is made necessary because of the many electrical consumers and subsequent amounts of energy, also because the battery’s service life is dependent on charging, discharging, and temperature. Battery management in a hybrid drive system controls charging and discharging of the storage cells; it monitors the battery and computes characteristic values, initiates any necessary actions, and communicates with the other systems in the vehicle [29-43].  attery monitoring B Battery monitoring is a key component of battery management, on the one hand for receiving information about which driving functions are possible, and, on the other hand, for ensuring battery functioning for as long as possible. Battery monitoring is also necessary for safety reasons, particularly in Li-ion batteries. The temperature, voltage, and current are measured using various sensors. The measured values are recorded both for the battery module as well as for each individual cell to some extent. Various characteristic values can be calculated using the measured values; these characteristic values are required for further evaluations, but other systems are also available, for example, the operating strategy via bus systems. The most important characteristic values are SOC, depth of discharge (DOC), state of function (SOF), and state of health (SOH). SOC indicates the current battery charge status as a percentage, based on the current capacity of the battery to the maximum possible or to the nominal capacity. Because capacity cannot be measured directly, the SOC is determined indirectly with the voltage and current measurements. This utilizes the fact that the steady-state voltage is dependent on the SOC among other factors, with the result that voltage metering can be implied to the SOC. If steady-state voltage measuring is not possible, for example, because the battery is now being charged, then the SOC can be implied by integrating the current in connection with an earlier steady-state voltage value. Because of the lack

of measuring quality in the vehicle, but also because the SOC is influenced by other parameters, it is necessary to verify the measurement with mathematical models. The reciprocal value for the SOC is the DOC, which is calculated from 100% SOC. The SOH is used as a parameter for battery aging. The SOH is the quotient from all the current, maximal possible capacity of the storage device and nominal capacity. An SOH is usually set at <1. The SOF is particularly important for other systems in the vehicle, as it provides information on whether a certain function can be supplied with the necessary amount of energy or output.  attery service life B The battery is a decisive cost factor in a hybrid drive system. In comparison with other application fields, such as in cameras or cell phones, the battery has to function for a comparatively long time. Chargeable batteries have a different characteristic service life depending on the type, distinguishing between calendrical and cyclical service life. The calendrical service life provides information about the temporal (days, months, years) service life of a battery. Degradation and evaporation influences, as well as wear, are the limiting factors in this context. The cyclical service life provides information such as how many charging and discharging cycles a battery contains. Because the cyclical service life also depends on the characteristics of the charging/discharging procedures, quantitative statements for an individual, specific application are difficult. As can be seen in Figure 29.58, the service life falls significantly measured against the penetrated energy depending on the discharge stroke; nevertheless, the basic behavior of the individual battery systems always remains the same in this point. Battery temperature is another factor that has a significant influence on the service life. Except high-temperature batteries, all batteries are sensitive to temperatures that are either too high or too low. The temperature of the battery is influenced by two factors. The first is the environment and the second

Internal Combustion Engine Handbook | 1003

6606_Book.indb 1003

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Ladungsdurchsatz in % der Nennkapazität

Chapter 29 Hybrid Drive Systems

NiMH Li-ion (MN oxide) Li-Ion (Ni-oxide) Lead acid – VRLA (AGM/gel) Lead-acid – liquid

Figure 29.58 Service life of different batteries dependent on the discharge level [29-44]. See color section page 1105.

in %

is the constant charging and discharging. The extremes are represented by a journey undertaken in summer underneath the blazing sun or one made in winter on a road covered in snow. To ensure the battery functioning all the same, the battery system should be equipped with air or coolant conditioning, thus allowing the battery to be cooled or even heated if necessary. If air is being used, then it may need to be dehumidified and removed of particles. Because the individual battery cells differ slightly, constant charging and discharging leads to different states of charge, that is, the SOC of the individual cells are not the same as that of the overall system; as a result, some cells are overcharged or exhaustively discharged in borderline situations. Furthermore, this can lead to negative, reciprocal effects between the individual cells, which shorten the service life of the battery. To prevent this from happening, the cells are monitored and adjusted to each other in a controlled manner if necessary.  harge and discharge control C Different charge and discharge procedures are suitable for individual battery types. Lead-acid rechargeable batteries are suited to charging at a constant voltage or with a voltage– current characteristic curve (VI characteristic curve). This involves charging at a constant current, until a certain charging voltage is adjusted. Afterward, a constant voltage is charged until the end. The current rating drops as the SOC rises. To accelerate the charging process, it is possible to switch back to charging with constant current if required, when falling below a preselected current. Li-ion batteries are ideally charged with constant current until the cell voltage is reached and then with constant voltage. In addition, the battery’s capacity is dependent on the SOC among other factors. The SOC range is displayed in Figure 29.59 for NiMH and Li-ion batteries, in which charging and discharging is expedient in light of the durability. As can be seen, a much higher proportion of the nominal capacity can be used for Li-ion batteries compared to NiMH batteries, in conjunction with the energy density, also higher in Li-ion batteries, which constitutes a major advantage. Limits are stored

in the battery management system, such as those shown in Figure 29.60, where the operating points of the battery from a hybrid vehicle are also marked.

Charging prohibited

d charging

Only reduce is possible

Full power

Only reduce

d charging

is possible

Discharging prohibited

Figure 29.59 Nominal and useful energy from batteries [29-45].

Current limit for discharging State of charge in %

DoD

Current limit for charging

Operating point in the FTP cycle

Battery current in A Discharging Charging

Figure 29.60 Battery operating points and limits based on the FTP cycle [29-46].

1004 | Internal Combustion Engine Handbook

6606_Book.indb 1004

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.6 Hybrid Drive System Transmissions

Optimal charging and discharging is rarely possible in a hybrid drive system, as the driver’s request is usually an obstacle to this. The battery management system must decide which compromises can be made between battery system durability and functional fulfillment.  ange estimation R To decide which electrical vehicle functions are now available and which are not, as well as initiating any measures for ensuring these functions, the vehicle’s energy management system requires information regarding the current capacity of the energy storage device from the battery management system. The capacity is made up of the available energy and maximum possible capacity load. The energy stored in the battery is calculated using the SOC and a recorded battery model. The maximum possible power is determined with the aid of maps, including dependence on the SOC and temperature. The energy content and/or power are decisive depending on the function. When calculating the range for the electric-only mode, then it is mostly the energy content that is of interest, whereas when starting up the combustion engine or with short boost modes during overtaking maneuvers, maximum available power is also important. If the recorded values reveal that a function is only partially or not at all possible, then the energy storage device must be filled or emptied by shifting the operating point of the combustion engine.

29.6 Hybrid Drive System Transmissions In a vehicle drive system, the transmission is tasked with adjusting the engine’s supply characteristic to the vehicle’s requirement characteristic. From the variety of technical options,

Combustion engine

Electrical machine(s)

Speed range Output Efficiency

Speed range Output Efficiency

Ratio

Ratio

Number of gears Transmission ratio

Number of gears Transmission ratio

Transmission

manual, conventional automatic, and CVT transmissions, as well as dual clutch transmissions, recently, have become widespread in gasoline and diesel drive system depending on region. Compared with drives in gasoline and diesel engines, a hybrid drive system has at least one other motor in the drivetrain. There are various options for the arrangement of the engine/motor plus transmission, depending on which functions (start/stop function, boost mode, operating point shift, electric-only drive, etc.) are scheduled to be implemented in the hybrid drive system. The transmission can combine both the combustion engine and the electric motor, it can be arranged between the two and thus only influence the combustion engine, or it can support the electrical machine. Other features include the particular speed and power ranges, as well as the operating-point dependent efficiency of the individual components (Figure 29.61). This calls for significant modifications in some cases compared to transmissions used in purely gasoline or diesel drive systems. For micro hybrid systems, that is, for systems with a low level of hybridization, modifications to the transmission are not usually required, whereas new transmissions are required for power-split systems. A special case is illustrated with series hybrid drive systems, where the electric motor is simply mechanically connected to the drive axle. Because of the electric motor’s good supply characteristic, a mechanical multistep transmission is usually omitted for this connection. In the following subchapters, we will focus on the integration and combination of the transmissions with the electrical machines used in hybrid drive systems. The actual function of the different transmission types can be checked in Chapters 29.6.1–29.6.3. Figure 29.62 illustrates the advantages and disadvantages of various transmission types in combination with hybrid drive systems.

Function Electric driving Start/Stop Boost mode

Arrangement ICE-E-G ICE-G-E ICE-GE

Figure 29.61 Technical influences upon the transmission.

Internal Combustion Engine Handbook | 1005

6606_Book.indb 1005

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Manual transmission

Automated Automatic manual transtransmission mission

CVT

Dual clutch Electric transCVT missions

Component costs Application complexity (varies greatly according to the functional design) Installation space / Package (Transmission only)

Efficiency Acceleration section Efficiency Constant travelsection Efficiency Energy recovery Option of all-electric driving Efficient shifting strategy

Figure 29.62 Advantages and disadvantages of various types of transmission in combination with hybrid drive systems.

29.6.1 Transmissions Without Integrated Electrical Machine

combustion engine and the transmission in mild hybrids. Thus, rendering is unnecessary to modify the transmission.

•• connected to the combustion engine via a belt drive,

 utomated manual transmissions A Automated manual transmissions are better suited to use in hybrid vehicles than manual transmissions. This is because specific shifting strategies can be implemented with the former. However, automated manual transmissions have until now exhibited a low level of shifting comfort, with the result that they have yet to make a breakthrough. A hybrid concept has no advantages or disadvantages in this regard.

The transmission does not have to combine the speed and torque range of both the combustion engine and the electric motor in all hybrid drive concepts. If the electrical machine is •• arranged between the combustion engine and the transmission, •• arranged between the transmission and the drive axle, •• or if the combustion engine and the electrical machine operate on different vehicle axles, it is still necessary to adjust the speed/torque range at the transmission input shaft to the requirements. Therefore, it is possible to use conventional transmissions, that is, the type usually employed with combustion engines.  anual transmissions M Manual transmissions are, from an energy-efficiency perspective, unsuitable for hybrid drive systems above the micro hybrid level. Although they boast the highest efficiency levels of all mechanical vehicle transmissions, the transmission ratio is freely selected by the driver, and therefore a low-energy operating strategy is difficult to implement. For the same reason, an electric-only drive in combination with manual transmissions is not expedient. Manual transmissions can only be used when the level of hybridization is so low that the electrical machine does not make a meaningful contribution to the driving mode, for instance, in micro or tenuous mild hybrid systems. In micro hybrid vehicles, electrical machines are linked with the combustion engine the same as the alternator, or the existing electrical machines (starter and alternator) are simply complemented by one intelligent control logic. The electrical machine is usually positioned between the

 utomatic transmissions A Automatic transmissions are particularly suited to parallel hybrid drive systems. While it is true that the driver is able to influence the selection of gear with modern automatic transmissions, and in doing so prevent optimum energy management, the gear is usually selected automatically, meaning that an efficient strategy can also be implemented. The electrical machine is installed either between the combustion engine and the automatic transmission or directly onto the transmission input shaft. Additional clutches are required between the combustion engine and the electrical machine and/or between the transmission and the electrical machine. In return, it is sometimes possible to eliminate the hydrodynamic torque converter and subsequently improve transmission efficiency.  ual clutch transmissions D Dual clutch transmissions are similar to automatic transmissions and automated manual transmissions in that they are also suitable for parallel hybrid drive systems. The electrical machine is likewise installed between the combustion engine and the transmission. If the gear is selected by the driver, this brings the same disadvantages as with manual transmissions. Transmission efficiency is higher than that found in automatic transmissions.

1006 | Internal Combustion Engine Handbook

6606_Book.indb 1006

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.6 Hybrid Drive System Transmissions

CVT transmissions Continuously variable transmissions (CVT) are just as well suited to use in hybrid drive systems as automatic transmissions are because of their continuously variable transmission ratios. However, the operating point is always a compromise between the optimal operating point of the electrical machine and that of the combustion engine.

29.6.2 Transmissions with Integrated Electrical Machine

Although electrical machines boast an ideal supply characteristic and their efficiency is very high compared to combustion engines, transmission ratios are still expedient or even necessary sometimes for them. Consequently, for electrical machines as with combustion engines, there are speed limits and operating ranges with relatively low efficiency (Figure 29.63).

Various transmission solutions are presented hereinafter in ascending complexity. Figure 29.64 depicts the Toyota hybrid system (THS). Both electrical machines and the combustion engine are linked together via a planetary gear set, specifically the combustion engine with the planet carrier, the first electrical machine with the sun gear and the output, as well as the second electrical machine with the internal gear. With this arrangement, a part of the mechanical power from the combustion engine is always converted into electrical power, because the combustion engine is supported by the sun gear. This power can be added again with the corresponding exchange losses from the second electrical machine.

Electrical machine

Output

Combustion engine

Electrical machine

Torque in Nm

Figure 29.64 Diagram of the transmission in a Toyota Prius.

Working point of the electrical machine at full-load acceleration Speed in rpm Figure 29.63 Operating range during full load acceleration in the efficiency diagram [29-48].

Hence for transmissions with integrated electrical machines, there are requirements in terms of

However, when driving steadily or at maximum power, this does lead to efficiency losses. Dual drive developed by Nexxtdrive is another hybrid drive concept. It involves a combination of various planetary gear sets and two electrical machines (see Figure 29.65 and Figure 29.66). The combustion engine drives the planet carrier, and three gearwheels are located on each of its three axles. One electrical machine is directly linked to the gearwheel of the third group via a sun gear, and the second electrical machine is connected to an additional transmission ratio via another sun gear [29-49]. These different transmission ratios have the advantage that one of the electrical machines can be designed as high speed and the other as high torque. Output occurs via the third gear set and a sun gear.

•• power distribution and power conflation •• operating point adjustments to the combustion engine •• operating point adjustments to the electrical machine(s) •• mechanical-only power section (from the combustion engine to the wheel) •• electrical-only power section (from the electrical machine to the wheel) •• starting up the combustion engine via the electrical machine •• reverse gear. The entire spectrum of transmission types are available to provide solutions, but concepts with planetary gears are particularly well suited, as they can be shifted with no interruption in tractive force and provide two degrees of freedom per gear set. In conjunction with electrical machines, these types of transmissions as also known as electronically controlled continuously variable transmission.

Combustion engine Electrical machine Electrical machine

Output

Figure 29.65 Concept image of the DualDrive.

Figure 29.67 illustrates a further development of the THS for the Lexus GS450h. In comparison to the earlier described planetary gear set of the THS, this time a Ravigneaux set (specific combination of two planetary gear sets) is shifted between the electric motor and the output shaft. As a result, the power of the electric motor can be coupled with two different transmission ratios (in this instance: Gear reduction 1.9 and 3.9) and subsequently with a higher level of efficiency.

Internal Combustion Engine Handbook | 1007

6606_Book.indb 1007

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Figure 29.66 Sectional view of the DualDrive [29-49]. See color section page 1105.

Generator

Electric motor

Combustion engine

Output

Power-split transmission

Two-stage creeper gear

Shifting uses both of the two clutches, K1 and K2. Even with this transmission, a proportion of the mechanical power is always converted into electrical power. This disadvantage is avoided in Two-Mode transmissions (developed by GM-DC-BMW) among others. In addition to the combined operation with different transmissions, both the electric-only drive system and the mechanical direct drive of the combustion engine are possible. A diagram of the Two-Mode transmission is shown in Figure 29.68. Figure 29.69 shows a section through the transmission. The examples listed illustrate the wide range of possible approaches.

Figure 29.67 Diagram of the transmission of a Lexus GS450h [29-50].

Another conceivable idea is that of combining two CVT transmissions, to constantly operate both the combustion engine and the electrical machine at their optimum points. However, limits are set to the extent of transmission complexity, in that the benefit in the shape of a more favorable operating point must be greater than the potential transmission efficiency loss.

29.6.3 Special Transmission Designs

In addition to transmissions with mechanical power transmission, there is also the option of electrical power transmission. An example of such a transmission was introduced by Volkswagen among others (Figure 29.70 and Figure 29.71).

1008 | Internal Combustion Engine Handbook

6606_Book.indb 1008

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.6 Hybrid Drive System Transmissions

Electrical machine

Electrical machine

Internal gear

Internal gear

Internal gear

Planetary

Planetary

Planetary

carrier

carrier

carrier

Sun gear

Sun gear

Sun gear

Combustion engine

Output

Figure 29.68 Diagram of the Two-Mode transmission [29-51].

Figure 29.69 Sectional view of the Two-Mode transmission [29-51].

Such transmissions have a generator and an electric motor arranged opposite from each other, both sharing a common stator. This stator can be smoothly adjusted on its longitudinal axis, and thus enables a variable transmission ratio. The combustion engine is flanged-mounted on the external rotor of the generator, and the output to the wheels is in turn flange-mounted on the external rotor of the electric motor. Three operating ranges are distinguishable depending on the position of the stator: 1. The stator is located entirely on the combustion engine side. In this position, it can work alongside the external rotor as either a generator or a starter for the combustion engine, because energy can be both fed to the stator as well as extracted from it. Power is not transmitted to the output.

Figure 29.70 Sectional view of the MEGA (Magnetisch-Elektrische GetriebeAutomat) transmission [29-52]. See color section page 1106.

Internal Combustion Engine Handbook | 1009

6606_Book.indb 1009

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Output rotor (Engine)

Input rotor (Generator)

Stator support

Crankshaft Variable stator with coil

Sprocket, Mechanical Power transfer to differential

Figure 29.71 Sectional model and schematic diagram of the MEGA transmission [29-53].

Electrical machine

Sealing

Open differential

Connection, coolant Cooling jacket

Flange, side shaft

Rotor bearing

Flange, cardan shaft

Figure 29.72 Electrical rear axle concept [29-54].

2. The stator is located entirely on the output side. If the stator is located on the output side, the vehicle can either be driven in electric-only mode while being supplied with electrical energy or energy can be recuperated when rolling/braking. 3. The stator is located in an intermediate position. Each intermediate position relates to a transmission ratio between the combustion engine and the output. Another concept combines the electrical machine with the axle differential. One electrical machine influences each of the outputs from the axle differential (Figure 29.72). Mechanical direct drive is still possible. Because the electrical machines can be operated as generators or electric motors independently of each other, torque vectoring is also possible in addition to the standard hybrid drive functions. The torque delivered from the combustion engine can therefore, depending on how the electrical machines are actuated, be distributed individually between

the two wheels. Electric-only drive can also be realized with an energy storage device and an extra clutch.

29.7 Energy Management System A vehicle’s energy management system is tasked with utilizing the available energy in an ideal way in the operation of the vehicle, as well as controlling the energy flows to ensure that all functions can be ensured in the desired form (Figure 29.73). As per [29-55], the energy management system balances the required energy against that which is generated and compensates the energy generated, stored, and required. Besides the use of combustion engine and electric motor in a hybrid vehicle, two energy storage devices are also available. Of the two, only the electrical storage device is able to both deliver and receive energy. The energy is supplied to the vehicle in the form of fuel, but also by charging the battery via an external energy source in a plug-in hybrid

1010 | Internal Combustion Engine Handbook

6606_Book.indb 1010

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.7 Energy Management System

Consumers

Energy sources

Electric driving Boost mode Light Infotainment Air conditioner

Regenerative braking Shifting the operating point

Figure 29.73 Energy management tasks.

vehicle. Because of the high energy density of liquid, and gaseous fuels compared to batteries, electrical energy only represents a small proportion of the total energy. Thanks to the combustion engine and a connected electrical machine, fuel can also be converted into electrical energy. Functions such as increasing the operating point, regenerative braking, and, perhaps in future, exhaust emission energy recuperation add to the options for producing electrical energy. From the mentioned items, a hybrid drive system results in a significantly higher degree of freedom in terms of supplying energy than a drive system designed solely for a combustion engine. Consumers are an obstacle in this instance. Besides the classic components, such as control units, air conditioning system, infotainment, and lighting, a hybrid vehicle also has an electrical machine, with which the drive functions, boost mode, and electric drive are implemented as necessary (Figure 29.74). It is evident that the energy management system selected also has an effect on fuel consumption. Can the strengths of the individual drive components be utilized? The energy management system is closely linked to the general operating strategy, with which the driver’s request is spread between the two drive machines. The functions, which

can be implemented/approved, also depend on the energy management system. Figure 29.74 illustrates an example of the approval of functions depending on the battery SOC for a parallel hybrid. However, the energy management system can differ enormously depending on the type of hybrid drive system. The compensation between generated, stored, and required energy can be improved still further in perspective with an analysis of the driving behavior in the sporty or defensive driving mode [29-57]. Even information concerning the surrounding area collected using sensors, GPS, and car-to-car communication helps to achieve an ideal energy management system to record the respective driving situation better, for example, to adjust the battery SOC to a traffic light stop or overtaking maneuver.

29.7.1 Start/Stop

Shutting off the combustion engine whenever it is not required is one approach for saving energy and in turn fuel. Start–stop systems (also known as: stop–start systems) invariably shut off the combustion engine when the vehicle is idle, that is, for instance, at traffic lights or in a traffic jam and when the driver actuates the brake pedal, as well as when the driver disengages

Output / Current / Moment

Energy recovery

Battery: Charging Electrical machine: Generator

Shifting the operating point Electric driving Battery: Discharging Electrical machine: Engine

Starting

Boost mode (short)

Boost mode (long) Variable limit range

Figure 29.74 Example of an energy management system [29-56].

Internal Combustion Engine Handbook | 1011

6606_Book.indb 1011

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

the gear in a vehicle with manual transmission. Compared to a conventional drive system, the number of engine starts and stops subsequently rises significantly, approximately by a factor of ten [29-58]. This must be taken into consideration when designing the service life of the components involved [29-59]. It must be ensured that there is enough energy for the next start prior to each stop. The current energy demands of the individually switched-on consumers, radio, rear window heating, etc., must also be taken into account. In comparison to the vehicles without a start-stop system, consumers do not necessarily have to be switched off during an engine stop. Furthermore, for reasons of comfort, an engine stop should not take place too soon after an engine start. An engine stop should be omitted where necessary. The start/stop function is the most elementary function of a hybrid drive system.

29.7.2 Regulating the Generator

Despite regenerative braking, potential external charging of the battery or exhaust gas energy recuperation from the combustion engine, most of the electrical energy is produced by a generator. Provided that the hybrid drive concept allows for it, it is sensible to postpone electricity production into combustion engine phases that are economical in terms of fuel consumption. This involves thrust phases and steady-state operating conditions. By contrast, the generator should be deactivated when setting off or during heavy accelerating. Therefore, the combustion engine is unburdened in the particularly low-efficiency/power lower speed-load range. The transition between intelligent generator regulation (Figure 29.75) and regenerative braking or shifting the operating point is, however, fluid.

Delay

Acceleration

In combustion engines, only a relatively small part of the fuel energy applied is converted into mechanical work and is subsequently used to move the vehicle. The by far greater part of it is lost in the form of heat (Figure 29.76). In addition, the kinetic energy of the vehicle is also converted into heat during braking etc. Energy recuperation, for example, the conversion of exhaust gas heat or kinetic energy into an energy that is once again useful to the vehicle is therefore one option for making a noticeable reduction to fuel consumption, regardless of whether or not this involves a hybrid drive system. Yet hybrid drive systems do have the advantage that their components, specifically the electrical machines, power electronics, and battery are suitable for converting and storing energy. In the generator operation, the electrical machines are able to convert the vehicle’s kinetic energy into electrical energy and thus replace or at least relieve the classic brake. Because high power surges occur for short periods during vehicle braking, a compromise must be made between the capacity of the electrical machine and battery on the one hand, and the proportion of the converted energy on the other hand. Otherwise, the individual hybrid components would need to be considerably more powerful and would consequently be much heavier and more expensive. Besides this conflict of aims, the integration of the electrical machine to the powertrain and the vehicle axle coupled to it are other important influencing factors on high energy recuperation. To achieve a high yield during vehicle braking, the combustion engine and transmission need to be decoupled from the axle and the electrical machine, as they would also convert a part of the vehicle’s kinetic energy into heat. However, this is not possible for every hybrid concept, and/or leads to higher costs because of the additional clutches.

Constant travel

Generator performance

Speed

Constant travel

29.7.3 Energy Recuperation

Time Without regulation With regulation Time Increased generator performance

Generator switched off

Generator switched off (as required)

Figure 29.75 Intelligent generator regulation [29-60].

1012 | Internal Combustion Engine Handbook

6606_Book.indb 1012

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.7 Energy Management System

Fuel energy Air conditioner Alternator

Cooling system

Steering assistance Water pump Miscellaneous Oil pump Transmission Differential Rolling resistance

Exhaust gas energy

Aerodynamic drag Accelerating

Driving

Figure 29.76 Distribution of fuel energy.

Moreover, varying wheel load distribution during braking leads to a situation where braking power is greater at the front axle than it is at the rear axle. And what is more, the rear wheels may never be blocked for driving-dynamic reasons. To achieve optimal regenerative braking, the electrical machines would therefore have to be effective on both axles. If this is not possible, a connection to the front axle is the next best option. It should not go unmentioned that because of the high braking power requirement, but also for legal reasons, a mechanical brake is still required. The interaction of the two braking systems and the energy recuperation via the brake pedal to the driver require fine-tuning. The energy management system must ensure that the electrical storage device is also able to receive the accumulated energy; otherwise, there would be a need to brake firmly using the mechanical brake and kinetic energy would be converted into heat. Various methods are conceivable for converting the heat energy in the exhaust gas, for example, via a small thermal engine or a thermoelectric generator. Particularly the latter, so-called TEGs are now the subject of research. It is hoped that they will lower fuel consumption in the NEDC by approximately 0.2 l/100 km [29-61]. The advantage of TEGs compared to other concepts is that they convert heat directly into electrical energy and have no moving parts. Even though the potential of exhaust gas heat is very promising, present exhaust gas energy recuperation concepts are either very cost intensive or are not yet ready for series production.

29.7.4 SOC Regulation

To guarantee the various electrical driving functions, but also to ensure the durability of the battery, the battery SOC must be regulated. Depending on the type of energy storage device used, for example, Li-ion or Ni-MH, both very high and very low states of charge should be avoided. Because the service

life is also dependent on the number of charging/discharging cycles, unnecessary, powerful charging and discharging of the battery must be avoided. At the same time, a long electrical drive support, as well as, for example, regenerative braking over long journeys is also desirable. To enable all the hybrid drive functions while simultaneously sparing the battery, an SOC target value needs to be determined first from the ambient conditions. This target value is then used to regulate the individual consumers and current suppliers. Options are available for increasing the battery’s SOC, for example, increasing the operating point of the combustion engine, starting the motor to produce electricity with the generator, as well as regenerative braking. The SOC can then be reduced with increased electrical drive support. Here too, all or only some of the functions are available depending on the hybrid concept.

29.7.5 Energy Distribution Management

The energy distribution management system comprises the overvoltage and undervoltage protection for the on-board power supply, that is, it monitors target voltages and implements fault measures. To switch consumers on or off in an expedient manner, first, they must be classified. Besides prioritizing along safety aspect lines, it also expedient to classify them according to duty cycle and possible power regulating [29-62]. Consequently, safety functions such as ESP, power steering, or lighting may never be disabled by the energy management system, whereas convenience devices such as air conditioning or navigation tools can be omitted in some circumstances. Classifying them according to typical duty cycle is also highly expedient as an additional decision aid. For instance, an indicator is usually only active for short periods, but a control device is always on. Finally, the control type is also important: can the consumer only be switched on and off, or are intermediate steps possible?

Internal Combustion Engine Handbook | 1013

6606_Book.indb 1013

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

29.7.6 On-Board Power Supply System

The electrical energy and power requirements of a hybrid drive system are significantly greater than those of conventional drive systems. Even the power requirements of a passenger car without a hybrid drive system have grown to such an extent in recent years that a standard 12-V on-board supply system is stretched to its limits. The additional electrical machines in a hybrid drive system cause the power requirements to soar once more. The result is that the standard 12-V on-board supply system no longer meets the requirements in part because of the high flows. Therefore, except a micro hybrid, a second supply system with a higher voltage level is inserted in addition to the 12 V power supply (Figure 29.77). Current converters are responsible for adjusting the electrical energy to the components’ varying requirements. The voltage value of the additional level is primarily determined by the power to be transferred. Voltage levels of 144 V are suitable for mild hybrids with electrical machines that have a power range of 20 kW, but voltages of up to 650 V can be required for full hybrids. The maximum possible voltage level is restricted both by the semiconductors available and by safety considerations. In the context of this restriction, the conflict of aims regarding the highest possible voltage for the electrical machines, converters, and lines, as well as the lowest possible voltage for the battery, needs to be resolved. To keep voltage levels stable and avoid disturbances, the respective voltage level is either supported directly from the battery or capacitors are used to smooth short-term disturbances.

Subsequently, a suitable choice of operating point can have an influence on fuel consumption and emissions. Besides the efficiency levels of the different drive systems, however, some restrictions need to be taken into account. For example, the energy reserve in the battery is so low that some of the drive functions cannot be used for as long as may be required. The decisive factor upon the operating strategy is the hybrid concept selected and level of hybridization. Thus, the driver request in a micro hybrid is almost always assumed by the combustion engine because of its low electrical power. The operating strategy simply involves the question: when is the combustion engine stopped and started? And during which drive conditions is the electrical machine switched on and off by the generator? A plug-in hybrid vehicle represents the other extreme: in this instance, the vehicle can be supplied with energy externally via an electrical branch and its energy storage devices are sometimes powerful enough for the vehicle to travel longer distances solely with electrical energy. On the other hand, the complexity of the operating strategy is reduced in a series hybrid drive system. In this case, the combustion engine is not responsible for direct traction tasks, and it simply generates the electrical energy for the drive electrical machine in conjunction with another electrical machine. In this instance, the extent to which the combustion engine follows the dynamic, that is, provides the traction energy required at that moment or simply works at its optimum point and primarily feeds the electrical energy to a battery, is important.

29.8.1 Efficiency

29.8 Operating Strategies The operating strategy specifies how the driver request will be met by the two drives (combustion engine and electric motor). Unlike vehicles with gasoline or diesel engine drive systems, hybrid drive systems have no clear correlation between the driver power request and the power of the combustion engine.

When determining the operating strategy, it is helpful to consider the various levels of drive engine efficiency. Compared to electrical machines, gasoline and diesel engines have, depending on the operating point (load/speed), very low levels of efficiency of approximately 30%. Figure 29.79 illustrates the specific fuel consumption of a gasoline and a diesel engine.

Variable high-voltage level

Traction battery

Voltage level Fixed highvoltage level

Traction battery

Voltage level

Figure 29.77 Example of an on-board supply system with highvoltage intermediate circuit.

1014 | Internal Combustion Engine Handbook

6606_Book.indb 1014

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.8 Operating Strategies

Driver request

Driver request Operating strategy

Regulating the combustion engine

Regulating the transmission

Regulating the combustion engine

Conventional drive system

Regulating the electrical machine(s)

Regulating the transmission

Regulating the energy storage device

Hybrid drive system

Torque in Nm

Torque in Nm

Figure 29.78 Comparison of the complexity between a conventional drive system and a hybrid drive system [29-63].

Speed in rpm

Speed in rpm

Figure 29.79 Map of a combustion engine (left: gasoline engine, right: diesel engine).

Both engines display the characteristic, shell-shaped consumption contour lines, where in each case the optimum point lies in the upper load range, at medium speeds. Up to low loads, the specific fuel consumption and subsequently the efficiency in the entire speed range drops sharply. In contrast, the efficiency levels of the electrical machine lie at a significantly higher level, in the optimum point above 95%. Figure 29.35 shows the maps for a synchronous and an ASM. The two machine types exhibit slightly different

characteristics. Because the level of efficiency falls dramatically with these engines at low speeds and close to zero load, this range should likewise be avoided. When comparing these two types of engines, the level of efficiency when storing and converting electrical energy must also be taken into account. An efficiency of between 80% and 95% can be applied to converters, compared to an efficiency of 80%–90% for storing electrical energy (see Figure 29.80).

Internal Combustion Engine Handbook | 1015

6606_Book.indb 1015

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Primary energy Fuel

Conversion into mech. energy Combustion engine (20) - 35%

Storage

Conversion into electrical energy

Electrical machine

Inverter

(50) - 95%

80 - 95%

Braking energy

Conversion into mech. energy Electrical machine

Storage HV battery

(50) - 95%

80 - 90% Figure 29.80 Efficiency chain [29-64].

29.8.2 Energy Balance

Besides pollutant emissions, energy balance is decisive when choosing an operating strategy. The combustion engine’s Willans lines [29-65] (Figure 29.81) are an important decision-making aid here, where the hourly fuel consumption is plotted over the mean pressure curve or torque at various speeds. Analogous to this, the output power over the input power for the electrical machines can be illustrated in a Pin/Pout diagram [29-66]. Furthermore, if the efficiency chain for the generation and storage of electrical energy, as well as regenerative braking is considered, then the most favorable drive variant can be selected for each operating point: Erecuperated = η generator ⋅ η converter ⋅ η battery ⋅ Ekinetic Ebattery = η VKM ⋅ η generator ⋅ η converter ⋅ η battery ⋅ Efuel Ekinematic electrical machine = η converter ⋅ η battery ⋅ η motor ⋅ Ebattery

Specific fuel consumption

Speed 6 Speed 5 Speed 4

Speed 3

Speed 2 Speed 1

Torque Figure 29.81 Willans lines for a combustion engine.

29.8.3 Fuel Consumption

The fuel consumption of hybrid drive systems compared to pure gasoline and diesel drive systems can be reduced by implementing one of four measures:

•• On the one hand, the combustion engine can be shut off during phases in which it is not required, for example, when the engine is idle. •• On the other hand, operating ranges where the efficiency of the combustion engine is disproportionately low, for example, low-load areas, can be avoided. It is particularly in urban traffic and congestion that fuel is saved this way. •• Furthermore, the powerful electrical machines can be used to transfer the production of required electrical energy to favorable combustion engine operating phases. •• Ultimately the electrical machine is able to complement the combustion engine at full throttle, meaning that the combustion engine can be designed as smaller. This support, however, is only limited to a short period by the batteries. Even if the electrical storage device can be charged externally and fuel consumption falls, this does not necessarily lead to a reduction in the vehicle’s energy consumption. If the fuel consumption is reduced by charging the hybrid drive system’s battery externally, then the vehicle’s overall energy balance needs to be taken into account. The vehicle is not necessarily consuming less energy during this process.

29.8.4 Exhaust Emissions

Aside from fuel consumption, exhaust emissions from a hybrid drive system can also be lowered in comparison to gasoline and diesel drives. However, the reciprocity between fuel consumption and exhaust emissions needs to be taken into account. Both positive and negative correlations arise. If the combustion engine is shut off and therefore does not consume any fuel, then at first glance, there will also be no pollutants emitted. In a cold environment and during frequent start/stop operation, it may, however, be the case that the exhaust gas aftertreatment systems do not reach their operating temperature, with the result that, because of the bad conversion rate, exhaust emissions are higher in total than if the combustion engine was being run warm at first. One way of reducing pollutants is to lower the dynamic of the combustion engine, also called phlegmatization. Quick speed and load cycles leads to disproportionately high pollutant emissions in combustion engines, because the control times of the individual engine actuators cannot be adjusted quickly and accurately enough.

1016 | Internal Combustion Engine Handbook

6606_Book.indb 1016

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.9 Current Hybrid Vehicles

29.8.5 Driving Performance

29.8.6 Approaches for Determining an Operating Strategy

Basic operating strategy approaches depend on the type of hybrid drive system. In a series hybrid, the drive axle is then mechanically connected to the electrical machine. Hence, the operating point of the electrical machine is directly linked to the driver request. In return, the operating point of the combustion engine can be chosen freely, provided that it supplies sufficient electrical energy with the coupled generator. With the aid of the battery, the combustion engine can for instance only be operated at one speed-load point, which covers the average energy demand, or, it follows the current power requirement of the electrical machine and therefore only taxes the battery very slightly. Various intermediate solutions are possible within these two extremes. For instance, the combustion engine is able to work at several, discrete speed-load points, or it is able to alter the load continuously at a fixed speed. With a parallel or mixed hybrid drive system, both the combustion engine and the electric motor are able to service the driver request. One basic strategy is to drive as “electrically” as possible in the lower load area, to extend the full throttle of the combustion engine with the electrical machine, and to supply the required energy using only the combustion engine at higher vehicle speeds. One example is shown in Figure 29.82. Because various options exist for an operating strategy for each of the many, possible hybrid concepts, and the vehicle

Boost mode

Combustion engine Torque

Driving performance is mainly dependent on the motor but also on the combination of combustion engine and electrical machine capacity. This depends on the type of hybrid drive system and level of hybridization. Therefore, in series hybrids, the electrical machine is solely responsible for driving performance, provided that it is supplied with sufficient electrical energy. By contrast, with parallel hybrids, driving performance is the product of the combustion engine and electric motor combined; however, in this case, the level of hybridization is a crucial factor. The combustion engine is responsible for driving performance in micro hybrid vehicles, because of their low electric share. With mild and full hybrids, the electrical machine determines the driving performance in the lower load range and the combustion engine determines it in the constant load range. The extent to which the driving performance of a vehicle with a hybrid drive is better or worse than a conventional drive system therefore depends on the hybrid concept chosen. A general statement is not possible. The fact that a hybrid drive is heavier than a conventional drive system because of the additional components: electrical machine and battery, turns out to be a disadvantage when accelerating close to full throttle, for example, at high speeds on a hilly interstate section. Yet the hybrid drive system has advantages in the lower speed and cycle range, because of the favorable torque characteristic of the electrical machine, meaning that better driving performance is possible in these areas.

Electric driving Speed Energy recovery Combustion engine drag torque

Figure 29.82 Operating strategy of a parallel hybrid [29-67].

drive must meet legal emissions standards, customer expectations, cost frameworks, etc., in this instance, it is expedient to use a simulation program such as Advisor or Velodyn [29-68] and efficient optimization processes. The overall vehicle simulation can be used to estimate the fuel consumption, exhaust emissions, and driving performance for each concept. To find the optimum now for each of the many parameters such as the electrical machine, storage device, transmission, and arrangement, modern model formation and optimization methods [29-69], as well as DoE [29-70] are necessary.

29.9 Current Hybrid Vehicles Even though hybrid drive systems as we know them today have been in series production since 1997, in relation to gasoline and diesel drive systems, they are still at the beginning of their development. Toyota now offers the widest range of models. Alongside the first global series production vehicle, the Prius, the Camry Hybrid, Estima Hybrid, and Highlander Hybrid, the RX 400 h, GS 450 h, and LS 600 h are also sold under the Lexus brand name (see Figure 29.83). Honda, also a pioneer of series hybrid vehicles, has the Honda Accord Hybrid and Honda Civic Hybrid in its range. Honda became the first automobile manufacturer to offer a hybrid drive system in the USA when it released the Insight. Also, in the USA various SUVs are offered as a hybrid version; these include the Nissan Altima Hybrid, Chevrolet Silverado and Tahoe, GMC Sierra, GMC Yukon, Ford Escape, Ford Mariner Hybrid, Mazda Tribute, and GM Saturn Vue. Some of the vehicles mentioned are identical models, or at least have the same hybrid components. The Chrysler Aspen and Dodge Durango were only granted a short production phase. Both vehicles were removed after just two months. The remaining vehicle manufacturers will only enter the hybrid vehicle market from 2009 onward. Mercedes-Benz has

Internal Combustion Engine Handbook | 1017

6606_Book.indb 1017

1/19/16 9:00 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Figure 29.83 Timeline of hybrid vehicles [29-71].

announced a hybrid version of the S-Class with Li-ion batteries for 2009. Mercedes, GM, and BMW have also developed a joint hybrid drive system, which will be used by GM as early as 2008 and sometime later by Mercedes and BMW. Vehicles equipped with a start/stop system are now offered by BMW, specifically the 1 Series and 3 Series, but also by PSA (C2 and C3).

29.9.1 Systems

There are now very few hybrid systems in series production. For the sake of completeness, we now turn our attention to micro hybrids, even though these vehicles are not usually associated with hybrid drive systems.

For start/stop systems, PSA relies on a starter generator manufactured by Valeo S.A. (see Figure 29.84). The functions of a starter a generator are combined in an electrical machine. A lead-acid battery is used as an energy storage device. BMW use a system developed by Bosch (see Figure 29.85), which consists of two electrical machines, a lightly modified starter, and an alternator. A lead-acid battery is also used here as the storage device. A simple battery management system controls the battery SOC. In combination with the engine electronics, both systems allow the combustion engine to start very quickly and stop comfortably, two important characteristics, which ensure that the customer accepts the systems and does not perceive them as irritating. Particularly long engine starts can irritate the driver.

Figure 29.84 Start/stop system manufactured by PSA [29-72]. See color section page 1106.

1018 | Internal Combustion Engine Handbook

6606_Book.indb 1018

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.9 Current Hybrid Vehicles

Onboard supply 12 Volt Signal lines

Engine control unit with start/stop software option

Neutral gear sensor

DC/DC converter 12 Volt

Wheel speed sensor

Battery sensor

Crankshaft sensor

Start/Stop starter

Efficiency-increased generator with regenerative braking

Above the micro hybrid systems, that is, the electrical branch of the drive system has a noticeable proportion here, the market is dominated by Toyota and their luxury brand Lexus. They use different variations of planetary gearing, each with two electrical machines (synchronous three-phase motors) in their hybrid vehicles (see Figure 29.86 and Figure 29.88). The electrical machines can operate as a generator or an electric motor depending on the drive mode. The combustion engine is connected to the planet carrier, one electrical machine to the sun gear, and the other to the internal gear. The electrical machine at the sun gear operates solely as a generator, while the other electrical machine operates both as an electric motor and a generator. The Lexus RX400h has a special feature. The four-wheel drive is realized via an

Figure 29.85 Start/stop system manufactured by Bosch [29-73]. See color section page 1106.

additional electric motor. The advantage of this construction is that it does not need an additional drive shaft to the rear axle, thus saving both on space and weight. Figure 29.87 depicts an electric motor with connection to the rear axle. Until now, Toyota has been using Ni–MH batteries in all Toyota and Lexus hybrid drives. The batteries are air-cooled, and in the Prius at least, this air is extracted from the vehicle interior. Toyota has already announced their intention to switch to Li-ion batteries in future. A power-split hybrid drive system has also been developed by a consortium of vehicle manufacturers, GM, Mercedes-Benz, and BMW. Until now the so-called Two-Mode transmission has only been used in GM vehicles. Series production rampups at Mercedes and BMW are planned for 2009.

Figure 29.86 Sectional view of the Prius and RX400h transmissions [29-74].

Internal Combustion Engine Handbook | 1019

6606_Book.indb 1019

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

Rear Electric Motor (MGR) Couter Gear Ring Gear

Speed Gear

Differential Gear Unit

Figure 29.87 Electric motor and transmission on the rear axle of the RX400h [29-75].

Electrical machine (generator)

Mechanical oil pump

Power-split

Electrical machine Two-stage creeper gear

Electrical oil pump

Hydraulic system Figure 29.88 Sectional view of the GS450h transmission [29-76].

Unlike its rival Toyota, Honda uses a crankshaft starter generator (Figure 29.89), which is a brushless DC motor. The system carries the designation IMA. Two gasoline engines are available in the Honda hybrid system. The first hybrid system consists of a four-cylinder engine with a CVT transmission, and the other system is composed of a six-cylinder engine with an automatic transmission. In contrast to its rival Toyota, Honda has made some modifications to the combustion engine. As a result, both engines have a sophisticated valve actuation system. There are four different operating modes: combustion engine off, valve actuation times for low and high speeds, and closed valves. The reciprocity between the electrical machine and combustion engine is shown for the six-cylinder gasoline engine in Figure 29.90. Three cylinders are shut off in the lower load range at low speeds. The three cylinders are supported by the electrical machine in the medium load range and lower speed

Figure 29.89 Four-cylinder petrol engine with crankshaft starter generator and CVT transmission [29-77].

1020 | Internal Combustion Engine Handbook

6606_Book.indb 1020

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.9 Current Hybrid Vehicles

Torque in Nm

Six cylinders and electric motor are active

Six cylinders active Three cylinders and electric motor are active

Three cylinders active

Figure 29.90 Interaction between combustion engine and electric motor [29-77].

Speed in rpm

range. All six cylinders are only switched on once the higher load range is reached at high speeds. When the combustion engine is in the full throttle range, support from the electrical machine is possible across the entire speed range.

The hybrid drive system can also be used as a generator. In USA, a standard socket is available for this purpose.

29.9.2 Vehicle Design

Only a handful of vehicles such as the Toyota Prius or Honda Insight have been designed to date exclusively for the hybrid drive system and manufactured with such a drive. The vehicle design of the other hybrid vehicles is mainly informed by standard drive concepts with gasoline and diesel engines. The hybrid drive system is in these cases simply “implanted” into an existing vehicle concept. Besides the hybrid drive’s electrical machine, it is particularly the energy storage device and power electronics that are relevant for the vehicle package. When selecting the installation location, any necessary coolant lines and power lines also need to be taken into account. Furthermore, the additional components alter the vehicle’s center of gravity and consequently its skid and crash behavior. Safety mechanisms need to be planned for the energy storage device and power cables under high voltage in the event of a crash. If the vehicle is also offered with other drive systems, that is, without a hybrid drive, then the additional components must also be positioned from the production point of view. Figure 29.91 15-kW electrical machine manufactured by Continental [29-78].

Honda also uses Ni–MH batteries as an energy storage device. The Chevrolet Silverado also uses a crankshaft starter generator. A 15-kW electrical machine manufactured by Continental is installed between the combustion engine and an automatic transmission. If the car is stationary and the brake pedal is also actuated, then the system shuts off the combustion engine. In addition, the electrical machine curbs the powertrain, which has meant that the automatic transmission’s torque converter could be downsized. The hybrid system works on a 42-V basis and uses a lead-acid battery. Figure 29.92 Toyota Prius [29-75].

Internal Combustion Engine Handbook | 1021

6606_Book.indb 1021

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

There is very little freedom in terms of the installation location of the electrical machines. Their location is determined with the selection of the hybrid concept. It is only in series hybrids that there is a certain degree of freedom with the arrangement of the electric motor, because it is not mechanically connected to the combustion engine. The crankshaft generator only has a small impact on the vehicle package. It is positioned directly between the combustion engine and the transmission. The transmission is modified in such a way that, with the electrical machine, it occupies the same space as the powertrain without hybrid drive. This is primarily achieved by downsizing the transmission’s torque converter and giving the electrical machine a very slender design. If the electrical machines are designed to be located in the engine compartment, then the combustion engine would need

Power electronics

to occupy less space accordingly. One advantage here is that a slightly smaller combustion engine is used in hybrid drive systems, with the result that appropriate freedoms are achieved. Unlike electrical machines, energy storage devices can be positioned relatively freely in the vehicle. In contemporary hybrid vehicles, the energy storage devices are usually installed in the area underneath the rear seat bench. For the GMC Sierra, see Figure 29.93, for the Lexus RX400h the energy storage devices are found directly underneath the seat bench, Figure 29.94, with the Honda Civic, Figure 29.97, in the area around the backrest, and in the Prius, Figure 29.92, they are installed behind the seat bench, right above the rear axle. One exception to this rule is the Mercedes S-Class, where the Li-ion batteries are located in the engine compartment. The energy storage device is very compact because of the selection of the hybrid

115V Power outlet

Energy Storage Device Electrical machine

Figure 29.93 GMC Sierra Hybrid with a crankshaft generator [29-79].

Figure 29.95 Li-ion batteries for the S-Class from Mercedes-Benz [29-80].

Figure 29.94 Battery of a RX400h under the rear seat bench [29-75].

1022 | Internal Combustion Engine Handbook

6606_Book.indb 1022

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.9 Current Hybrid Vehicles

concept and new battery technology (see Figure 29.95) and it is located in the engine compartment. In addition, the energy storage device is cooled by the air conditioning system, which is also located in the front end. Vehicles have already been announced for the coming years, which see installation of the battery in the underbody among other places. The third large component in a hybrid drive system is the power electronics, and until now they have either been located in the engine compartment or, as in the case of Honda, close to the battery. They also require cooling, meaning that in

addition to the power cables for the electrical machine and battery, as well as the data cable for the control device, there must also be space for the transport of a coolant. The power electronics must be positioned so that no undesired vehicle parts become live in the event of a crash. A similar case to batteries, the specific volume of the power electronics will improve in future. Whereas for batteries, this further development will presumably be used to increase the amount of energy that can be stored, and the aim of the power electronics is rather to decrease the installation space.

Figure 29.96 Two-Mode hybrid system in a Chevrolet Tahoe [29-81].

Figure 29.97 Honda Civic [29-77].

Internal Combustion Engine Handbook | 1023

6606_Book.indb 1023

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

29.10 Future Development Only a handful of vehicle manufacturer have released hybrid drive systems to date. This situation is likely to change dramatically in the next few years. All the famous vehicle manufacturers have showcased close-to-production prototypes at trade fairs and announced their market launch [29-82]. Consequently, this represents the first generation of hybrid drive systems for many vehicle manufacturers (not including development results prior to 1998 that are no longer applicable: Audi Duo, 3L Lupo, etc.). Further developments following on from this can be organized into the topic blocks: •• technical development •• customer expectations •• cost reductions. Technically speaking, all the hybrid drive assemblies display huge potential for optimization because of their early stage of development. Despite modern development and test methods, some of the flaws of the individual hybrid systems only become apparent when they are used on a large scale. With energy storage devices, we will see a medium term shift from Ni–MH rechargeable batteries to Li-ion rechargeable batteries, thus resulting in a greater electric range, less weight, and lower costs. Nevertheless, at today’s starting point, the electric range will not come close to that of hydrocarbonbased drive systems for a very long time to come. In future, it will also be possible to charge the energy storage device via an external power supply. These so-called plug-in hybrid vehicles can then be charged at a power outlet, thus enabling a local emission-free short-drive operation in combination with larger energy storage devices. The reduction in electrical losses is another development field, focusing on the electrical machines, power electronics, and energy management system. Although electric motors already boast very high levels of efficiency, specially adapted machines are still able to deliver single-digit percentage improvements. Further developments in energy management will control the energy flows more accurately and thus further reduce unnecessary charging and discharging operations to the energy storage devices. Furthermore, the proportion of recuperated braking energy will continue to grow in future hybrid drive systems. Variable, electrical auxiliary units will continue to prevail thanks to the high-voltage networks available. The power consumption of the individual devices will be increasingly demand-based. As mentioned in Section 29.3.2, the fuel consumption benefits of hybrid drive systems are very much dependent on the particular driving style, with the result that they can sometimes revert to a negative. In this instance, there will be a vehicle-typical alignment, irrespective of the legal driving cycle. Small vehicles are very efficient in short-range traffic, but use a lot of fuel comparatively during long journeys and at high vehicle speeds, whereas larger vehicles are optimized for long-range traffic. The consideration of typical use according to vehicle class is consequently gaining in importance.

Vehicle packaging will lead to standardizations for the individual drive types. The hybrid drive will no longer be seen as an “add-on,” but will enjoy equal status as a drive variant in future vehicle concepts. To some extent, individual functions and characteristics, which are taking off in hybrid drive systems will in future be integrated into conventional powertrains (start/stop function, electrification, etc.). Evaluating customer expectations of this drive system or deducing which potential functions have an additional benefit to the customer is difficult because hybrid drive systems are still not very widespread. What characteristics does the customer demand? Which examples is the customer prepared to pay for? These questions must be answered for future hybrid drive systems. Thus, the customer will decide the extent to which the combustion engine can be phlegmatized and, under some circumstances, only be operated at isolated load points. Technically speaking this does preclude such a solution. Presumably different solutions will prevail depending on the vehicle class. Costs are a factor, which cannot afford to be neglected by vehicle manufacturers. This does not just include the unit costs. Because of the quick introduction of this drive type, the new process flows and divisional structures could boast huge savings potential for vehicle manufacturers. The unit costs will involve the exact design of the individual components. An oversize energy storage device is not only expensive, but is also larger and heavier, which has an adverse effect on both the logistics during vehicle production and later in the vehicle because of the increased weight. Knowledge of the maximum useful charging/discharging stroke of the particular energy storage device is a critical cost factor in this context.

29.10.1 Gasoline Hybrid Drive System

The gasoline hybrid drive is the focus of drive system development in the short and medium term. A gasoline engine is much cheaper than its diesel counterpart and has a higher outright CO2 savings potential, since its CO2 emissions are greater than those of a diesel engine [29-83]. For these reasons, new hybrid drive technologies will initially enter series production with gasoline engines and only afterward in combination with cost-critical diesel engines. The reciprocity of both drive forms will continue to be optimized. In combination with gasoline engines, all variants from micro hybrid up to full hybrid will be available in the short-term.

29.10.2 Diesel Hybrid Drive System

The diesel hybrid drive is the combination of two cost-intensive technologies, the diesel engine and the hybrid drive system [29-84]. At the same time, however, when compared to gasoline hybrids it represents a lower CO2 emitting drive solution, meaning that this technology should not be overlooked. In sales markets that already have a high percentage of diesel passenger cars, such as in Europe, mild and full hybrid diesel hybrid drives will already exist in the medium term, whereas at first, only gasoline hybrid vehicles will be seen in the other markets. Lawmakers will play a pivotal role in

1024 | Internal Combustion Engine Handbook

6606_Book.indb 1024

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.10 Future Development

this. To what extent will lawmakers influence the weighting of exhaust emissions and fuel consumption? Local regulations in built-up areas, which may enforce certain technologies, can bring about quick changes here. It is particularly the case for delivery vehicles, on purely economic grounds, that relevant legal ordnances could propel the diesel hybrid system to a speedy breakthrough. As micro hybrid systems are not so cost critical, these will most likely make the breakthrough in the short term.

29-6. Norbye, J.P. and Dunne, J. “… and a Commuter Car with Hybrid Drive,” Popular Science, 1969. 29-7. Rixmann, W. “Toyo Kogyo zeigte neue Verwendung des Kreiskolbenmotors Bauart NSU-Wankel,” ATZ, December 1970. 29-8. Christian, M. “Verringerte Schadstoffemission durch Hybridantrieb,” ATZ, January 1974. 29-9. Yaegashi, T. “The History of Hybrid Technology,” AutoTechnology, Vieweg Verlag/GWV Fachverlags GmbH, Wiesbaden, 2005. 29-10. VW. “Volkswagen-Taxi mit Hybridantrieb,” 1977, Wolfsburg. 29-11. www.audi.de: “Audi Q7 hybrid concept 15 Jahre Pionierarbeit,” 2007.

29.10.3 Electric-Only Drive

Electric-only drives do not count as hybrid drive systems. But because hybrid drive systems are often seen as a bridging technology to electric vehicles, particularly to those with a fuel cell, we will turn our attention to them briefly here. The short range is the major downside of an electric drive with an electrical energy storage device. Batteries and rechargeable batteries are unable to deliver sufficient energy density in the foreseeable future to achieve the sorts of ranges motorists have come to expect. Advancements in battery technology are certainly within sight but the projected power values are well below those of liquid fuels. One approach for enabling longer distances is to generate part of the electricity in the vehicle, for example, with a combustion engine in conjunction with a generator or a hydrogen-operated fuel cell. This concept is designated as an “electric drive with range extender” and can also be classified as a series hybrid drive. Regardless of this, a market for electric vehicles may yet grow in built-up areas because of increased local emissions requirements and reduced range demands. The use of “city cars” in conjunction with innovative opportunities for battery charging, for example, via an electrical connection in the parking lot, is conceivable here, thus leading to a serious reduction in the vehicle range required by the customer. However, it should not be forgotten that the emissions are simply being transferred and are not necessarily being reduced with an electric drive vehicle. The electric drive system is only as clean as the power plant that produces its electricity. It is particularly with respect to the use of fuel cells and by implication hydrogen as a fuel that locally emission-free is often also mistaken for globally emission-free. Like a battery, hydrogen is only an energy source.

Bibliography

29-1. “Electric Vehicle-Technology and Expectations in the Automobile Age,” The Johns Hopkins University Press, Baltimore and London, 2004. 29-2. Pieper, H. “Regelungsvorrichtung für mit Dynamomaschinen gekuppelte Explosionskraftmaschinen.” Patent specification no. 21202, Imperial-Royal Austrian Patent Office 1905. 29-3. Dr. Ing. h.c. F. Porsche AG, historical archive, 2007. 29-4. Rajashekara, K. “History of electric vehicles in General Motors,” Industry Applications Society Annual Meeting, 1993, Conference Record of the 1993 IEEE. 29-5. Car Craft: “An electric Car that makes its own electricity,” August 1969.

29-12. Press information from 08.11.1993: “Abschluß des VolkswagenHybrid-Versuchs in Zürich,” VW, Wolfsburg, 1993. 29-13. UNECE: “Agreement concerning the Adoption of Uniform Technical Prescriptions (Rev.2),” Regulation No. 85 Amendment 4, 2005. 29-14. Voß, B., Mehler, O., and Lintz, S. “Development of Hybrid Electric Vehicles for Mass Production Part 2: Electrical Components, Auxiliaries, Brake System, Control,” ATZ Volume 109 Vieweg Verlag, 2006. 29-15. Kolbenschmidt Pierburg Gruppe: “Oil pumps for combustion engines—conventional and variable,” Neuss, 2005. 29-16. Wandt, H.-P. “Energiemanagement und Regelungsstrategien bei Hybridfahrzeugen, in Energiemanagement und Bordnetze,” Expert Verlag, Renningen, 2004. 29-17. Bady, R. and Biermann, J.-W. “Hybrid-Elektrofahrzeuge—Strukturen und zukünftige Entwicklungen,” 6th Symposium “Elektrische Straßenfahrzeuge,” Technische Akademie Esslingen, 2000. 29-18. Hohenberg, G. and Spurk, P. “Comparing the hybrid propulsion to the conventional drive,” CTI Hybrid Drivetrains and Transmissions, Germany, 2007. 29-19. Brechmann, G., Dzieia, W., Hörnemann, E., Hübscher, H., Jagla, D., and Klaue J. “Elektrotechnik—Tabellen, Energieelektronik, Industrieelektronik,” Westermann, 3. Auflage, Braunschweig, 1994. 29-20. Paul, M., Hofmann, W., and Frei, B. “Schleppverluste bei permanenterregten Synchronmaschinen und deren Reduzierung,” EMA 2008, Aschaffenburg, 2008. 29-21. Böckl, M. and Rius-Sambeat, B. “Auslegung des elektrischen Antriebs und Auswirkungen auf die Betriebsstrategie bei einem Parallelhybrid,” “Hybridantrieb—die Zukunft des Automobilantriebs?,” Berlin, 2005. 29-22. Soon-O Kwon, Jeong-Jong Lee, Geun-Ho Lee, and Jung-Pyo Hong: “Torque Ripple Reduction Control of Permanent Magnet Synchronous Motor for Electric Power Steering Using Harmonic Current at Loaded Conditions,” EVS 24 Towards Zero Emission, Norway, 2009. 29-23. Brauer, M., Brendel, B., and Holl, E. “ELFA®—Innovative Serienhybridantriebe für Citybusse in Solo- und Gelenkbusausführung,” EMA 2008, Aschaffenburg, 2008. 29-24. Gröter, H.-P. “Weiterentwicklung bei Hybridantrieben,” Hybrid- und Brennstoffzellen-Elektrofahrzeuge: Energiemanagement-Aufgaben und Strukturen, DGES, Ingolstadt, 2005. 29-25. Wallentowitz, H. and Gnörich, B.: “Entwicklungstrends in der KFZ-Antriebstechnik,” Hybrid- und Brennstoffzellen-Elektrofahrzeuge: Energiemanagement-Aufgaben und Strukturen, DGES, Ingolstadt. 29-26. Mathoy, A. “Die Entwicklung bei Batterien und Antriebstechnik für Elektromobile,” Bulletin SEV/VSE 1/2008. 29-27. Schüttler, J., Werner, U., Vinogradski, M., and Orlik, B. Stromregelung einer zweisträngigen Transversalflussmaschine in Sammlerbauweise, DFMRS 2004, Bremen. 29-28. Press Kit “MICHELIN ACTIVE WHEEL,” 2008 Paris Motor Show, Paris, 2008. 29-29. Ulrich, K. “Elektromotor-Getriebe-Kombination für kompakte PKW,” E-Motive Elektrifizierter Fahrzeugantriebsstrang, Hannover, 2008. 29-30. Schröder, D. “Leistungselektronische Schaltungen—Funktion, Auslegung und Anwendung,” 2. Auflage, Springer, Berlin, 2008.

Internal Combustion Engine Handbook | 1025

6606_Book.indb 1025

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 29 Hybrid Drive Systems

29-31. SAFT: Source for energy density. 29-32. www.aral.de. 29-33. Brunner, T., “BMW CleanEnergy—Fuel Systems, Liquid Hydrogen Vehicle Storage,” ZEV Technology Symposium, Sacramento, 2006. 29-34. Friedrich, J. “Anforderungen an elektrische Antriebssysteme für Kraftfahrzeuge und resultierende Anwendungsspektren,” E-MOTIVE Elektrifizierter Fahrzeugantriebsstrang, Hannover, 2008. 29-35. Habib, A. “High-Power Electrochemical Storage Devices and Plug-In Hybrid Electric Vehicle Battery Development,” Presentation from the U.S. DOE Office of Vehicle Technologies “Mega Merit Review,” US advanced Battery Consortium, Bethesda, Maryland, 2008. 29-36. Hauck, B. “Elektronische Überwachungs- und Steuergeräte zum Erhalt der aktuellen Qualität vielzelliger elektrochemischer Speichersysteme,” Kaiserslautern, 2003. 29-37. Sauer, D.U. “Optionen zur Speicherung elektrischer Energie in Energieversorgungssystemen mit regenerativer Stromerzeugung,” Solarzeitalter-Politik, Kultur und Ökonomie Erneuerbarer Energien, 4/2006. 29-38. Barsacq, F., Liska, J.-L., and Genin, P. “High Power Lithium-ion technology for full hybrid automotive applications,” Tagung Hybridantrieb—die Zukunft des Automobilantriebs?, Expert Verlag, Germany, 2005. 29-39. Steiger, W. “Der elektrifizierte Fahrzeugantrieb-Chancen und Herausforderungen aus Sicht eines Automobilkonzerns,” E-Motive Elektrifizierter Fahrzeugantriebsstrang, Hannover, 2008. 29-40. Köhler, U. and Liska, J.-L.: “Status and Trends of Li-ion Battery technology for automotive applications,” 7th International ATA Conference on engines, Sardinia, 2008. 29-41. Truckenbrodt, A., Mohrdieck, C., and Noreikat, K.E.: “Plug-in hybridspromise, hype or the solution?,” Tag des Hybrids, Aachen, 2007. 29-42. Kötz, R. “Doppelschichtkondensatoren—Technik, Kosten, Perspektiven,” Kasseler Symposium Energie Systemtechnik, Kassel 2002. 29-43. Czarkowski, U. “Modular Battery Management System for Lithium Ion Batteries for Plug-in Hybrid Vehicles,” 5th Symposium Hybrid Vehicles and Energy Management, Braunschweig, GZVB, 2008. 29-44. Kümpers, J. and Schmitz, C. “Nickel-metallhydrid-Batterien für Hybridfahrzeuganwendungen,” Tagung Hybridantrieb—die Zukunft des Automobilantriebs?, Expert Verlag, Germany, 2005. 29-45. Hackenberg, U. “Optionen zur Reduzierung von CO2-Emissionen,” ATZ/MTZ Konferenz—Energie CO2—Die Herausforderung für unsere Zukunft, Munich, 2008. 29-46. Ullrich, M. “Energiespeichersysteme für Hybrid- und Brennstoffzellenfahrzeuge,” DGES-Fachtagung Hybrid- und Brennstoffzellen-Elektrofahrzeuge: Energiemanagement-Aufgaben und Strukturen, Ingolstadt, 2005. 29-47. Sauer, D. U. Wo die Batterietechnik steht, Mobility 2.0, Ausgabe 01.2011. 29-48. Toyota Motor Corporation, 1995–2015. 29-49. Möller, F. and Vocht, W. “DualDrive® E-CVT Getriebe- Ein innovatives E-CVT Getriebe—nicht nur für den Einsatz im Hybrid-Fahrzeug,” 5th International CTI Symposium Innovative Vehicle Transmissions, Berlin, 2006.

29-53. Schöttel, M. “VW revolutioniert Hybridantrieb,” AutomobilProduktion, December 2006. 29-54. Lindemann, M., Freimann, R., and Cebulski, B. “Konzept eines aktiv momentenverteilenden Achsdifferentials mit Hybridfunktionalität” in H. Schäfer (Ed.), Neue elektrische Antriebskonzepte für Hybridfahrzeuge, Expert Verlag, Renningen, 2007. 29-55. Schöllmann, M., Olk, J., and Rosenmayr, M. “Trends bei der Batterieüberwachung,” Hybrid- und Brennstoffzellen- Elektrofahrzeuge: Energiemanagement-Aufgaben und Strukturen mit Sensoren, DGES, Ingolstadt, 2005. 29-56. Hofmann, L. “Antriebs- und Steuerungskonzept des VW Golf TDI Hybrid,” DGES Hybrid- und Brennstoffzellen-Elektrofahrzeuge: Energiemanagement-Aufgaben und Strukturen, Ingolstadt, 2005. 29-57. Reiser, C., Zellbeck, H., Härtle, C., Reisemann, M., Klaiß, T., and Vroegop, S. “Incorporating customer’s driving behaviour in vehicle development,” The 8th International Stuttgart Symposium, Stuttgart, 2008. 29-58. Auer, F. “Von der Lichtmaschine zum Micro-Hybrid—Eine Einführung in Generator-Regler-ICs für Standard-Batterien, Start-Stop-Systeme und Micro-Hybrid-Antriebe in PKWs,” EMA 2008 Elektromobilausstellung, Aschaffenburg, 2008. 29-59. Hochkirchen, T. “How to Use Designed Experiments to Understand Real-World Usage Driven System Requirements,” Design of Experiments (DoE) in Engien Development III, expert Verlag, Renningen, 2007. 29-60. Ertl, C., Honeder, J., and Schinnerl, M. “Die Motorsteuerung des neuen 4-Zylinder Motors in der BMW 1er Serie,” 6. Steuerungssysteme für den Antriebsstrang von Kraftfahrzeugen, Berlin, 2007. 29-61. Breitling, T., Siegert, R., Steffens, D., and Baumgärtner, W. “Potenziale des Energiemanagement für den Realverbrauch,” Thermoelektrik—Eine Chance für die Automobilindustrie, Berlin, 2008. 29-62. Bäker, B., Kutter, S., and Morawietz, L. “Energiemanagement— vom 12 V Verbraucherbordnetz zum elektrischen Antriebsstrang,” ATZ/MTZ Konferenz Energie, CO2—Die Herausforderung für die Zukunft, Munich, 2007. 29-63. Scholz, N. and Kücükay, F. “Modular Simulation Environment for Structural Analysis of Hybrid Drives,” 4th Symposium Hybrid Vehicles and Energy Management, GZVB, Braunschweig, 2007. 29-64. Männel, R. and Reimann, W.: “Verbrauchseinfluss der elektrischen Energie,” ATZ/MTZ-Konferenz—Energie/CO2—die Herausforderung für unsere Zukunft, Munich, 2007. 29-65. Pischinger, R., Klell, M., and Sams, T. “Thermodynamik der Verbrennungskraftmaschine—Der Fahrzeugantrieb,” Zweite, überarbeitete Auflage, Springer Verlag, Vienna, 2002. 29-66. Böckl, M. and Rius-Sambeat, B. “Auslegung des elektrischen Antriebs und Auswirkungen auf die Betriebsstrategie bei einem Parallelhybrid,” The 1st: Hybridantrieb—die Zukunft des Automobilantriebs? conference, Berlin, 2005. 29-67. Bäker, B., Kutter, S., and Morawietz, L. “Energiemanagement—vom 12 V Verbraucherbordnetz zum elektrischen Antriebsstrang,” ATZ / MTZ-Konferenz—Energie/CO2—die Herausforderung für unsere Zukunft, Munich, 2007. 29-68. Wolter, T.-M. and Nasdal, R. “Integration of GALOP into HEV Control,” 5th Braunschweiger Symposium Hybridfahrzeuge und Energiemanagement, Braunschweig, 2008.

29-50. Killmann, G., “Toyota Hybrid Vehicles, Technology Evolution from 1997 to 2007,” “Hybrid Drivetrains and Transmissions,” 4th Symposium Hybrid Vehicles and Energy Management, Braunschweig, 2007.

29-69. Lindemann, M., Große-Siestrup, L., Link, M., and Neßler, A. “Einsatz von Optimierungsverfahren in der Hybrid-Antriebssimulation,” Hybridfahrzeuge, Ed. B. Voß, Expert Verlag, 2005.

29-51. Kaehler, B., Kersting, K., Brouwer, M., and Christ, T. “Entwicklungskriterien, Analysemethoden und Beurteilung von leistungsverzweigten Hybridgetrieben, am Beispiel eines Two-Mode Hybridantriebs,” 16th Aachen Colloquium Automobile and Engine Technology, Aachen, 2007.

29-70. Ploumen, S., Kok, D., Nessler, A., and Gühmann, C. “Anwendung in Modellbildung und Simulation—Applications in Modelling and Simulation,” Design of Experiments (DoE) in der Motorenentwicklung, Ed. K. Röpke, Expert Verlag, 2003.

29-52. Steiger, W., Böhm, T., and Schulze, B.-G. “Directhybrid—a Combination of Combustion Engine and Electric Transmission,” 15th Aachen Colloquium Automobile and Engine Technology, Aachen, 2006.

29-71. IAV GmbH. Ingenieurgesellschaft Auto und Verkehr, E. v. Engineer Society Automobil and Traffic, Berlin.

1026 | Internal Combustion Engine Handbook

6606_Book.indb 1026

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

29.10 Future Development

29-72. Chabot, L., Simon, O., and Kernen, V. “Start-Stop-Technologie— Funktionelle Auslegung der Systemarchitektur,” 14th Aachen Colloquium Automobile and Engine Technology, Aachen, 2005.

29-89. Killmann, G., et al. “Der Verbrennungsmotor im Toyota Hybridsystem,” 22nd International AVL Conference “Engine & Environment,” September 9–10, 2010, Graz.

29-73. Bosch press photo: 1-SG-14942.

29-90. Kirsten, K. “How much and what kind of variabilities in valvetrains of combustion engines are necessary in times of down-sizing, hybrids and range-extender?,” 11th Stuttgart International Symposium, Automotive and Engine Technology, February 22–23, 2011.

29-74. Killmann, G. “Toyota Hybrid Systems—Technology Evolution from 1997 to 2007,” Hybrid Vehicles and Energy Management, Braunschweig, 2007. 29-75. Schuermans, R. “Toyota Motor Europe Technology Evolution from 1997 to 2007,” CTI Hybrid Drivetrains and Transmissions, Berlin, 2007. 29-76. Fujikawa, M., Ito, M., Hattori, H., and Noda, K. “Toyota’s new hybrid transmission for RWD vehicles,” CTI Transmission Symposium, 2006. 29-77. Böttcher, J. “Transmissions for Hybrids Matching transmissions to mild Hybrid applications,” CTI Hybrid Drivetrains and Transmissions, Berlin, 2007.

29-91. Krüger, M., et al. “Betriebsstrategien eines dieselelektrischen Hybridfahrzeuges aus motorischer Sicht,” 19th Aachen Colloquium, October 4–6, 2010. 29-92. Wehlen, T., et al. “CO2-Reduzierung durch elektrifizierte Antriebsstränge,” TAE, 9th Symposium Ottomotorentechnik, December 2–3, 2010. 29-93. Brauchrowitz, E., Graf, H., Kessler, F., and Lichtenberger, M. The Hybrid Drive System in the BMW Active Hybrid 7, ATZ, 09/2010, Volume 112.

29-78. Warburg, N., Mailänder, E., Saatkamp, T., Reckziegel, C., and Stutz, M. “Lifecycle Assessment of the ISAD System,” ATZ 10/2002 Volume 104, Vieweg Verlag, 2002.

29-94. Klima, B., Huss, A., and Nöst, M. Integrated Methodology for Simulation and Measurement of the Diesel Hybrid Potential, ATZ 11/2010, Volume 112.

29-79. Kozlowski, F. and Spazierer, J. “Der ISAD-Starter-Generator: Eine Innovoation im Antriebsstrang zur Reduktion des Treibstoffverbrauchs geht in Serie,” Hybridfahrzeuge und Energiemanagement, Braunschweig, 2004.

29-95. Schöttle, M. Technik-Porträt Toyota Prius III, ATZ 111 (2009) No. 11.

29-80. Continental press photo: img_2006_02_28_genf_ hybrid_1_en. 29-81. Nitz, L. “Tahoe and GMC Yukon Hybrids—Integration von Fahrzeugund Powertrain-Technologien zur Verbesserung des Kraftstoffverbrauchs und der Fahrleistungen,” 16th Aachen Colloquium Automobile and Engine Technology, Aachen, 2007. 29-82. Olschewski, I. and Freialdenhoven, A. “Technologiepotentiale und Herausforderungen von Automobilstandorten vor dem Hintergrund des globalen Wettbewerbs,” 16th Aachen Colloquium Automobile and Engine Technology, Aachen, 2007. 29-83. Buschmann, G., Mayr, B., Link, M., and Knobel, C. “Hybrid— Konkurrenz oder Unterstützung für Verbrennungsmotoren?,” Hybridfahrzeuge, HdT 2005. 29-84. Buschmann, G., Nietschke, W., and von Essen, C. “Welchen Beitrag können alternative Kraftstoffe und die Hybridtechnik zur CO2-Absenkung leisten?,” 8th Symposium Entwicklungstendenzen bei Ottomotoren, Leipzig, 2006. 29-85. Atkins, A. “CO2-Reduzierung—Der mechanische Hybridantrieb von Ricardo,” 19th Aachen Colloquium, 4-6 October 2010. 29-86. Brandt, M., et al. “Downsizing und Hybridisierung: Konkurrierende Systeme oder die Kombination für zukünftige Antriebsstränge?,” 19th Aachen Colloquium, 4–6 October 2010. 29-87. Fischer, M., et al. “Konzeptvergleich hybrider Antriebe durch Simulation,” 22nd International AVL Conference “Engine & Environment,” September 9–10, 2010, Graz. 29-88. Fraidl, G., et al. “Herausforderungen und Lösungen für ottomotorische Range Extender—von Konzeptüberlegungen bis zu Praxiserfahrungen,” TAE, 9th Symposium Ottomotorentechnik, December 2–3, 2010.

29-96. Schneider, E., Müller, J., Leesch, M., and Resch, R. Synthesis of an Eight-speed Automatic Transmission for Hybrid Drives, ATZ 12/2010, Volume 112. 29-97. Wachtmeister, G., Höhn, B.-R., Wirth, C., Habersbrunner, G., and Ziegler, A. Concept for Hybrid Vehicles with Simplified Diesel Engines, ATZ 05/2010, Volume 112. 29-98. Mohr, M., Götz, M., Fellmann, M., and Brehmer, U. Hybridisation of Powertrains for Construction Machinery, Special Edition, ATZ offhighway, April 2010. 29-99. Yong-Seok Kim, et al. Sonata Hybrid—The First Full Hybrid Electric Vehicle from Hyundai, ATZ, Feb. 2011, Volume 113. 29-100. Atkins, A. and Feulner, P. The Ricardo Mechanical Hybrid Drive, MTZ 02/2011, Volume 72. 29-101. Maiwald, O., et al. Simulation Environment for the Analysis of Different Hybrid Powertrain Configurations, ATZ 01/2010, Volume 112. 29-102. Kim, S.-K., Park, J.-S., Lee, J.-S., and Lee, C.-W. Hyundai-Kia Develops Liquefied Petroleum Gas Hybrid Powertrain, MTZ 02.2010, Wiesbaden, February 2010. 29-103. Morris, G., Criddle, M., Dowsett, M., and Quinn, R. Concept for Low-Cost Low-Voltage Hybridisation, MTZ 09.2010, Wiesbaden, September 2010. 29-104. Passerini S., et al. “The Electrification of the Powertrain with Lithium Ion Technology,” 32nd International Vienna Motor Symposium, May 5–6, 2011. 29-105. Klausner M., et al. “Technical Challenges for Automotive LithiumIon Batteries and Possible Solutions,” 32nd International Vienna Motor Symposium, May 5–6, 2011.

Internal Combustion Engine Handbook | 1027

6606_Book.indb 1027

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 1028

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30 Alternative Vehicle Drives and APUs (Auxiliary Power Units) 30.1 Reasons Behind Alternatives With only a few exceptions, modern vehicles are powered by gasoline or diesel engines and the appurtenant fuels. In a few countries, ethanol is mixed with gasoline or alternatively the gasoline engine is run on ethanol. Rape methyl ester (RME) is available for diesel vehicles, but there is insufficient quantity. The exceptions take the form of vehicles fueled with CNG (compressed natural gas) or LPG (liquefied petroleum gas). Hybrid drive systems, a handful of electric vehicles and the particular “plug-in” version of the hybrid drive system, as well as fuel-cell vehicles are being used to a greater extent in countries, such as China, Germany, and the USA. Future forecasts from Mercer [30-1] estimate that in 2011 up to 4.5 million vehicles, roughly 6.7% of world production, will have alternative drive systems. Gasoline engine fuel currently occupies a world market share of somewhat less than 90% and diesel more than 10%. Due to the rising fuel costs and high efficiency of diesel engines, as well as development progress in terms of reducing particulate and NOx emissions, this drive type is also of particular interest. The rationales for alternative energy sources either can be found in local circumstances or are orientated around the availability of independent energy sources in a particular country or region. In the past, sufficient crude oil was available as the original energy form for gasoline and diesel fuel, but the search for alternatives has expanded enormously in recent years. The essential reasons for this are as follows: •• Many states are attempting to make their energy use less dependent on the dictates of the oil-producing countries. In

addition, the production of crude oil is becoming increasingly more complex and costly. •• Carbon dioxide emissions need to be reduced significantly. Part of the blame for global warming of the atmosphere is attributed to this carbon dioxide. Carbon dioxide is produced when carbon-containing energy sources are burned. Discussions among lawmakers are particularly intense in Europe. Thus, 120 g CO2/km is considered the target quantity of average emissions for all newly registered passenger cars in the EU this year. There are numerous subitems to this regulation, which were decided in the EU Parliament at the end of 2008. A diesel share of 40% would equate to a fuel consumption of 4.8 l/100 km. Further tightening of these laws is expected. This value is expected to be 95 g CO2/km in 2020. •• Regionally, there is a desire for totally emission-free vehicles. The factors mentioned previously must be taken into account when evaluating the use of alternative energies for vehicle drive systems. Figure 30.1 provides an overview of the arguments for and against the various energy sources. A comprehensive approach was outlined in [30-2] and [30-3]. Sunfuel (biomass-based fuel) and Synfuel (natural gasbased fuel) have been developed in cooperation between the automobile industry (Volkswagen and DaimlerChrysler) and the mineral oil industry (Shell). The first industrial applications have already begun. Figure 30.2 details a corresponding prognosis for the use of different fuels. The type of energy used is of critical importance to the assessment of vehicle drive systems. It is absolutely necessary to take into account the complete energy chain (well to

Internal Combustion Engine Handbook | 1029

6606_Book.indb 1029

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Suitability Availability Cost effectiveness Infrastructure CO HC NOx Particulates CO2

Gasoline (reference)

Diesel with DPF

SynFuel produced from natural gas with DPF

○

○

○

○

○

○

+

○

○

○

○

○

○

○

○

○

–

+

○/–1

○

○

–

–

○

○

+

+

Methanol (fossil fuel/renewable)

Ethanol (renewable)

RME with DPF

SunFuel produced from biomass with DPF

Hydrogen (renewable) FC/ICE

CNG

LPG

–

–

–

–

○

–

–

–

–

–

– –/– –

–//–/+3

–/+3

–/+3

–/+3

– –/– –

–

○/+2

–

–

–

–

○

○

○

○

○

○/+2

○

○

○

+

○

○

○

+

FC = Fuel cell ICE = Internal-combustion engine DPF = Diesel particulate filter 1 Distribution infrastructure / production infrastructure 2 With/without mineral oil tax relief up to 2020

○

○

○

○

+//++

○

○

○

○/–1 ○

○

–

–

–

○

○

○

++

++

++

– –/– –

– –/– – ++/+4 ++/+4 ++/○ 5 ++/+4 ++/++

3

With/without mineral oil tax exemption up until 2009 and agricultural subsidies From lubrication oil contamination 5 With performance-optimized adjustment 4

Figure 30.1 Evaluation of selected fuels for the period up to 2010 (Based on: Gasoline EU 4)—own data and Volkswagen AG research.

wheel). This includes the original recovery of crude energy, preparation and processing (refining), transportation, and consumption in the vehicle. It is absolutely essential that, in addition to the further development of the vehicle technology provoked by political decision makers, the framework for the available energies is also specified.

30.2 Hybrid Vehicles As a general definition, hybrid drive systems are vehicles with two different drive systems and two different energy storage devices. Appropriate electronic management systems make it possible to achieve a number of advantages over a conventional drive system. One such advantage is the reduction of fossil energy consumption, because the combustion engine is operated at or very close to its maximum efficiency working point, the acceleration is boosted by the electric motor, and/ or the electric motor is used when setting off. This means that downsizing the combustion engine is also possible. Partial recovery of braking energy helps improve overall efficiency.

Emissions can be kept low by the overall optimization of the driveline, close to zero in the electric mode. A further advantage, particularly in electric operating mode, is the low noise output. In contrast to the merely niche market when looking at electric vehicles, hybrid vehicles have, however, already achieved a sizable number of units. The design of hybrid vehicles can be classified as per Figure 30.3 [30-4]–[30-6]. The series, parallel, and mixed hybrid types are represented. Hybrid vehicles allow a series of new technologies in the overall vehicle, such as energy and drive management, exhaust gas optimization, heat management, and driving performance optimization. Multiple components are still in the development phase, including high-temperature electronics (silicon carbide), new transmissions, heat recovery, advanced lithium-ion batteries, etc. Similarly, a new crossover technology is emerging thanks to hybridization, which includes further synergies. Therefore, individual vehicle manufacturers [30-7] even at this point are no longer talking about alternative drive systems, rather of a specific further development of the gasoline and diesel engine according to the motto, “in future there will be a little bit of hybrid XE ” in every vehicle.”

Figure 30.2 Fuels and drive systems [30-2], [30-3].

1030 | Internal Combustion Engine Handbook

6606_Book.indb 1030

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.2 Hybrid Vehicles

Series hybrid

Parallel hybrid

Mixed hybrid

Transmission

Transmission

Transmission

Planetary gear

ICE

ICE

Energy from ICE is converted twice

Energy from ICE is relayed directly (via clutches) to the wheel

Battery

Electric motor

Electric motor K2

Battery

Battery Generator

Electric motor

K1

Electric motor

ICE Energy from ICE is relayed directly (via planetary gears) to the wheel

Parallel hybrids allow the vehicles to be driven by both drivelines simultaneously. Series hybrids invariably have an electric motor drive system, consequently electricity must be generated on board, or supplied externally. This can be achieved with the most diverse array of units, for example, with a gasoline or diesel engine, gas turbine, Stirling engine, or fuel cell. The advantage is that the unit that is generating electricity can be operated to the maximum possible levels of efficiency and the lowest possible levels of emissions. The power-split hybrid drive system represents another variety known as a mixed hybrid, even though it is not the farthestranging hybrid variety. Conventional transmissions (manual and automatic transmissions) are replaced here by a one-stage or multiple-stage planetary gear with direct coupling of at least two electrical machines. The extremely compact unit constitutes a continuously variable transmission (ECVT), which can be used to implement all hybrid functions, such as start/stop, regenerative braking, as well as partial or allelectric drive. Figure 30.4 [30-8] shows the concept-specific characteristics of hybrid drive systems. The reasons behind the use of alternative-drive vehicles are analyzed from a marketing perspective in [30-8]. The

Figure 30.3 Definition of hybrid drive systems [30-4].

political and legal framework parameters at first show various differences: Europe

• Fleet consumption reduction from 2012: 120 g CO2/ km taxation of high CO2 emissions in some cases by individual countries.

USA

• Fulfillment of the ZEV (zero emission vehicles) mandate from 2010: Mercedes-Benz requires 2% ZEVs (FC—fuel cell, EV—electrical vehicles) and 2% AT-PZEVs—advanced technology partial zero emission vehicle (hybrids) or 4% ZEVs (law) • Tightening of the CAFE (corporate average fuel economy) limit value to 30 mpg (around 8 l/100 km) up to 2010 (draft legislation with high probability) • Regulations for reducing the CO2 emissions of passenger cars and light utility vehicles in California or in the whole of the USA

Japan

• Vehicle-class-dependent CO2 tax (law) from 2010: (JAMA/KAMA (Japan/Korea Automobile Manufacture Association)—Proclamation: Promissory fulfillment of the limit values from 2005) • Consumption reduction to 140 g CO2/km from 2009 (JAMA/KAMA voluntary commitment) • State funding for alternative drive systems through tax benefits (hybrid: up to 73 % off vehicle acquisition tax, up to 75 % off annual vehicle tax) and direct incentives.

Series hybrid (full hybrid)

Parallel hybrid (full hybrid)

Starter/Generator (mild hybrid)

Electric-only axle drive

Combined CE elec. axle drive

Electrical machine rigidly coupled to CE

+ Starter/Generator function + Electric-only driving + Energy recovery with full power + Electricity supply at 0 km/h + Optimal energy management + Comfortable engine start even at v > 0 km/h – Heavy weight (triple nominal capacity installed) – Boost mode electric-only due to overload – High costs – Heavy weight – Elaborate packaging

+ Starter/Generator function + Boost mode + Energy recovery with both electric motors + Electric-only driving + Optimum energy management possible + Immediate CE start with v > 0 km/h + Relative small PM electric motors + Electricity supply even at 0 km/h – High costs – Heavy weight – Complicated design – Longer drive train – Power-limited energy recovery

+ Starter/Generator function + Boost mode (ISG) + Simple design + Low costs + 42 V-output possible + Also has belt drive (RSG) – CE cannot be decoupled – Limited energy recovery (engine drag torque) – No electric-only mode – Low-fuel consumption savings

Figure 30.4 Technical evaluation of hybrid drive systems according to [30-8].

Internal Combustion Engine Handbook | 1031

6606_Book.indb 1031

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

China

• Even now, a huge energy demand is looming in China, with the result that alternative fuels (not crude-oil based) and economical vehicle concepts (hybrid) are very important there.

But even traffic conditions are the basis for different technology requirements Europe

• Balanced mix of freeways (high speeds) and urban traffic • Dynamic driving profile

Hybrids Gasoline hybrid (possible) Diesel hybrid (possible) Diesel (preferential)

USA

• Significant proportion of highway journeys • Characteristically steady traffic flow in the moderate speed range

Gasoline hybrid or diesel Diesel engines would be ideal, yet their market share is only rising very slowly.

Japan

• Urban traffic with a high proportion of stop-andgo traffic • Higher speeds are rare • No diesel fuel for passenger cars • Short average distances

Hybrids! Gasoline hybrids are a logical consequence, also borne by a technological passion

China

• High proportion of urban traffic • Low domestic crude oil deposits • Higher speeds rare over longer distances

Alternative fuels Hybrids

Estimates regarding the market share of hybrid vehicles vary significantly. The study carried out by Mercer mentioned above estimates a figure of 1.1 million vehicles in 2011 with regard to passenger cars, SUVs, pickups, and vans. An optimistic study published in the United States [30-9] shows that around 50% of all new vehicles coming onto the US market in 2022 will be hybrids. This would represent at least 6 million vehicles a year. Regardless of estimates, it is clear that hybrid technology is on the rise. Market leaders, Toyota and Honda, are both on their second or third generation of hybrid vehicles. The Toyota Prius, RX 400 h, and other Toyota models are full hybrids, while the Honda Civic is a mild hybrid. As is illustrated in Figure 30.5, around a quarter of a million new hybrid vehicles were purchased in the first seven months of 2008. Toyota and Lexus alone sold around 430,000 globally in 2007, and approximately 1.675 million between 1997 and 2008 [30-10]. The Toyota Hybrid System THS II is the successor of the THS I (Prius), which first went into series production in December 1997 [30-11]. The sudden success of hybridization is attributable to the strategy change at Toyota. The Prius II has a significantly improved driving performance (mileage) compared to the Prius I. Therefore, the THS II was marketed along these lines. More than 200,000 THS I vehicles have been sold worldwide. The system design of the Prius II [30-12] can be recognized from Figure 30.6. Figure 30.7 shows a comparison of the characteristics of THS II compared to THS I.

250.000

200.000

150.000

100.000

50.000

0 Global

USA

Japan

Canada

UK

Netherlands

Figure 30.5 New hybrid vehicle registrations 01/07/2008 [30-9].

1032 | Internal Combustion Engine Handbook

6606_Book.indb 1032

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.2 Hybrid Vehicles

Hybrid battery

System design Power control device

High-voltage power circuit

Generator

Inverter

Engine

Motor

(power output one and a half times better)

Power distribution

Reduction gear unit

Wheel -

Electric path Mechanical path

Hybrid transfer

Engine

Electric motor

System

Battery

Figure 30.6 System design for the Toyota Prius II. THS II

THS I

Type

1.5-l gasoline engine (high expansion rate)

1.5-l gasoline engine (high expansion rate)

Maximum output (kW)

57/5000

53/4500

Maximum torque (Nm)

115/4200

115/4200

Type

AC sync

AC sync

Maximum output (kW)

50/1200–1540

33/1040–5600

Maximum torque (Nm)

400/0–1000

350/0–400

Maximum output (kW)/ Vehicle speed (km/h)

82/85 or higher

74/120 or higher

Maximum output (kW) at 85 km/h

82

65

Maximum torque (Nm)/ Vehicle speed (km/h)

478/22 or lower

421/11 or lower

Maximum torque at 22 km/h

478

378

Type

Nickel-metal hydride

Nickel-metal hydride

Maximum output (kW)

21

21

Figure 30.7 Comparison of characteristics of THS II compared to THS I.

30.2.1 Fuel Consumptions in the Official Test Mode Test mode

THS II Prius 2

THS Prius 1

10–15 M

35.5 (km/l)

2.82 l/100

31 (km/l)

3.23 l/100

US Combined

65.8 (mpg)

3.57 l/100

57.6 (mpg)

4.08 l/100

EEC

104 (g CO2/ km)

4.33 l/100

120 (g CO2/ km)

5.0 l/100

Worthy of particular mention are the improvements to the gasoline engine, power, and torque increases to the electric drive system at simultaneously higher voltages for THS II compared to THS I. The voltage was increased from 274 to 500 V. The AC synchropermanent motor displays optimal values in terms of weight and volume employed. The generator has also been improved, for example with a speed increase from 6500 to 10,000 rpm. The same can also be seen in the improvement of the nickel-metal hydride battery, whose power density increased from 880 to 1250 W. The fuel consumptions achieved

in the drive cycles in the official driving styles are displayed in the table, the different data are converted into l/100 km. The route to power and torque increases was rigorously pursued when the Prius 3 was unveiled in Detroit in 2009. The output of the gasoline engine is now 73 kW, the electric motor 60 kW, and there is a combined output of 100 kW. However, the fuel consumptions achieved during practical fuel consumption vary significantly from the official test. A comparison carried out by AutoMotorSport [30-13] elicited the following values:

30.2.2 Practical Fuel Consumption Standard Test Route: AMS Vehicle type

l/100 km

g CO2/km

Audi A3 1.6 FSI

6.6

158

Honda 1.3 IMA

5.4

130

Toyota Prius II

5.5

133

Volkswagen 1.9 l TDI

4.9

131

Internal Combustion Engine Handbook | 1033

6606_Book.indb 1033

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

1.3 liter i-DSI, i-VTEC

1.3 liter i-DSI engine

Nickel metal hydride (Ni-MH) battery

Motor assist DC brushless motor

Regeneration

MT

Power control unit (PCU)

The low acceleration of the Honda in comparison to the other vehicles should be considered. The Honda IMA concept works contrary to the Toyota Prius II as a mild hybrid [30-14] with a 15 kW electric motor, which is positioned between the engine and the transmission. The DC motor works at 158 V; the energy storage device used is a nickel-metal hydride battery with 158 V and 5.5 Ah. Figure 30.8 shows the general system design. Meanwhile the latest version of the Honda Insight is now on the market, with CO2 emissions of 101 g CO2/km. The ISAD system is also arranged between the engine and the transmission. Figure 30.9 [30-15] shows a particularly spacesaving solution for the Chevrolet Silverado. The powertrain extension is less than 5 mm. Some solutions favor a V-belt for connecting the starter/generator with the crankshaft.

Flexplate

Rotor

Figure 30.8 The Honda IMA concept [30-14].

in addition to the combustion engine and the arrangement of the power-split transmission with electric motor and generator familiar from the Toyota Prius, also has an additional electrical machine at the rear axle, which generates the all-wheel drive. When compared to the Toyota Prius II, the RX 400 h also has a higher system voltage in addition to the second drive motor. Furthermore, expanding the speed range has more than doubled the output of the front electric motor. 1. 2. 3. 4.

Generator Power-split Electric motor Two-stage motor reduction

Startor

1

2

3

4

Hydraulic converter

Transmission

Figure 30.9 Mild hybrid system for the Chevrolet Silverado [30-15]. See color section page 1107.

In the Lexus GS 450 h, Toyota has combined hybrid technology with rear-wheel drive. The GS 450 h is equipped with 250 kW (340 PS) of power and can go from zero to 100 km/h in under 6 seconds. Its average consumption is given at 7.9 l of premium per 100 km. The Lexus RX 400 h, Figure 30.10 [30-16],

Figure 30.10 Section through the hybrid transmission of the Lexus RX 400 h [30-17].

The capacity of the nickel-metal hydride battery has also been effectively doubled. Toyota introduced the Lexus LS 600h in 2007. The V8 engine and electric motor deliver a combined ≥ output of 330 kW (≥ 445 PS). The system in the LS 600 h is designed to provide the same power and torque performance as a 12-cylinder gasoline engine. The power achieved in the Lexus vehicles are possible in part thanks to the operation of the electric motors. Unlike the Prius, the electric motors in the Lexus work at up to twice the speed and start to reach their limits only at around 13,000 rpm. The current Prius has an operating voltage of slightly over 500 V, in the Lexus this figure is 650 V.

1034 | Internal Combustion Engine Handbook

6606_Book.indb 1034

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.2 Hybrid Vehicles

Citroen C3

Micro hybrid

BMW Mild hybrid

Chevrolet Silverado



Honda Civic IMA



Honda Insight

Full hybrid

Saturn

Cadillac Chevrolet Silverado



Chevrolet Tahoe



GMC Yukon



Ford Mercury Mariner



Lexus LS 600 h



Lexus GS 400 h



Lexus RX 400 h



Mazda Tribute



Mercedes S-Class (Lithium ion)



Nissan Altima



Toyota Estima



Toyota Prius

Passenger car homologous * Gasoline Diesel FEV hybrid vehicle benchmark (7 passenger cars, gasoline and diesel)

2



In Figure 30.12 [30-19], an assessment is given in terms of the reduction in CO2 emissions and additional system costs. These carry even less weight when, due to the hybridization, other characteristics that have a high value to the customer can be achieved more easily, for example, park assist. Because of ongoing discussions on the subject of CO2, but also due to the technical possibilities already described, vehicle manufacturers are currently working flat out to introduce new hybrid vehicles. Special mention must go to the consortium of BMW/DaimlerChrysler/GM [30-20] and that of Volkswagen, Audi, and Porsche. The diesel engine is also suitable for hybrid variants, as has already been demonstrated in various concept vehicles [30-4].

CO emissions [g/km]

The following list gives details of other hybrid vehicles currently in series production:

In the meantime, all vehicle manufacturers have announced hybrid vehicles and/or plug-in versions for series production, source www.hybrid-autos.info. Previous experience has shown how important battery technology is. The returns for the batteries used up till now are still far too high for larger quantities. Reference [30-18] compares CO2 emissions in the New European Driving Cycle. Figure 30.11 shows a comparison of the results from eight hybrid vehicles.

VEH weight class [kg]

* (German Federal Motor Transport Authority)

Figure 30.11 CO2 emissions in the NEFC [30-18].

CO2 emissions in the NEDC [g/km] Gasoline MPI CNG in future

Starter-generator, belt-driven Crankshaft starter generator in future

Mild hybrid

Diesel in future *)

*) With particulate filter, consumption reduction thanks to downsizing, friction reduction

Hybrid drive system with electric range (gasoline-based)

Cost increase

Figure 30.12 CO2 emissions and costs of future drive systems [30-19].

Internal Combustion Engine Handbook | 1035

6606_Book.indb 1035

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Figure 30.13 details the approach for developing standard components for the two-mode hybrid concept [30-20]. The two-mode hybrid system is a full hybrid, for which two electric motors or the combustion engine as well as both drives can be driven simultaneously depending on the driving situation. Furthermore, the two-mode concept can provide both a simple and a double power-split driving range. The cooperation provides an individual module development for different vehicle concepts.

Figure 30.13 Two-Mode concept GM/DaimlerChrysler/BMW.

The method described by the consortium is, according to DaimlerChrysler, superior to the previous solutions and is also protected by patents. As is visible from the detailed descriptions on the subject of hybrid concepts, there is scarcely an area that is not affected by hybridization. Figure 30.14 explains the impact of the individual assemblies and functions. Further cross-sectional tasks can be assigned to hybridization, such as the topics of acoustics and oscillations (NVH) and electromagnetic compatibility. We can assume that the hybridization of vehicles will press ahead in future, especially due to the fact that downsizing in particular will maintain or even surpass driving performances (mileages).

30.2.3 Plug-In Hybrid Vehicles

Plug-in vehicles are also hybrids (parallel or series). They are, however, also able to supply electrical energy externally and are therefore very flexible with battery costs that are still moderate (in comparison to an all-electric vehicle). Chevrolet unveiled the Volt study as early as 2007. A combination of a small combustion engine with an electric motor and a lithium-ion battery. It has been announced that this version will enter series production in 2009/2010. Volkswagen has developed another promising concept with the Twindrive. As per Figure 30.15, this concept also uses a combination of a diesel combustion engine (77 kW) with an electric motor (60 kW) and a lithium-ion battery [30-22]. Its fuel consumption is highly interesting as it equates to 2.5 l/100 km in the NEDC for diesel fuel at 8 kWh of electrical energy. Toyota has also developed plug-in hybrids and is carrying out corresponding trials in Japan, Europe, and the United States. It is the opinion of the authors of this book that plug-in concepts can be the solution in the medium term, since they also allow multiple options for energy generation. In 2008, FEV GmbH debuted [30-23] a 20-kW Wankel engine with a 75-kW electric motor, and other vehicle manufacturers such as Toyota and GM are also focusing on the plug-in drive system.

30.3 Electric Drive System Electric drive systems were, in fact, used at an extremely early stage in the development of motor vehicles. Because of the inadequate energy storage device (the battery), they have up to now, unlike gasoline and diesel engines, become established only in niche applications. The electrically powered vehicle itself can, of course, be “emission-free” in operation, but overall

Combustion engine

Chassis

Transmission

Body

Electrical machines

Braking and steering system

Energy storage device

Heating/Ventilation/Air con

Power electronics

Cockpit and HMI

Conduit system

Overall vehicle / Safety

Energy/Drive system management

Other subsections Figure 30.14 Vehicle areas affected by hybridization [30-21].

1036 | Internal Combustion Engine Handbook

6606_Book.indb 1036

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.3 Electric Drive System

Figure 30.15 Plug-in Twindrive from Volkswagen as a front-drive vehicle [30-22]. See color section page 1107.

assessment must depend on how the electricity is generated. The drive system of electric vehicles comprises [30-24]: •• Electric motor with electronic control system (inverter) and cooling system •• Traction battery including battery management system and the necessary charger •• Any necessary transmission, including differential gear •• System for the transmission of power to the drive wheels

•• Special motors •• DC brushless motor •• Transverse flux motors •• Switched reluctance motors. Figure 30.16 shows a comparison of different electric motors based on the parameters selected by the author. The switched reluctance motor appears to be of particular interest.

•• Electrical steering power and braking force assistance •• Heating and air-conditioning system •• Chargers (fixed or onboard). A range of different electric motors are available for the drive system. The selection criteria are as follows: low weight, high efficiency, compact design, low costs (manufacture and maintenance), and high torque across the broadest possible motor speed range. The possible variants can be seen in the following list: •• DC motors •• DC series motors •• Variable speed DC motors •• Three-phase motors •• Asynchronous motors •• Synchronous motors (a)  Permanently excited synchronous motors (b)  Separately excited synchronous motors

Efficiency

DCM

ASM

SESM

PESM

SRM

TFM

–/+

+

+

++

+

++

Maximum speed

––

++

+

++

++

––

Volume

–

+

+

++

+

+

Weight

–

+

+

++

–

+

Cooling system

––

+

+/–

++

++

+

Manufacturing costs

–

++

+/–

+

++

+/ –

Costs

+

++

+

+

++

–

DCM: DC machine; ASM: Asynchronous machine; SESM: Separately excited synchronous machine; PESM: Permanently excited synchronous machine; SRM: Switched reluctance machine; TFM: Transverse flux machine

Figure 30.16 A comparison of electric motors used in electric vehicles.

The front and/or rear wheels are frequently driven by the central electric motor for the purpose of transmission of the torque to the drive wheels; in individual cases, such as buses the electric motors, are, in fact, installed in the wheels. A single-stage transmission with a fixed transmission ratio is generally sufficient for the powertrain, as a result of the high torque generated by the electric motor and because the electric motors can be overloaded for short periods.

Internal Combustion Engine Handbook | 1037

6606_Book.indb 1037

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Battery type

Electric vehicle battery (high energy) Specific energy Wh/kg

Specific power W/kg

Energy density Wh/l

Power density W/l

Costs EUR/KWh

Lead-acid

 35

200

 90

510

100–150

Nickel-cadmium

 50

200

150

600

225–350

Nickel-metal hydride

 70

140

200

400

225–300

Lithium-ions

100

200

250

500

?

Specific energy Wh/kg

Specific power W/kg

Hybrid vehicles (high performance) Energy density Wh/l

Power density W/l

Costs EUR/KWh

Lead-acid

32

  430

 68

  910

100–150

Nickel-cadmium

35

  700

100

2,000

225–350

Nickel-metal hydride

40

1,200

100

3,000

225–300

Lithium-ions

70

2,000

150

4,200

?

Figure 30.17 Battery capacity data.

The central reason for the low numbers of electric vehicles in use is the limited performance and cost of the batteries. The traction battery is most important component of the electric drive. The vehicle’s range depends on the energy content. The available electrical output determines vehicle performance. Figure 30.17 [30-25] provides an overview of possible traction batteries. Questions regarding which battery will be used in the near feature always have the same answer: The Li-ion battery, which was introduced to series production in hybrid vehicles in 2008. Despite future widespread use and competition between battery manufacturers, costs will initially be in the higher price ranges. Even if the given target values are achieved individually, the energy storage devices can only used in narrow range of 50% of the actual capacity for durability reasons (NiMH, Li-ion). Furthermore, optimizing the energy density causes the power density to suffer and vice versa. It can be seen that the state of California’s demand for zero emission vehicles and the problems associated with increased CO2 output have driven battery developments, nickel-metal hydride and lithium-ion batteries in particular. The range of products available will most certainly increase as battery performance rises. The gap in terms of energy content between these batteries and gasoline/diesel fuel is, however, very high. When compared to lead-acid batteries, this factor is ≅350 and compared to lithium-ion batteries it is still ≅120. Lithiumion batteries also have high technical requirements and must first prove their power capability once they are subjected to day-to-day customer use. Figure 30.18 shows several versions of electric vehicles from the Norwegian company, Think, who are the market leaders in electric vehicles. The THINK CITY features the following data:

THINK CITY TECHNICAL DATA Seats

2+2

Dimensions

Length 3.12 m, width 1604 m, height 1548 m

Curb weight

1113 kg

Permissible total weight

1397 kg

Luggage space

350 l

Maximum loading

284 kg

Maximum speed

100 km/h

Acceleration

0–50 km/h in 6.5 s

Range

170 km

FRAMEWORK AND BODY Framework

Steel, galvanized

Chassis

Mc’Pherson Front, Trailing link rear axle

Body

Thermoplastic (polyethylene)

Roof

ABS plastic

BATTERIES AND MOTOR Batteries

Zebra Li-ion

Energy content

28.3 kWh

Charging device (internal)

220 V–16 A/10 A (32.2/2 kW)

Charging time

4–6 hours (80 % battery output), chargeable at any 220 V power outlet

Type of motor

3-phase asynchronous motor

Maximum motor output

30 kW

Voltage

114 V

Tires

165/65X14

1038 | Internal Combustion Engine Handbook

6606_Book.indb 1038

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.4 Energy Storage Devices

motors represents an interesting alternative Figure 30.19. Quick charging times and reducing battery costs are still important development goals at present. Other models are already available to buy commercially or are taking part in large field tests, transport for London are using the Smart ED, the range from the Indian company REVA, and the Electric Mini, which can be leased in California for US$ 1000. The technical data of the Electric Mini are as follows: asynchronous four-pole motor with 150 kW, maximum torque 220 Nm, lithium-ion battery with 35 kW/h, charging duration at 220 V = 8 h (110 V = 20 h) and a range of 250 km. Subaru have also been offering an electric version of the Stella model since July 2009. The future of electro traction is largely dependent on the battery performance and costs. These are currently estimated at US$ 20,000 for the Mini.

Figure 30.19 Wheel hub drive Mitsubishi [30-27].

30.4 Energy Storage Devices

Figure 30.18 Current and future electric vehicles from Think Global, including the Think City, Think Ox, and Think Open [30-26].

Mitsubishi unveiled a new concept vehicle in 2006, which has a range of 150 km using lithium-ion batteries (159 kg), and this figure can rise to 250 km with additional battery improvements. The vehicle, which is fitted with wheel hub

UltraCaps (super capacitors) are seen as potential energy storage devices, particularly in conjunction with hybrid vehicles and have even aroused interest in Formula 1. While flywheel energy storage devices are currently used mainly in larger vehicles (buses and trams) [30-28], [30-29], [30-30], the use of super capacitors (high-performance capacitors) is also conceivable for passenger cars. Numerous potential applications exist, due to the fact that super capacitors are able to store power and then deliver it again very quickly. These uses range from catalytic converter preheating, initial movement boosting, and power storage in electric and hybrid vehicles up to and including use in potential regenerative braking [30-31]. Figure 30.20 shows a comparative technological assessment of the performance of a number of storage devices.

Internal Combustion Engine Handbook | 1039

6606_Book.indb 1039

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

1 hour

1000

10 min

20 sec

2 sec

Batteries 100

Energy Density [Wh L–1]

2 ms 10

Supercapacitors 1

Capacitors

0.1

0.01 0.01

0.1

1

10 100 Power Density [kW L–1]

30.5 Stirling Engine The Stirling engine (invented as early as 1918) has again and again been discussed as a potential power unit for motor vehicles. It functions on the basis of continuous external combustion or heat input. This thermal energy is transferred via a heat exchanger to the working gas in the cylinder. Using a displacer piston, the gas is transferred backward and forward between a chamber with a constant high temperature and a chamber with a constant low temperature. Internal pressure changes periodically as a result. The pressure changes are converted to mechanical energy by a working piston and a corresponding crankshaft drive. As shown in [30-32], the process’s theoretical cycle (closed cyclical process with continuous heat input) can be described using two isotherms and two isochores. Figure 30.21 shows the theoretical cyclical process of the Stirling engine as a p-V and T-s diagram. In the engine process, the cycle is implemented on a clockwise basis, and in the refrigeration set or heat pump, on a counterclockwise basis. The individual elements in the theoretical cyclical process are as follows: 1 to 2: Isothermal compression; following adiabatic compression, the working gas is recooled to its initial temperature in

Isochore

T

p

4

2 to 3: Isochoric heat absorption; absorption of heat in a regenerator; 3 to 4: Isothermal expansion; after adiabatic expansion, the working gas is reheated in the heater to its initial state, input of heat from an external continuous combustion process being necessary:   Useful work is yielded in this subcycle; 4 to 1: Isochoric heat removal; removal of heat in the regenerator. The ideal process described can be achieved only if the working and displacer pistons move discontinuously. The efficiency of the ideal process equates to the Carnot efficiency. The Carnot efficiency can form the basis for the assessment of the efficiency of combustion engines:

η = 1 − T1/T3 = 1 − Tmin/Tmax The most important advantages of Stirling engines are low emissions, the ability to use any suitable heat sources (which can be generated using a range of different energy sources), an extremely good efficiency at the optimum working point (even in the partial-load operating range, given regulation of

Isochore 4

3

Actual process

Actual process 2

1 V

Figure 30.20 Technological assessment of energy storage devices.

a radiator, the heat being emitted to the environment or to a fluid requiring heating;

Ideal process

2

10000

Isotherme

Isotherme

3

1000

1

Ideal process

s

Figure 30.21 Stirling process in p-V and T-s diagram.

1040 | Internal Combustion Engine Handbook

6606_Book.indb 1040

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.6 Gas Turbine

swept volume), low vibrations, and low noise; Figure 30.22 shows a design provided by the STM company.

(because of the size of the heat exchangers), and high production costs, with the result that, unlike the situation in stationary machines, the Stirling engine has up to now failed to become established in mobile applications.

30.6 Gas Turbine High-temperature gas turbines can be operated using an extremely large range of different fuels, that is, different forms of energy. In some cases, chemical conversion processes that would otherwise be necessary for use in conventional engines can actually be omitted. Gas turbines powering motor vehicles have, for example, been operated directly using pulverized coal. Advantages for the overall efficiency of the conversion chain from the primary energy source up to the vehicle drive system derive from this. The design of a motor vehicle gas turbine derives from the special requirements of automobile operation. Figure 30.23

Figure 30.22 Stirling engine (25 kW) manufactured by STM.

The disadvantages are as follows: Poorer throttle reaction, higher load regulation complexity, greater space requirements Air inlet

15 °C; 1 bar Radial compressor

Combustion chamber

4.25 bar, 204 °C 4,25 bar 1040 °C

4,45 bar 660 °C

Compressor 2,2 bar 863 °C

Heat exchanger Work turbine

1.04 bar 725 °C

1 bar 280 °C

Exhaust gases

Exhaust gases

Output to the transmission

Figure 30.23 Design of a twin-shaft vehicle gas turbine.

Internal Combustion Engine Handbook | 1041

6606_Book.indb 1041

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

•• Three-shaft turbine. The twin-shaft gas turbine constitutes a good compromise between complexity and performance. The torque curve, for example, is significantly better than in the case of the single-shaft machine; load regulation is accomplished via regulation of working gas temperature and/or via adjustable guide vanes on the turbine and compressor. Despite the emissions advantages, multifuel capability, lower vibrations, and relatively good torque curve, excessively restricted suitability for lower power ranges, nonetheless, derives from the higher fuel consumption, the high-temperature-proof ceramic systems necessary, and the high level of noise generation. Poorer throttle reaction compared to the reciprocating piston engine has also resulted in the lack of use of gas turbines as a direct drive system in series-produced motor vehicles.

30.7 The Fuel Cell as a Vehicle Drive System If the hybrid concepts, including plug-in versions, described in Chapter 30.2 are seen as an extension of the gasoline or diesel engine drive system, then the fuel cell and all-electric vehicles are the only real alternatives. The electric vehicles, which are only fed from the battery, have as yet been unable to break onto the market due to the low storage capability of the battery and long charging times. The fuel cell as an onboard energy producer, which in turn stores the generated energy in a battery or uses it to drive an electric motor, shows significantly more potential. The motivation for the development of the fuel cell vehicle is based on the following factors: •• Zero emissions in drive mode (tank-to-wheel = TTW)

•• High level of comfort •• Freedom from crude oil dependency, for example, hydrogen can be produced from various different sources. This also applies to the combustion engine, if the fuel is not produced from crude oil. •• Generates electricity for the ever-increasing number of electrical consumers in the vehicle. A series of advantages can be identified from this list, however, fundamental further development steps are needed in subsections, and hydrogen production, distribution, and storage in the vehicle is gaining growing significance. In the most recent years of development, it has emerged that the method of generating hydrogen from gasoline or diesel fuel onboard with the use of a reformer is not expedient. The same applies to the direct methanol fuel cell. Research and development work for the fuel cell as a vehicle drive system has been carried out intensively for more than 15 years. Virtually all major vehicle manufacturers as well as numerous research institutes and automotive suppliers are working on the fuel cell drive system, batteries, and hydrogen storage technology [30-34]. Although a number of different fuel cell types exist, the PEM = polymer electrolyte membrane is considered particularly promising for use in mobile systems.

30.7.1 Design of the PEM Fuel Cell

Figure 30.24 [30-35], [30-36] shows the functional principle of the PEM fuel cell. The cell consists of the electrolyte membrane,

te

as

W

at

he

O2

H2 H2O H2

Waste heat

•• Twin-shaft turbine (gas generator shaft and drive shaft are separated)

•• Good acceleration flexibility thanks to the electric motor

Electrolyte

•• Single-shaft turbine (gas generator set and work turbine on a single shaft)

•• Very high efficiency TTW, around twice as high as a gasoline engine

Waste heat

[30-33] shows the principle of this in schematic form. The twin-shaft design, in which a separate work turbine is installed downstream of the gas generator set consisting of the compressor and compressor turbine, produces the elevation of torque needed for a vehicle to set off. Adjustable guide vanes upstream from the work turbine are also indicated in this illustration. During operation, positioning of these vanes can be used to modify the passage cross section and, thus, vary mass flow in such a way that the required output is achieved at maximum permissible turbine inlet temperature. Thus, delivering minimum fuel consumption. Reaction time at acceleration can be shortened by opening the vane cross section briefly, while, in coasting operation, the gas flow can be directed onto the rotor blades by turning the guide vanes in the opposite direction so that a braking torque is generated. Besides the magnitude of operating temperatures, the heat exchanger is the most important element in achieving good fuel consumption. The common types of gas turbine (open design) that are conceivable for use in motor vehicles differ in the number of shafts and stages used. The following are possible:

Anode

O2 Cathode

Electrical output

Electrical load

Figure 30.24 Concept representation of a fuel cell.

1042 | Internal Combustion Engine Handbook

6606_Book.indb 1042

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.7 The Fuel Cell as a Vehicle Drive System

the catalysts applied to it by coating and electrodes, as well as the bipolar plates, which are responsible for the supply of gases and takeoff of the electricity. The membrane is an approximately 0.1-mm-thick film made from a sulfonated fluorocopolymer, which separates the oxygen and hydrogen reaction gases and permits diffusion only of the protons. These are the result of the oxidation of hydrogen at the anode. On the cathode side, oxygen from the air evaporates to form steam with the protons that have migrated through the electrolyte. The difference in potential can be converted into electrical work in an external circuit. The individual cells are now arranged repeatedly one after the other and consequently form fuel cell stacks of around 25 kW. Multiple stacks are readily capable of delivering 150 kW or more. The design of the PEM fuel-cell stack can be seen in Figure 30.25 [30-34]. It is very importance to dispose of water resulting from the reaction of hydrogen and oxygen. The efficiency of the fuel cell stack is dependent on various factors, particularly the properties of the membrane and catalytic converter. In 2003, the power-to-weight ratio was 1.3 kW/kg

and the power-to-volume ratio was 1.7 kW/l. In addition to the subject of water management already mentioned, heat removal and oxygen supply by the corresponding compressors are extremely important. At the outset, fuel-cell systems were not capable of cold starts and even at temperatures above 0°C, the starting time lasted several minutes. By contrast, it is now possible to start operation at −25°C in a few seconds.

30.7.2 The Fuel Cell in the Vehicle

Figure 30.26 shows the placement in a front-wheel drive A-Class passenger car from Daimler AG. Space requirements are also fundamentally influenced by the cooling, hydrogen storage system, and the battery. An overview of the individual components in Figure 30.27 reveals the complexity that needs to be conquered for implementation in the vehicle [30-36]. If we consider the line-up shown in Figure 30.28 of the passenger car and utility vehicle drive systems from Daimler AG, then we get a glimpse of the monumental efforts involved

Brennstoffzellen-Stack mit Endplatten und Verbindungen Bipolar-Element mit Fließkanälen für Wasserstoff

Sauerstoff

Wasserstoffversorgung (Anode)

Wasserstoff

Technische Daten (Volllastbetrieb) Betriebstemperatur

85 °C

Betriebsdruck

2.5 bar

Stromdichte

1.5 A/cm2

Zellspannung

0,6 V

Power density Stack-Modul

1000 W/kg bzw. 1300 W/l

Einzelne Brennstoffzelle Polymer-ElektrolytMembran (PEM)

Figure 30.25 Design of a PEM fuel cell stack [30-34].

FC stack

Radiator

Electric drive system / Transmission

Sauerstoffversorgung (Kathode)

Bipolar-Element mit Fließkanälen für Sauerstoff

Compressed hydrogen tank

Battery

Fuel cell system

Power management system

Figure 30.26 Components of a fuel-cell drive system.

Internal Combustion Engine Handbook | 1043

6606_Book.indb 1043

1/19/16 9:01 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Vehicle control

System controller

Data via CAN bus Hydrogen recirculation

Hydrogen tank

H2

Luft

Pressure regulation

Air filter

DC

Fuel cell

12 V

Direct current

Air compressor

12 V Battery

DC/DC Battery

Air + water

Coolant

Coolant

Temperature control valve

Vehicle cooling

Four phases of commercialization for fuel cell vehicles “Fit for Daily Use”

Field tests

“Ramp-Up”

Cold-start capability Increased reliability Simplified system design

Since 2002

Figure 30.27 Fuel-cell drive system using hydrogen as the fuel [30-36].

Commercialization

Cost efficient Technology and materials

around 2015

in moving from the laboratory stage into preproduction and customer trials. The route that Daimler AG has taken/are still planning highlights the difficulties involved in introducing new technologies (Figure 30.29). In addition to intensive test drives (the first vehicles were tested in 1994) with various models (bus, small utility vehicles, and passenger cars) improvements to the system itself need to be carried out simultaneously, with the result that, in regard to the system, initial use may begin somewhere between 2015 and 2020. In the meantime, a considerable number of kilometers in terms of mileage and operating hours have been covered with the various vehicles in many parts of the world. At the same time, the fuel side

Target

Comparable to combustion engines

before 2020

Figure 30.28 Fuel-cell vehicles, Daimler AG launch strategy [30-36], [30-37].

also needs to be resolved, that is, hydrogen production and distribution including the filling station network. If we look at the technological progress achieved, it is very clear that development is a long way off reaching the point of saturation. Figure 30.30 shows possible advancements in stack technology with a power-to-weight ratio potential of 1.5 kW/kg and 2.0 kW/l. In the Mercedes-Benz vehicle unveiled in 2005 and the research vehicle F 600 HY Genius [30-37], the newly designed fuel-cell stacks delivered 30% more output, 16% less fuel consumption, and used 40% less space than the fuel-cell vehicle 2003. They were also able to be started in the matter of a few seconds at −25°C.

1044 | Internal Combustion Engine Handbook

6606_Book.indb 1044

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.7 The Fuel Cell as a Vehicle Drive System

Fuel cell system

Tank system

Electric motor

Mileages

Concept vehicle

Stack system

Type*

Volume

Pressure

Maximum output

vmax

Range**

Necar 1 (Based on: MB 100)

 50 kW

CH2

150 l

300 bar

30 kW

 90 km/h

>130 km

Necar 2 (Based on: MB V-Class)

 50 kW

CH2

2 × 140 l

250 bar

45 kW

110 km/h

>250 km

Necar 3 (Based on: MB A-Class)

 50 kW

MeOH

38 l

–

45 kW

120 km/h

>300 km

Necar 4 (Based on: MB A-Class)

 70 kW

LH2

100 l

  9 bar

55 kW

145 km/h

>450 km

Necar 4a (Based on: MB A-Class)

 75 kW

CH2

3 × 140 l

350 bar

55 kW

145 km/h

>200 km

Necar 5 (Based on: MB A-Class)

 75 kW

MeOH

45 l

–

55 kW

145 km/h

>400 km

Nebus (Based on: MB O 405 N)

250 kW

CH2

7 × 150 l

300 bar

2 × 75 kW

 80 km/h

>250 km

BZ-Sprinter (Based on: MB Sprinter)

 75 kW

CH2

308 l

300 bar

55 kW

120 km/h

>150 km

A-Class (HW1)

55 kw

CH2

110 l

350 bar

65 kw

140 km/h

170 km

B-Class (HW23***)

80 kw

CH2

170 l

300 bar

55 kw

120 km/h

150 km

**CH2—compressed hydrogen, LH2—liquid hydrogen, MeOH—methanol **cross comparison not possible due to different driving cycles *** 2009

Figure 30.29 Technical data from fuel cell concept vehicles [30-36].

Mark 9 2000, 80 kW • Gravimetric power: • Volumetric power:

Technology study TC 2003 (1.3 kW; 1.7 kW/l) 0.9 kW/kg 1.0 kW/l

• Power-to-weight ratio (potential) 1.5 kW/kg • Power-to-volume (potential) 2.0 kW/l

Intensive studies with H2 pressure storage systems with 700 bar of pressure are being carried out alongside and the lithium-ion battery is being integrated as a newer battery type. Research still needs to be carried out on the cold start and freezing suitability, cooling (fuel cell output), service life, system component reliability, and cost reduction requirements [30-38]. Besides Daimler AG, various other vehicle manufacturers and research institutions have also developed fuel cells and fuel-cell vehicles. Fuel-cell research is receiving intensive support from the government in China. The vehicles used during the 2008 Beijing Olympics ran with relatively few

Figure 30.30 PEM fuel-cell stacks based on hydrogen [30-37].

problems and have since been shipped from China to California. One particularly interesting solution has been proposed by the research team at Volkswagen [38-39]. This special fuel cell consists of a film that is stable at high temperatures made from polybenzimidazole (PBI), which is interspersed with phosphoric acid [HT-PEM] Figure 30.31 and works at an operating temperature of 160°C (the conventional membrane is designed for 80°C). The advantages of the HT PEM are shown in Figure 30.31. It requires less catalytic material, is more compact, and has a longer service life.

Internal Combustion Engine Handbook | 1045

6606_Book.indb 1045

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Reduced system complexity  No humidification  No water management  Reduced cooling  Reduced catalytic converters – degradation

Figure 30.31 Concept and advantages of high-temperature PEM [30-39]. See color section page 1107.

 High tolerance to gas impurity

140

•• reduces dependency on crude oil.

60 48

40

•• H2 storage in the vehicle

35

•• cold-start capability not yet sufficient.

20

23

21

20

en

H gin yb e rid 20 C 10 C N -H G 2 fro FC m EU C -H co 2 al fro m EU na FC tu ra C lg -H as 2 fro m EU bi om as s

G

e in ol as G

gi en el ie s

10

FC

ne

20

20

10

10

H

H

yb

yb r

rid

id

0

D

When comparing different fuels and drives systems, in addition to mere figures, such as WTW, (well-to-wheel), efficiency, and emissions, the availability of raw materials (crude oil, biomass, coal, nuclear power, water, etc.) also needs to be taken into consideration. This is true both geographically as well as politically. Energy supply for the mobility sector must concern itself intensively with other energy forms and resulting drive systems. Hydrogen can be produced from the most diverse energy sources, from crude oil to nuclear power and even biomass. Moreover, fuel-cell vehicles can use many of the technologies tested in hybrid vehicles, such as batteries, electric motors, AC compressors, and power electronics. In addition to the requirements of availability, there are other demands such as the lowest possible emissions in terms of WTW. Incidentally, this also applies for the carbon dioxide per kilometer. The required 120 g CO2/km in 2012 for the mean value of new vehicles coming onto the market in Europe is one of the greatest challenges at present and alternative fuels need to be considered now.

81

ne

•• overall efficiency lower than that of all-electric vehicles, if battery performance reaches its targets

80

gi

•• infrastructure

100

WTW CO2/km

•• weight

129

120

Disadvantages at present are as follows: •• costs

Tank-to-Wheel Well-to-Tank

140

en

•• it is possible to produce H2 from renewable sources, even to cover energy peaks

e

•• no CO2 emissions during operation •• extensive range compared to all-electric vehicles: 1 kg H2 equals around 100 km, target is 5 kg = 500 km

in

•• highly efficient

ol

The fuel-cell drive system certainly has the opportunity of being used in future alongside other drive alternatives. Advantages are as follows:

There are numerous studies on the subject of “efficiency for fuels and drive systems.” The most comprehensive wellto-wheel study can be found online at http://ies.jrc.ec.europe. eu/wtw. The values given in Figure 30.32 reveal a few of the examples for the technologies in 2010. According to this study at least, the combination of cheap hydrogen production and fuel-cell drive systems are far superior to the gasoline and diesel fuel based on crude oil. However, there are also concepts for combustion engines, which demonstrate minimal CO2 emissions, see Chapter 30.1.

as

30.7.3 Assessment of the Fuel Cell in Comparison to other Drive Systems

Figure 30.32 WTW CO2 emissions from different drive systems and fuels.

Why has the use of the fuel cell as a vehicle drive system been consistently delayed? This delay can be attributed to two reasons: Firstly, it is due to the fact that the necessary further development stages already mentioned need to be implemented particularly on the cost side. Secondly, that major political entities, such as the USA, EU, and Japan, need to agree on an energy policy that provides broad assurances for vehicle developers, fuel producers, and customers alike that the hydrogen will be economical. In any rate, the present

1046 | Internal Combustion Engine Handbook

6606_Book.indb 1046

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.9 Generating Electricity using an Auxiliary Power Unit = APU

development of fuel-cell vehicles has also exerted massive pressure on the further development of conventional vehicles.

30.8 Synoptic Evaluation of Alternative Energies and Drive Systems Alternative energies and drive systems have the chance to achieve larger market shares in the medium term. The precondition for this medium-term success is that they achieve equivalent vehicle performances for the customer and can be sufficiently optimized for the other criteria of complexity, convenience, and cost to permit their acceptance from customers. Another fact that cannot be forgotten is that present-day vehicles (engine, transmission, powertrain management) are also undergoing continuous improvement. New developments must prevail in the face of this future capability.

Heating/Cooling and power generation of thermoelectric heating equipment Heating/Cooling (Peltier effect)

Power generation (Seebeck effect) Heat supply

Heat absorption

Heat removal

Heat removal

Figure 30.33 Potential thermoelectric applications [30-40].

30.9 Generating Electricity using an Auxiliary Power Unit = APU The constantly growing need for electrical power and air conditioning in motor vehicles (passenger cars and utility vehicles) has led to significant improvements in the efficiency of vehicle generators and air conditioners. At the same time, however, this has revealed the limits in terms of efficiency, installation space, and energy requirements. The acronym APU refers to a device, which produces electrical energy for the auxiliary units. As a result, alternatives are currently in the research and development phase. Possible solutions are described in the following: Thermoelectrics, the SOFC fuel cell used as an APU (auxiliary power unit) and a free-piston linear generator used to generate electricity.

30.9.1 Thermoelectrics

The capability of thermoelectric elements has been known since they were made famous by Peltier in 1894. Their application in motor vehicles primarily allows functions (see also the “Energy management” chapter) such as heating and cooling and the generation of electricity by supplying heat. Figure 30.33 shows the relevant effects. The Peltier effect (heating and cooling) is shown on the left-hand side, while the generation of electricity is illustrated on the right-hand side (Seebeck Effect). It can be seen from Figure 30.34 that thermoelectrics can also be applied to large series production [30-40]. It shows an illustration of air-conditioned vehicle seats, of which almost one million units had been sold by the end of 2007. Three cases for further application in motor vehicle exist: •• implementation as a vehicle air conditioner •• generation of electrical energy from exhaust gas heat •• battery cooling.

MTM – seat technology

Heat exchanger

cooled or heated air outlet to the seating upholstery

Peltier element

used air outlet

Air distribution

Perforated leather distribution layer Electrical control device

Switch

Figure 30.34 Air conditioning in vehicle seats.

Its application as an air conditioner for plug-in vehicles and electric vehicles appears to be particularly interesting, even if there is still enormous development work to be done in order to optimize the overall system into the vehicle, taking into account the possibilities of thermoelectrics.

30.9.2 The Fuel Cell as an APU

While the volume production application of the fuel cell as a vehicle drive system has been delayed somewhat due to costs and a lack of hydrogen infrastructure, the use of the fuel cell as an APU seems to have an earlier chance at volume production application in vehicles with a high proportion of electrical consumers. The customer benefits are persuasive when compared to the existing onboard technology and the additional costs are of less consequence for smaller outputs.

Internal Combustion Engine Handbook | 1047

6606_Book.indb 1047

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

BMW introduced an APU based on a PEM (proton-exchange membrane) as early as 1999. This solution could be excellently integrated into a vehicle already run on hydrogen. However, costs and complexity rise considerably if the PEM fuel cell is designed to be run on liquid fuels, such as gasoline or diesel. In this instance, hydrogen firstly needs to be generated from the fuel using a complex reformer and a multi-stage gas purification system. This is why several companies are working on APU systems based on high-temperature fuel cells for diesel and gasoline vehicles, mostly with the SOFC (solid oxide fuel cell). It has an operating temperature of around 700–800°C and allows a much simpler reformer to be used, such as one along the lines of the POX (Partials Oxidation) concept. Figure 30.35 details a schematic comparison of a system with a PEM and SOFC fuel cell.

Gasoline autothermic “steam reformer”

800 °C High-temperature shift reactor

Output

5 kW

Fuel

Gasoline

Service life (continuous)

>5000 h

Service life (start cycles)

>5000

Efficiency

>35 %

Starting time

<10 minutes

Weight

<4 kg/kW

Volume

2 l/kW

Solid oxide fuel cell stack

“Partial oxidation” gasoline reformer

> 900 °C

PEM FC

Consequently, the SOFC gives the APU the advantage of being able to use commercially available fuels, yet the complexity of the system can be kept in check. The design of a SOFC fuel cell APU is illustrated in Figure 30.36. The targeted performance data of an APU for use in passenger cars are as follows:

SOFC

800 °C

Low-temperature shift reactor

“Preferential” oxidation

PEM fuel cell stack

Figure 30.35 Comparison of fuel preparation for PEM + SOFC [30-41].

88 °C

2x30-cell SOFC stacks Cathode air heat exchanger

230 Volt exit

Reformer Figure 30.36 Delphi solid oxide fuel cells [30-42].

1048 | Internal Combustion Engine Handbook

6606_Book.indb 1048

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.9 Generating Electricity using an Auxiliary Power Unit = APU

The targeted performance data of an APU for use in utility vehicles are as follows: Output

1–3 kW

Fuel

Diesel

Service life (continuous)

>25,000 h

Efficiency

>25%

Starting time

<60 minutes

Weight

<30 kg/kW

Volume

<80 l/kW

The utility vehicle sector illustrates an interesting case for a diesel-operated SOFC-APU. First and foremost when large amounts of power are required in the stationary mode without a combustion engine for stationary air conditioners. US legislation (California) should be observed here in particular, and it requires that the engine is shut off during quiet/waiting periods. Initial approaches can be found in [30-22]–[30-24]. This involves preparing the diesel fuel in the same way as the gasoline version using a POX reformer and guiding it to a SOFC fuel cell. A reheater combusts the residual gas and at the same time heats up process air for the fuel cell. A prototype of a SOFC-APU for diesel fuel is shown in Figure 30.37.

Charge cycle

Valves

Combustion chamber

Stator

Rotor

Figure 30.37 Prototype of a SOFC-APU for diesel fuel manufactured by Webasto [30-43], [30-44].

Springback space

Regulation valve

Figure 30.38 Internal combustion engine with linear motor concept.

Internal Combustion Engine Handbook | 1049

6606_Book.indb 1049

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 30 Alternative Vehicle Drives and APUs (Auxiliary Power Units)

Volume production application of SOFC systems will largely depend on whether law makers makes a relevant demand, whether the starting time can be significantly reduced, and whether the objectives for the thermal number of cycles within the system can be achieved.

30.9.3 Internal Combustion Engine in Combination with a Linear Generator (Free-Piston Linear Generator)

Another solution for efficiently generating onboard electrical energy is a combustion engine designated as a two-stroke engine with direct injection, which induces coils through the forwards and backwards motion of the piston via a permanently magnetic rotor and this subsequently produces electricity. The regulation of the gas exchange via slits in the cylinder wall stands out due to the low constructive effort, however, it does entail HC emissions in the exhaust gas are explained here. The application of conventional valves solves this problem, but at the same time increases the constructive effort. Great requirements are made of the control and regulation technology, since it involves a purely mechatronic system [30-45], [30-46], [30-47], [30-48]. The solution shown in Figure 30.38 (DLR—Institute of Vehicle Concepts (FK)) reveals a system that uses a gas spring with adjustable spring characteristic instead of a second combustion chamber [30-49]. In addition to the inbuilt variable compression, there is the possibility of the displacement variation, whereby load control is possible via the stroke adjustment. Thus, minimizing throttle and friction losses leads to a significant efficiency increases in the partial-load range. A compact construction can also be realized, since two systems need to be switched one after the other for the mass balancing. Electrically synchronizing both systems presents another challenge. The combustion cylinder, gas spring, and linear generator have already been put into practice on test benches. The next steps will have to be a suitable vehicle design, which can be integrated into a hybrid vehicle as a drive unit.

Bibliography

30-1. Behlmer, A. 2004. Zünglein an der Waage, Automobil Industrie 1–21. 30-2. Drescher, I., Steiger, W. 2008. Volkswagen Fuel and Powertrain Strategy FISITA, F-2008-06-139, proceedings by ATZ. 30-3. Weber, Th. Innovative Vehicle concepts towards accident and emission free driving, plenary lectures, FISITA 2008, proceedings by ATZ. 30-4. Köhle, S. 2004. Entwicklungsziele, Fahrzeugbeschreibung und erste Messergebnisse des VW Bora mit Hybridantrieb, Hybridfahrzeuge und Energiemanagement, GZVB, Braunschweig ISBN 3-937655-00-X. 30-5. Antony, P. et al. 2007. Vieweg Handbuch Kraftfahrzeugtechnik, Vieweg, Wiesbaden, ISBN 978-3-8348-0222-4. 30-6. Noreikat, K.E. et al. Hybride Fahrzeugantriebe. Die Evolution zum Mehrwerthybrid, VDI-Bericht 1565, Düsseldorf. 30-7. Borgmann, K., et al. 2006. Effiziente Dynamik als Lösung des Zielkonflikts zwischen Kundenwunsch und Gesetzesforderung, In: Proceedings Technical Congress, VDA, Frankfurt.

30-8. Miska, J. 2004. Hybridfahrzeuge—Produktausprägungen als Antowrt auf Kundenanforderungen, GZVB e.V. Braunschweig, ISBN 3-937655-00-X. 30-9. Kemper, H. et al. 2008. Elektrische Energiespeichersysteme für zukünftige Hybridfahrzeuge, Aachen Colloquium, Conference proceedings. 30-10. Information from Toyota—Deutschland January 2009. 30-11. Harada, I. 2000. Entwicklung eines neuen Toyota Hybrid Fahrzeuges, Proceedings Motor und Umwelt, AVL Graz. 30-12. Muta, K. et al. 2004. Development of New Generation Hybrid System THS II—Drastic Improvement of Power Performance and Fuel Economy, SAE SP 1833, USA, ISBN 0-7680-1369-0. 30-13. Hack, G. et al. 2004. Die Liter-Leistung. www.auto-motor-sport.de 6/2004 and Honda press material. 30-14. Brachmann, Th. 2004. Development of Powertrain for the Civic IMA, a Gasoline-Electric-Hybrid-Vehicle, Honda RD-Center, Offenbach, Germany. Iigima, T. 2006. Development of Hybrid Systems for 2006 Compact Sedan, SAE International SP_2008, Warren, USA. Brachmann, Th. 2009. Today’s and future Hybrid-Fuel Cell Vehicles of Honda, 6th Symposium Hybrid and Energy Management! GZVB. ISBN 967-3-937655-20-8. 30-15. Hövermann, M. 2006. Die Continental-Strategie für Mild und Full-Hybrid-Systeme, In: Proceedings 3rd Braunschweiger Hybridsymposium, ISBN 3-937655-06-9. 30-16. Adachi, M. et al. 2006. Developement of a new Hybrid Transmission for RWD Car, SAE International SP_2008, Warren, USA. 30-17. Wandt, H.P. 2005. Das Antriebskonzept des Hybrid Pkw Toyota Lexus RX 400 h, lecture at TU Braunschweig. 30-18. FEV Motorentechnik GmbH, Hybrid Technology, Aachen 2008, http://www.fev.com. 30-19. IAV GmbH, Berlin 2009. 30-20. BMW, Daimler Group, GM: Hybrid Cooperation, Special publication, 27th International Vienna Motor Symposium 2006. 30-21. Nietschke, W. 2006. Hybrid-Hype oder technologische Chance?, Proceedings MTZ Konferenz Motor, Stuttgart. 30-22. Hofmann, L. et al. twin Drive—Ein Schritt in Richtung Elektromobilität, Innovative Fahrzeugantrieb 2008, VDI-Berichte 2030, VDI-Verlag, ISBN 978-3-18-092030-6. 30-23. FEV Motorentechnik GmbH, Electrical Vehicle technology …, Aachen, 2009. 30-24. Wüchner, E. 2007. Elektroantriebe, Vieweg Handbuch Kraftfahrzeugtechnik, Wiesbaden, ISBN 978-3-8348-0222-4. 30-25. Köhler, U. 2005. Batterien für Elektro- und Hybridfahrzeuge in Hybrid-, Batterie- und Brennstoffzellen-Elektrofahrzeuge, Expert Verlag, ISBN 3-8169-2433-6. 30-26. www.thinkmobility.com. 30-27. Mitsubishi Motors 2006. 30-28. Noreikat, K. 2007. Stirlingmotor, Dampfmotor, Gasturbine und Schwungrad, Vieweg Handbuch Kraftfahrzeugtechnik, Vieweg, Wiesbaden ISBN 978-3-8348-0222-4. 30-29. Täubner, F. et al. 2000. Ergebnisse aus Prototypen neuer Schwungradspeicher, VDI-Bericht 1565, Innovative Fahrzeugtechnik, Düsseldorf, ISBN 3-18-091565-X. 30-30. Dietrich, T. 2000. Ultracapacitors—Power für innovative Automobilanwendungen, VDI-Bericht 1565, Innovative Fahrzeugtechnik, Düsseldorf, ISBN 3-18-091565-X. 30-31. Michel, H. et al. 2004. Ultracap-Module—Leistungsfähige Energiespeicher für Hybridfahrzeuge, Hybridfahrzeuge und Energiemanagement, GZVB Braunschweig, ISBN 3-937655-00-X. 30-32. Noreikat, K.E.. 2007. Handbuch Kraftfahrzeugtechnik, Vieweg, Wiesbaden, ISBN 978-3-8348-0222-4.

1050 | Internal Combustion Engine Handbook

6606_Book.indb 1050

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

30.9 Generating Electricity using an Auxiliary Power Unit = APU

30-33. Seiffert, U. et al. 1989. Automobiltechnik der Zukunft, VDI Verlag Düsseldorf, ISBN 3-18-400836-3.

30-52. Nietschke, W., Fickel, F., Kümmel, S. 2011. Inductive Energy Transfer for Electric Vehicles, ATZ 04/2011, Jrg. 113.

30-34. Mohrdieck, Ch. 2006. Die Brennstoffzelle—Emissionsfreier Antrieb der Zukunft, Brennstoffzellentechnologie-Nachmittag, Hamburg.

30-53.Rummisch, E. 2011. Elektrische Straßen- und Hybridfahrzeuge, Expert Verlag, ISBN 978-3-8169-2734-1. Knödel, U. et al. 2010. Design and Implementation of Requirement-driven Electric Drives, ATZ 06/2010 Volume 112.

30-35. Lamm, A. 2005. Alternative—Innovative Ansätze zur Senkung von Verbrauch und Emissionen, lecture at University of Stuttgart. 30-36. Truckenbrodt, A. 2005. Brennstoffzellen als Antrieb für mobile Systeme, Vieweg Handbuch Kraftfahrzeugtechnik, Vieweg, ISBN 3-528-33114-3. 30-37. Lamm, A. 2005. Fuel cell presentation, Aachen Colloquium. 30-38. u.a. Steiger, W., 2006. Nachhaltige Antriebssysteme—Chancen und Herausforderungen, 10. Handelsblatt Jahrestagung Automobiltechnologien Munich. Dietrich, P. et al. 2002. Erste Ergebnisse des HyPower VW-Bora, VDI Conference Innovative Fahrzeugantriebe Dresden. Schubert, U., et al. 2002. Entwicklung der Brennstoffzelle bei GM/Opel, VDI Conference Innovative Fahrzeugantriebe, Dresden. Matsumoto, T. et al. Development of Fuel-Cell-Hybrid Vehicle, Toyota SAE-Paper 2002-01-0096, Warren, USA. Lee, K.C. et al. Hyundai Santa Fe FCV Powered by Hydrogen, SAE Paper 2002-01-0093, Warren, USA. Honda: New Honda Fuel Cell Stack Operates at Low Temperatures, Auto Technology 6/2003. 30-39. Steiger, W., et al. Second Generation PEM Fuel Cells—Initial Findings MTZ 12/2007 Volume 68.

30-54. Atkins, A.; Feulner, P. 2011. The Ricardo Mechanical Hybrid Drive, MTZ 02/2011, Volume 72. 30-55. Köll, L. et al. 2011. Analysis of the Configuration of an Electric Vehicle, ATZ 02/2011, Volume 113.

Further Literature on the Subject of Hybrids and Fuel Cells GZVB, Hybridfahrzeuge und Energiemanagement, February 2004, Braunschweig, ISBN 3-937655-00-X. GZVB, Hybridfahrzeuge und Energiemanagement, February 2006, Braunschweig, ISBN 3-937655-06-9. GZVB, Hybridfahrzeuge und Energiemanagement, February 2007, Braunschweig, ISBN 978-3-937655-10-9. GZVB, Hybridfahrzeuge und Energiemanagement, February 2008, Braunschweig ISBN 978-3-937655-15-4. VDA, Technical Congress 2006, Frankfurt.

30-40. Bell, Lon broader Use of Thermoelectric Systems in vehicles expert verlag, 2009 ISBN 978-3-8169-2877-5.

VDA, Technical Congress 2007, Sindelfingen.

30-41. Zizelman, J., et al. 2002. Solid Oxide Fuel Cell Auxiliary Power Unit—A Development Update, SAE, Warrendale.

SAE International, Advanced Hybrid Vehicle Powertrains 2006, SP_2008, Warrendale, USA.

30-42. Delhi, USA, 2008. Biyendolo, J.M. 2007. European market Potential for a Delphi Solid Oxide FuelCell based truck Auxiliary Power Unit Hochschule für Technik und Wirtschaft des Saarlandes.

VDI, Innovative Fahrzeugantriebe, VDI-Bericht 1852, November 2004, ISBN 3-18-091852-7.

30-43. Boltze, M., Wunderlich, Chr. The SOFC-APU for long haul trucks—A promising early market application. Ninth Grove Fuel Cell Symposium 09/2005, London. 30-44. Boltze, M., Wunderlich, Ch. Bordstromversorgung mittels SOFC-APU. 4th Specialist Conference Innovative Fahrzeugantriebe, Dresden, 2004; VDI-Bericht 1852, Düsseldorf, ISBN 3-18-0918527. 30-45. Deutsches Patent- und Markenamt: Offenlegungsschrift DE 19943993A1. Rotthäuser, S. et al. 1996. Der Freikolbenmotor—Eine Informationsschrift. 30-46. Blaringan, P. van: Advanced internal combustion electrical generator, Proceedings of the 2001 DOE Hydrogen Program Review, Livemore, CA 91556. 30-47. Arshad, W.M. et al. 2002. Integrated Free-Piston Generators:An Overview, Royal Institute of Technology, Stockholm. 30-48. Aichlmayr, H.T. et al. 2002. Miniature free-piston homogeneous charge compression ignition engine-compressor concept—Part I: performance estimation and design considerations unique to small dimensions, Chemical Engineering Science 57, Perganon 4161–4171.

VDA, Technical Congress 2008, Ludwigsburg.

VDI, 27th International Vienna Motor Symposium 2006, FortschrittBerichte, Reihe 12, 622, VDI-Verlag, Düsseldorf 2006, ISBN 3-18-362212-2. VDI, Innovative Fahrzeugantriebe, VDI-Berichte 1975, November 2006, ISBN 3-18-091975-2. VDI, Innovative Fahrzeugantriebe, VDI-Berichte 2030, November 2008, ISBN 978-3-18-092030-6. VDI, Brennstoffzelle, VDI-Berichte 2036, May 2008 Braunschweig ISBN 978-3-18-092036-8. Braess/Seiffert. Handbuch Kraftfahrzeugtechnik, 5. Auflage, Vieweg, 2007, ISBN 978-3-8348-0222-4. Fickel, C. et.al. Fuel Cell Hybrid Concept Vehicle for Emission-free Mobility in City Traffic, ATZ 04/2011, Volume 113. Kim, S.H., Ahn, P.B.K., Lim, P.T.W. Development of Hyundai-Kia’s Fuel Cell Stack, MTZ 01.2010, Wiesbaden, January 2010. Kurzweil, P. 2003. Brennstoffzellentechnik, Vieweg Verlag, 2003, ISBN 3-528-03965-5. Mettner, M. 2005. Hybrid-, Batterie- und Brennstoffzellen-Elektrofahrzeuge, Expert Verlag, ISBN 3-8169-2433-6.

30-49. Gräf, M. et. al. 2007. Investigation of a high efficient Free Piston Linear generator with variable Stroke end variable Compression Ratio, WEVA Journal, pp. 116–120, Vol. 1.

Stan, C. 2008. Alternative Antriebe für Automobile, 2. Auflage, Springer Verlag, Berlin-Heidelberg, ISBN 978-3-540-76372-7.

30-50. Mohrdieck, Ch.: Elektrifizierung des Automobils—Die technischen und wirtschaftlichen Herausforderungen, Innovative Fahrzeugantriebe 2008, Dresden, VDI-Verlag Düsseldorf ISBN 978-3-18-092030-6.

Winterhagen, J. 2010. Last Chance for the Fuel Cell? MTZ 01.2010, Wiesbaden.

Voß, B. 2005. Hybridfahrzeuge, Expertverlag, 2005, ISBN 3-8169-2501-4.

30-51. Haraguchi, T. 2011. Reducing Fuel Consumption by Improved Vehicle Efficiency, ATZ 04/2011, Volume 113.

Internal Combustion Engine Handbook | 1051

6606_Book.indb 1051

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 1052

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

31 Energy Management in the Engine and Vehicle The search for potential means of reducing fuel consumption for operating motor vehicles is one of the key research and development levels in automotive engineering. This is particularly true, because a reduction in fuel consumption is accompanied by a direct lowering of CO2 emissions. The fleet average of all newly registered passenger cars in Germany was 159.3 g CO2 per km in January 2009. This is roughly equivalent to the European Union (EU) fleet average during 2006. The average carbon dioxide output of passenger car fleets is limited to 120 g/km across Europe in 2012. The compromise reached in the EU provides a relaxing of these targets with a transition phase. From 2012, 65% of new cars will have to stick to a specific manufacturer limit value. This is

subject to the average weight of the fleet sold. It is understood to be 75% from 2013 and 80% from 2014. The limit value for a manufacturer’s entire fleet will apply from 2015. A further reduction of CO2 emissions is planned from 2020 and will see an upper limit of 95 g/km. As can be seen in Figure 31.1, there are only a few vehicles that meet these requirements now. A penalty payment is planned if the mentioned limit values are exceeded. The first gram that the manufacturer-specific and weight-dependent limit value exceeds will be billed at €5, a second extra gram will be billed at €15, and a third gram at €25. Manufacturers will have to pay €95/gram when exceeding the limit by four grams or more. According to The German Association of the Automotive Industry (Verband

CO2 [g/km]

Regression line for 2006 (Weighted by volume)

EU fleet average 2006

Curb weight [kg/veh.]

Figure 31.1 Fleet CO2 average in the EU (source BMW).

Internal Combustion Engine Handbook | 1053

6606_Book.indb 1053

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 31 Energy Management in the Engine and Vehicle

der Automobilindustrie e.V.—VDA) calculations, this amount is twenty-four times greater than that which the coal and steel industries, for instance, have to pay in the emissions trading system. Hence, a significant disparity exists to the disadvantage of automobile manufacturers. There shall be exceptions for small volume and niche manufacturers. Vehicles with a CO2 output under 50 g of carbon dioxide per km receive so-called super credits, which can be deducted multiple times across the fleet. One key area in the development of engine and vehicle technology to satisfy these requirements is effective energy management in the vehicle. In this instance, it is important that only the correct type and amount of energy is provided that is actually needed by the consumers. Therefore, only the as much energy should be provided as is necessary to satisfy the functions. This sounds trivial at first, but because an energy-related overall analysis of the vehicle is often missing, a number of energy bypasses have been developed that need to be rearranged. Besides a requirement-based allocation of energy, the following are still important for good efficiency: •• to avoid energy losses as much as possible •• to recover accrued energy losses. If we consider the energy flows in the vehicle, then in the first instance, the energy content of the fuel is in the form of

Primary energy

Conversion in the motor

Mechanical energy at the crankshaft

Chemical conversion (oxidation) of primary energy products (diesel, gasoline, biofuels, hydrogen)

Crude oil

Natural gas

Available energy form following conversion in the motor

Solar energy (biomass, hydrogen)

chemical energy. All the functions, which are called on during the operation of the vehicle, feed on this energy. If we want to increase the efficiency of the energy flows in the vehicle effectively, this knowledge is similarly of critical importance, just as the consideration of the total energy flow system in the vehicle. Therefore, all cause-effect relationships must be considered. Figure 31.2 shows potential energy flows in the vehicle. Primary energy can be divided into fossil fuels, such as natural gas and crude oil, or alternative energy sources, such as solar energy. Conversion in conventional combustion engines takes place through the chemical reaction of the fuel with oxygen in the air. This results in a series of available energy forms in the vehicle, which can be used more or less expediently. Central to this is the available mechanical energy, which is partly used to propel the vehicle, but can be converted into other energy forms as well. This equates to approximately one third of the chemical energy of the fuel. A large proportion of the fuel energy (roughly another third) is found again as exhaust gas energy. Because this exhaust gas energy displays a high thermal potential, and is greater than that of the ambient air, this exergetical energy potential (according to the second law of thermodynamics) is convertible into other energy forms. The last third of the chemical energy largely lost to the coolant. With regard to the increasing power requirement of vehicles through electrically driven components, it is increasingly important to no longer meet these requirements with

Use of energy in the vehicle by type of energy

Mechanical energy

Use of energy in the vehicle by function (examples)

Propulsion, e.g. accelerating

Air conditioning, e.g. cooling, heating Exhaust gas enthalpy

Electrical energy

Lighting

Infotainment, e.g. radio, telephone Energy in the coolant

Pneumatic energy Convenience, e.g. electric seat adjustment

Radiation/Convection

Thermal energy

Safety, e.g. pneumatic closing systems

Figure 31.2 Energy flow in the vehicle.

1054 | Internal Combustion Engine Handbook

6606_Book.indb 1054

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

31.2 Requirement-Based Energy Management

mechanical energy as before, but to exploit the previously unused energy sources. Hybrid vehicles have the advantage of also being able to use a large electrical machine here. Thus, a large amount of (kinetic) braking energy can be recovered as electricity using the technology. This valuable energy can then be utilized via an intelligent energy management system to supply the auxiliary units, support the combustion engine with electrical energy, or it can even be used to realize electric driving shares.

31.1 Losses During Energy Conversion If we consider the efficiency of an engine, then theoretically there is great potential for reusing the “waste energy.” Converting this “waste energy” into electrical energy would bring about the greatest benefits, because the “electrification” of the motor vehicle continues to forge ahead unchanged. A few examples of the recovery into “useful” energy forms or avoiding energy losses include: •• Braking energy. The actual potential that is currently regarded as recoverable in the motor vehicle emissions group (MVEG) cycle for a mid-size vehicle weighing approximately 1500 kg is approximately 700 kJ. This figure already includes all losses such as rolling resistance and aerodynamic drag, as well as the efficiency of the electrical machine. This equates to a fuel consumption saving of approximately 0.5 liters/100 km. To exploit this energy as much as possible, additional electrical machines are required in the vehicle, such as the ones installed in full hybrid vehicles, for instance, or electrical machines accomplished as wheel hub machines. •• Generator control. Generators (alternators) are usually driven via a belt drive from the crankshaft. This means that mechanical energy is diverted at the output (crankshaft) of the engine and therefore reduces the energy flow to the drive wheels. Braking energy is also recoverable using this technique with appropriate generator regulation. This occurs in the following way: in coasting or braking phase (vehicle is rolling, foot is off the gas pedal, and fuel supply is switched off), as much energy as possible is converted into electrical energy, that is, the generator absorbs as much power as possible. One requirement, however, is that the state of battery charge is less than 100%. Therefore, the battery may only be partially charged in this case to ensure that it is ready for energy absorption in coast mode. This in turn postulates regulation of the state of battery charge, which must be integrated into an electrical energy management system. •• Electrical energy management system. The first step in the direction of electrical energy management for current on-board supply systems has been carried out with the regulation of the state of battery charge. Fixed-cycle batteries are a basic requirement for this. The following elements are important:

•• Battery status sensor. Accurate information regarding the state of battery charge is the basis of the energy management system. This is gained from the battery current, voltage, and temperature variables and is determined using a battery sensor. This can be used to calculate the battery charge and wear state, as well as the prognosis of the necessary minimum charge state for the next start. •• Idle speed adjustment. With accurate information regarding the state of battery charge and the projection of the next driving situation, it is possible to reduce the speed in the idle mode. This delivers advantages in terms of fuel consumption and pollutant emissions. If the state of charge is critical, or the generator is required to perform at full power, then the idle speed must be raised again. •• Distribution of energy flows. If the battery state of charge is assessed as critical, then the power in the on-board supply system must be distributed so that all safety-critical consumers are supplied. Further battery discharge must be avoided with corresponding power increases from the generator. •• Battery charging. The generator’s control voltage is adjusted such that the desired charge voltage is set at the battery poles to minimize voltage losses. If the target state of charge is met, then the generator voltage is lowered to reduce fuel consumption. The battery may no longer be unconditionally fully charged, so as to be able to feed in the recovered energy. An adequate state of charge is determined depending on the relevant environmental and driving conditions. •• Energy diagnosis of the on-board supply system. The state of battery aging, driving profile, and typical customer energy requirements can be derived from the storage of battery data. It is useful to amend the operating strategy of the electrical energy management system.

31.2 Requirement-Based Energy Management One example of a nonelectrical energy management system used to increase efficiency is the requirement-based control of the oil flow using a regulated oil pump, such as that seen in the Audi A4 and in most of the engines from the BMW Group since 2007. This regulates the volumetric flow of the oil pump. The advantages are that only as much oil is supplied as the engine actually requires and that the oil pressure is adjustable in several stages, according to a stored engine map. In the new European driving cycle (NEDC), this leads to a saving of approximately 5 g CO2/km. This example can be applied to other components, such as the fuel pump, air conditioning compressor, and so on. More and more vehicles are being offered with intelligent energy management systems. The following options are usually apparent:

Internal Combustion Engine Handbook | 1055

6606_Book.indb 1055

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 31 Energy Management in the Engine and Vehicle

Requirements

Coordination

Implementation

On-board supply system

Mileage

Electronics Electronics Braking energy etc.

mechanical

Idling Air conditioning compressor

Priority-orientated

electric

Comfort

Fuel consumption

Power steering

Energy supply

Thermal component protection etc.

Figure 31.3 Energy management coordinator (as per BMW).

•• Requirement-based and condition-dependent switch-on and shutoff of electrical consumers •• Specification and implementation of a maximum saving potential achieved by switching off the electrical consumers, with convenience sacrifices, which are still accepted by the driver. •• Automation of all control interventions. This is for example achieved via an energy management system coordinator, as is implemented in the new BMW 7 Series from model year 2009 (Figure 31.3). A priority list ensures that the “correct” decision is always made in terms of efficient energy use, thus preventing energy losses and unwelcome fluctuations in the on-board voltage. This is necessary as recovered energy is supplied on the one hand, yet on the other hand, certain consumers such as the fuel pump and air conditioning fan are sensitive to voltage fluctuations.

31.3 Generating Electricity in Vehicles The constantly growing demand for electrical power and air conditioning in motor vehicles, combined with the necessity of taking some of the strain off the mechanical energy path in the vehicle for consumption reasons, has led to a considerable improvement in the efficiency of the vehicle generators, air conditioning systems, and electrical consumers. At the same time, however, this has revealed the limits in terms of efficiency, installation space, and energy requirements.

In addition, more and more functions in the vehicle are supplied with electrical energy, meaning that a separate energy path, which is not fed by the mechanical energy from the engine, will become more important in the future. Figure 31.4 shows the increasing demand for electrical power in the vehicle. This demand can only be met to a limited extent by conventional systems (generator and battery). Therefore, approaches are required to create an energy source that is independent of the combustion engine. At the same time, efforts are underway to move away from the 14 V on-board supply and reach on-board supplies with higher voltages. The higher costs of such an on-board supply system are an obstacle to this approach.

Installed electrical output [kW]

•• Optimized operation of the electrical consumers, which must be designed for extreme requirements, but whose maximum output may not necessarily be required depending on the driving situation, driver request, use, and so on.

Full-size car

Mid-size car

Year Figure 31.4 Electrical energy requirement in the vehicle.

1056 | Internal Combustion Engine Handbook

6606_Book.indb 1056

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

31.3 Generating Electricity in Vehicles

Therefore, alternatives are currently being examined for developing, so-called auxiliary power units (APUs), which allow the onboard generation of electrical energy, regardless of the central driveline. A series of developments are in sight, of which, for example, the fuel cell can be regarded as having the best prospects (see Chapter 30: Alternative Vehicle Drives and APUs). Furthermore, the utilization of waste heat by directly converting it into electrical power appears very worthwhile. This is reported to be the main focus of subsequent implementations.

α , absolute temperature T, electrical resistance ρ and thermal conductivity κ : ZT =

a=



Uthermo ∆T

hTE =

(31.1)

The Seebeck coefficient is used to determine the dimensionless coefficient ZT for a thermoelectric material depending on

hot

(31.2)

This ZT value is a common variable for evaluating the capability of a thermoelectric material at a certain temperature T and is directly integrated into the calculation of energyrelated conversion efficiency during the conversion of heat flowing from Thot to Tcold.

31.3.1 Thermoelectric Generator (TEG)

Converting heat into electricity by using a TEG is based on the effect discovered by Thomas Seebeck in 1821. This effect describes the generation of an electrical current Utherm between the contact points of two conductive materials, provided they are exposed to a temperature ∆T = Thot − Tcold difference. The reverse of the Seebeck effect is known as the Peltier effect (named after J. Peltier from 1834) and describes the formation of a temperature difference provided a voltage is applied. The extent of the producible voltage per Kelvin for a certain material is described with the Seebeck coefficient α (Figure 31.5):

a 2T rk



( 1 + ZT )1/2 − 1 ⋅ hCarnot T ( 1 + ZT )1/2 + cold

(31.3)

Thot

As is evident from this formula, the efficiency is approaching that of Carnot efficiency, if the ZT value is infinitely large. The challenge of developing a highly-efficient thermoelectric material, such as the one identifiable from the definition of ZT, lies in combining the intrinsic properties of good electrical conductivity with bad thermal conductivity. The ideal value is achieved with semiconductor materials such as bismuth telluride (Bi2Te3) for low temperatures, as well as lead telluride (PbTe) or silicon germanium (SiGe) for higher temperatures (see Figure 31.6). These semiconductor materials are so heavily doped that their transport properties equate to those of pure metals. In

cold

ZT value [-]

Figure 31.5 Thermoelectric effect and technical implementation in the module. See color section page 1108.

Thermoelectric generator [°C]

Figure 31.6 ZT values for various commercially available materials.

Internal Combustion Engine Handbook | 1057

6606_Book.indb 1057

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 31 Energy Management in the Engine and Vehicle

technical applications, p and n-doped semiconductor materials are combined as one thermoelectric element (Figure 31.5), where for the characteristic of a temperature difference, the free electrons (for the n-doped material) and the free holes (with the p-doped material) flow from the hot to the cold side and thus generate a voltage (see Figure 31.7).

Figure 31.7 Electron flow in a thermoelectric material.

Up till now, the maximum ZT values were approximately 1, but because of the use of nanotechnology, it has recently

been possible to produce materials in the laboratory with considerably higher ZT values of up to ZT = 1.5. For integration in the vehicle, as can be seen from the above equations, the temperature and good heat transfer are essential elements for high efficiency in the TEG. Both these criteria are present in engine exhaust gas. Integrating a TEG into an exhaust gas system requires intensive system adjustments, depending on the level of electrical recovery power required. The maximum exhaust gas back pressure and the maximum temperature, with which the thermo electric materials can be applied, constitute the orientation framework for the design of the overall system. If the cold side of the thermoelectric module is realized via a coolant link, then the maximum injection of heat into the coolant constitutes an additional limiting variable for overall efficiency. As the engine load points rises, both the exhaust gas temperature and the exhaust gas mass flow increase. To prevent the module from overheating, as well as to avoid excessively high exhaust gas pressure from forming, because of the hot gas heat exchangers at high loads, a bypass must be integrated running parallel to the TEG hot gas path. Using a standard bismuth telluride material, this arrangement, for example, in a BMW experimental vehicle achieved approximately 200 W TEG power at a vehicle speed of 130 km/h. The increase in the material coefficient ZT to approximately 1.2–1.5 would, for example, be at a vehicle speed of 100–130 km/h, offset the entire on-board requirement of 600–700 W for a mid-size vehicle, which equals a fuel saving of approximately 5% (Figure 31.8).

BI2 Te 3 thermoelectric generator, Pmax= 200 W

Figure 31.8 Design of a TEG (BMW). See color section page 1108.

1058 | Internal Combustion Engine Handbook

6606_Book.indb 1058

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

31.4 Heat Management

Heat management provides another opportunity for efficient energy use in the sense of a CO2 reduction in the vehicle. The objective is to take the pressure off the mechanical energy path through heat management, particularly by reducing friction during the powertrain warmup phase or by decoupling the APUs. One such an example is the water pump. The water pump is usually fixed driven in the engine belt drive. There is no variability of the volumetric flow from the respective requirement. Because the design of the volumetric flow for the water pump is guided by low speeds (low engine speed, high engine performance, low vehicle speed, and low cooling effect), high volumetric flows are produced at high speeds, even though this is not required. Assistance is provided by a requirement-dependent control of the water pump. In addition, this must be taken from the coupling that is dependent on the engine speed, for example, by an electric drive system. This has the additional advantage that in the case of the engine warm-up phase, the coolant flow can be stopped, which means that the engine operating temperature can be achieved quickly. Measurements elicited a fuel consumption benefit of 2% and a performance advantage of 2 kW. Furthermore, heat management offers the potential for a significant consumption reduction through the efficient control/regulation of thermal energy flows in the powertrain, especially in connection with the coolant and oil supply. Heat management is not one single measure but its potential, and above all else, it taps into the interaction between other methods for reducing fuel consumption, such as downsizing, direct injection, exhaust-gas recirculation (EGR) cooling, and reducing friction. The key factor is the regulation and circuitry of the coolant circuit in the engine and neighboring systems. The design of one such coolant system is shown in Figure 31.9.

The EGR, vehicle heating, and engine-oil cooler are integrated into the system via pumps and regulating valves. The map thermostat is an important element in the coolant chain, with an integrated throttle thermostat in the bypass line, as well as the electrically operated solenoid valves. Different operating temperatures are necessary for optimum engine operation at different operating points. These can be realized using a map-controlled thermostat. The increase in the coolant temperature, thus possible, reduces friction losses, which has a positive effect upon fuel consumption. The cold start and the warmup phases are critical in terms of fuel consumption for the engine, because higher friction losses occur here. For this reason, it is expedient for the coolant to be brought to the operating temperature as quickly as possible at critical engine stages (Figure 31.10). This is especially the cylinder head and the entire crankshaft drive together with the cylinder sleeves. This can be achieved by stopping the coolant during the engine start or warmup phase, by transferring heat to the oil or coolant, for example, from the exhaust gas energy, or by using a heat accumulator.

Under-consumption [%]

31.4 Heat Management

NEDC, environment 25 °C

Vehicle Engine

Automatic transmission

Oil temperature at the start [°C] Figure 31.10 Underconsumption as a function of oil temperature (source BMW).

to the heater KFT DT

EGR

Coolant radiator

Low-temperature radiator

Kühlluft

EOC

Charge air Exhaust gas Main cooling circuit Low-temperature circuit

LLK

from the heater

Figure 31.9 Coolant circuit for optimized heat management (source MTZ). See color section page 1108.

Internal Combustion Engine Handbook | 1059

6606_Book.indb 1059

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 31 Energy Management in the Engine and Vehicle

Requirement-based engine cooling has as a further aspect active engine cooling using a variation of the cooling air inflow. The maximum cooling air requirement is, in addition to load conditions, another function of engine performance. The cooling air requirement is regulated using valve systems conditional on these parameters. According to BMW, an approximately 2% fuel consumption saving is possible with this method. In addition, this provision counteracts engine compartment cooling. This provision also has a positive effect on the drag coefficient and acoustics.

Bibliography

31-1. Liebl, J. “Wärmemanagement—ein weiterer Schlüssel zu Efficient Dynamics,” ÖVK Vortragsreihe, Graz, 2008. 31-2. Böhme, J., Fröhlich, G., Dornhöfer, R., and Grigo, M. “Der neue 1,8-l-TFSI-Motor im Audi A4,” ATZ/MTZ Extra, 2007. 31-3. Eifler, G., Burkard, M., and Kawert, F. “Demand-Controlled Engine Cooling Circuit,” MTZ, 2005 31-4. Edwards, S., Müller, R., Feldhaus. G., Finkeldei, T., and Neubauer, M. “The Reduction of CO2 Emissions from a Turbocharged DI Gasoline Engine through Optimised Cooling System Control,” MTZ, 2008. 31-5. Liebl, J. “Der BMW-Weg zur CO2-reduzierung,” 13. Internationaler Kongress Elektronik im Automobil, Baden-Baden, 2007. 31-6. Krist, S., Mayer, J., and Neuendorf, R. “Aerodynamik und Wärmehaushalt, Der neue BMW 5er,” MTZ/ATZ Extra, 2003. 31-7. 2009. “Thermoelektrik—Eine Chance für die Automobilindustrie,” edited by Jänsch, D., Expert Verlag: Renningen. 31-8. Dillann, G. et al. “Effizientes Fahren, Der neue BMW 7er,” ATZ Extra, 2008. 31-9. Schmidt, M. “Elektrische Energiemanagementstrategien zur CO2-Reduktion—Technische Voraussetzungen und Auswirkungen auf das Bordnetz, CO2—Die Herausforderung für unsere Zukunft,” ATZ/MTZ Energy Conference, Munich, 2007. 31-10. Hübner, W. and Lindemann, U. “Energiemanagement—Analyse und virtuelle Abbildung energetischer Zusammenhänge im Fahrzeug, CO2—Die Herausforderung für unsere Zukunft,” ATZ/MTZ Energy Conference, Munich, 2007.

bei Erhalt des thermischen Komforts, CO2—Die Herausforderung für unsere Zukunft,” ATZ/MTZ Energy Conference, Munich, 2007. 31-12. Liebl, J. “Energiemanagement—ein Beitrag zur effizienten Dynamik, Der Antrieb von morgen,” MTZ Engine Conference, Stuttgart, 2006. 31-13. Thumm, A. “Thermomanagement reduziert Verbrauch und Emissionen, Der Antrieb von morgen,” MTZ Engine Conference, Stuttgart, 2006. 31-14. Böhm, T. “Energiemanagement für Hybridantriebsstränge, Hybridfahrzeuge und Energiemanagement,” Braunschweiger Symposium, Braunschweig, 2006. 31-15. Treffinger, P. and Friedrich, E. “Unkonventionelle Nutzung von Abgaswärmeströmen im Fahrzeug, 8th Stuttgart International Symposium,” Automotive and Engine Technology, Stuttgart, 2008. 31-16. Steinberger, T. Wärmemanagement des Kraftfahrzeugs VI. ExpertVerlag: Renningen, ISBN 978-3-81-69-2820-1. 31-17. Warnecke, M., Schoemaker, M., Bank, D., and Soukhojak, A. “TESS— Wärmeenergiespeicher für Kraftfahrzeuge,” Fifth Emission Control 2010, Dresden, June 2010. 31-18. Albrecht, M. et al. “Auto-Start-Stopp-Funktion für Fahrzeuge mit Automatikgetriebe als Beitrag zur EfficientDynamics Strategie der BMW Group,” Sixth MTZ Specialist Conference: Der Antrieb von morgen. Hat der Verbrennungsmotor eine Zukunft? January 25, 2011 and January 26, 2011. 31-19. Bartosch, S. et al. “Abwärmenutzung in Antrieb von heute und morgen—Voith Abwärmenutzungssysteme,” Sixth MTZ Specialist Conference: Der Antrieb von morgen. Hat der Verbrennungsmotor eine Zukunft? January 25, 2011 and January 26, 2011. 31-20. Metzner, F.-T. et al. “Innovatives Thermomanagement am Beispiel des neuen Volkswagen Touareg,” Nineteenth Aachen Colloquium, October 4–6, 2010. 31-21. Neumeister, D. et al. “Thermomanagement von Hybrid- und Elektrofahrzeugen,” Nineteenth Aachen Colloquium, October 4–6, 2010. 31-22. Neugebauer, S. et al. “Effizient und dynamisch—die Entwicklung des AGR Thermoelektrischen Generators bei der BMW Group,” Twenty-Second International AVL Conference “Engine and Environment,” September 9, 2010 and September 10, 2010, Graz. 31-23. Sauer, J. et al. Fire and Ice—Wärmemanagement im Zeichen von Efficient Dynamics 31-24. Stotz, I. et al. “Prognose Thermomanagement,” Nineteenth Aachen Colloquium, October 4–6, 2010.

31-11. Heckenberger, T., Edwards, S., and Kroner, P. “Potenziale im Thermomanagement von Fahrzeugen zur Reduktion des CO2-Ausstoßes

1060 | Internal Combustion Engine Handbook

6606_Book.indb 1060

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

32 Forecast Automobiles have been around for more than 150 years and are almost entirely driven by reciprocating piston engines. Both gasoline and diesel engines developed at a rapid pace, and developmental progress is a long way from being exhausted. Upon closer examination, we find that the pace of development and the accompanied advancements at all levels have increased drastically in recent years. Competitors such as the Stirling engine, gas turbine, Wankel engine, steam engine, and even the electric motor with its associated energy storage devices have yet to have, and will fail to have soon, a serious chance of displacing the reciprocating piston combustion engine. Some time ago, the fuel cell and the electric drive system set about joining the race. To evaluate the fuel cell’s prospects, we should not compare them directly with today’s reciprocating piston engines, and this often occurs by mistake instead, and we need to examine the developmental potential of both systems. Reciprocating piston engines have potential for improvement in the areas of fuel consumption and pollutant emissions, performance and torque, driveline weight and required installation space (packaging), not to mention costs. However, the fuel cell has the additional problem that now there is no infrastructure for the supply of hydrogen and that the production of hydrogen as an energy supply from both fossil fuels and renewable energy sources is uneconomical and energy losses are very high. During electrolysis, including the subsequent compression of the hydrogen for instance, efficiency is less than 50%. Now we primarily address the potential fuel savings from competing gasoline and diesel engines.

32.1 Gasoline Engine The introduction of direct injection in diesel passenger car engines in 1989 with exhaust gas detoxification enlarged the diesel engine’s volumetric fuel consumption lead by further

15–20 vol% to approximately 35 vol% [32-6]. After direct injection was also increasingly introduced into series production for gasoline engines a few years ago, this procedure has yielded savings of up to 7% in fuel consumption in the extra urban driving cycle (EUDC) test, depending on the engine displacement. The combustion process now in series production primarily involves the wall-guided or air-directed method. Since 2006, jet-directed methods, which utilize multihole valves and increased injection pressures of more than 200 bar, are in series production. This allows fuel consumption improvements in the driving cycle of over 20 vol% (based on engines with intake manifold injection) thanks to charge stratification and consequently also lowers wall heat losses. Multiple injections stabilize the combustion process. This will lower the advantage in fuel consumption held by diesel engines with direct injection at partial load to just 10 vol%. At low partial loads, the gasoline engine with direct injection has already caught up and overtaken the diesel engine in terms of fuel consumption, in relation to mass. Further development steps, which positively influence fuel consumption, some of which are already in series production, include: •• Fully variable valve trains in series production deliver a consumption improvement of around 8–12%. The additional cost for mechanical systems is, however, approximately 10% [32-10]. Electromechanical and electrohydraulic systems have also been in development for some time. Camshafts may be dispensed with in electromechanical systems. However, they have not yet made the breakthrough. Electrohydraulic systems have been in series production for two years. •• Pulse turbocharging using an air pulse valve: Similar effects can be obtained as with fully-variable valve actuation, but with relatively less effort, and controlled intake manifolds can be dispensed with because of the increased charge over the entire speed range. Other functions such as hot and cold charging are possible. Pulse turbocharging is still in the preliminary development stage.

Internal Combustion Engine Handbook | 1061

6606_Book.indb 1061

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 32 Forecast

•• Cylinder shutoff for engines with a high number of cylinders drastically reduces fuel consumption and pollutant emissions. The consumption saving achieved thus is 20 vol% at low load. •• Crankshaft starter generators allow consumption savings in the 10–15% range. •• Variable displacement delivers a consumption reduction of 8–20 vol% in the partial-load range. Engine internal mechanical systems are in development. •• Fuel consumption can be reduced by approximately 5 vol% with a dual spark plug engine and sequential dual ignition. •• High-frequency ignition, plasma ignition, and multiple spark ignition reduces CO2 output by up to 10 vol%. •• Variable geometric compression ratio in conjunction with supercharging yields up to 30 vol% reduction in fuel consumption. •• Downsizing by high supercharging at 25-bar average pressures, while simultaneously decreasing the displacement by reducing the number of cylinders, for example, from six to four or from four to two cylinders in conjunction with downspeeding, and high supercharging can be implemented particularly effectively with two-stage turbocharging. Such approaches also reduce the engine weight and required space (packaging). The potential savings in fuel consumption are 20–25 vol%. •• Turbocharging as a fuel consumption approach, that is, with a relatively small supercharger allows fuel consumption to be reduced by approximately 15 vol%. •• Reducing friction loss. Over the last 20 years, friction loss has been reduced by a remarkable 25 vol%. Nevertheless, substantial potential remains as the theoretical case of a frictionless gasoline engine would yield an approximately 20 vol% gain in efficiency. If we could attain one half of this, fuel consumption would be reduced by approximately 10 vol%. The use of rolling bearings instead of plain bearings and new methods for treating the sliding surfaces with nanocoatings, for instance, look like promising approaches. •• Energy and thermomanagement with hot cooling under partial load. Savings of 5–10 vol% are conceivable. •• Cooling the exhaust gas by integrating the exhaust manifold in the cylinder head shortens the warmup time and renders enrichment of the mixture unnecessary at full load. The passenger cell also heats up quicker thanks to the quickerheating coolant, which likewise reduces fuel consumption. •• Fuel and lubrication oil optimization can yield potential savings of 5–10 vol%. •• Hybridization: Microhybridization (i.e., start/stop systems) reduces fuel consumption by around 4 vol%, mild hybridization (belt-drive starter generator) by around 10 vol%, and strong hybridization (i.e., full hybrid with power-split) by around 14 vol%. •• Electrically driven auxiliary units. The electrical drive systems must not remain limited to engine units.

•• Waste heat utilization: Exhaust gas heat has the greatest potential. It can be utilized with the aid of a thermoelectric generator and the Rankine process among others. The thermal generator works according to the Seebeck effect [32-9] and delivers the generated electrical energy to the battery. The increase in fuel consumption is around 6 vol%. This method is anticipated to enter series production in around three to five years. •• Mechanical supercharging should be avoided because, despite also having advantages over exhaust gas turbocharging, it does, however, perform worse in terms of fuel consumption during comparison tests. The reasons behind this are the high power loss because of the mechanical drive of the turbocharger and the thermodynamic benefit of the exhaust gas for the exhaust-driven turbocharger. •• Homogeneous charge compression ignition (HCCI) methods have a potential of 15 vol%. •• Coasting: Decoupling the combustion engine from the powertrain during the coasting operation can double the coasting distance. This leads to a fuel consumption reduction of 10% [32-13]. If all the cited fuel consumption improvements were implemented, the gasoline engine would shortly be able to produce fuel! On a more serious note, it is clear that consistently and rationally combining complementary measures could likely reduce fuel consumption by 40 vol% in ten to fifteen years. This does not include savings in fuel consumption from advances in transmission technology. Depending on the basis of comparison, this could lead to a further 15 vol% reduction in fuel consumption. This is also applies to the diesel engine.

32.2 Diesel Engine The potential to lower fuel consumption in a diesel engine is much lower because direct injection has already found its way into nearly every engine with a consumption reduction of 15–20 vol%. However, there are still many possibilities for improvement. •• Variable valve actuation could also be adopted for diesel engines. The gain in fuel consumption would be less, because diesel engines use quality control and therefore are already dethrottled under partial load. The reduction in fuel consumption would be 5–10 vol%. •• Pulse turbocharging using an air pulse valve. The potential savings in fuel consumption are also substantial in this instance at around 20 vol%. It would allow the glow element to be omitted or the compression ratio to be reduced, which permits a higher supercharging ratio and, hence, performance. When using hot charging and cold charging, pollutant emissions are also reduced. •• Cylinder shutoff is possible; the reduction in fuel consumption is, however, less than in gasoline engines and cyclic irregularity is greater.

1062 | Internal Combustion Engine Handbook

6606_Book.indb 1062

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

32.3 Concluding Observations

•• Crankshaft starter generators have effects similar to those in gasoline engines. •• A variable geometric compression ratio would not be worthwhile, as the latest diesel engines are already operated in the optimum efficiency range (ε < 16). •• Reducing friction loss: Friction loss lowers efficiency in diesel engines by around 26 vol%. Cutting this in half would reduce fuel consumption by 13 vol%, which would be slightly greater than with gasoline engines. Measures similar to those used for gasoline engines are conceivable. •• Downsizing is possibly the same as for gasoline engines. In development are peak pressures during combustion of more than 200 bar in passenger cars and 250–300 bar in trucks for the purpose of greater specific performance. To counter the weakness when setting off, electrically supported turbochargers are being developed that increase the average pressure at setting off speed. Variable twin turbocharging (high supercharging) for instance, which first entered series production in diesel engines in autumn 2004, is another promising approach. The consumption gains would be 20–25 vol%. •• Vario valves (partial-load/full-load nozzle): Optimizes the bore hole diameter for the injections jets between partial load and full load. The consumption benefit is greater than 5 vol%. This delivers dramatic improvements in terms of pollutants, combustion noise, torque, and specific power. However, series production is not yet in sight because of the considerable complexity. Instead injection pressures are increasing all the time, which could make vario nozzles redundant. Injection pressures of 2500–3000 are already under development and are anticipated sometime between 2013 and 2020. The consequences would be: higher specific power, significant improvements to fuel consumption, European Union (EU) 6 without DeNOx catalytic converters, and component protection by limiting the cylinder pressure and the exhaust gas peak temperature. •• Energy and thermal management: Hot cooling plus fuel and lubrication optimization deliver fuel consumption reductions of 5–10 vol%. •• Thermodynamic optimization could reduce fuel consumption by 5 vol%. •• Water injection reduces NOx and soot emissions, and tends to diminish fuel consumption. •• Turbocompound supercharging: By exploiting the exhaust energy better than pure exhaust gas turbocharging, fuel consumption can be lowered by 5 vol%, and power can be increased by 10 vol%. •• Hybridization: The absolute lowest fuel consumptions can be achieved with the hybridization of the diesel engine. However, the costs are significantly higher than those for gasoline hybrids from a present-day perspective. Nevertheless, all well-known automobile manufacturers are understandably working on this system to solve the problem of cost optimizations.

•• Fuel and lubrication oil optimization, drastic shortening of the warmup time, electrically driven auxiliary units, waste heat utilization, and mechanical supercharging behave in a similar way in diesel engines as they do in gasoline engines. •• Transition from aluminum to steel pistons. Consumption savings achieved by lowering the weight and reducing friction are 2–5%.

32.3 Concluding Observations Even taking into account the imprecision of the above assumptions, we can still conclude that the fuel consumption of gasoline and diesel engines will grow closer in the future under the evaluated conditions. It can be expected that in the future, the diesel engine will have a volumetric fuel consumption advantage of only approximately 15–20%. The result of this will be that the current extraordinary significance of the diesel engine will recede somewhat again, because its production costs are greater than those of the gasoline engine by a factor of 1.5–2. The much greater need for platinum in pollutant after treatment could turn out to be a competitive disadvantage for the diesel engine in the long term if it cannot find a replacement material. The savings in fuel consumption for gasoline and diesel engines in mobile applications will be so considerable that in comparison with fuel cells that naturally also have a certain amount of developmental potential, the lead will continue to widen. The relatively high production costs of the fuel cell needs to be taken into account here, which are still around 10–20 times greater now. The argument that fuel cells are free of pollution in contrast to reciprocating piston engines no longer applies to gasoline engines because it is already solved, and the problem will be solved at the latest for diesel engines with the introduction of homogeneous and cold combustion (e.g., HCCI), and further-developed fuels [32-1]. The extensive introduction of the soot particulate filter has also made a significant contribution. Moreover, research work is now being conducted at Volkswagen under the term combined combustion system (CCS) (Figure 32.1) [32-7] and at Mercedes-Benz under the name Diesotto (diesel/otto) [32-8], with the goal of combining the advantages of the diesel engine with those of the gasoline engine. Finally, it would seem appropriate to point out one potential measure, which, as far as the author is aware, has, astonishingly, not been identified or has only been poorly identified. Increasing the oxygen content of the fuel during combustion could enable completely new developments, which would give the combustion engine a new push in development terms. Extracting nitrogen oxide from the air and/or fuels with a high oxygen content and or water injection could be expedient [32-11]. It can be safely deduced from all these examples that the reciprocating piston engine has another exciting and successful decade still to come [32-5].

Internal Combustion Engine Handbook | 1063

6606_Book.indb 1063

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Chapter 32 Forecast

Self-igniting in some cases

High-pressure injection

Supercharging

Supercharging Direct injection

Whirl chamber

Intake manifold injection

Bibliography

32-1. Mayer, K.P. 2002, Verbrennungsmotor versus Brennstoffzelle—Potenziale und Grenzen für den Automobilantrieb, Proceedings of the Thirteenth International AVL Congress, September 6, 2001 and September 7, 2001. 32-2. Takimoto, M. “Ausblick auf zukünftige Fahrzeugantriebe und deren Herausforderung für Toyota,” 10th Aachen Colloquium Automobile and Engine Technology, 2001. 32-3. Heil, B., Weining, H. K., Karl, G., Panten, D., and Wunderlich, K. “Verbrauch- und Emissionen-Reduzierungskonzepte beim Ottomotor,” MTZ 62 (11):900–915, Part 1, 2001. 32-4. Heil, B., Weining, H. K., Karl, G., Panten, D., and Wunderlich, K. “Verbrauch- und Emissionen-Reduzierungskonzepte beim Ottomotor,” MTZ 62(12):1022–1035, Part 2, 2001. 32-5. van Basshuysen, R. and Schäfer, F. 2006. Lexikon Motorentechnik, 2. Auflage, Vieweg Verlag: Wiesbaden. 32-6. van Basshuysen, R. 2008. Gasoline Engine with Direct Injection, Processes × Systems × Development × Potential. Vieweg+Teubner: Wiesbaden.

ne gi

Direct injection

(High)pressure injection

en

Di es

el

Homogeneous in some cases

e lin so Ga

en gi ne

CSS

Figure 32.1 From diesel and gasoline engines to the CCS [32-7]. This is the long-cherished dream of development engineers. The measures described will make it possible to fall below the limit of 95 g/km CO2 required by EU law makers before 2020.

32-7. Pfeil, J., Fischer, J., Kettner, M., and Spicher, U. “Influence of Spray Propagation on the Gas Phase in Direct Injection Gasoline Engine,” Tenth International Symposium on Flow Visualization, 2002. 32-8. Herden, W. and Vogel, M. 2004. Perspektiven alternativer Zündsysteme. Diesel-Benzindirekteinspritzung. Expert-Verlag: Essen. 32-9. 2009. “Thermoelektrik,” edited by Jänsch, D et al. “Volume 94 of Haus der Technik Fachbuch Reihe Technik,” Expert Verlag: Renningen. 32-10. Flierl, R. et al. “Univalve—A Fully Variable Mechanical Valve Lift System for Future Internal Combustion Engines,” MTZ 72(5):380–385, 2011. 32-11. European Patent Specification EP 0 917 620 B1. patent:DE19632179, “Brennkraftmaschine mit erweitertem Arbeitszyklus.” Clercq, Ludo De Ir, Brussels Belgium, 1998 32-12. Königstein, A. “Der Ottomotor der Zukunft. TAE,” Nineth Symposium Ottomotorentechnik, December 2, 2010 and December 3, 2010. 32-13. Müller, N. et al. “Coasting—Next Generation Start/Stop Systems,” MTZ 72(9):644–649, 2011.

1064 | Internal Combustion Engine Handbook

6606_Book.indb 1064

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Free torque 2. Order Engine Long Axis

Force 2. Regulation on the motor bearing

The page number given in each figure caption refers to the location in the text where the figure is discussed.

Full load without AGW Full load with AGW without height offset Full Load with high offset AGW

Eradication Point

Thrust without AGW Thrust with AGW without height offset Thrust with height offset AGW

Eradication Point

Thrust without height offset Thrust with height offset

Full load without height offset Full load with height offset

Eradication Point

Eradication Point

Engine Speed [min–1]

Engine Speed [min–1]

Figure 6.46 System behavior of balance shafts with and without height offset [6-12]. See page 70.

Internal Combustion Engine Handbook | 1065

6606_Book.indb 1065

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 6.68 Mechanisms for turning the cylinder head (source: MOT). See page 80.

14:1

Figure 6.70  Variable compression ratio conrod (source: MTZ/ Pischinger). See page 81.

8:1

Figure 6.69 Longitudinal section through the Saab Variable Compression engine (source: MOT). See page 80.

1066 | Internal Combustion Engine Handbook

6606_Book.indb 1066

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Vertical shift of cylinder block

Vertical momentary combustion chamber due to secondary piston

Piston with adjustable compression height

Conrod bearing in eccentric crank pin

Eccentric crankshaft bearing (VCR principle)

Power transmission with gear-wheel drive

Second movable articulation point of the conrod (1)

Second movable articulation Second movable articulation point of the conrod (2) point of the conrod (2)

[°C]

Figure 7.8 Temperature distribution at a piston for a gasoline engine. See page 86.

301.0 290.0

Figure 6.71 Schematic of a variable compression (source: MOT). See page 81.

[°C]

375.0 359.0

280.0

343.0

269.0

327.0

259.0

311.0

248.0

295.0

237.0

279.0

227.0

263.0

216.0

247.0

206.0

231.0

195.0

215.0

184.0

199.0

174.0

183.0

163.0

167.0

153.0

151.0

142.0

135.0 Figure 7.9 Temperature distribution at a piston with cooling channel for a diesel engine. See page 86.

Internal Combustion Engine Handbook | 1067

6606_Book.indb 1067

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 7.33 Stress analysis for a conrod with an angular split, with a trapezoidal small end (half model, Federal-Mogul). See page 98.

Figure 7.71 Pressure differential map, intake manifold pressure— atmospheric pressure of a turbocharged, quantity-controlled 1.8-L passenger SI engine with fuel direct injection. See page 128.

Figure 7.72 Blow-by gas flow map, intake (without ventilation) of a turbocharged, quantity-controlled 1.8-L passenger SI engine with fuel direct injection. See page 128.

Figure 7.94 Section of the water jacket for coolant flow simulation [7-64]. See page 146.

1068 | Internal Combustion Engine Handbook

6606_Book.indb 1068

1/19/16 9:02 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

high

low

Figure 7.95 Strength analysis at the cylinder head [7-67]. See page 147.

Figure 7.117 Cylinder head of the V12 BMW engine with direct injection and Valvetronic. See page 159.

780.0

1.0

702,4

0.9

624.8

0.8

547.2

0.7

469.6

0.6

391.9

0.5

314.3

0.4

236.7

0.3

159.1

0.2

81.5

0.1

3.9 1020.7 1306.6 1592.4 1878.3 2164.1 2450.0 2735.9 3021.7 3307.6 3593.4 3879.3 Rotational speed [U/min]

Normalized amplitude

Frequency [Hz]



0.0

Figure 7.127 Campbell diagram of an engine run-up. See page 164.

Internal Combustion Engine Handbook | 1069

6606_Book.indb 1069

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 7.148 Concept representation of the UNIAIR system. See page 173.



solenoid status

Solenoid Switching

0

20

40

LVO

60

80 100 120 angle camshaft [°]

140

160

180

200

EVC

UniAir Valve Control Modes

10

valve lift [mm]

8 6 4 2 0

0 full lift

20

40

60

80 100 120 angle camshaft [°]

140

160

180

200 no lift

Figure 7.149 Valve lift modes, UNIAIR system. See page 173.

1070 | Internal Combustion Engine Handbook

6606_Book.indb 1070

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Belt force [N]

Color Section

1400 1200

Series

1000

ID²

800 600 400 200 0 800

1200 1600 2000 2400 2800 3200 3600 4000 4400 4800 5200 5600

Angular displacement [°Cam]

Crankshaft speed [1/min]

2 1,8 1,6 1,4 1,2 1

Series

0,8 0,6

ID²

0,4 0,2 0 800

1200

1600

2000

2400

2800

3200

3600

4000

4400

4800

5200

5600

Crankshaft speed [1/min]

Figure 7.156 Effects of the ID² system. See page 176.

Figure 7.193 Temperature distribution inside a valve seat insert at the exhaust port. See page 195.

Internal Combustion Engine Handbook | 1071

6606_Book.indb 1071

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Qth in Ltr\min at 1000 min-1

Color Section

Figure 7.209 Pump sizes corrected by the translation, as function of the engine power in SI engines. See page 202.

Engine power in Kw

2 4

3 1

Figure 7.211 Schematic pump display. See page 203.

Calculated

Measured

Pulsation Figure 7.215 Measured and calculated pressure progression in the crescent of a Trochocentric pump. See page 205.

1072 | Internal Combustion Engine Handbook

6606_Book.indb 1072

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

System pressure [bar] Theoretical delivery Eff. delivery Engine consumption System pressure

Figure 7.223 Engine consumption curve, system pressure, and theoretical and effective displacements. See page 208.

Engine speed [1/min]

System pressure [bar] Theoretical delivery Engine consumption System pressure

Eff. delivery Engine consumption, two-stages System pressure two-stage regulation

Engine speed [1/min]

Figure 7.227 Two-stage regulation. See page 210.

Internal Combustion Engine Handbook | 1073

6606_Book.indb 1073

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Power consumption, oil pump [W]

Color Section

Power consumption, continuous pump Power consumption, regulating pump Power savings Potential at pressure decrease For the consumption cycle relevant speed range

Figure 7.229 Potential of power savings. See page 211.

Engine speed [min -1 ]

NEFZ- Geschwindigkeitsverlauf

NEFZ speed progression 0 100° 100°

140

90° 90° Außerstädtischer Extra-urban Fahrzyklus driving cycle

100

80° 80° 70° 70° 60° 60°

80 StadtfahrStadtfahrStadtfahrStadtfahrUrban Urban Urban Urban driving 1 driving zyklus cycle 1 zykluscycle 2 zykluscycle 3 zykluscycle 4 2 driving 3 driving 4

60

50° 50° 40° 40° 30° 30°

40

20° 20° 20 0

Temperatur [°C] [°C] Temperature

Speed [km/h] Geschwindigkeit [km/h]

120

10° 10° 0

200

400

600

Zeit [s] [sec] time

Geschwindigkeit [km/h]

800

1000

Temperatur [°C]

Figure 7.237 NEDC cycle according to 80/1268/EEC. See page 214.

Five-speed transmission 5-Gang-Getriebe

Six-speed transmission 6-Gang-Getriebe

4,11

4,02

Axle ratio Gear I Gear II Gear III Gear IV Gear V Gear VI

3,45

Ratio [--]

0° 1200

3,54

2,38

2,03

1,59

1,34

1,17 1,0

1,0 0,80

0,80

Figure 7.238 Transmissions in various vehicles by different Original Equipment Manufacturers (OEMs); mean values displayed in circles. See page 215.

1074 | Internal Combustion Engine Handbook

6606_Book.indb 1074

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Vehicle measurement, continuous pump, highway driving Pressure downstream filter

Temperature downstream filter

Temperature oil sump

Pressure [bar]

Temperature [°C]

KW Rotational speed [rpm]

Engine speed

time [s]

Vehicle measurement, regulating pump, highway driving Pressure downstream filter

Temperature downstream filter

Temperature oil sump

Pressure [bar]

Temperature [°C]

KW Rotational speed [rpm]

Engine speed

time [s]

Figure 7.239 Progression of oil sump temperature during a highway tour with a regulating pump (top) and a constant pump (bottom) of the same nominal dimension. See page 215.

Basis continuous pump Regulation to pressure level Regulation to two pressure stages Map control Figure 7.240 Achievable fuel savings in NEFZ cycle. See page 216.

Internal Combustion Engine Handbook | 1075

6606_Book.indb 1075

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Ground gear set 90°C oil temp.

Sintered gear set 90°C oil temp.

Figure 7.256 Calculated flow within the pump housing of a vane pump. See page 224.

Figure 7.249 Differences in the structural noise between a sintered (top) and a ground (bottom) toothing. See page 221.

Kavitatonsgebiet Cavity area

Figure 7.257 Calculated distribution of flow velocities in the gear set of an external gear pump. See page 224.

Figure 7.251 Speed and frequency analysis of the oil pressure pulsation of a gear pump. See page 222.

1076 | Internal Combustion Engine Handbook

6606_Book.indb 1076

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Mass/cylinder [g]

Color Section

Solid material Cylindrically hollow Profile hollow Assembled camshaft

Cylindrically hollow (stepped)

Cylinder spacing [mm]

Figure 7.274 Reducing masses in camshafts. See page 233.

Cam angle

Figure 7.276 Valve stroke, speed, and acceleration plotted against the cam angle for a roller cam follower valve train with hydraulic valve lifters. See page 235.

Cam angle

Figure 7.277 Theoretical valve stroke and Hertzian pressure (kinematic and dynamic) for a roller cam follower drive train with hydraulic valve lifters. See page 235.

Theoretical valve stroke Theoretical valve speed Theoretical valve acceleration

Theoretical valve stroke Kinematic pressure Dynamic pressure

Profile hollow

Internal Combustion Engine Handbook | 1077

6606_Book.indb 1077

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Cam angle Theoretical valve stroke Cam contour radius Hydrodynamically-effective speed

Cam angle Theoretical valve stroke Kinematic contact force Dynamic contact force, n = 6000 1/min

Figure 7.278 Cam contour, theoretical valve stroke, and hydro-dynamically effective speed plotted against the cam angle at contact between cam and flat tappet. See page 236.

Figure 7.279 Theoretical valve stroke, kinematic contact force, and dynamic contact force plotted against the cam angle for a rocker arm valve train with hydraulic valve lifters. See page 237.

1078 | Internal Combustion Engine Handbook

6606_Book.indb 1078

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

hoch/high

niedrig/low

Figure 7.356 Comparison of distribution of sealing pressure: Left: a stopper with constant thickness. Right: the optimized stopper with variable thickness. See page 278.

Figure 7.302 Reduction of belt forces, compared to a round crankshaft sprocket. See page 248.

Figure 7.303 Reduction of vibration amplitudes of the camshaft, compared to a round crankshaft sprocket. See page 248.

1 2 3 4 5 6 7 8 9 10

Oil screen Delivery pump Coarse filter Overpressure valve Rocker arm shaft Transmission Coupling Oil injection nozzles Fine filter Crankshaft

Figure 8.42 Single-cylinder engine with oil circuit. See page 340.

Figure 7.311 Belt tensioning systems for starter–alternator drives. See page 251.

Internal Combustion Engine Handbook | 1079

6606_Book.indb 1079

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 8.88 Engine section 125/200. See page 356.

Vmax [km/h] 3.2 3.1

< 130

United Nations

2.2 < 115 < 100 < 60 < 50

2.1 1.1

1.3 1.2 Displacement

Figure 8.107 Classification: WMTC cycle. See page 362.

6000

Figure 8.139 Compact valve train. See page 378.

Overall force over the crankshaft angle (V4/75 Var_30)

6000 5000

Overall force [N]

Overall force [N]

5000 4000 3000 2000

4000 3000 2000 1000

1000 0

Overall force over the crankshaft angle (Inline 4)

0

FMAG front

90

180 270 Crankshaft angle [degrees]

FMAG rear bottom

FMAG rear top

360

0

0

FMAG front

90

180 270 Crankshaft angle [degrees]

FMAG rear bottom

360

FMAG rear top

Figure 8.123 Engine suspension stresses. See page 372.

1080 | Internal Combustion Engine Handbook

6606_Book.indb 1080

1/19/16 9:03 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Coolant tank

Water temperature sensor activates monitoring beeper and limp home mode when temperature exceeds 100 °C (212 °F)

Blend hose from cylinder head to coolant tank Water flows to oil cooler

Waterpump housing including thermostat (operates at 87 °C/188 °F) and waterpump impeller Ride plate (operates as radiator)

Water flows to ride plate

Oil cooler Water return from oil cooler Water return from ride plate

Figure 8.140 Closed coolant circuit. See page 379.

System with integrated water separator

Figure 8.142 Intake system with integrated water separator. See page 380.

Figure 8.143 Surface velocity of the plastic intake collector. See page 381.

Figure 10.71 CS shifter as a chain timing device for the Porsche Boxter [10-39]. See page 450.

Internal Combustion Engine Handbook | 1081

6606_Book.indb 1081

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Connector Control device

Anchor plate

Figure 10.74 CS shifter arrangement in a six-cylinder engine according to the swing motor principle [10-34]. See page 451.

Camshaft

Forked lever

Electro-magnet

Valve spring

Background

Eccentric shaft

Work curve Fork lever in detail

Roller cam follower

Valve

Valve with different strokes

Figure 10.104 Design of the UniValve system [10-55]. See page 463.

Figure 10.115 Single actuator of the electromechanical valve train. See page 468.

10 Target curve

9

Dimensioning

Valve stroke [mm]

8

Measurement

7 6 5 4 3 2 1 0

0

20 40 60 80 100 120 140 150 180 200 220 240 260 280 300 320 340 360 Opening time [°CS]

Figure 10.108 Opening characteristic of the valve. See page 465.

1082 | Internal Combustion Engine Handbook

6606_Book.indb 1082

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Valve mount Impulse artery valve Intake manifolds

Speed [m/s]

Collector

978,71

Figure 10.137 Pulse turbocharger valve with aspiration pipe and valve mounting. See page 483.

87,74 76,77 65,80 54,84 43,67 32,90 21,94 10,97 0,00

Figure 10.117 Simulation of the turbulence with different valve lifts of 1 and 10 mm. See page 470.

Throttle Valve

Collector

Figure 10.138 Integrated aspiration module. See page 483.

Injector

Impulse charger Valve

Figure 10.133 Position of the collector, injector pulse turbo loader valve in the intake module. See page 480.

Internal Combustion Engine Handbook | 1083

6606_Book.indb 1083

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Turbocharger turbine Combustion chamber

Fuel

4 Throttle

3 2

Turbocharger compressor

1 Oil supply

Electrically driven compressor

Combustio chamber air

High pressure connection

Figure 11.47 Schematic design of a turbocharger test rig. See page 508.

Fuel return High-pressure control valve (PCV) Fuel feed pump (ITP)

High-pressure pump element





Volume flow control valve (VCV)

Figure 12.40 High-pressure pump, 1 Fuel resupply pump; 2 Volumetric flow control valve; 3 High-pressure pump element; 4 High-pressure control valve; 5a Fuel feed; 5b Highpressure connection; 5c Fuel return flow. See page 546.

Figure 12.45 Common-rail piezo injector. See page 548.

1084 | Internal Combustion Engine Handbook

6606_Book.indb 1084

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

3

Compensation reservoirs 4 1

5 2 6

Fill-level limiting valve

1 High pressure connection 2 Fuel return 3 Electrical connection to the engine control unit (ECU) 4 Piezo-Actuator 5 Valve piston 6 Valve head 7 Control piston 8 Nozzle needle 9 High-pressure chamber nozzle 10 Nozzle injection holes

7

8

Figure 12.64 Tank with an internal venting system. See page 560.

9 10

Figure 12.46 Sectional view of a piezo injector. See page 549.

Piezo actuator not controlled

Piezo actuator controlled 7

4

4

1

2

2 5

8 F1x

F1 3 6

1 2 3 4 5 6

High-pressure inlet Control chamber High-pressure chamber Valve head Fuel-return Nozzle needle

6

F2

F2x

7 Piezo-Actuator 8 Valve piston F1 Force on the control piston F2 Force acting on the nozzle needle

Figure 12.47 Injector function. See page 549.

Internal Combustion Engine Handbook | 1085

6606_Book.indb 1085

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Returns

Fine filter

Radiator

Common Rail

Flow

High-pressure pump

Fuel pump

Engine Tank

Ejector pump

Figure 12.65 Principle of a diesel fuel supply system for common rail. See page 561.

Engine control unit

Electronic Fuel Control Unit

Fuel pump with EC-Motor



Figure 12.72 Principle of an on-demand regulated gasoline supply system. See page 564.

Figure 14.23 2D section through a shear layer calculated with LES and DNS [14-33]. See page 603.

1086 | Internal Combustion Engine Handbook

6606_Book.indb 1086

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Catalytic converter temp

Figure 15.46 Temperature windows of different NOx exhaust gas treatment systems. See page 639.

3 separate areas of vaporized fuel in

Striation technique

1 large area at 1000 bar

Figure 15.55 Pressure chamber results of a multihole nozzle at different injection pressures under constant chamber conditions—constant injection mass for all pressures. See page 643.

Internal Combustion Engine Handbook | 1087

6606_Book.indb 1087

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Injection process Injection rate [-]

Combustion

Mixture formation

Injection

single

Ignition stable Double injection Triple injection

double

Inflammation characteristics

Variation ZZP

triple

12 °CA

Analysis Fuel conversion

Fuel conversion 8 °CA after ZZP

Air ratio [-] rich

1.0

fast inflammation

6 °CA

slow inflammation

lean

Acceptance time ZZP

Acceptance time 8 °CA after ZZP

Ignition timing [°CA after ZOT]

Figure 15.58 Multi-injection in stratified operation at 2000 min , pmi = 3.0 bar (simulation results) [15-52]. See page 646. –1

1. Injection

2. Injection

AnteilPortion der Zweiteinspritzung of secondary injection an derofGesamtmenge total volume in in %%

-360

TDCFiring

35

360

Einfacheinspritzung: Single injection 33 °CA 33 °KW 26.99 26.36

28.37

30

26.97

28.38

28.04

28.59

27.00

25 30.07

20

29.24

27.81 28.00 28.40

15

28.90

31.98

10

30.09

30.36 33.08

30.30

50 % conversion point °CA after TDCHD 5

80

70

60

50

40

30

20

Time of secondary / °CA /before TDC Zeitpunkt der zweiteninjection Einspritzung °KW vor OTHD HD

10

0

Figure 15.59 Development of the 50% fuel conversion point at a variation of the fuel mass distribution in a homogeneous and stratified component and variation of the injection timing of the stratified injection at constant full load (n = 1500 min–1, pme = 19.5 bar, and ignition permanently at the knocking limit). See page 646.

1088 | Internal Combustion Engine Handbook

6606_Book.indb 1088

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 15.65 Sectional view of the 1.2-L three-cylinder two-stroke engine by Orbital [15-84]. See page 652.

Load case 1 Rotational speed: 700 U/min

Load case 2 Rotational speed: 1700 U/min

Load case 3 Rotational speed: 4000 U/min

Ambient air: 90°C / approach flow: 0.5m/s Figure 16.7 Virtual validation using simulation—analysis of different load states. See page 660.

Internal Combustion Engine Handbook | 1089

6606_Book.indb 1089

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Speed sensors

Position sensors

Position sensors

Cover

Internal plug Insertable seal and silicone gel

Flexible printed circuit board

LTCC substrate

Alu floor plate

Pressure regulator, contact bar Vehicle plug

Integrated pressure sensors

PCB cover

Figure 16.16 Main components of a modern MTM. See page 665.

communication cycle dynamic segment

channel 1

static segment

1

2

channel 2

4

3 C1

A1

5

7

6 E1

A2

D1

12 9 8 E2 D2 C2 t

1

2 A4

4

3 B1

C1

5 B2

7

6 A2

E1

8 9 10 F1 C2

A3 t

Figure 16.28 Communication cycle of the data transfer. See page 673.

1090 | Internal Combustion Engine Handbook

6606_Book.indb 1090

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Service life, energy density

Color Section

Figure 17.24 Development of battery systems from conversion in lead-acid to intercalation in nickel-metal hydride and Li ions. See page 706.



Safety

Lead-acid

- Low costs but reduced service life expectation - Average overall performance

Costs

Service life

Availability

Output

DLC

- High service life expectation - High costs, low energy retention

NiMH

Process capacity

Cold Start

- Good overall performance

Li-Ions

Environment Lead-acid

1000000

DLC

Energy

Figure 17.25 Assessment of various storage systems for the use in motor vehicles. See page 707.

NTC resistor temperature characteristics

100000

Resistance [Ohm]

Li-Ions

NiMH

- Good overall performance, in conjunction with high energy densities

type A type B type C type X2 type X1

10000 1000 100 10

–50

1

0

50 100 temperature [°C]

150

200

Figure 18.1 Typical characteristics of temperature sensors (NTC). See page 711.

Figure 20.5 CFD simulation of the coolant circulation in the front end of a passenger car. See page 737.

Internal Combustion Engine Handbook | 1091

6606_Book.indb 1091

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Cumulative HC emissions [g]

0.6

Raw emissions 400 cpsi, 0.05 mm 600 cpsi, 0.04 mm 800 cpsi, 0.03 mm 1000 cpsi, 0.025 mm

0.5 0.4 0.3 0.2 0.1 0 0

20

40

60

80

Figure 21.57 Cumulative hydrocarbon emissions during the first 100 seconds of the FTP cycle (catalytic converter dimensions: ∅ 98.4 × 74.5 mm) [21-27]. See page 777.

100

Time [s]

0.3 FTP-72

Cumulative THC emissions [g]

0.25

US-06

0.2

0.15

0.1

0.05

0 Ceramic 3.8 ltr.

TS metallite 2.8 ltr.

LS metallite 2.7 ltr.

0.2

Emissions [g/km]

Figure 21.63 Reduction in the total carbon emissions of a 2.7-l V6 diesel engines with various catalytic converter systems [21-30]. See page 779.

LSPE metallite 2.5 ltr.

HC CO/10 NOx

0.15

0.1

0.05

0 unheated

heated

Figure 21.69 Comparison of the emission results (heated/ unheated) in the Euro II-Test for the BMW. See page 781.

1092 | Internal Combustion Engine Handbook

6606_Book.indb 1092

1/19/16 9:04 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 24.4 Distribution of breaking load cycles of an exhaust manifold. See page 878.

Figure 24.7 Optimization of the geometry of a connecting rod. Left: plot of stress in initial state. Right: plot of stress in optimized state. See page 879.

Figure 24.5 Distribution of acoustic velocities on an engine surface. See page 878.

Internal Combustion Engine Handbook | 1093

6606_Book.indb 1093

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

CAD model

Structure of the FE model

Mechani- OperaThermo- Thermal tional dynamics FE calcu- cal FE lation calculation strength

Report

Figure 24.8 Partial processes of piston calculation. See page 880.

Figure 24.12 Temperature of an aluminum piston for an SI engine in a passenger car at rated capacity. See page 882.

Figure 24.13 Temperature of an aluminum piston for a diesel engine in a passenger car at rated capacity. See page 883.

1094 | Internal Combustion Engine Handbook

6606_Book.indb 1094

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 24.14 Temperature of an steel piston for a diesel engine in a truck at rated capacity. See page 883.

Vertical forces SI piston 75 Fz (pressure) 60

Fz (total) Fz (mass)

Power [kN]

45

30

15

0

Intake

Compression

Work

Exhaust

-15 0

60

120

180

240

300

360

Crank angle [°]

420

480

540

600

660

720

Figure 24.15 Vertical piston force of an SI engine piston (maximum output operating point). See page 884.

Internal Combustion Engine Handbook | 1095

6606_Book.indb 1095

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Lateral forces SI piston 12 Lateral (2000 rpm) 10

Lateral (5000 rpm) Lateral (7000 rpm)

8

Ignition pressure [bar]

Lateral force [kN]

Ignition pressure (5000 rpm) 6 4 2 0 -2 -4

Intake

Compression

Work

Exhaust

-6 0

60

120

180

240

300

360

420

Crank angle [°]

480

540

600

660

720

Figure 24.16 Piston-side force of an SI engine piston, in dependence on the speed. See page 884.

Figure 24.17 Thermal deformation of a diesel piston (fifty times enhanced representation). See page 885.

Figure 24.18 Thermal stresses (third principal stress) before and after relaxation. See page 885.

1096 | Internal Combustion Engine Handbook

6606_Book.indb 1096

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Gas force

Gas force

Cylinder wall pressure-end Section, running plane

Section, pin plane

Figure 24.19 Mechanical Von Mises stress with twenty-five times scaling of mechanical deformation at maximum ignition pressure (CDC). See page 886.

Gas force

Figure 24.20 Von Mises stress in the wristpin with twenty-five times scaling of mechanical deformation at maximum ignition pressure (CDC). See page 886.

Figure 24.21 Mechanical circumferential stress in the recess of a diesel piston. See page 887.

Internal Combustion Engine Handbook | 1097

6606_Book.indb 1097

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

log S

200°C 300°C 400°C

0

1

2

3

4

5

6

log N

7

8

Figure 24.22 Wöhler diagram. See page 887.

 Figure 24.23 Safety factor. See page 888.

1098 | Internal Combustion Engine Handbook

6606_Book.indb 1098

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 24.26 Distribution of coefficients of heat transfer in the vicinity of the exhaust ports in a four-cylinder, five-valve engine. See page 893.

Stratified combustion

Measurement/calculation comparison CFD modeling

LIF measurements

24° before TDC

18° before TDC

Concentration distribution, reaction products

20° before TDC

Flame luminosity

Figure 24.30 Combustion with charge stratification—comparative assessment calculation versus measurement [24-31]. See page 897.

Internal Combustion Engine Handbook | 1099

6606_Book.indb 1099

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

14 1/(ms) 7

7

7 70

1/(ms) 14

7

Int.

Flame arrival time, evaluated for intensity threshold

5

Figure 25.11 Flame core generation and observation using a spark-plug sensor. The result illustrates the symmetry or asymmetry of the flame core and its predominant direction of propagation. See page 910.

Figure 25.12 Arrangement of a micro-optic sensor system in the cylinderhead gasket for tomographic flame reconstruction. See page 911.

900 100

with swirl

10 0

–20

–10

TDC

10

20

30

40

50

Rel. intensity

without swirl

°CA

Figure 25.13 DI gasoline engine: flame tomography shows the local position of bright, soot-producing diffusion flames. Swirling flow produces a significant improvement. See page 911.

1100 | Internal Combustion Engine Handbook

6606_Book.indb 1100

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Discharge side

Intake side

Lateral extension

Squish area influence

Figure 25.14 Flame propagation: tomography with sensor system installed in the cylinder-head gasket. The isolines indicate the progress of the flame front against time. The influence of internal flow on flame propagation is clearly visible. See page 911.

5 Out

Out Degree CA

In

In -4

35 Out

Out %

In

In 0

Figure 25.15 Flame tomography supplies comprehensive documentation of flame propagation and knocking spot distribution. See page 912.

Internal Combustion Engine Handbook | 1101

6606_Book.indb 1101

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

1 bar –2

Swirl: low

4000 1/min 20

1 22,4

25

10

Int. var.

Channel nr.

15

medium 35 5

°CA ATD

25

1 40

high 20 % Frequency 0

500 Off –60 °CA 100

–60 °CA 100

Diffusion flame Flame brightness [rel. unit]

On Injector

FSN = 0.0: Signal pattern of an ideal premixed combustion

Figure 25.16a Determination of knocking spots by a fan-type sensor—presentation of results: single-cycle and derived knocking spot statistics. See page 912.

FSN = 0.8: Smoke measurement indicates that improvement is necessary; the flame pattern hints where to start

0

Premix flame

Scaling of the flame brightness

Figure 25.16b Determination of flame pattern by a fan-type sensor. The polar diagram shows when and in which combustion-chamber sector flam anomalies occur. The flame brightness is evaluated from the unfiltered intensity signal. See page 913.

Visio sensor channel nr

Visio spark plug sensor channel directions

EX

EX

IN

IN

No Flame



irregular ignition location Flame

Deg CA

Figure 25.17 Sensor system and sample signal for determining the ignition sites during irregular combustion. See page 913.

1102 | Internal Combustion Engine Handbook

6606_Book.indb 1102

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Local peak temperature

736°C

5 mm

°C

746°C

9 mm

°C

Crank angle resolved dual spot measurement

580 560 540 520

100 °C

Max. valve temperature [°C][degC] Maximal ExValve Temperature



Figure 25.18a Thermal image of a spark plug in full-load operation The screw-in depth of the spark plug exhibits measurable temperature differences. See page 913.

500

valve stem

480 460 440

valve seat

420

100 200 300 400 500 600 700 800 900 1000 Cycle [-]

cycle to cycle temperature of two exhaust valves, stationary engine operation Figure 25.18b Measuring the radiation temperature at exhaust valves. Spark-plug sensor system with multichannel optics, use in normal engine operation. See page 914.

-68° CA

-68° CA

-38° CA, 2000 1/min, 3 bar p mi

-38° CA, 3000 1/min, 1 bar p mi

Figure 25.20 Gasoline DI: fuel distribution in the injection process and after deflection from the pistons. The stability of distribution states is determined from individual images using image statistics. Green/red: fuel vapor with increasing stability. Blue/ white: fuel droplet with increasing stability. See page 915.

Internal Combustion Engine Handbook | 1103

6606_Book.indb 1103

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Efficiency of generator at I= 40A

Efficiency of combustion engine

Torque [%]

Torque, Efficiency [%]

Efficiency

Overall efficiency of electrical energy

Figure 26.11 Efficiency map of a combustion engine (gasoline engine), efficiency characteristic curve of a generator, and overall efficiency map for the provision of electrical energy in the lower image (without belt losses). See page 922.

Output

Torque

Vehicle Electric motor Diesel engine Gasoline engine

Vehicle Electric motor Diesel engine Gasoline engine

Speed

Speed

Figure 29.12 Supply identifier of combustion engines and electric motors. See page 983.

1104 | Internal Combustion Engine Handbook

6606_Book.indb 1104

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Key

Combustion engine Clutch Transmission Generator Electric motor

Figure 29.16 Principle of a series, parallel, and power-split hybrid drive system (from left to right). See page 986.

Electrical energy storage device

Ladungsdurchsatz in % der Nennkapazität

Fuel tank

NiMH Li-ion (MN oxide) Li-Ion (Ni-oxide) Lead acid – VRLA (AGM/gel) Lead-acid – liquid

DoD

in %

Figure 29.58 Service life of different batteries dependent on the discharge level [29-44]. See page 1004.

Figure 29.66 Sectional view of the DualDrive [29-49]. See page 1008.

Internal Combustion Engine Handbook | 1105

6606_Book.indb 1105

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Figure 29.70 Sectional view of the MEGA (Magnetisch-Elektrische GetriebeAutomat) transmission [29-52]. See page 1009.

Figure 29.84 Start/stop system manufactured by PSA [29-72]. See page 1018.

Onboard supply 12 Volt Signal lines

Engine control unit with start/stop software option

Neutral gear sensor

DC/DC converter 12 Volt

Wheel speed sensor

Battery sensor

Crankshaft sensor

Start/Stop starter

Efficiency-increased generator with regenerative braking

Figure 29.85 Start/stop system manufactured by Bosch [29-73]. See page 1019.

1106 | Internal Combustion Engine Handbook

6606_Book.indb 1106

1/19/16 9:05 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

Flexplate

Rotor

Startor

Hydraulic converter

Transmission

Figure 30.9 Mild hybrid system for the Chevrolet Silverado [30-15]. See page 1034.

Figure 30.15 Plug-in Twindrive from Volkswagen as a front-drive vehicle [30-22]. See page 1037.

Reduced system complexity  No humidification  No water management  Reduced cooling  Reduced catalytic converters – degradation  High tolerance to gas impurity

Figure 30.31 Concept and advantages of high-temperature PEM [30-39]. See page 1046.

Internal Combustion Engine Handbook | 1107

6606_Book.indb 1107

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Color Section

hot

cold

Figure 31.5 Thermoelectric effect and technical implementation in the module. See page 1057.

BI2 Te 3 thermoelectric generator, Pmax= 200 W

Figure 31.8 Design of a TEG (BMW). See page 1058.

to the heater KFT DT

EGR

Coolant radiator

Low-temperature radiator

Kühlluft

EOC

Charge air Exhaust gas Main cooling circuit Low-temperature circuit

LLK

from the heater Figure 31.9 Coolant circuit for optimized heat management (source MTZ). See page 1059.

1108 | Internal Combustion Engine Handbook

6606_Book.indb 1108

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index Symbols 3D-CFD Model, 603 32-bit processor, 676 α -methylnaphtalene, 816 λ control, 29 A Abort reaction, 593 Absolute accuracy, 961 Absolute muffler, 950 Absorber, 76 Absorption, 76, 951 Absorption section, 951 Absorption-type damper, 430 Acceleration free, 805 Acceleration enrichment, 518, 522 Acceleration knock, 831 Acceleration ramp, 465 Acceleration resistance, 918, 920 Accelerator pump, 516, 519 ACEA, 850, 917 ACEA specification, 851 Acid sulfurous, 818 Acoustic decoupling, 289 Acoustic energy, 952 Acoustic field transformation, 956 Acoustic intensity, 946, 955 Acoustic pattern, 952 Acoustic power, 946 Acoustics, 269, 271, 878 Acoustic velocity, 946 Activated carbon canister, 528 Activated charcoal tank, 731 Activation energy, 797 Actual 0.2% offset limit, 298 Actuator, 467, 729 intelligent, 723

AC voltage ignition, 572 AdBlue, 818 Addition of fuel, 798 Additive ash, 803 Additive dosing, 804 Additive package, 835 Additives, 846 Additives for diesel fuel, 819 Additives for gasoline, 835 ADI austempered ductile iron, 162 Adjusting air nozzle, 517 Adjustment angle, 236 Aerodynamic drag, 918, 919 AFP steel, 161 Agglomeration, 789, 796 Aging stability, 843 Air baffles, 125 Airborne noise, 946, 950 Airborne noise route, 952 Air conditioning compressor, 954 Air cooling, 125 Air-Entrainment, 533 Air Expenditure, 23, 415, 434, 479 Air filter, 270, 863 Air filter element, 269 Air-fuel ratio, 40, 42, 813 Air-fuel-ratio, 598 Air–fuel ratio, 40, 42 Air-Fuel Ratio, 24 Air gap, 574 Air injection, 5 Air injector, 527 Air intake system, 268 Air mass, 533 stoichiometric, 40 Air mass flow, 965 Air mass meter, 527 Air mass sensor, 717 Air path, 269 Air pulse valve, 1061, 1062

Air pumping, 953 Air requirement, 813 stoichiometric, 24 Air spring, 420 Air stroke valve, 471 Al bearing alloy, 259 Alcohol, 824 Alcohol component, 825 Alcohol-diesel fuel mixture, 823 Alcohol fuel, 839 Alignment error, 250 Alkane, 589 Alkanes, 589 Alkene, 589 Alkenes, 589 Alkines, 589 Alloy, 150 Alloys supereutectic, 93 Al-Si alloy, 120 Altitude compensation, 523 Aluminum, 147 Aluminum piston, 87 Aluminum screw, 298 Aluminum-silicon alloy, 83, 93 AMT, 698 Anergy, 39 Angle sensor, 964 Angular Velocity, 17, 383 Anti-foaming agents, 820 Anti-jerk function, 685 Anti-jerk regulation, 687 Antiknock additive, 4 Anti-knock quality, 826 Antinoise, 959 Antinoise system, 959 Antioxidant, 836, 848 Apex seal, 382 API, 850 API classification, 852

Internal Combustion Engine Handbook | 1109

6606_Book.indb 1109

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

API TC, 859 APU system, 1048 ARAL (ARomates/ALifates), 833 Arc Phase, 569 Aromate, 824 Aromate content, 833, 834 Aromatics, 589, 590 Arrangement longitudinally symmetrical, 68 pressure-side, 867 suction-side, 867 Array technique, 956 asbestos-free, 281 Aspirated engine, 11 Assembly techniques, 294 Assessment of the Fuel Cell, 1046 ASTM standard D 3306, 861 ASVP (Air Saturated Vapor Pressure), 833 Asymdukt® piston, 88 Asynchronous machine, 991, 992, 1037 Asynchronous motor, 705, 1037 ATC Code of Practice, 856 ATIEL, 845 ATIEL Code of Practice, 856 Atkinson cycle (cycle), 231 Atomization quality, 529 Attachment shells, 950 Attenuation, 947 Autoignition, 12, 574, 578, 592, 595, 928 Automated manual transmission, 698 Automated quality control, 291 Automatic multi-stage transmission, 664 Automobile manufacturer specification, 855 Automotive industry (ACEA) European, 818 Auxiliary Power Unit = APU, 1047 Auxiliary unit, 932, 954 Auxiliary unit drive, 249 Auxiliary units, 950 Average pressure, 40 Axial bearing, 226, 227, 253 Axial Scale, 596 Integral, 596 Axle drive electric only, 1031 B Background, 463 Balance equation, 43, 45, 46 Balance shaft, 952, 959 Balancing mass, 66, 69 Ball calibration, 163 Barium NOx adsorber, 772 Base circle, 226, 227 Base circle diameter, 232 Base circle fault, 227 BASF test engine, 817 Basic material balance, 44 Battery capacity, 934 Battery charging, 1055 Battery management, 1003, 1037

Battery monitoring, 1003 Battery service life, 1003 Battery status sensor, 1055 Bead, 277, 288 Bead force, 283 Beamforming technique, 956 Bearing, 227, 251 open, 225, 226 Bearing brackets, 4 Bearing clearance, 256 Bearing damage, 267 Bearing diameter, 232 Bearing end, 257 Bearing engineering, 256 Bearing failure, 267 Bearing journal displacement path, 255 Bearing loading, 236, 254 Bearing location, 257 Bearing material, 258 Bearing metal, 259 Bearing play, 252, 257 Bearing version, 262 Beat frequencies, 954 Beehive spring, 188 Bell housing, 953 Bell-shaped curve, 19 Belt drive, 243 Belt-driven starter-alternator, 251, 703 Belt drives, 243 Belt pulley bolt, 292, 296 Belt tensioning system, 174, 246 Bending moment, 70, 233 Benz Karl, 3 Benzene content, 826, 833 Bergsträsser, 187 Bernoulli equation, 516 BET surface, 789 Beveled-edge oil control ring, 103 Bevel edge ring, 102 Bio diesel, 814, 821 Biofuel, 918 Biomass, 840, 943 BLDC motors, 674 Blind-hole nozzle, 553 Block height, 80, 188 Blow-by, 120 Blow-by gas, 230 Blow-by gas measuring technology, 965, 969 BMW, 1035 Boiling curve, 826, 832 Boiling curve (distillation), 832 Bonding technology, 120, 122 Booming noise, 952 Boost, 703 Boost mode, 985, 1031 Bore ratio, 79 Borghi-Diagram, 597 Boundary friction, 389 Boxer engine, 12 Brake, 962 Braking energy, 1055 Braking energy recuperation, 922

Branching reaction, 592 Brayton George Bailey, 3 Breakdown, 569 Break-in oil, 850 Brush honing, 124 Bulkhead surface area, 918 Burn-through speed, 511 Bush chain, 240 Butane, 824 Bypass, 724, 726 C CAD methods, 145 CAFE, 1031 Calculating Charge Cycles, 448 Calculation kinematic, 235 Calculation example inertial torque, 65 Calculation of cooling circuits, 890 Calculation of oil circuits, 890 Calculation of rotational oscillation, 73 Calorific value, 46, 591, 816, 819, 941 bottom, 42, 819 top, 819 Cam, 225 Cam angle, 235, 1077 Cam contour, 231, 236, 423, 454, 1078 Cam flank, 226, 234 Cam follower, 166, 225, 229, 230, 231, 232, 234, 235, 417 Cam follower valve train, 166 Cam lift, 227 Cam lobe, 225 Cam material, 229 Campbell diagram, 164, 954 Cam profile, 233, 234 Cams, 225, 226, 227, 232, 233, 234, 235 Cam segment, 231 Camshaft, 155, 224, 225, 226, 227, 229, 230, 232, 233, 417, 418 assembled, 227, 229, 232, 235 below head, 225 CamInCam, 230 forged, 227 hollow, 226 hollow cylinder, 232 hollow-profiled, 232 injection-molded plastic, 231 machined, 227 on rolling bearings, 230 solid, 226 special design, 229 variable, 230 Camshaft adjustment, 30, 469 Camshaft angle, 423 Camshaft bearing, 155, 225, 226, 227, 232 Camshaft bearing cap bolt, 292 Camshaft Drive, 419 Camshaft loading, 233, 234 Camshafts from cast iron, 227

1110 | Internal Combustion Engine Handbook

6606_Book.indb 1110

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Camshaft shifter, 448, 449 dual effect, 230 Camshaft shifter system, 236 Camshaft timing devices, 226, 446 Camshaft tube, 230 Cam tip, 226 Cam width, 232 Cancellation, 959 Carbon dioxide, 753, 917 Carbon dioxide emissions, 1029 Carbon monoxide, 753 Carbon residue, 819 Carburetor fuel CF, 824 Carburetor icing, 835 Carburetors Electronically controlled, 521 Carburettor electrical, 522 Carnot Sadi, 36 Carnot efficiency, 1040 Carnot process, 36 Car program, 815, 826 Cassette-type recorders, 951 Cast, 807 Cast cams, 228 Cast camshaft, 227, 229 Cast conrod, 294 Casting, 160 Casting model, 151 Casting process, 150 Cast iron, 227 Cast iron with nodular graphite, 161 Cast manifold, 300, 301 Cast tappet, 232 Catalytic converter, 531 monitoring the, 691 Near the engine, 303 Catalytic converter concept, 761 Catalytic converter deactivation, 764 Catalytic converter design, 760 Catalytic-converter protection function, 522 Catalytic converter substrate metallic, 775 Catalytic converter system, 762 Catalytic converter temperature, 531 Cavitation, 220, 861 Cavity content, 802 CCMC, 850 CEC, 851 Cell filter ceramic monolithic, 792 Center bolt, 296 Center of gravity of combustion, 624 Centipoise, 843 Central tube, 867 Centrifugal casting, 91 Centrifugal force, 230 Cetane, 816 Cetane number, 590 CFD, 224, 599, 869 CFD simulation, 44, 737, 1091 CFPP test, 818

CFR single-cylinder knock test engine, 829 CFR test engine, 817 Chain, 225 Chain design, 240 Chain drive, 238 Chain explosion, 592 Chainlike Hydrocarbons, 589 Chain tensioner, 243 Chain tensioning system, 177, 231 Chain value, 241 Change of contact, 84 Channel structured, 777 Characteristic Time Scale Model, 604 Characterizing Features, 15 Charge cycle, 36, 312, 415, 448, 929 Charge Cycle, 7, 11 Charge cycle calculation, 889 Charge cycle energy, 425 Charge cycle loop, 929 Charge cycle work, 926 Charge dilution, 627, 628 Charge mass, 929 Charge movement, 630 Charge pressure, 726 Charge Pressure Curve, 496 Chargers Mechanical, 506 Charge stratification, 927 Charging homogeneous, 530 stratified, 530 Chatter mark, 382, 385 Chevrolet Silverado, 1034 Chilled casting, 227 Chilled-cast iron cams, 229 Chilled-cast iron camshaft, 228 Chlorine ion content, 860 C/H ratio, 813 Chrome plating, 107 Chromium ceramic coating, 107 Chromium-diamond coating, 108 Classification, 806 Clean air pipe, 269, 270 Clearance adjustment mechanical, 234 Climbing resistance, 918, 919 Closed-deck, 112 Closed-deck design, 120 Closing ramp, 234 CNG, 823, 836, 1029, 1030 CO2, 1035 CO2 debate, 934 CO2 emissions, 917, 940, 941, 1035 Global, 943 CO2 limit value, 918 CO2 loop, 821 CO2 tax, 1031 Coalescence, 868 Coating catalytic, 798 Coaxial-Vario-Nozzle, 553 CO concentration, 29

Coefficient stoichiometric, 41 Coefficient of friction, 298 Coefficient of heat transfer, 890 Coefficient of rolling resistance, 920, 921 Coefficient of thermal expansion, 192 CO emission, 31 Coherence, 959 Coil ignition system Inductive, 570 Coking, 819 Cold knocking noise, 951 Cold start, 1045 Cold Start, 576, 579 Cold start aid, 585 Cold start aids, 581 Cold-start behavior, 817 Cold starting, 832 Cold-start strategy, 763 Cold start support, 581 Collared screw, 297 Combusted, 43 Combustion, 589, 609, 896 Diesel Engine, 602 Gasoline Engines, 602 knocking, 625, 828 normal, 828 Combustion center, 930 Combustion chamber, 154, 508, 601 Combustion chamber deposit, 831, 835 Combustion chamber design, 140 Combustion chamber die, 149 Combustion chamber plate, 152 Combustion chamber pressure sensor, 719 Combustion chamber recess shape, 617 Combustion chamber shape, 632 Combustion chamber variant, 139, 140, 141 Combustion delay, 44 Combustion duration, 44 combustion engine first, 1 Combustion engine, 9, 10, 1040 Combustion function, 43, 44 Combustion gas temperature, 930 combustion-generated, 805 Combustion model, 896 Combustion Model, 599 Combustion noise, 950, 952 Combustion process, 10, 35, 44, 139, 611, 928 flameless, 798 spray-guided, 471 Combustion product, 24, 40, 43 Combustion progress, 930 Combustion rate, 611 Combustion simulation, 45 Combustion speed, 44, 840 Combustion system, 617 Common rail, 925 Common-Rail, 562

Internal Combustion Engine Handbook | 1111

6606_Book.indb 1111

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Common Rail System, 546 Common-Rail-System, 537 Comparative process, 37 open, 40 Comparative stress according to Von Mises, 886 Comparison process engine, 39 Component calculation, 287 Component deformation elastic, 275 Component optimization, 879 Component protection, 33 Composite material, 281 Compressed natural gas, 1029 Compression geometrical, 79 isentropic, 36 isothermic, 36 variable, 79, 80 Compression Efficiency, 503 Compression height, 83, 84 variable, 80 Compression ignition homogeneous, 619 Compression ratio, 17, 37, 79, 625, 926 effective, 17 geometric, 17 Variable geometric, 1062 Compression Ratio, 16 Compression ring, 102 L-shaped, 102 Compression ring set, 85 Compression stroke, 225 Compression volumes, 16 Compressive load, 234 Compressor Electrical, 506 Compressor efficiency, 510 Compressor map, 489, 492, 509 Computing time, 881 Concave radius, 234, 235 Conchoid, 28 Condensation nucleus counter, 807 Condensation nucleus counter CPC, 805 Cone spray valve, 532 Configuration, 11 Connecting rod, 49 Conrod, 880 Conrod angular travel , 50 Conrod bearing, 254 Conrod bearing force, 56 Conrod bolt, 292 Conrod bolt connection, 294 Conrod bolts, 294 Conrod eye, 83, 85 Conrod pattern, 61 Conrod ratio, 50, 952 Constant pressure combustion, 40 Constant pressure cycle, 37 Constant pressure turbocharging, 489 Constant vacuum carburetor, 522 Constant vacuum carburettor, 517 Constant volume combustion, 40

Constant volume cycle, 37 Construction step, 144 Consumption, 444 Consumption map, 938 Consumption Maps, 19, 28, 29 Consumption potential, 921, 928, 930 Consumption reduction, 237 Consumption saving, 939 Consumption savings in NEFZ cycle, 214 Contact and opposite side, 87 Contact corrosion, 861 Contact force, 233, 235 dynamic, 237, 1078 kinematic, 237, 1078 Contact pressure, 880 Contact radius, 234 Contact surface, 233 Contact width, 233 crowning, 233 Contamination, 866 Continuous knocking, 829 Continuously variable transmission (CVT), 700 Control desmodromic, 6 Control deviation, 683 Controlled intake manifold, 270, 449 Control mechanism, 312 Control piston, 85, 87 control sleeve inline fuel injection pumps, 540 Control valve, 237 Conversion, 762 Conversion of energy, 570 Conversion of pollutants, 776 Coolant, 860 Coolant circuit, 146, 1059, 1108 Coolant concentrate, 861 Coolant flow simulation, 146, 1068 Coolant radiator, 738, 739 Coolant temperature, 933 Coolant thermostat, 743 Cooling, 842 Cooling air flow, 125 Cooling-air inlet, 951 Cooling by coolant, 737 Cooling channel, 87 Cooling channel piston, 89 Cooling fins optimization, 125 shape, 125 Cooling module, 742 Cooling system, 12, 735 Feature, 735 Cooling water flow rate, 46 Coordinate measurement unit, 153 Copper alloy, 259 Copper loss, 989 Cordierite, 793 Core-casting machines, 148 Core package process, 149 Core plug, 151 Cores, 151 Corner edge seal, 385

Corporate Average Fuel Economy, 1031 Corrosion, 818, 847, 868 Corrosion inhibitor, 819, 836, 847, 848 Corrosion protection, 843, 860 Corrosion resistance, 109 Cosworth low-pressure sand-casting process, 148 Countermass, 61 Counterweight, 61 Counting criterion, 796 Cracking component, 814 Crankcase, 58, 119 Crankcase scavenging, 4 Crankcase venting, 126 Crank diagram, 63 Crank gear, 49, 52 Crank gear force, 53, 58 Crank pin, 55 offset, 71 Crank pin force, 56 Crankshaft, 58, 159 cast, 162 manufacturing, 160 vibration resistance, 163 Crankshaft angle, 50 Crankshaft angle , 50 Crankshaft drive, 12, 49, 52, 79, 950 desaxised, 51 shifted, 51 Crankshaft material, 161 Crankshaft rumbling, 952 Crankshaft starter generator, 921, 922, 988, 1063 Crankshaft start generator, 468 Crankshaft throw, 55 CRC F-28, 831 Creep resistance, 148 Crescent-Type Oil Pumps, 205 Critical plane, 888 Critical plane method, 888 Cross-flow cooling, 143 CRT filter system, 799 Cruise control, 725 Cummins engine, 5 Current converter, 995 Curve, 27 Curved radius, 225 Curve representation, 27 CVS bag analysis, 80 CVT, 938 cw value, 919, 920 Cycle consumption, 29 Cyclical process, 35 Cyclical processes ideal, 37 Cyclical process work, 36 Cyclical service life, 1003 Cyclic irregularity, 60, 952 Cylinder, 15 wet, 121 Cylinder arrangement, 11, 13 Cylinder charge, 23 Cylinder charge dilution, 627 Cylinder cooling, 124

1112 | Internal Combustion Engine Handbook

6606_Book.indb 1112

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Cylinder cutout, 155, 170, 472, 475 Cylinder engineering, 120 Cylinder head, 137, 139 casting process, 147 Cylinder Head, 80 Cylinder head assembly, 196 Cylinder head bolt, 147, 155, 292 Cylinder-head cladding, 950 Cylinder head cooling, 143 Cylinder head cover, 950 Cylinder head development, 137, 146 Cylinder head geometry, 196 Cylinder head prototype, 144 Cylinder head shape, 138 Cylinder liner, 279 Cylinder output, 21 Cylinder peak pressure, 83 cylinder power indicated, 21 Cylinder pressure characteristic, 623, 626 Cylinder running surface, 120, 123 machining, 123 Cylinders, 119 Cylinder shutdown, 79 Cylinder shutoff, 922, 931 Cylinder warping, 120, 275 D Daimler Gottlieb, 3 DaimlerChrysler, 1035 Damage accumulation of the piston, 885 Damage index, 887 Damage mechanism, 803 Dam charcoal number, 597 Damping, 72, 76, 950 Damping filter, 273 Danger classification, 817 DC B, 15 T, 15 DC brushless motor, 1037 DC/DC converter, 704 DC machine, 1037 DC motor, 722, 1037 Brushless, 1037 Deactivation thermal, 785 Dead volume, 16 Deaxising, 52 Decoupling system, 289 Deep-bed filter medium, 867 DEF, 851 Deformation thermal, 886 Degree of absorption, 948 Degree of fraction separation, 865 Degree of radiation, 947 Degree of separation, 792, 863 Degree of transmission, 947 Delta-Control, 457 Denitrification, 803

Density, 826 Density sensor, 816 Deparaffination, 845 Deposit, 861 Deposition effects, 794 Deposition mechanism, 794 Deposits, 847 Depth filters, 796 Depth of Discharge, 1003 Design, 139 engineering, 83 stopper-less, 280 Design cylinder pressure, 885 Design of the PEM Fuel Cell, 1042 Desmodromic system, 225 Desulphurization temperature, 788 Detector, 972, 973 Paramagnetic, 973 Detergent and dispersant additives, 819 Detergents, 835, 836, 847 Detergents/Dispersants, 847 De-throttling, 926 Dethrottling, 237, 472 Diameter aerodynamic, 790 DI combustion process, 928 Die, 149 Die-cast aluminum oil pan, 953 Die casting, 91, 149 Die casting process, 120, 151 Diesel Rudolf, 3, 36 Diesel combustion, 607 Diesel engine, 922, 925 Diesel Engine, 10 Diesel Engines, 5 Diesel exhaust unit, 561 Diesel four-stroke combustion systems, 612 Diesel fuel supply system, 561 Diesel fuel tank, 557 Diesel hybrid drive system, 1024 Diesel oxidation catalytic converters, 782 Diesel racing fuel, 823 Diesel supply unit, 562 Diesel tank pump, 562 Diesel-water emulsion, 823 Diethylene glycol, 860 Differential gear, 68 Differential gear ratio, 936 Diffusion, 45, 794 Diffusion battery, 806 Diffusion flame, 598 Diffusion speed, 795 Digitizing, 153 Diisobutylene, 841 Diluting agent rotating, 807 Dilution, 807 Dilution Controlled Combustion System, 620 Dimethyl ether, 823 DIN EN 228, 826

DIN EN 590, 815 Direct cooling, 12 Direct-current motor, 722 Direct drive, 937 Direct injection, 11, 526, 894, 922, 924, 925, 926, 952 Direct-injection, 7, 139, 471 Direct Injection Air-supported, 527 Direct Numerical Simulation, 603 Dirt storage capacity, 867 Discharge port variant, 142 Dispersants, 847 Displacement Variable, 1062 Displacement line, 491 Displacement magnet, 447 Displacement pump, 562, 565 Distance-related fuel consumption, 921 Distillation atmospheric, 845 Distribution bimodal, 791 Distribution of lubrication film, 255 Distributor injection pump, 541, 542 Disulfide sulfur, 834 Dividing characteristics, 796 DNS, 603 DOHC, 225, 231 DOHC engine, 226 Double-beveled oil control ring, 103 Double-layer capacitors, 706 Double overhead camshaft, 225 Double overhead camshaft = DOHC, 164 Double stopper, 279 Double trapezoid ring, 102 Downdraft carburetor, 517, 518 Downsizing, 479, 920, 922, 1030, 1062, 1063 Downspeeding, 924, 1062 DPF, 898 Drag coefficient, 918 Drive-by-wire, 725 Drive control software, 678 Drive element, 227 Drive flange, 226, 227 Drive machines, 9 Drive moment, 233 Driven machines, 9 Drive past test legal, 952 Driver behavior, 939 Drive shaft, 952 Drive sprocket, 226 Drivetrain, 77, 685 Driving Electric-only, 1031 Driving mode, 935 Driving resistance, 918 Driving speed, 921 Driving style, 938 Drop in viscosity, 844, 856 Droplet size, 532

Internal Combustion Engine Handbook | 1113

6606_Book.indb 1113

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Droplet size distribution, 609 Droplet-wall interaction, 895 Dry friction, 389 Dual cam adjustment, 449 Dual carburetor, 517 Dual clutch transmission, 664 Dual spark coil, 571 Ductility, 148 Dummy-head recording, 957 Duplex chain, 240 Durability of screw connections, 296 Dust capacity, 864 DVPE, 833 DVPE (Dry Vapor Pressure Equivalent), 833 Dynamics, 232 Dynamics calculations, 236 Dynamic tire radius, 920 Dynamometer, 962 E Easy-change disposable fuel filter, 868 eBooster, 500 Eccentric, 226 Eccentric adjustment, 460 Eccentricity, 383 Eccentric shaft, 383, 464 ECE cycle, 922 ECE-Tank, 560 EC motor, 722 EDC, 558 EDC engine control system, 32 EDC-System, 557 EELQMS, 856 Effect thermophoretic, 805 Efficiency, 42 effective, 17, 22, 46, 415 indicated, 22, 46 mechanical, 22 thermal, 36, 38, 45, 79 Einsteinian relation, 795 E-Gas, 725 EGR, 531, 728 EGR flap valve, 730 EGR valve, 528 Ejector diluting agent, 807 Elasticity Test, 498 Elastomer coating partial, 278 Elastomer head gasket, 276 Elastomer material, 283, 284 Elastomer seal, 281, 285 Elastomer sealing lip, 276 Elastomer sealing system, 284 Electrical energy, 921, 922, 1104 Electrical energy management, 1055 electrical output, 921 Electric coolant pump, 934 Electric double-layer capacitor, 1001 Electric drive system, 703, 943, 1036 Electric driving, 984 Electricity production, 943 Electric motor, 1037

Electric-only driving, 1031 Electric vehicle, 703, 943, 1037, 1042 electrohydraulic valve train, 930 Electro membrane, 1042 Electronic diesel control, 32 Elementary analysis, 813 Elementary EC carbon, 789 Emission, 972, 975 Emission map, 29 Emission noise, 952 Emission requirements, 557 Emissions factor, 804 Emitec-SCRi-System, 779 Emitted noise emission value, 952 Emitted noise measuring procedure, 948 Emptying method, 437 Encapsulation tight-fitting, 950 Enclosure off-engine, 950 spaced, 950 Enclosure element, 951 Energy and drive management, 1030 Energy balance, 45, 46 Energy chain, 1029 Energy conversion, 1055 Energy density, 997 Energy flow, 46 Energy management, 1054 Intelligent, 1055 Optimal, 1031 Energy management system, 1010 Energy recovery, 936, 943, 1031 Energy recovery with both, 1031 Energy recuperation, 1012 Energy storage device, 1036, 1039 Energy storage devices, 1039 Energy Storage Devices, 996 Energy storage systems, 704, 934 Engine, 15 Engine acoustics, 945 Engine bearing, 952 Engine block, 950 main dimension, 111 material, 118 Engine Block, 6 Engine block design, 120 Engine block types, 116 Engine braking system, 225, 233 Engine characteristics, 15 Engine compartment, 950 Engine compartment capsule, 286 Engine component, 83 Engine cooling, 737, 743 Engine enclosure, 951, 952 Engine hood absorber, 951 Engine hood lining, 951 Engine intake air, 863 Engine Knock, 595 Engine map, 27, 28 Engine mounting, 953, 954 Engine mounting bracket, 954 Engine noise, 949, 950, 952 mechanical, 950, 952

Engine noise emissions, 950 Engine oil, 870 Engine oil for four-stroke engine, 848 Engine oil for two-stroke engine, 858 Engine operation un-throttled, 530 Engine order, 959 Engine test, 287, 855 Engine test bed, 961, 962, 965, 970 Engine working point, 931 Enthalpy free, 42 molar free, 41 Enthalpy difference, 46 Enthalpy flow, 46 Environmental awareness, 917 EPEFE, see also 22.1.2.2, 815 Epitrochoids, 383 Equal pressure stage, 517 Equidistant, 383 Equilibrium chemical, 40, 44 thermodynamic, 42 Equilibrium constant, 594 Equilibrium temperature, 800 Erosion Behavior, 578 Ethanol, 840, 1029, 1030 Ether, 824 Ethylhexyl nitrate, 817 ETOH, 836 euATL, 500 European driving cycle, 919 European Engine Lubricant Quality Management System, 856 EU standard EN 589, 836 Evaporation heat, 523, 840 Evaporation loss, 843 Evaporative emission, 731 Exciter force, 75 Exciter work, 73, 75 Exciting force, 73, 75 Exergy, 39 Exergy loss, 39, 42 Exhaust average specific heat, 46 Exhaust closes, 433 Exhaust Counter-pressure Level, 431 Exhaust emissions, 618, 817, 917, 961, 967 Exhaust gas, 973, 975 Diluted, 974 Undiluted, 973 Exhaust gas cooling, 740 Exhaust gas exergy loss, 42 Exhaust gas limit, 827 Exhaust gas mass flow, 965 Exhaust gas measuring technology, 961, 965 Exhaust gas particulate measuring technology, 974 Exhaust gas power, 508 Exhaust gas recirculation, 30, 723, 728 Exhaust-gas recirculation, 449 Exhaust gas recirculation rate, 30, 31, 531

1114 | Internal Combustion Engine Handbook

6606_Book.indb 1114

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Exhaust gas recirculation valve, 728, 729 Exhaust gas system, 430 Exhaust gas temperature, 33, 300 Exhaust gas temperature map, 33 Exhaust gas threshold current, 746 Exhaust gas treatment, 897 Exhaust Gas Treatment for gasoline Engines, 760 Exhaust gas turbocharging, 7, 493 two-stage, 506 Exhaust Manifold, 299, 878 Exhaust measurement in the combustion chamber, 750 Exhaust measuring technology, 746 Exhaust muffler, 950 Exhaust muzzle noise, 949 Exhaust opens, 433 Exhaust port valve seat insert, 195, 1071 Exhaust side adjustment, 236 Exhaust slot, 384 Exhaust system, 965 Exhaust system seal, 282 Exhaust Treatment in Diesel Engines, 782 Exhaust treatment systems such as particle filters, 818 Exhaust turbochargers, 444, 489 Exhaust valve, 138, 419, 950 Expansion isentropic, 36 isothermic, 36 thermal, 234 Experiment design, 893 Explosion thermal, 592 Explosion limit, 824 Explosion Limit, 593 External air pipe, 269 External Gear Pump, 206 Extraction isokinetic, 805 Extrusion casting, 91 F FAME, 814 Fan drives, 741 Fans, 741 Fast-off detection, 700 Fatigue limit, 187 FCKW, 823 FEM, 879 FEM calculations, 236 Fermentation, 840 Ferrotherm® piston, 90 FIA analysis, 824 Fiber rope filter, 793 Fiction loss, 248 Filerability, 818 Filing capacity, 103 Filling level measuring, 566 Filling method, 437

Filling process, 148 Filter, 537 open, 778 Filter candle, 793 Filter characteristics, 796 Filter depth, 795, 802 Filter element, 269 Filter fabric, 864 Filter felt, 794 Filter fineness, 865, 866 Filter fleece, 794 Filter head, 868 Filtering, 564 Filter medium, 792 Filter paper, 794 Filter surface area, 867 Filter testing machine, 804 Filtration, 268, 863, 870 Filtration velocity, 865 Final drive systems, 706 Fine dust, 789 Fines, 918 Finger follower, 146, 168 Finite element analysis, 287 Finite element method, 875 Finite Elements, 6 Finite elements mesh, 881 Finite volume method, 891 Fire land, 84 Firing channel, 385 Firing sequence, 69, 71 Firing sequences, 71 First-order inertial force, 61 oscillating, 63 First ring land, 84 Fischer-Tropsch synthesis, 814, 834 Fixed air funnel carburetor, 517, 519, 523 Flaked graphite, 227 Flame Non pre-mixed, 596, 598 Partial pre-mixed, 598 Pre-mixed, 596, 597 Flame front, 43 Flame ignition, 3 Flamelets-Model, 604 Flame path, 930 Flame propagation, 623 Flame Propagation, 596 Flame speed, 627 Flame Speed, 591 Flame Type, 596 Flange concepts for tube manifolds, 304 Flap valve, 729 Flash point, 817 Flat-based tappet contact, 236, 1078 Flat seal, 280 Flat-tube coolers, 740 FlexRay, 673 FlexRay bus, 673 FlexRay fieldbus, 673 Flexural load, 233 Float artery valve, 517 Float chamber, 516, 519, 522

Floating axle, 464 Flow dynamic, 532 static, 532 Flowability, 818 Flow calculation, 889 Flow Cross Sections, 435 Flow determination, 436 Flow factor for the valve, 436 Flow function, 438 Flow improver, 818 Flow measuring technology, 965 Flow noises, 950 Flow process stationary, 46 Flow property, 847 Flow pump, 562, 565 Flow range linear, 532 Flow simulation, 146, 889 Flow turbulence, 891 Fluctuation cyclical, 624 Fluid or viscous friction, 389 “Flying lawn chair” by Baumm, 382 Flywheel, 952 Flywheel bolt, 292, 296 Flywheel energy storage device, 1039 FM method, 622 Foam, 950 FON, 829 Force acting on the conrod, 55 on crankshaft drive, 52 on piston, 54 on wristpin, 54 Forced-feed lubrication, 858 Ford Henry, 4 Forged conrod, 294 Forging, 91, 160 Forked lever, 225 Formula analytical, 880 Formula 1 piston, 89 Fossil fuels, 917 Four-cycle engine, 803 Four-stroke process, 11 Four-valve cylinder head, 141, 154 Four-valve technology, 930 Free burning temperature, 574 Free-piston linear generator, 1047, 1050 Freezing suitability, 1045 Frequency range, 952 fresh charge mass theoretical, 23 Fresh oil lubrication, 858 Friction, 154, 842 Crankshaft, 396 Frictional power, 880 Frictional Power, 389 Friction loss, 225, 847 Friction map, 933 Friction Mean Pressure, 22

Internal Combustion Engine Handbook | 1115

6606_Book.indb 1115

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Friction modifiers, 835, 836, 847 Friction vibration, 385 Front octane number, 832 Frost protection, 860 FTP, 926 FTP 75 Test cycle, 748 Fuel, 10, 24, 589, 1029 Gaseous, 10 liquid, 10 reformulated, 834 solid, 11 sulfur-free, 799 Fuel and drive strategy, 1030 Fuel atomization quality, 529 Fuel cell, 704, 1031, 1042 Fuel Cell as an APU, 1047 Fuel cell concept vehicles, 1045 Fuel cell drive system, 1042, 1044 Fuel cell in the vehicle, 1043 Fuel cell stack, 1043 Fuel cell vehicle, 1044 Fuel cell vehicles, 703 Fuel charge, 224 Fuel Chemistry, 589 Fuel conditioning, 965, 968, 969 Fuel consumption, 917, 921 effective specific, 19 specific, 19 Fuel Consumption, 7, 19, 79, 917, 919, 927, 934, 942 Fuel consumption lead, 1061 Fuel consumption measurement, 965 Fuel consumption measuring technology, 961 Fuel consumption potential, 919, 1061 Fuel consumptions, 1033 Fuel consumption saving, 934 Fuel economy oil, 849 Fuel filters, 866 Fuel hydraulic circuit, 890 Fuel injection (IDI) Indirect, 614 Fuel mass flow, 965 Fuel rail, 483 Fuel return, 523 Fuel saving potential, 936 Fuel-saving potential, 926, 936 Fuel savings, 472 Fuel supply systems, 557 Fuel Supply Systems, 560 Fuel Tanks, 557 Full bead, 277, 283 Full-flow burner, 797 Full flow dilution system, 974 Full hybrid, 703, 987, 1031 Full-load, 28, 32, 531 Full-load curve, 27 Full-load enrichment, 928 Full lubrication, 842 Full-skirt piston, 89 Function model-based, 681 Function architecture torque-based, 680

Function verification virtual, 889 Fundamentals of Thermodynamics, 35 G Gap corrosion, 861 Gap width, 106 Gas composition, 44 Gas constant general, 40 Gas diesel engine, 623 Gas engine oil, 850 Gas escaping, 83 Gas exchange, 145, 225, 269 Gas exchange control system, 11 Gas exchange cycle, 79, 231 Gas force, 52, 225, 235, 952 Gas fuel, 836 Gasket testing, 287 Gasoline, 824, 1030 alternative, 836 Gasoline direct injection, 529, 927, 928 Gasoline engine, 922, 926 Gasoline filters, 866 Gasoline fuel in-tank pump, 565 Gasoline hybrid drive system, 1024 Gasoline supply systems, 563 Gasoline tank, 557 Gasoline vapor, 731 Gas port, 138 Gas pressure, 225 Gas-switching element, 139 Gas torsional force, 75 Gas turbine, 1031, 1041 Gas turbine engine, 6 Gas work, 20 Gauge pressure tank diagnosis pump, 732 Gear pump, 216 Gear rattle, 950 Gear set design, 216 Gear shift speeds, 939 Gearshift strategy, 699 Gearwheel oscillations, 950 Generating electricity, 1056 Generator, 922, 1104 Thermoelectric, 1013 Generator control, 1055 Generator Operation Mode, 494 Generator whistling, 952 German monopoly administration for spirits, 814 GH2, 836 Glass fiber, 870 Glow discharge Glow Discharge, 569 Glow Plug, 582 Glow System, 581 glow tube ignition uncontrolled, 3 Glycerides and glycerins, 821 Glycols, 860 GM, 1035

Gray-cast camshaft, 232 Gray cast iron, 147 Gray cast iron running surface, 122 Gray cast iron sleeve, 122 Gray cast structure, 227 Greenhouse gas emissions, 821 Greenhouse gases, 821 Grooved bearingsTM, 263 Gross reaction equation, 591 G-Rotor pump, 562 GTL (gas-to-liquid), 836 GTL Gasto Liquid, 823 Gümbel-Holzer-Tolle procedure, 74 H H2–O2 System, 593 Half bead, 277, 283 Half-shell manifold, 302 Hammering, 952 Hard anodizing, 92 Hardening radii inductive, 163 H bridge circuitry, 674 HCCI, 594, 928 HCCI Process, 594 HC/CO conversion, 772 HC emission, 16, 30 HC emissions map, 30 HCF, 886 HC-storage catalytic converter, 764 Head contact area, 296 Head gasket, 275 Heated catalyst electrical, 763 Heat engine, 35 Heating flange, 584 Heating Flange, 584 Heating system sequential, 798 Heat loss, 46 Heat management, 934, 1059 Heat management method, 934 Heat Range, 574 Heat removal, 35, 125 Heat stress, 803 Heat supply, 35 Heat throughput, 736 Heat transfer, 12 Heat transfer coefficient, 736 Heat transfer model, 600 Heat transfer process, 35 Heavy gasoline, 3 Heavy oil operation, 623 Height offset of the balance shaft, 69 Height profiling, 278 Helical compression spring, 187 Helical slide honing, 123 Helical toothing, 237 Helmholtz resonator, 273, 428, 951 Hertzian pressure, 225, 232, 234, 235, 1077 dynamic, 235, 1077 kinematic, 235, 1077 HFRR test, 820

1116 | Internal Combustion Engine Handbook

6606_Book.indb 1116

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Higher NOx emissions, 16 High Pressure Direct Injection, 526 High-pressure fuel injector, 530 High-pressure fuel pump, 530 High Pressure Injection, 528 High pressure line, 537 High-pressure pump, 529, 546 High-pressure rail, 548 High pressure regulator, 529 High-pressure sensor, 716 High-pressure storage tank, 838 High pressure system, 537 High-speed engine, 14 High-temperature aging, 772, 773 High-temperature battery, 999 High-temperature electronics, 1030 High-temperature fuel cell, 1048 High-temperature high-shear viscosity, 849 HNBR material, 243 Hole nozzles, 552 Hollow casting cylindrical, 232 Hollow drilling, 232 Holography acoustic, 956 homogeneous self-ignition, 928 Homogenous Charge Compressed Ignition, 928 Honda IMA concept, 1034 Honda MMC process, 122 Honeycomb cooler, 3 Honeycomb stopper, 277 Hooke’s Law, 287 Horizontal draft carburetor, 518 Hot cooling, 1063 Hot driving, 832 Hot gas generator, 508 Hot gas simulation, 291 Hot operation, 522 Hot water corrosion, 861 Housing-type fuel filter, 868 HTD profile, 244 HTHS, 849 HVOF layers, 108 Hybrid Parallel, 1030, 1031 Series, 1030, 1031 Hybrid battery, 703 Hybrid concept, 934 Hybrid drive system, 983, 1029, 1030, 1039 Fuel consumption, 1016 power-split, 987 Power-split, 1031 Hybrid engine, 10 Hybrid propulsion system, 703 Hybrid vehicle, 934, 1030 Hydraulic element, 167 Hydraulic pump, 952 Hydraulic tappet, 466 Hydraulic valve, 451 Hydraulic valve clearance compensation element (HVA), 227

Hydraulic valve lifter (HVA), 234 Hydrocarbon aromatic, 833 Hydrocarbon (alkane) paraffinic, 814 Hydrocarbon indexing, 752 Hydrocarbon (PAH) polycyclic aromatic, 818 Hydrocarbons, 753 synthetic, 814 Hydrofinishing, 845 Hydrogen, 7, 838, 943, 1030, 1043, 1044 Hydrogen combustion, 593 Hydrogen direct injection, 386 Hydrogen engine oil, 850 Hydrogen infrastructure, 1047 Hydrogen production, 1046 Hydrogen rotary piston engine, 386 Hydrogen storage system, 1043 Hydrogen storage technology, 1042 Hydrogen sulfide (H2S), 834 Hydrogen treatment, 818 Hydrothermatik® piston, 88 Hydrothermik® piston, 87 Hypocycloids, 383 HYZEM driving cycle, 920 I Icing, 523 Idle revolutions, 520 Idle-speed control, 724 Idle-speed control in SI engines, 724 Idling air nozzle, 517 Idling fuel consumption, 921 Idling nozzle, 517 Idling rpm control, 523 Ignitability, 816 Ignition, 569, 609 Ignition coil, 571 Ignition delay, 580 Ignition delay time, 594 Ignition jet method, 622 Ignition lag, 609, 816, 817 Ignition map, 32 Ignition pressure, 275 Ignition system, 12, 930 Ignition systems alternative, 572 ILSAC certification, 852 Immersion lubrication, 4 Impaction, 794 Impact load damping, 724, 725 Impregnating agent, 867 Impression of comfort, 959 Impulse counting procedure, 963 Impulse turbocharger valve, 479 Inclusion, 188 Increase azeotropic, 825 Indexing method, 391 Individual Pump Systems with a Line, 538 Individual valve system, 147

Inertial force, 52, 60, 61, 62, 63, 235, 952 first order, 63 oscillating, 52, 61 rotating, 52, 61 Inertial torque, 61, 64 oscillating, 65 rotating, 65 Infinite variability, 470 Influence of ignition angle, 624 Injected Fuel Metering, 532 Injection mechanical, 6 Injection map, 32 Injection nozzles and nozzle-holder assemblies, 551 Injection orifice design, 554 Injection pressure, 533, 926 increase, 551 Injection process, 535 Injection pump cam, 227 Injection rate, 534 Injection system adjusting , 555 Injection System, 534 Injection time, 535, 556 Injection timing system, 537 Injection valve, 537 Inlet closes, 434 Inlet opens, 434 Inlet timing, 231, 449 Inlet valve, 138 Inline engine, 3, 12 Inline filter, 867 Inline fuel injection pumps, 539 Inner contour profiled cavity, 232 Input inertance, 948 Insert, 276 Insertion insulation increment, 947 Insert technique, 120, 121 In-situ exhaust measurement, 749 In-situ measuring technology, 805 Installation space, 803 Installation thickness, 277 Installed flexure tension, 106 Installing in the Cylinder Head, 202 Insufficient lubrication, 229, 868 Insulation, 947 Intake Air Duct, 518 Intake air volumetric flow measurement, 965 Intake manifold, 950 Intake manifold charging model, 681 Intake manifold injection, 11, 895 Intake manifold injection systems, 523 Intake manifold pipe variables, 428 Intake manifold pipe cross section, 428 Intake manifold pipe length, 428 Intake muffler, 950 Intake opening, 269 Intake opening noise, 949 Intake port variant, 142

Internal Combustion Engine Handbook | 1117

6606_Book.indb 1117

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Intake system, 268 Intake valve lift curve, 462 In-tank unit, 867 Integrated powertrain management (IPM®), 701 Integrated starter/alternator (ISG), 695 Interception, 794 Intercooling, 490, 740 Interior high-pressure forming, 227 Interior noise, 952 Internal cooling, 928 Internal efficiency, 623 Internal Gear Pump, 203 Internal noise levels, 950 Internal pressure simulation hydraulic, 290 Intervention Mechanism, 595 Ionic Current Measurement, 585 Iron loss, 989 IRP, 212 ISAD system, 161, 1034 Isentropic exponent, 37, 490, 926 Isentropic relationship, 40 Isooctane, 40 Isoparaffin, 824 IT material, 281 IV. Quadrant, 512 J jet-directed method, 928 Jet direction, 531 just-in-time, 6 K Karlovitz number, 597 Key oil pump value, 218 Kinematics, 423 Kinematics calculation, 235 Kink lever, 456 Klimov-Williams Criteria, 598 Knight sleeve valve engines, 4 Knocking control, 688 Knocking sensitivity, 79, 830 Knocking signal, 688, 689 Knock sensor, 829 Kolmogorov Axis, 596 KS Lokasil® process, 122 L Laboratory testing procedures, 290 Lacquer formation, 847 Lambda closed-loop control, 523 Lambda control, 684 Lambda-control, 521 Lambda sensor, 7, 713 Lambda window, 683 Lancaster balancing, 952, 959 Lane angle of ascent, 920 Lapping, 124 Large-Eddy Simulation, 603 Laser double-pulse holography, 956 Laser-induced incandescence, 976

Laser texturing, 123 Laser vibrometry, 956 LASP lubricating oil, 800 Lateral exhaust, 385 Lattice gas theory, 892 Law of thermodynamics, 40 Law of thermodynamics , 40 Layer approach, 677 Layered-metal head gasket Metaloflex®, 276 LCF, 886 Lead-acid, 1038 Lead-acid accumulator, 706 Lead-acid battery, 998 Lead as anti-knocking agent, 835 Lead compound, 824 Leaded bronze, 259 Leak Rate, 532 Lean-burn concept, 926 Lean-burn operation, 926 Lean operation, 530 homogeneous, 531 Length reduction, 73 LES, 603 Lever-type sensor, 566 Lexus GS 450 h, 1034 Lexus RX 400 h, 1034 LH2, 836 Lifetime filter element, 867 Lifetime fuel filter, 867 Lift valve, 416 Lightweight design, 286 Lightweight metal pistons, 84, 85 Li-ion, 1038 Li-Ion battery, 706 Li-Ion cells, 707 Linear motor, 1049 Linear regulator, 673 Linear valve, 481 Line contact, 233 Liquefied petroleum gas, 1029 Liquid gas, 837 Liquid hydrogen, 838 Liquid storage tank, 838 Lithium-Cobalt-Nickel-Manganese oxide, 707 Lithium-Cobalt oxide, 707 Lithium-ion, 1038 Lithium-ion battery, 1000, 1030 Lithium-Iron-Phosphate, 707 LNG, 836 Load crankshaft, 60 Load-adjustment system, 13 Load-bearing capacity of the lubricating film, 236 Load change valve, 225 Load controlling, 444, 471 Load cycle loss, 79 Load distributions, 188 Load influence, 632 Load on valve seat inserts, 190 Load range lower, 32 Load step, 477

Locating bearing, 253 Locus diagram, 62, 64 Longitudinal flow cooling, 143 Longitudinal frame concept, 114 Long-time fatigue, 886 Loop scavenging, 441 Loop scavenging cylinder, 313 Loss analysis, 632 Loss of contact, 236 Lost-foam process, 150 Lost-Motion Element, 455 Loudness, 957 Low-pressure die-casting, 120 Low-pressure sand-casting process, 148, 151 Low-pressure system, 536 Low-sped engine, 14 Low speed concept, 925 Low-temperature behavior, 818 Low temperature oxidation, 594 Low-temperature pump viscosity, 849 Low-voltage impactor ELPI, 806 LPG, 836, 1029, 1030 l regulation, 683 LS 600 h, 1034 Lubricant, 842 Lubricating film pressure maximum, 255 Lubricating gap smallest, 255 Lubricating oil ash, 792, 803 Lubrication, 842 Lubrication properties of the gasoline, 835 Lubricity additives, 820 Lubricity improvers, 835 Lubrifier modifiers, 836 M Mach number, 509 Macro-sealing, 283 Magnesium, 118 Magnesium ion, 861 Magnetic Passive Position Sensor, 567 Main bearing cap bolt, 292, 293 Main bearing force, 58 Main bearing pin force, 58 Main injection, 32 Main nozzle, 516 Manifold absolute pressure sensor, 715 Manifold and turbocharger module, 303 Manifold components, 304 Manifold design, 299 Manifolds Air Gap Insulation, 302 MAN-M process, 622 Manual starter, 520 Manual transmission automated, 700 Map-controlled thermostat, 934 MAPPS, 567

1118 | Internal Combustion Engine Handbook

6606_Book.indb 1118

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Marginal condition thermal, 881 Masking, 462 Mass moving, 232 oscillating, 51 static, 232 Mass air-flow sensor, 717 Mass balancing, 66, 67, 68, 952 Mass criterion, 796 Mass reduction, 73 Mass spectrometry, 749 Material balance, 41 Materials for piston rings, 109 Mating material tribologic, 123 Maximum pressure, 37 Maximum pressure limit, 38 maximum speed, 937 Maximum speed, 937 Maybach Wilhelm, 3 Mean friction pressure, 389 Mean pressure, 491, 923, 928 effective, 16, 21, 415, 487 indicated, 20 indicated, 21, 479 Mean pressure curve, 493 Mean stress correction according to Smith or Haigh., 887 Measurand, 961, 963 Mechanical, 963 Thermodynamic, 964 Measurement, 961, 963, 964, 965, 966, 969, 972, 973, 975, 976 Dynamic, 965 opacimetric, 805 Stationary, 965 Measurement data evaluation, 976 Measurement data processing, 976 Measurement principle capacitive, 716 Measurement techniques, 976 Measuring technology, 961, 972, 973 Blow-by, 965, 969 Blow-by gas, 965 Fuel consumption, 965 Medium-pressure sensor, 716 Medium-speed engine, 14 MEOH, 836 Mercaptan sulfur, 834 Mercaptan (thioalcohol), 834 Mercedes-Simplex-Motor, 3 Meshing, 881 Metal abrasion, 857 Metal-containing additive, 820 Metal deactivator, 836 Metal-elastomer gasket, 285 Metal-elastomer head gasket, 276 Metal-elastomer seal, 281 Metal foam, 792 Metal head gasket, 120 Metal hydride storage tank, 838 Metaloseal, 282 Methanol, 1030

Method gravimetric, 789 MF system, 104 Micro hybrid, 987 Micro-hybrid, 703 Middle distillate, 814 Mild hybrid, 703, 987, 1031, 1032 Mileage, 942, 943 MIL-L, 851 Miller cycle (cycle), 231 Miller Process, 471 Miller-Process, 477 Miner rule, 888 Minimum Ignition Energy, 569 Minimum injection functionality, 535 Minimum requirement, 815, 826 Minimum start speed, 580 Minimum wrap angle, 250 Minimum wristpin play, 85 Mixed friction, 253, 389 Mixed hybrid, 1030, 1031 Mixed octane number, 829 Mixing pipe, 517 Mixture lean, 25 rich, 25 uncombusted, 43 Mixture enrichment, 521 Mixture formation, 11, 520, 608, 894 Mixture Formation in Diesel Engine, 532 Mixture generation external, 623 Mixture homogenization, 619 Mixture Ignition, 576 Mixture leanness, 926 Mixture preparation homogeneous, 623 stratified, 623 Mobility, 795 Mobility analyzer differential, 805 Mobility diameter, 790 Modal analysis, 956 Model multidimensional, 44 Phenomenological, 599, 601 Thermodynamic, 599 zero-dimensional, 43 Zero-dimensional, 599 Model calculation, 35, 42 Model Development, 598 Model manufacturing, 151 Modified Uniontown Method, 831 Modular graphite, 227 Modulation, 947, 952 Module, 286 Multi-functionality of, 287 Moles specific, 40, 43, 44 Molybdenum coating, 108 Moment dynamic, 235 resulting, 233 transferable, 235

Momentary combustion chamber, 16 MON, 829 Monoethylene glycol, 860 Monolithic design, 120 Monopropylene glycol, 860 Monotherm® piston, 90 Mounting system, 275, 285 Movement equation, 73 Muffler, 803 Muffler system, 430 Multi-body simulation, 236 Multi-function layer design, 278 Multigrade, 869 Multi-grade oil, 849 Multi-layer steel gasket, 276 Multiple injections, 534, 535 Multiple power driver, 291 Multi-point fuel injection, 831 Multi-rib V-belt, 249 Multi-stage charging principle, 926 Multi-stage regulation, 210 Multi-valve engine, 137, 225 Multi-valve technology, 137 Multizone model, 35 Muzzle noise, 950 MVEG, 214 N Natural bending frequency, 953 Natural frequency, 73, 236 Natural gas compressed, 823 Natural gas or synthetic gas, 823 Natural oscillation mode, 74 Natural torsion frequency, 953 Navier-Stokes equation, 891 NEDC, 214, 919, 939 Needle bearing, 236 Needle guide doubled, 554 Needle nozzle, 517 Needle seat, 532 NEFZ cycle, 212, 213 Neutralization of acids, 847 Newtonian liquid, 843 Newton’s shear stress law, 845 Nickel-cadmium, 1038 Nickel-metal hydride, 1033, 1034, 1038 Nickel-metal hydride battery, 998, 1033 NiMH, 1038 NiMH battery, 706 Nitric oxide formation, 592 Nitriding, 108, 163 Nitrite amine phosphate, 860 Nitrocarburizing, 108 Nitrogen oxide emission, 728 Nitrous oxide, 754 Nitrous oxide mass emissions, 748 Noble metal electrode, 577 NO formation, 35 Noise characteristic, 952 Noise emission, 220, 817, 945 Noise-insulation materials, 959

Internal Combustion Engine Handbook | 1119

6606_Book.indb 1119

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Noise radiation, 947, 950, 953 Noise reduction, 952 Noise-reduction provisions, 950 Noise source, 949 Noise suppression, 803 Noncombusted, 43 Nonferrous-metal deactivator, 847 Nonslip condition, 804 Non-woven cotton fabric, 950 Non-woven fiber fabric, 951 Normal balance, 67 NOx Adsorber, 786 NOx emission, 30, 31 Noxious diesel, 782 NOx reduction, 767 NOx storability, 770 NOx storage catalytic converter, 769, 927 Nozzle carburetor, 3 Nozzle-holder assembly, 554 Nozzle-holder combination, 555 Nozzle-hole cross section geometrical, 556 Nozzle hole diameter, 534 Nozzle needle conical, 517 Nozzle-needle stroke, 555 Nozzles, 519 NSU Spider, 382 NTC, 593 O OBD II legislation, 731 OCPTM-air injector, 528 OCPTM injector, 528 Octane number, 590, 824, 829 Octane number requirement, 830 Odor improvers, 820 OEM, 851 Offset limit, 299 Offsetting, 85 OHC, 6, 225, 226 O–H–C-balance, 35 OHC equilibrium, 43 OHC-System, 592 OHV, 6, 226, 231 Oil naphtene-based, 844 paraffin-based, 844 Oil carbonization, 86 Oil consumption, 143, 411, 969 Oil consumption measurement technology, 968 Oil control ring, 85, 103 Oil cooling, 741 Oil dilution, 857 Oil feed geometry, 256 Oil mist separation integrated, 230 Oil oxidation, 847 Oil pan, 950 Oil pan fixing bolt, 292 Oil pan screw, 297 Oil program, 815, 826

Oil pump, 202, 933 Oil requirement, 143 Oil separation rate, 230 Oil slotted ring, 103 Oil stripping system three-part, 104 Olefin, 824 Olefin content, 834 On-board power supply, 921 Opacity measurement, 975 Open-deck design, 120 Opening characteristic, 465, 1082 Opening ramp, 234 Opening stress, 106 Opening time variation, 453 Operating characteristics, 15 Operating map, 28 Operating optimum, 938 Operating point, 494 Operating point shift, 80, 922, 923 Operating principle, 11 Operating strategy, 1014 engine, 28 Operating time difference, 63 Operational strength, 881, 886 Operational strength calculation, 879 Operational vibration analysis, 956 Opposed piston uniflow scavenging, 441 Optimization of geometry, 879, 1093 Optimization procedure, 877 Order, 947, 954 Order curve, 955 Origin, 818 OSEK operating system, 677 Otto Nikolaus August, 1 Outlet timing, 231 Output maximum, 18 Ovality, 106 Oval sprocket, 247 Oval sprocket technology, 247 Overall acoustic level, 272, 946 Overall efficiency, 921 Overall process analysis, 891 Overall transmission ratio, 924, 937 Overdrive, 924 Overhang, 256 Overhead camshaft, 225 Overhead valve = OHV, 164 Overhead valves = OHV, 225 Overlapping, 449 Overlay electroplated , 261 sputtered , 262 Overrun, 523 Overrun fuel cut-off, 522 Overrunning, 536 Overshooting, 188 Overspeed design, 937 Oxidation catalytic converter, 799 Oxidation catalytic converters, 818 Oxidation inhibitor, 847 Oxygen, 1043

Oxygen compound, 839 Oxygen storage unit, 762 P Package model, 602 Panel resonator, 951 Panic valve, 209 Paraffin, 824 Parallel crank gear, 80 Parallel hybrid, 703, 1031 Parameter fractal , 790 psycho-acoustic, 957 Partial flow dilution system, 974 Partial load, 78, 433 Partial load mode, 425 Partial-load range, 931 Partial lubrication, 842 Particle concentration, 866 Particle definition, 789 Particle erosion, 866 Particle filter catalytic, 808 Particle filter regeneration, 724 Particle Filters, 788 Particle mass, 818 Particle measuring technology, 805 Particle property, 789 Particles, 788 Particulate emission, 32 Particulate formation, 32 Particulate measurement, 974 PASS, 975 Passage of dust, 863 Passenger car fuel-cell drive system, 1043 Passenger cars, 1053 Past accelerating, 272 Patent application, 447 Peak-and-hold, 669 Peak combustion temperature, 728 Peak pressure, 16 PEM, 1042, 1048 PEM fuel cell, 1042, 1048 PEM fuel cell stack, 1043 Pencil coils, 572 Pencil stream valve, 532 Pendulum-slider pump, 207 Penetration, 796 Performance effective, 415, 487 Performance class, 850 Performance increase, 237 Performance test, 804 Permanently-excited synchronous motor, 706 Phase diagram, 72 Phase direction diagram, 75 Phase rotation angle, 225 Phases of the spark, 569 Phasing, 462, 463 Phlegmatization, 1016 Phosphating, 92

1120 | Internal Combustion Engine Handbook

6606_Book.indb 1120

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Photoacoustics, 975 pH value at 20 °C, 860 Piezo actuator, 467, 959 Pin nozzles, 552 Pipe, 227 Pipe Flow, 438 Piston, 880 cooled, 89 forged, 87 heat treatment, 91 installation play, 85 oil-cooled, 86 operating condition, 84 race car, 88 running surface protection, 92 temperature distribution, 86, 1067 wear protection, 92 weight optimization, 85 Piston acceleration, 51 Piston alloy, 93, 881 Piston material, 881 Piston annulus, 383 Piston boss, 85, 880 Piston calculation, 879 Piston cooling, 83, 87 Piston designs, 85, 87 Piston diameter, 83 Piston displacement, 78 Piston Displacement, 15, 16 Piston force, 53, 54 Piston groove temperature, 87 Piston groove wear, 109 Piston head, 86, 90 Piston loading, 83 Piston machines, 9 Piston manufacture, 91 liquid casting, 91 squeeze casting, 91 Piston mass, 85 Piston material, 83, 86, 93 Piston movement, 50, 85 Piston noise, 880 Piston pin deaxising, 53 Piston rattle, 55 Piston ring, 101 Piston ring design, 104 Piston ring first Piston ring, 86 Piston ring manufacturing, 107 Piston ring parameter, 106 Piston ring set, 105 Pistons, 83 Piston-side force, 52, 54, 883 Piston skirt, 85 Piston slapping, 85, 950 Piston speed mean, 18 Piston Speed, 17, 50 Piston stroke, 15 Piston temperature, 86, 882 piston travel, 15, 16 Piston travel, 50 Piston weight, 85 Plain bearing, 230 planimetry, 20 Plasma process, 794

Plasma spatter layers, 108 Plastic oil pan, 126 Plateau honing, 123 Play, 234 Play adjustment, 155 Pleated construction, 867 Pleating star, 867 Plug-in hybrid, 703 PM filter catalytic converter, 778 PM-Metalit®, 779 Pneumatic screwdriver, 298 Point contact, 233 Poise, 843 Polar diagram, 57 Polar diagram of forces, 254 Pollutant component, 29 Pollutant emissions impermissible, 564 Pollutant formation, 611 Pollutant reduction gasoline engine, 756 Polymer electrolyte membrane, 1042 Polystyrene casting cluster, 150 Pore size, 795 Porosity, 148 Port closing, 428 Port deactivation, 726, 727 Port development, 139 Port fuel injection engine, 623 Portion of water, 866 Port liner, 385 Position control, 721 Positive Displacement Superchargers, 491 Postmortem analyses, 786 Post start-up phase, 520 Potassium catalytic converter, 771 Potential chemical, 41 Pour point, 860 Pour point depressant, 847 Power indicated, 21 Power curves, 20 Power density, 997 Power electronics, 704, 995 Power hyperbola, 27 Power-MOSFET, 482 Power pack test bench, 962 Power split, 703 Power station mix, 943 German, 943 Power steering pump, 954 Power-to-weight ratio, 19, 922 Powertrain test bench, 962 POX reformer, 1049 Practical fuel consumption, 1033 Pre-chamber system, 614 Precision-casting, 157 Preforms, 122 Preinjection, 32 Preliminary catalytic converter, 692 Preload force, 281, 298

Press fit friction, 227 Pressure, 844 Hertzian, 233 Pressure build-up dynamic, 253 Pressure loss, 802 Pressure measurement, 961, 964, 965, 972 Dynamic, 965 Pressure oil supply, 230 Pressure regulating, 565 Pressure regulating valve, 867 Pressure sensor, 529 Pressure switch, 716 Pressure wave supercharging, 480 Pressurized carburetor, 518 Prestrainer, 868 Prethrottle actuator, 726 Prevention of lacquer formation, 847 Prime mover, 962 Process, 35 electrostatic, 794 gravimetric, 805 irreversible, 45 Rotation-angle-controlled, 294 Seiliger, 38 Process control, 38 Process efficiency, 35, 36 Process work, 42 Product calculation, 288 Product development virtual, 879 Profile steel ring, 104 Project management, 144 Protection from deposit, 843 Prototype manufacturing, 152 Prototypes, 144 Provision concerning emitted noise, 948 Provision for noise reduction, 950 Pulsating turbines impingement, 512 Pulse turbocharger-aspiration module, 483 Pulse turbocharger- control device, 482 Pulse turbocharger valve, 481 Pulse turbocharging, 471, 472, 475, 479, 1061, 1062 Pumpability, 818 Pump Angle Time Signal, 541 Pump limit, 509 Pump nozzle, 225 Pump-nozzle injection system, 562 Pump Nozzle System, 544 Pump nozzle technique, 156 Pump threshold temperature, 849 Pump-up, 227, 236 Push rod, 225, 422 Pushrod, 140, 225, 422 Push rod drive, 455 Push rod ratio, 15 p-v diagram, 36 p-V-Diagram, 20

Internal Combustion Engine Handbook | 1121

6606_Book.indb 1121

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

PVD layer, 108 PVD process, 261 Q Quadrocyclane, 841 Quality agreement, 298 Quality assurance, 152 Quality control, 13, 927 Quantity balance, 45 Quantity control, 13, 78 Quantity of heat removed, 36 Quick start glow system, 583 R Racing fuel, 840 Radial bearing, 252 Radial Compressor, 488 Radial force, 55 Radial force characteristic, 56 Radial piston distributor pump, 543 Radial piston pump, 543 Radial pressure distribution, 106 Radiator protector, 739 Radius roller, 163 Rail, 547 Ram induction turbocharging, 489 Ramp, 234 Ramp function, 686 Ram tube charging, 270, 426 Range extender, 1025 Rape methyl ester, 1029 Rapeseed oil, 814, 821 Rate of deposition, 796 Reaction kinetics, 35 Reaction mechanism, 35, 752 Reaction speed, 45 Reaction torque, 56 Reaction work reversible, 42 Recharging effect, 428 Reciprocating displacer engine, 9 Reciprocating piston compressor, 443 Reciprocating piston engine, 10 classification, 10 Recovered alloy, 147 Rectangular filter, 866 Rectangular ring, 102 Recuperation, 703 Reducing agent, 818 Reducing friction loss, 1062, 1063 Reduction in pollutants Diesel engine, 758 Reduction of friction, 922 Reduction of fuel consumption, 931, 941 Reduction of Mass, 232 Reed valve, 442 Reed valve control, 442 Reference fuel i-octane, 829 Reference fuel n-heptane, 829 Reflection-type muffler, 430 Reflection-type sound damper, 430

Reformate, 824 Regeneration, 796 Regeneration additives, 798 Regeneration aids, 820 Regeneration help, 797 Regeneration process, 797 Regenerative braking, 985, 1030, 1039 Regulating principle, 208 Regulating pump, 211 Regulation circuit, 237 Relaxation, 278 Reluctance machine, 993, 1037 Reluctance motor Switched, 1037 Removal efficiency, 866 Repeatability, 961, 968 Replacement component, 824 Reproduction reaction, 592 Required Ignition Voltage, 575 Requirement, 735 Requirement-Based Energy Management, 1055 Residual compressive force, 189 Residual exciter moment curve, 74 Residual exhaust gas, 40, 43 Residual gas scavenging, 480 Residue, 797 Resistance characteristic curve, 511, 512 Resistance to thermal shock, 148 Resonance, 954 Resonance charging, 426, 428 Resonance oscillation, 626 Resonator, 270 Response, 924 Response Behavior, 498 Response time, 801 Restricted exhaust component diesel engine, 755 gasoline engine, 753 Return rail, 156 Reversed head scavenging, 441 Reynold’s number, 516 Rib and pipe system, 737 Ring break, 107 Ring carrier, 87 Ring carrier material, 89 Ring carrier piston, 89 Ring gap, 106, 107 Ring set, 104 Ring type, 103 Ring with inner chamfer, 102 Ring with inner shoulder, 102 Rise in viscosity, 844, 856 RME, 1030 Ro 80, 382 Road-surface/tire noise, 949 Rocker arm, 164, 166, 225, 417, 420, 457, 463 Rocker arm actuation, 156 Rocker arm components, 453 Rocker arm module, 454 Rocker lever valve train, 167 Rod force, 54, 55 Roller bearing steel, 232

Roller cam follower, 166, 234, 236, 421, 455, 456, 463 Roller cam follower design, 455 Roller chain, 240 Roller lever, 234 Roller tappet, 234 switchable, 170 Roller test bench, 962 Rolling bearing, 230 Rolling circle, 383 Rolling contact, 225, 228, 229, 232, 235, 236 Rolling fatigue strength, 225 Rolling resistance, 918, 920, 921 RON, 829 Roots blower, 4 Rotary axis, 383 Rotary compressor, 444 Rotary-disk valve, 416 Rotary displacer engine, 9 Rotary piston engine, 6, 10, 382, 384 Rotary piston engines, 3 Rotating body, 383 Rotational irregularity, 952 Rotational oscillation, 72 Rotational oscillation damper, 77 Rotational oscillation state, 76 Rotational oscillation system, 72 Rotation angle technique, 296 Rotation-controlled tightening, 299 Rotor, 385 Rotor-type pump, 10 Roughness, 952, 957 Round filter element, 866 RSG, 251 Rubber vibration damper, 77 Run-down ramp, 226 Running-resistance curve, 27 Running surface coating, 107 Run-up ramp, 226 RVP = Reid Vapor Pressure, 833 RX 400 h, 1032, 1034 S SAE range, 844 SAE viscosity class, 848 Safety factor, 887 Safety strategy, 692 Sampling, 805 Sand-casting process, 122, 148 Sauter mean diameter, 528 Scales;Turbulent, 596 Scattered light measurement, 976 Scavenging, 439 Scavenging air supply, 443 Scavenging blowers, 443 Schnürle reverse or loop scavenging, 4 SCR catalytic converter, 818 active, 769 SCR catalytic converters, 818 passive, 768 Screaming noise, 953

1122 | Internal Combustion Engine Handbook

6606_Book.indb 1122

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Screw DIN EN ISO 24014, 299 Shafted, 296 Threaded, 296 Screw Compressor, 488 Screws Amenable to assembly, 292 Screw tightening Torque-controlled, 298 SEA statistical energy analysis, 959 Seal, 275 metallic, 282 Seal compatibility, 843 Sealing, 383, 842 Sealing gap fluctuation dynamic, 276, 277 Sealing pressure, 281 Sealing system, 275, 288 acoustically decoupled, 285 Seat contact area, 190 Seat-hole nozzle, 553, 554 Seating velocity, 234 Seat width, 195 Secondary airborne noise, 950 Secondary air injection, 764 Secondary emission, 801 Secondary piston motion, 85, 880 Secondary piston motion calculation, 880 Secondary radiation, 950 Second order inertial force oscillating, 63 Second-order inertial force, 61 Segmented camshaft, 229 Seiliger Myron, 37 Seizure, 296 Selective catalytic reduction, 768 Self-balance, 68 Self-diagnosis, 691 Semi-downdraft carburetor, 518 Semi-surface gap, 574 Sensitivity to water, 825 Sensor, 868 Sensors, 468 Sensor system, 280 Separately-excited synchronous motor, 706 Separating valve, 386 Separation system active, 230 Sequential supercharging, 923 Sequential Turbocharger, 503 Serial hybrid, 703 Serpentine drive, 249 Serpentine stopper, 277 Service life Calendrical, 1003 Service life analysis, 236 Service life prediction, 164 Service life test, 290 Severity, 831, 957 Shaft centrally symmetrical, 68

Shaker, 291 Shaker cooling, 90 Shaped bore, 886 Shaped bore of piston boss, 881 Shear speed, 845 Sheet metal manifolds, 300 Sheet-metal oil pan, 950, 953 Shell Model, 595 Shifter unit hydraulic, 237 Shifting, 52 Shifting the operating point, 984 Short-time fatigue, 886 Shoulder/bevel ring, 103 Shoulder ring, 103 Shrink-fit thermal, 227, 230 Shrink-fit conrod, 85 Shunt resonator, 273 Shutting off the fuel, 522 Sideband, 947 Signal redundant, 726 Signature analysis, 954 Signs of coking, 797 Silent chain, 241 Silicon carbide, 1030 Silicon crystal layer, 122 Simulation, 290, 598, 875 Cooler, 736 one dimensionally, 436, 438 thermodynamic, 881 Simulation of flow CFD, 222 Simulation process integrated, 163 Simulation system, 43 Simulation tool acoustic, 274 Single-barrel carburetor, 517 Single-cylinder crank gear, 61 Single-grade oil, 849 Single injection pump, 225, 538 Single power driver, 291 Single-shaft machine, 1042 Single-shaft turbine, 1042 Single spark ignition coil, 571 Single valve drive model, 460 Single-zone model, 42 Sintered aluminum, 452 Sintered cams, 228, 229, 235 Sintered conrod, 294 Sintered material, 228, 229 Sintered metal plate, 793 Six port induction, 386 Size distribution of the particles, 790 Sleeve, 225 Sleeve concept, 121 Sleeve support surface, 121 Slide support, 454 Sliding contact, 225, 227, 229, 232, 235, 236 Slippage, 250 Slot control, 11 Smoke emission, 975

Smoke-emission value measurement, 975 Smoke meter, 975 SMPS method, 805 Snugged down, 296 Snugging moment, 298 Sodium-nickel-chloride battery, 999 SOFC, 1048 SOFC-APU, 1049 SOFC fuel cell, 1047, 1048, 1049 Software structure, 676 SOHC, 231 Solenoid valve needle, 543 Solid bearings, 263 Solid-borne engine noise, 948 Solid-borne noise, 945 Solid-borne noise contribution, 956 Solid-borne noise oscillation, 952 Solid-borne noise path, 952 Solid-borne noise transmission, 953 Solidification, 148 Solid Oxide Fuel Cell, 1048 Solid particle, 804 Solubilizers, 823 Solvent refinement, 845 Sommerfeld number, 252 SON, 829 Soot, 818 Soot combustion, 797 Soot emission, 975 Soot ignition temperature, 798 Soot regeneration, 809 Sound design, 950 Sound Engineering, 957 Sound package, 959 Sound pressure, 946 Sound volume, 957 Source of solid-borne noise, 952 Spark failure, 576 Spark ignition, 569 Spark-Ignition Engine, 10 Spark plug, 138, 154, 573 Design, 573 Spark plug position, 931 Spark Position Advanced, 574 Normal, 574 Spatial velocity, 804 Special seals, 280 Specific fuel consumption, 923 Specific power output, 19 Spectrum, 947, 950 Speed critical, 73 hydro-dynamic, 236, 1078 hydro-dynamically effective, 235 Speed fluctuation, 60, 72 dynamic, 72 static, 72 Speed measurement, 963, 964 Speed reduction, 495 Speed sensor, 718 passive, 718 Spiral turbocharger, 7, 488 Spiral V filter, 867

Internal Combustion Engine Handbook | 1123

6606_Book.indb 1123

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Split-pin crankshaft, 71 Split stream valve (Split Stream), 532 Sports engine, 158 Spray cooling, 87 Spring characteristics, 188 Spring force, 186, 235 Sprocket, 241, 245 Sprue, 149 Sputter bearing, 263, 265 Squeeze Casting, 122 SS50 system, 104 Stage-regulating pump, 210 Standard honing, 123 Standardized Mass Flow, 509 Standardized Volumetric Flow, 509 Standard pressure, 40 Start automatic, 520 Start behavior, 579 Start control, 33 Starter-alternator, 251 Starter-alternator system, 251 Starter/Generator, 1031 Starter/Generator function, 1031 Starting control, 519 Start quality, 581 Start reaction, 592 Start/Stop, 1011, 1031 Start-stop system, 251 State change, 37 State equation thermal, 40, 44 State of Charge, 1003 State of charge regulation, 1013 State of Function, 1003 State of Health, 1003 State variable intensive, 45 Stationary engine, 229 Stator, 384, 385 Status vector, 437 Steam engine, 3 Steel cams, 228, 229 Steel camshaft, 229 forged, 229 solid material, 229 Steel die, 149 Steel spring cylindrical, 420 Steel strip ring, 104 Stepping motor, 722 Stiffness, 225, 236, 953 Stirling engine, 10, 1031, 1040, 1041 stoichiometric full-load operation, 928 Stopper, 277, 289 coined, 277 topographic, 277, 278 Stop/start, 984 Stop/Start, 703 Stop/start system, 939 Storage capacity, 792 Storage time, 804 Strainer, 867 Strain under Load, 187 Stratified charge, 727 Stratified-charge concept, 930

Stratified charging, 530 Street evaluation number, 831 Street octane number, 831 Strength analysis, 147, 1069 Stress mechanical, 231 thermal, 231 thermo-mechanical, 879 Stribeck curve, 253 String oscillation, 953 Strip apex, 383 Strip method, 390 Stroke Changeover, 169 Stroke function, 15 Stroke offset, 71 Stroke ratio, 79 Stroke variation, 457 Structural resonance, 953 Structural stiffening provision, 953 Substance absorbed, 791 Substance component, 40 Substitute combustion characteristic, 42, 43 Substitute Combustion Process, 599 Substitute fuel, 590 Substrate sintering, 766 Sudden load, 234 Sulfate, 818 Sulfate ash content, 846 Sulfate decomposition, 774 Sulfate regeneration, 774 Sulfur content, 818, 826, 834 Sulfur dioxide, 818 Sulfurization reaction, 799 Sulfur Poisoning, 773 Summer quality, 833 Sump pump, 217 Sunfuel, 1029 SunFuel, 1030 SuperCap, 1001 super capacitors, 1039 Supercharger design, 443 Supercharging, 12, 39, 80, 922 Mechanical, 492 Super diesel, 814, 820 Supplementary air, 158 Supplied ignition, 12 Support element hydraulic, 166 Surface hydrophobic, 868 Surface contact, 233 Surface corrosion, 861 Surface gap, 574 Surface pressure, 225 Surface ratio, 79 Surface roughness, 123 Surface temperature, 300 Surface tension, 820 Surface treatment, 108 Surface-volume-ratio, 16 Surplus force, 937 Suspension attachment member, 952 Swash plate, 530

Swept volume variable, 78, 79 Swing motor, 450 Swirl effect, 146 Swirl generator, 230 Swirl plate, 726 Switched reluctance motors, 1037 Switching pushrod, 455 Switching regulator, 673 Switching tappet, 227 Synchronous belt, 225 double-sided, 244 Synchronous belt dynamics, 247 Synchronous belt profile, 244 Synchronous belt screaming, 953 Synchronous machine Permanently excited, 1037 Separately excited, 1037 Synchronous machine (SM), 992 Synchronous motor, 1037 Synfuel, 836, 1029 SynFuel, 1030 Synthetic material, 453 Synthetic material aspiration pipe, 483 System acoustically decoupled, 288 Cam-edge-controlled, 537 Central Pressure Reservoir, 545 closed, 37, 40, 42 electro-mechanical, 467 Electronic on-demand regulated, 566 nozzle-time controlled, 537 On-demand regulated, 564 with Injection-Synchronous Pressure Generation, 537 System block diagram for CR, 550 System integration, 482 System limit, 46 T Takeoff, 226 Takeoff element, 227 TAME, 826 Tandem turning, 107 Tangential force, 55, 105 average, 58, 59 Tangential force characteristic, 55, 58, 59 Tangential force curve, 75 Tank diagnosis, 732 Tank-to-wheel, 1042 Tank ventilation valve, 730 Tank venting system, 560 Tappet, 225, 417, 422 Tappet chamber, 466 Targeted motion, 235 Target value generator, 469 Temperature coefficient More negative, 593 Temperature distribution, 880 Temperature map, 880, 882 Tensile strength, 229, 232

1124 | Internal Combustion Engine Handbook

6606_Book.indb 1124

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Tension mechanical, 881 thermal, 881, 886 Tension and deformation, 880 Tensioning pulley automatic, 247 Tertiary Butyl Alcohol, 824 Test aerosol, 804 Test bench, 962, 963 dynamic, 961 Dynamic, 962 Stationary, 962 Test bench automation system, 976 Test bench measuring technology, 961 Test bench overall system, 961 Test bench rollers, 962 Test bench system, 961, 962 Testing equipment servo-hydraulic, 290 Test method engine, 855 Tetrahedron element, 881 The , 40 The filling and emptying method, 43 The Fuel Cell as a Vehicle Drive System, 1042 Thermal capacity specific, 881 Thermal conductivity, 193, 881 Thermal dilator, 805 Thermal efficiency, 923, 926 Thermal piston expansion, 883 Thermal transition, 43 Thermal transmission, 45 Thermodesorber, 805 Thermo-mechanical fatigue, 887 Thermostatic valve, 868 The Steam Drive, 5 The Steam Motor, 5 Thickening, 847 Thick-wire bonds, 664 Think City, 1038 Thinking ahead, 939 Thiophene sulfur, 834 Threaded connector, 291 Three-mass flywheel, 78 Three-material bearing, 4, 263 Three-phase A.C. inverter, 704 Three-phase A.C. motor, 704 Three-phase motor, 1037 Three-shaft turbine, 1042 Three valve concept, 454 Three-valve cylinder head concept, 155 Three way catalytic converter, 761 Throttle-free load control, 472 Throttle losses, 466 Throttle valve, 517, 521, 530, 531, 723 Throttle valve actuator, 521, 723 Throttle valve control, 687 Throttling, 78 Throw, 63 Throw angle, 64 Throw sequence, 68, 71 Thrust washer, 253, 263

THS II, 1032, 1033 Tightening Rotation-angle-controlled, 293 Tightening technique 0.2% offset limit, 296 Tight seating, 257 Tilting movement, 52 Timing, 139, 444 fixed, 234 Timing chain, 231 Timing chain drive, 240 Timing chain screaming, 953 Tiptronic, 699 Tire noise, 952 tire-specific constants, 921 TMF, 887 Toluene, 833 Tonality, 957 Toothed gear, 225 Toothed V-belt drive, 249 Tooth engagement frequency, 953 Topology optimization, 877 Torque, 233, 444, 964 internal, 70 Torque balancing, 68 Torque curve, 18 Torque differential gear, 68 Torque measurement, 963 Torque motor, 722 Torque tightening, 298 Torque vector, 64 Torque wrench, 298 Torsional force, 58, 59 Torsional load, 233 Torsional rigidity, 72 Torsion break, 73 Torsion resistance, 229, 232 Torsion stiffness, 232 Total emitted noise, 950 Total O2 content, 826 Total pressure, 509 Total pressure ratio, 509 Toyota Hybrid System, 1032 Toyota Prius, 1032 Toyota Prius II, 1034 Traction battery, 704, 1037 Traffic noise, 949 Transesterification of rapeseed oil, 821 Transfer path analysis, 956, 957 Transition metal, 798 Transition period, 833 Transmission Power-split, 1034 Transmission function acoustic, 948 Transmission housing, 950 Transmission noise radiation, 953 Transmission ratio, 924, 936 Transmission route for solid-borne noise, 952 Transmitting force, 842 Transport process, 44 Transverse flux machine, 994, 1038 Transverse flux motor, 1037

Trapezoid ring single, 102 Trends, 1061 Trigger wheel, 226, 237 Trilobe camshaft, 227 Trimming control, 683 Triple-barrel carburetor, 517 Triple Flame, 598 Trochoid housing, 384 T-s diagram, 36 Tube, 227 Tube diameter, 232 Tube-runner manifold, 299 Tube-runner manifolds, 302 Tumble effect, 146 Tumble-flow, 436 Tunnel bearing, 225, 226 Turbine efficiency, 510, 511 Turbine geometry Variable, 728 Turbine Geometry Variable, 496 Turbine map, 490, 508, 510 Turbine pressure ratio, 510, 511 Turbocharger, 728 Turbocharger Characteristic Maps, 507 Turbocharger efficiency, 510 Turbocharger internals, 489 Turbocharger main equation, 493 Turbocharger mapping, 890 Turbocharger speed, 509 Turbocharger Test Rigs, 507, 508 Turbocharging Two-stage, 496 Turbocompound supercharging, 1063 Turbo lag, 728 Turbulence Intensity, 596 Twin bank engine, 12 Twincharging, 923, 925 Twin-shaft gas turbine, 1042 Twin-shaft turbine, 1042 Twin-shaft vehicle gas turbine, 1041 Twin Turbo System, 503 Twisting, 102 Two-cycle diesel engine, 313 Two-cycle engine, 803 Two-cycle piston, 88 Two-mass fly wheel, 77, 78, 952 Two-mass oscillator, 685 Two-mass rubber vibration damper, 77 Two-material bearing, 263 Two-Mode concept, 1036 Two-Mode hybrid system, 1036 Two-phase spray model, 609 Two-spring, nozzle-holder assembly, 555 Two-stage carburetor, 518 Two-staged ignition, 593, 594 Two-stage system, 464 Two-stroke engine, 312, 1050 Two-Stroke engine Charge Cycle, 439 Two-stroke process, 11

Internal Combustion Engine Handbook | 1125

6606_Book.indb 1125

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Two-stroke SI engine, 6, 650 Two-valve cylinder head, 138 Two-zone model, 43 Type test, 804 U Ultracap, 1001 UltraCaps, 1039 Underbody catalytic converter, 762 Undershield, 951 Underspeed design, 937 UNIAIR system, 172 Uniflow scavenging, 313, 314, 441 Unit-Pump-System, 539 UniValve, 463 UniValve-Valve train, 464 Unregulated exhaust component diesel engine, 756 gasoline engine, 754 Updraft carburetor, 518 Urea consumption measuring technology, 965, 970 Useful work, 36 Using CAD, 145 Utility vehicle fuel cell drive system, 1043 V Vacuum distillation, 845 Valve, 225, 234, 419 opening time, 231 Valve acceleration, 225, 234, 235, 465, 1077 Valve actuation, 11, 226 electro-mechanic, 468 fully-variable, 230, 234, 456 hydraulic, 467 hydraulic variable, 466 partially-variable, 230, 234 variable, 444 Valve actuation dynamic, 879 Valve angle, 138, 139 Valve area, 416 Valve break, 466 Valve clearance, 227, 234 Valve clearance compensation, 423 mechanical, 234 Valve contouring switching, 454, 470 Valve cover fixing bolt, 292 Valve cross section, 140 Valve diameter, 139, 234 Valve disk, 140 Valve drive friction, 229 Valve drive kinematics, 233 Valve flow number, 436 Valve guide, 196, 201 material, , xiii Valve guide clearance, 197 Valve guide load, 196 Valve guide material, 201 Valve lever, 417, 421

Valve lifter, 225 defeatable, 169 hydraulic, 164, 165 mechanical, 164 Valve lift modes, 173, 1070 Valve linkage, 225, 236 Valve opening area, 416 Valve opening time, 234 Valve overlap, 24, 234, 236 Valve play adjustment hydraulic, 227, 236 mechanical, 168, 227 Valve play compensation hydraulic, 168 Valve seat, 194, 234 Valve seat insert contour, 194 Valve seat ring, 189 material, 191 Valve setting down velocity, 469 Valve shaft end, 236 Valve shroud module, 286 Valve spring, 186, 225, 420 Valve spring force, 424 Valve stem seal, 197 Valve stroke, 225, 227, 234, 235, 1077 kinematic, 227 theoretical, 236, 237, 1078 Valve stroke change, 227, 231 Valve stroke curve, 234, 235 Valve stroke height, 234 Valve stroke variation, 459 Valve timing, 137, 225, 227, 229, 234, 236, 467 variable, 137 Valve train, 164, 234, 416, 952 electro-mechanic, 468, 1082 fully-variable, 172 mechanically fully-variable, 231 variable, 169, 172, 227 Valve train component, 143, 146, 164 Valve train configuration, 226 Valve train design, 142 Valve train system partially-variable, 231 Valve train type, 234 Valvetronic, 458, 461 Valve velocity, 225, 235, 1077 Vane pumps, 206 Vane shifter, 238 V-angle, 71 natural, 71 Vapor bubble formation, 832 Vapor cushion, 832 Vapor pressure, 826, 833 Variable valve actuation, 930 variable valve actuation times, 929 Variable valve train, 929 VarioCam Plus, 227 VARIOVALVE, 463 V crank gear, 62 Vehicle category, 920 Vehicle curb weight, 917 Vehicle mass, 920 Vehicle test bench, 962 Vehicle weight, 917, 919, 920

V-Engine, 12 Venturi section, 516 VF system, 104 Vibe function, 43 Vibe Function, 599 Vibration, 952, 959 Vibration damper, 77 Vibration generator, 959 Vibration nodes, 72 Vibratory testing system, 291 V.I. determination, 844 V.I. improver, 845, 846, 847 V.I. scale, 844 Viscose vibration dampers, 77 Viscosity, 817 kinematic, 849 Viscosity-temperature behavior, 843, 847 Viscosity/viscosity index (V.I.), 843 Viscous coupling, 743 Viscous damper, 77 Volatility, 825 Volatility class, 833 Volume specific, 40 Volume ratio, 79 Volumetric Efficiency, 23, 415 Volumetric flow measurement, 965 W Waisted-shank bolt, 293 Waisted-threaded screw, 293 Wall film, 895 Wall flow, 792 Wankel Engine, 10 Wankel engine oil, 858 Warm-up, 519, 520 Warm-up phase, 939 Washcoat, 784 Washer, 293 Waste gate function, 726 Water specific heat, 46 Water cooling, 124, 143 Waterfall diagram, 954 Water hardness, 860 Water jacket, 122, 124 Water jacket depth, 124 Water outlet, 868 Water removal, 868 Water sensor, 868 Wax Anti-Settling Additive (Wasa), 818 Wear, 225, 870 Wear at the flanks, 109 Wear characteristics, 161 Wear mechanism, 577 Wear protection, 836, 843 Wear-protection layer, 107 Web width, 276 Weight increase, 920 Weighting curve, 946 Weight reduction, 920 Weight trend, 921

1126 | Internal Combustion Engine Handbook

6606_Book.indb 1126

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

Index

Well-to-wheel, 1030, 1046 W-Engine, 12 Wet sump lubrication system, 386 Wheel hub drive, 1039 Wheel hub motor, 994, 1039 Wheel resistance, 920 Whirl chamber diesel engine, 533 Whirl chamber system, 615 Wiese scale, 830 Winding, 107 window, 299 Window method, 955 Winter quality, 833 Wire cross section, 188

Wöhler curve, 887 Work, 42 indicated, 36 Work area, 37 Work fluctuation, 60 Working point, 27 Work process calculation, 889 Woschni, 599 Wristpin, 880 designs, 94 flexure and oval distortion, 880 floating, 85 Wristpin bushing, 263 Wristpin circlip, 83, 94

Wristpin diameter, 83 Wristpins, 83, 85, 94 Wristpin steel, 95 WTW, 1046 X Xylene, 833 Z Zero Emission Vehicles, 1031 ZEV, 1031

Internal Combustion Engine Handbook | 1127

6606_Book.indb 1127

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

6606_Book.indb 1128

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

About the Editors Dr.-Ing. E .h. Richard van Basshuysen, VDI, wa s b or n i n 1932 i n Bingen/Rhein, Germany. Fol low i ng vo cat ion a l training with examination as a car mechanic, he studied at the technical college Braunschweig/ Wolfenbüttel from 1953 to 1955 and obtained a degree as a mechanical engineer. In 1982, he was granted the degree “Diplom-Ingenieur.” Between 1955 and 1965, he worked as a research associate for Aral AG in Bochum. In 1965, he joined NSU AG where he was employed on test management for engine and transmission development, including Wankel engine development, and was appointed deputy manager of vehicle testing. In this capacity, he was co-responsible for the development of the Prinz 4, NSU 1000 and 1200, RO 80 and K 70 cars. In 1969, NSU AG was taken over by what is today, Audi AG. As the development manager at Audi AG, he established the V8/A8 vehicle comfort class, became the manager responsible for engine and transmission development and, at the same time, the member of the Supervisory Board as the elected representative of the executive employees. His most important development was the worldwide first exhaustpurified diesel engine in a passenger car with direct injection and turbocharging which he pushed against major opposition within the company and the VW group. Because this engine

consumed 20 % less fuel than its predecessor as a chamber diesel engine and is an engine with high performance and very high torque, it has succeeded worldwide. In Europe, its market share has grown from approximately 12 % in 1989 to about 50 % in just over a decade. Following his active career in the automotive sector, Richard van Basshuysen founded an engineering consultancy firm in 1992 that he still manages. For twenty years, he was the publisher of the internationally-distinguished ATZ (Automobiltechnische Zeitschrift) and MTZ (Motortechnische Zeitschrift) technical/ scientific professional journals. He consults for international automotive manufacturers and engineering service providers and is author and publisher of technical /scientific textbooks which have been translated into English and Chinese. Since 2006, he and Prof. Dr. Ing. Fred Schäfer have been publishers and co-authors of www.motorlexikon.de. In addition, he was council member and board member on various bodies such as the German Association of Engineers (VDI) and the Austrian Association for motor vehicle engineering. He is author and co-author of more than 60 technical/scientific publications. For his development of the pioneering diesel engine with direct injection, in 2001 he was awarded the prestigious ErnstBlickle-Preis 2000 and the BENZ-DAIMLER-MAYBACH medal of honor of the VDI for “his outstanding engineering work in the development of the passenger car diesel engine with direct injection and his long-lasting commitment as publisher of ATZ/MTZ and member of the council of the VDI society vehicle and traffic engineering.” In 2004, the University of Magdeburg awarded him an honorary doctorate for his lifetime achievements.

Internal Combustion Engine Handbook | 1129

6606_Book.indb 1129

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017

About the Editors

Prof. Dr.-Ing. Fred Schäfer was born in 1948 in Neuwied a m R hei n, G er m a ny. After vocational training as a machine builder, he studied mechanical engineering at the State School of Engineering in Koblenz. Following this, he completed a course of studies at the University Kaiserslautern in the subject of motors and machines and was granted the “Dipl.-Ing.” degree. He received the Dr.-Ing. from the Institute for Motors and Machines at the University Kaiserslautern upon submission of his thesis on “Reaction-kinetic examinations of the hydrogen/methanol combustion in a spark-ignition engine.” His ongoing career led him to Audi AG in Neckarsulm, where he began as assistant to the development manager. Subsequent positions during his

ten years with the company were principal group leader for engine testing and manager of the engine design department, In 1990, he was appointed professor for motors and machine at the then technical college in Iserlohn which is now a part of the technical college Südwestfalen located in Iserlohn. One of his responsibilities was the management of the laboratory for combustion engines and turbo-engines. Prof. Dr.-Ing. Schäfer was active in many bodies of the college, including the college senate. In his function as vice dean for teaching and research, he was a member of the management committee of the mechanical engineering department. Prof. Dr.-Ing. Schäfer is also active in freelance research and development in engine technology. Amongst other things, he and Dr. van Basshuysen have been the publishers of the “Shell-Lexikon Verbrennungsmotor” magazine supplement until 2003, which was published in 2004 as a book with the title “Lexikon Motorentechnik.” He and Dr.-Ing. E.h. van Basshuysen are co-publishers and co-authors of www.motorlexikon.de and the “Internal combustion engine handbook.” Prof. Dr.-Ing. Fred Schäfer is a long-standing member of the VDI and SAE.

1130 | Internal Combustion Engine Handbook

6606_Book.indb 1130

1/19/16 9:06 PM

Downloaded from SAE International by Indian Institute of Technology - Chennai, Tuesday, May 02, 2017