2.61 — Spring 2004 Internal Combustion Engines
John Heywood Wai Cheng
2.61 — Spring 2004 Internal Combustion Engines
John Heywood Wai Cheng
MIT OpenCourseWare Cambridge, MA http://ocw.mit.edu/
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Contents Highlights of this Course ...................................................................... 1 Course Description .............................................................................. 1 Syllabus ............................................................................................ 2 Calendar............................................................................................ 3 Readings ........................................................................................... 5 Lecture Notes ................................................................................... 13 Overview and Basics of Engine Operation Engine Geometry and Performance Parameters Ideal Cycle Analysis and Combustion Stoichiometry Gas Exchange: 4-Stroke and 2-Stroke Spark-ignition Engine Combustion and Flame Propagation and Structure Knock SI Engine Emissions Emission Control Technology and Review Mixture Preparation and Engine Friction Engine Heat Transfer and 2-Stroke SIE Performance 4-Stroke SIE Performance Variable Valve Control and Gasoline DI Engines Diesel Overview and Diesel Combustion Future Engine Technology and Discussion Labs...............................................................................................110 Lab 1 Lab 2 Assignments ...................................................................................117 Problem Problem Problem Problem Problem Problem
Set Set Set Set Set Set
1 2 3 4 5 6
Exams ............................................................................................141 2002 Quiz 1 2002 Final Exam Project ...........................................................................................153 Project Description Exhaust Temperature Data, with EGR Emissions Data One of Many Possible Solutions Related Resources............................................................................174
Highlights of this Course This course features homework assignments, labs and an extensive reading list.
Course Description This course elaborates on the fundamentals of how the design and operation of internal combustion engines affect their performance, operation, fuel requirements, and environmental impact, study of fluid flow, thermodynamics, combustion, heat transfer and friction phenomena, and fuel properties, relevant to engine power, efficiency, and emissions, examination of design features and operating characteristics of different types of internal combustion engines: spark-ignition, diesel, stratified-charge, and mixed-cycle engines. The project section details the Engine Laboratory project. We have aimed this course for graduate and senior undergraduate students.
Course Meeting Times Lectures: Two sessions / week 2 hours / session
Level Graduate
1
Syllabus Text Heywood, John B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, 1988. ISBN: 007028637X. Grading Grade will be based on ACTIVITIES
PERCENTAGES
Quiz + Exam
45%
Homework
25%
Lab Report
20%
Design Problem
10%
Total
100%
Homework Homework Policy 1. The purpose of the homework and problems is to get each of you to think about and use the material we discuss in class. 2. Obviously, I want each of you to make a serious try at each problem. If discussions with other students help you get started, then contribute to and benefit from such discussions. 3. However, I expect that each of you will work independently on the details of the problem solutions which you hand in as your work. I regard it as dishonest to copy from any previously circulated solutions and present such work as your own. 4. If you have any questions about the above, please discuss them with me.
2
Calendar SES #
TOPICS
KEY DATES
Overview 1
Problem set 1 out Basics of Engine Operation Engine Geometry
2 Performance Parameters Engine Disassembly
Problem set 1 due
In Lab
Problem set 2 out
3
Ideal Cycle Analysis 4 Combustion Stoichiometry Fuel-air Cycle Model
Problem set 2 due
Fuel-air Cycle Results
Problem set 3 out
5
Gas Exchange: 4-Stroke 6 Gas Exchange: 2-Stroke Spark-ignition Engine Combustion 7 Flame Propagation and Structure Problem set 3 due 8
Recitation Problem set 4 out
9
Knock
10
SI Engine Emissions Emission Control Technology
11 Review
3
12
Quiz (Duration: 1-1/2 hours) Open Book
Problem set 5 out
Mixture Preparation 13 Engine Friction Engine Heat Transfer
Problem set 5 due
2-Stroke SIE Performance
Problem set 6 out
14
15
4-Stroke SIE Performance Variable Valve Control
16 Gasoline DI Engines Diesel Overview
Problem set 6 due
Diesel Combustion
Design problem out
17
Diesel Emissions 18 Emission Control Turbocharging 19 Diesel Performance 20
1-hour Lab Preparation
21
2-hour Lab Session
Design problem due
Engine Noise
Lab report due 1 week after session 22
22 Engine Dynamics Future Engine Technology 23 Discussion 24
Final Exam (Duration: 3 hours) Open Book
4
Readings The readings referred to the table below are recommended readings from the text: Heywood, John B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, 1988. ISBN: 007028637X. SES # 1
TOPICS
READINGS
Overview Essential Basics of Engine Operation Chapter 1 Essential
2
Engine Geometry Section 2.2 Essential Performance Parameters Chapter 2
3
Engine Disassembly In Lab Essential
4
Ideal Cycle Analysis Sections 5.1-5.4 Essential Sections 3.1-3.4 Combustion Stoichiometry
Sections 3.7, 4.1, 4.3 (Working fluid properties) Helpful Section 3.5
5
Sections 4.2, 4.4, 4.9.1 (Working fluid properties) Essential Sections 4.5, 4.6, 5.5 5
Fuel-air Cycle Model Helpful Section 4.7 Fuel-air Cycle Results Essential Sections 6.1, 6.2, 6.5
6
Gas Exchange: 4-Stroke Helpful Section 6.3 Essential Section 6.6 Sections 8.1, 8.2.1 (Incylinder flows) Gas Exchange: 2-Stroke Helpful Section 6.7 Sections 8.3-8.5 (In-cylinder flows) Essential
7
Sections 9.1, 9.2.2, 9.2.3, 9.3.1, 9.3.2, 9.4.1, 9.4.3
Spark-ignition Engine Combustion
Helpful
6
Sections 9.2.1, 9.3.3, 9.3.4, 9.4.2, 9.5.1 Flame Propagation and Structure 8
Recitation Essential Section 9.6.1
9
Knock Helpful Sections 9.6.2, 9.6.3 Essential
10
SI Engine Emissions
Sections 11.1, 11.2.1, 11.2.3, 11.3, 11.4.1, 11.4.3, 11.6.1, 11.6.2, 7.4 Essential
SI Engine Emissions (cont.)
11
Sections 11.1, 11.2.1, 11.2.3, 11.3, 11.4.1, 11.4.3, 11.6.1, 11.6.2, 7.4
Emission Control Technology Review
12
Quiz (Duration: 1-1/2 hours) Open Book Essential Sections 7.1, 7.3.1, 7.6
13
Mixture Preparation Table D4, p. 915 (Fuels) Helpful
7
Section 7.5 Essential Sections 13.1 – 13.3, 13.5 Engine Friction Helpful Section 13.6 Essential Sections 12.1, 12.2, 12.3 14
Engine Heat Transfer Helpful Sections 12.4, 12.7 Essential Sections 15.1, 15.2 2-Stroke SIE Performance Helpful Sections 15.3, 15.4 Essential Sections 15.1, 15.2
15
4-Stroke SIE Performance Helpful Sections 15.3, 15.4 Essential 4-Stroke SIE Performance (cont.)
Sections 15.1, 15.2 Helpful Sections 15.3, 15.4
8
16
Variable Valve Control Gasoline DI Engines
17
Diesel Overview Essential Sections 10.1, 10.2, 10.3, 10.7 Diesel Combustion
Helpful Sections 10.5, 10.6.1, 10.6.3 Sections 10.6.2, 10.6.5 (Diesel fuels) Essential Sections 11.2.4, 11.4.4, 11.5.2
18
Diesel Emissions
Helpful Sections 11.5, 11.6.4 Section 12.5 (Diesel heat transfer)
Emission Control 19
Turbocharging Essential Sections 15.5, 15.7 Diesel Performance Helpful Section 15.6
20
1-hour Lab Preparation
9
21
2-hour Lab Session Essential
22
Section 7.6, "Engine Noise." In Heywood, J. B. and E. Sher, The Two-Stroke Cycle Engine. Philadelphia PA: Taylor and Francis, 1999.
Engine Noise
Engine Dynamics 23
Future Engine Technology Discussion
24
Final Exam (Duration: 3 hours) Open Book
Corrections to the text Internal Combustion Engine Fundamentals Chon, Dale M., and John B. Heywood. "Performance Scaling of SparkIgnition Engines: Correlation and Historical Analysis of Production Engine Data."SAE (Society of Automotive Engineers) Technical Paper 2000-01-0565 (2000). Berckmüller, M., et al. "Potentials of a Charged SI-Hydrogen Engine." SAE (Society of Automotive Engineers) Technical Paper 2003-01-3210 (2003).
10
11
Corrections to text (current printing): (JBH 10/29/01)
p. 77
Table 3.2. The enthalpies of formation for C8H18 are for n-octane. For isooctane they are –224.1 and –259.3 MJ/kmol for gas and liquid C8H18 , respectively.
p. 89:
Middle of page: xCO 2 , xCO and x 02 should be x˜CO 2 , x˜CO , and x˜ O2 .
p. 122: Figure 4-10 is a repeat of Fig. 4-3 due to an editing error, though Fig. 4-10 is correctly labeled “burned mixture properties.” A correct Fig. 4-10 is attached. It is only slightly different: e.g., at 1000 K the burned mixture us for I 1.2 is 4% lower than the unburned mixture value, and hs is 1% lower than the unburned mixture value. These differences scale, approximately, with I . p. 151
Underneath Eq (4.65) insert: K is given by Eq. (4.63)
p. 152: Line 5. C mH nOr should be C nHm Or . p.188
In Eq. (5.66c), m is omitted. It should read: §T · §T · S3b S2 mcv ln¨ 3 a ¸ mc p ln ¨ 3b ¸ mcv ln D mc p ln E © T2 ¹ © T3a ¹
p. 306: Equation (7.18): The sign at the beginning of the second line of the equation (a minus sign) should be a plus sign. p. 388: Equation (9.27). The sign in front of the third term in the square bracket should be , not + : ªT c T 1 § J 1 ·º i.e., « ln¨ ¸» ¬ Tw Tw (J 1) bTw © J c1¹ ¼ p. 553: Equation (10.37). There should be a + sign between the two round brackets within the square bracket., i.e., 0.63 ª § 1 1 · § 21.2 · º W id (CA) (0.36 0.22Sp )exp«E A ¨ ˜ » ¹ ¨ © RT 17,190 ¸ ¹ ¼ © p 12.4 ¸ « » ¬
p. 620: The reference for Fig. 11-33 should be Yu, R.C., Wong, V.W., and Shahed, S.M., “Sources of Hydrocarbon Emissions from Direct Injection Diesel Engines,” SAE paper 800048, SAE Trans., vol. 89, 1980. (This is a new reference; make it reference 87 and add it to p. 667.) p. 679: In the inserted graph in Figure 12-5, the scale for thermal conductivity k g is not correct. The values should be multiplied by 5 x 105: e.g., the peak value of 10 x 10-8 = 10-7 W/m.K should be 10-7 x (5 x 105) = 5 x 10-2 W/m.K. p. 880
In Fig. 15-45, the units for pressure (middle left) should be kPa and not MPa. 12
Lecture Notes and Topics Lecture Session 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23
Topic Overview and Basics of Engine Operation Engine Geometry and Performance Parameters Ideal Cycle Analysis and Combustion Stoichiometry Gas Exchange (4-Stroke and 2-Stroke) Spark-ignition Engine Combustion and Flame Propagation and Structure No Lecture Available Knock SI Engine Emissions Emission Control Technology and Review No Lecture Available Mixture Preparation and Engine Friction Engine Heat Transfer and 2-Stroke SIE Performance 4-Stroke SIE Performance Variable Valve Control and Gasoline DI Engines Diesel Overview and Diesel Combustion No Lecture Available No Lecture Available No Lecture Available No Lecture Available No Lecture Available Future Engine Technology and Discussion
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2.61 Internal Combustion Engines Lab session: Engine disassembly Tuesday, February 10, 2004 2:00 – 4:00 P.M.
Please report to Ferran Ayala and Thane Dewitt in the Laboratory area of the Sloan Automotive Lab (first floor of Building 31). ________________________________________________________________________ The purpose of this session is to get some hands-on experience on the mechanical aspect of the engine. You should learn the mechanical construction of the engine, how the different components are arranged, and get a feel for the size and weight of the components. There are three engines. Two of them are Ford 4-cyliner, 16-valve engines for the compact size vehicles. The more modern one has a plastic intake manifold to reduce cost and NVH (Noise, Vibration and Harshness). The third engine is a Peugeot engine. The class will be divided into three groups; each group is responsible to disassemble one engine. However the whole class should participate in the initial ‘looking’ at each of the whole engine and the final study of all the parts for the three engines. Before disassembly, take a look at the engines as a whole, note the arrangement of the different components, the gas exchange circuit: intake, exhaust, EGR (Exhaust Gas Recirculation), and PCV (Positive Crankcase Ventilation) system, the coolant circuit and the fuel flow circuit. Understand the function of these components. When the engine is open, look at the valve train arrangement, the lubrication circuit, the EGR route, the coolant passage, and the piston/ crank/ balance weight arrangement. ________________________________________________________________________ Record the measurements indicated on the next page. The measurements may be shared by the group, but the comments and calculations should be done by each individual.
110
2.61 Internal Combustion Engines Lab session: Engine disassembly Tuesday, February 10, 2004 Engine Measurements Mean intake port length Mean exhaust port length Minimum intake port cross sectional area Minimum exhaust port cross sectional area Mean intake manifold runner length Mean exhaust manifold runner length Mean intake manifold cross sectional area Mean exhaust manifold cross sectional area Intake valve and minimum seat diameter Exhaust valve and minimum seat diameter Intake and exhaust valve stem diameter Maximum intake and exhaust valve lift Cylinder bore (at TDC and BDC) Stroke Connecting rod length (pin center to pin center) Piston top land diameter Piston top land height Piston middle land diameter Piston middle land height Overall piston height (max and min) Piston diameter (skirt center to skirt center) Piston top ring dimensions (thickness x height x outer diameter) Piston top ring groove diamensions (inner diameter x height) Piston middle ring dimensions (thickness x height x outer diameter) Piston middle ring groove diamensions (inner diameter x height) Calculations Cylinder displaced volume Bore/Stroke Connecting rod length/Crank radius Intake valve minimum seat diameter/Bore Exhaust valve minimum seat diameter/Bore Maximum intake valve lift/Intake valve diameter Maximum exhaust valve lift/Exhaust valve diameter Volume of one intake port plus one intake manifold runner/ displaced volume of one cylinder Volume of one exhaust port plus one exhaust manifold runner/ displaced volume of one cylinder Clearance Volume (Vc) Crevice Volume V1 Crevice Volume V2 Crevice Volume V3 Crevice Volume V4 Connecting rod lower end and upper end mass 111
2.61 Internal Combustion Engines Lab session: Engine disassembly Tuesday, February 10, 2004
Crevice Volumes
Top of Piston
V1 V2 V3 V4
Determine V1, V2, V3, and V4 in mm3, and as a percentage of the clearance volume
112
2.61 Internal Combustion Engines Lab session: Engine disassembly Tuesday, February 10, 2004
Balancing of Connecting Rod The connecting rod can be considered as equivalent to two masses concentrated at its ends (see diagram below), such that the sum of the masses is equal to the mass of the rod (Taylor).
cg
(A)
W1
W2 h
j
cg
(B) W1
W2
When doing a force balance of the piston motion, the mass of all the parts which are considered to reciprocate with the piston must be taken into account. These include the piston, the piston rings, the piston pin, and the equivalent mass of the upper end of the connecting rod. Roughly find the center of gravity of the connecting rod by balancing it, and then find the corresponding mass of its lower and upper ends by summing moments. What proportion of the total mass of the rod should be included in the piston mass calculation?
113
2.61 Internal Combustion Engines Laboratory Session Tuesday, April 27, 2004
Background This laboratory is designed to provide direct experience with the experimental methods used in internal combustion engine research and development. During the laboratory a set of engine performance and emissions data will be taken from an operating spark-ignition engine on a test stand. The data will be processed, analyzed and interpreted in the context of the course lecture material. Students will work in groups. Each group should process their experimental data as a group. Each individual will then write up a laboratory report as described below. Please read these notes carefully before you come to the laboratory. Objective The experiments are designed to examine the effect of changes in spark timing and relative AFR on spark-ignition engine performance and emissions characteristics. The results are going to be compared with those computed from the GMR simulation code. Experimental Method To eliminate effects of extraneous variables we will hold constant: x Design variables (compression ratio, valve timing, etc). x Inlet pressure (load) and speed. x Inlet temperature and exhaust pressure. x Oil and coolant temperature. The test procedure is the following: 1. Maintain engine speed at 1500 rpm and inlet pressure at 0.7 bar. Run the engine until it reaches steady state conditions (coolant temperature is approximately 600 C). 2. Set the relative AFR to stoichiometric. Retard and advance the spark timing about the value that gives the MBT according to the lab sheet and record all data listed on it. To determine GIMEP, COV and location of maximum pressure, record in-cylinder pressure data for approximately 150 cycles for each case.
Page 1 of 3 114
3. Repeat these measurements for different relative AFR’ s, both on the lean and rich side according to the lab sheet. Use the specified MBT spark timing for each value of the relative AFR. Analysis The following tasks are to be addressed in the lab report: 1. Effect of changes in spark timing: x Plot GIMEP, COV and location of maximum in-cylinder pressure versus spark advance and explain the shape of the curves. What is the location of maximum incylinder pressure when MBT spark timing is applied? x Plot NOx, HC emissions and exhaust gas temperature versus spark advance. Explain the effect of spark timing on each of them. 2. Effect of changes in relative AFR: x Plot the MBT spark timing variation with the relative AFR. What does the MBT spark-timing data tell you about the variation in burn rate with the relative AFR? Is there any significant variation in the location of maximum in-cylinder pressure when the relative AFR changes but the spark timing is always set to be the optimum? x Plot GIMEP and COV versus the relative AFR. Explain the effect of AFR on combustion variability. At what AFR is combustion most stable, and why? x Determine the indicated values for specific fuel consumption and fuel conversion efficiency and plot them versus the relative AFR. x Plot HC, CO, CO2 and NOx emissions versus the relative AFR. Calculate H2O and estimate H2 concentrations in the exhaust stream and include them also in graphs. Specify the background moisture for the measurement of each species. Plot also the exhaust gas temperature versus the relative AFR. Explain the trends of all curves you plot. x Determine the relative AFR from the exhaust gas composition measurements and plot it versus the measured value as recorded by the UEGO sensor. x Calculate the combustion efficiency for each point and explain the trend. 3. Comparison of experimental data with the fuel-air cycle results and the GMR simulation code (change geometry when possible according to given engine specs): x Plot on the same graph GIMEP based on the experimental data, the fuel-air cycle results and the GMR simulation code versus the relative AFR. x Draw the corresponding graph for the indicated fuel conversion efficiency.
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x Compare the experimental data with the corresponding calculated values based on the GMR simulation code of CO, CO2, NOx and H2O. Be careful to have the same background moisture for each species for compatible comparisons. Laboratory Report The laboratory report is due on Thursday, May 6, 2004 by 5 pm in Ferran Ayala’s office. We recommend that you process your data with other members of your laboratory group. Reports should be written separately by each individual, however. The report should include the following: x Brief introduction, which reviews the goals of the experiment and procedure. x Analysis. x Results. x Discussion and interpretation. x Experimental data (attached as one or more appendices). The questions the report should address are listed in the previous section. Word processor or neatly handwritten reports are requested. Excessive length and repetition of material in the laboratory handout are undesirable. Best marks are obtained with a logical report organization, a concise summary of what was done, and good physical explanations of the trends observed in the data. There may well be inconsistencies in the data; these should be identified and explained where possible.
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2.61 Internal Combustion Engines Problem Set 1 Tuesday, February 3, 2004 Due: Tuesday, February 10, 2004 1.
Explain briefly the following differences between a standard automobile spark-ignition (SI) engine and a truck diesel engine: (a) where the fuel is injected and why (b) how the load is varied at fixed speed (c) how the combustion process starts, develops and ends (d) how the in-cylinder pressure varies as a function of crank angle (draw qualitatively the pressure traces for both engines in the same graph showing their relative magnitudes and when approximately the combustion starts and ends in each case). (e) how the fuels are different and why
2.
At 450 after top dead center (ATDC) on the expansion stroke, the pressure in a SI engine cylinder is 1000 kPa. The bore and the stroke are 80 mm, the ratio of the connecting rod length to the crank radius is 3.5, the piston mass (including the pin and half the connecting rod) is 0.57 kg, the crankcase pressure is 100 kPa, the axial friction force on the piston (due to rings sliding against the cylinder liner) is 65 N and the engine speed is 2500 rpm. (a) Draw the piston and connecting rod free body diagrams, calculate all forces acting on them and compute the torque delivered to the crankshaft. (b) Answer part (a) for the same crank position during the intake stroke if the engine is working at half load conditions assuming that the friction force is the same. Hint: a good approximation to piston acceleration is the following: a = 2p 2 N 2 L(cos ? +
1 cos 2? ) R
N … engine speed L … stroke R … connecting rod length to crank radius ratio theta … crank angle measured from TDC (Bosch Automotive Handbook, 4th Edition, pp. 404 – 405) A more accurate expression can be found be differentiating twice the distance between the crank axis and the piston pin axis. 3.
An automobile engine should provide enough power to a car in order to overcome its resistances. These resistances consist of the rolling resistance arising from the friction of the tires with the road, the aerodynamic drag, the inertia as the car accelerates and the gravity if it is travelling up a hill. Ford Taurus (4-door) weights 1515 kg, and has a relatively flat torque – engine speed curve of 247 Nm for medium speeds. Estimate: (a) the road load power (power required to drive a vehicle on a level road at steady speed) when the vehicle speed is 50 and 100 km/h. Comment on the relative importance of the resistances as vehicle speed increases (b) the power required to drive the car up to 20% gradient hill at steady speed of 50 km/h
Page117 1 of 2
(c)
the minimum time required to accelerate the car from 40 to 80 km/h on a level road with the 3rd gear. You can assume that the ratio of the engine to the wheel speed is approximately 1x3.27 (which is the transmission torque ratio with the 3rd gear times the axle ratio). Make appropriate estimations for road roughness and car geometry (drag coefficient, frontal area and wheel radius). 4.
A 4-cylinder, 4 stroke diesel engine is being designed. A bore of 100 mm and a stroke of 120 mm have been selected and the operating speed is to be 1500 rev/min. A turbocharging system is envisaged which will supply inlet manifold air at 2.0 bar, 380 K. The volumetric efficiency is expected to be 90%. The indicated fuel conversion efficiency has been estimated at 55%, and an air-fuel ratio of 28:1 is to be used. The friction mean effective pressure is expected to be 2.2 bar. The calorific or heating value of the fuel is 42.5 MJ/kg. Estimate: (a) Mass flow rate of air into the engine (kg/s) (b) Mass of fuel burned per cylinder per cycle (mg) (c) The indicated work done per cylinder per cycle (kJ) (d) The brake mean effective pressure (kPa) (e) The engine shaft power output (kW) (f) The brake specific fuel consumption (g/kWhr) (g) The brake fuel conversion efficiency
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2.61 Internal Combustion Engines Problem Set 2 Tuesday, February 10, 2004 Due: Thursday, February 19, 2004 1.
Several velocities, time, and length scales are useful in understanding what goes on inside engines. Make estimates o the following quantities for a 1.6-liter displacement four-cylinder spark-ignition engine, operating at wide-open throttle at 2500 rev/min. (a) The mean piston speed and the maximum piston speed. (b) The maximum charge velocity in the intake port (the port area is about 20 percent of the piston area). (c) The time occupied by one engine operating cycle, the intake process, the compression process, the expansion process, and the exhaust process. (Note: The word process is used here not the word stroke.) (d) The average velocity with which the flame travels across the combustion chamber. (e) The length of the intake system (the intake port, the manifold runner, etc.) which is filled by one cylinder charge just before the intake valve opens and this charge enters the cylinder (i.e., how far back from the intake valve, in centimeters, one cylinder volume extends in the intake system). (f) The length of exhaust filled by one cylinder charge after it exits the cylinder (assume an average exhaust gas temperature of 425 C) You will have to make several appropriate geometric assumptions. The calculations are straightforward, and only approximate answers are required.
2.
Evaluate the maximum bmep, the bmep at maximum power and the maximum value of the mean piston speed (where applicable) for the engines described in the attached sheets: Gasoline (natural aspirated and turbocharged) (a) General Motors Vortec 4.2L DOHC I-6 (b) Audi AG 4.2L DOHC V-8 (c) BMW AG 3.2L DOHC I-6 (d) K13C with common rail fuel injection (e) Subaru 2.5L H-4 turbocharged Diesel (natural aspirated and turbochared) (f) 5.9L Cummins 600 OHV l-6 turbodiesel (g) Isuzu new V12 PE1-S (h) Mitsubishi 2.8L Specialty (hybrid, rotary, formula 1) (i) Mazda 1.3L Renesis rotary (j) Toyota 1.5L DOHC I-4 Hybrid Formula 1 engine: naturally aspirated, 3.0 liter, 10-cylinder, stroke to bore ratio is 0.48, (k) maximum brake power at 17000 rpm is 800 hp and the maximum brake torque at 13000 rpm is 350 Nm
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Explain briefly the differences in performance parameters between the various engine configurations: - SI: 2 valves per cylinder, 4 valves per cylinder, F1 racing engines - CI: naturally aspirated, turbocharged engines Also using the bsfc maps for the Isuzu H-Series diesel, the Audi 5-cylinder turbo diesel and the Toyota SI 3L V-6 engine, calculate the brake fuel conversion efficiencies at the following conditions: - Best efficiency - Typical automotive light load operation (half speed and 25% load) 3.
The Isuzu V12 DI diesel 12PEI-S engine described in the attached sheets is operating at 2000 rpm, full load. At this operating condition the total friction power (consisting of rubbing friction, pumping and accessory power) is 50 kW. Making an appropriate assumption for the relative airfuel ratio (remember that diesels operate lean overall) and using the manufacturer’s specifications, calculate the following: (a) mechanical efficiency (b) brake fuel conversion efficiency (c) indicated fuel conversion efficiency (d) volumetric efficiency The engine is working with light diesel fuel, which has lower heating value 43.2 MJ/kg and stoichiometric air-fuel ratio 14.5.
4.
Using the ideal constant volume combustion cycle model, draw an accurately proportioned cylinder pressure versus cylinder volume diagram for a lower compression ratio engine (say 8) and a higher compression ratio engine (say 12) that shows why the higher compression ratio engine has a higher indicated fuel conversion efficiency. In making this comparison on a p-V diagram, a critical question is “what should be held constant?” Possible choices are one or more of these: maximum cylinder volume, displaced cylinder volume, mass of air in cylinder, mass of fuel in cylinder, relative air/fuel ratio of mixture in cylinder, pressure in cylinder at start of compression, temperature of gas in cylinder at start of compression. (a) Explain your choice of which of the above you will hold constant in this comparison. (b) Carefully draw the two p-V diagrams on the same graph, indicating the relative magnitude of pressures and temperatures at start and end of compression, at start and end of expansion. (c) Using these p-V diagrams explain why the higher compression ratio engine is more efficient
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2.61 Internal Combustion Engines Problem Set 3 Thursday, February 19, 2004 Due: Tuesday, March 2, 2004 1.
A modern naturally-aspirated spark-ignition engine has an indicated fuel conversion efficiency K f ,i of 38 percent over its operating range. At 2000 rev/min, the mechanical friction mean effective pressure (mfmep) is 90 kPa and the accessory (water, oil and fuel pump) friction (afmep) is 20 kPa; both are independent of load. The engine operates at stoichiometric and MBT spark timing at all intake manifold pressures. The exhaust pressure pe is 1 atmosphere. (a)
Develop an expression for the brake fuel conversion efficiency as a function of K f ,i , mfmep, afmep, intake manifold pressure pi, exhaust manifold pressure pe, fuel heating value QHV, fuel/air ratio, intake system air density U a ,i , just downstream of the throttle and volumetric efficiency K v based on U a ,i .
(b) (c)
2.
Evaluate the engine’s mechanical efficiency and brake fuel conversion efficiency, at an intake manifold pressure of 0.4, 0.7, and 1.0 bar Assuming that the mechanical and accessory friction values remain the same at idle, determine the intake manifold pressure at idle
An internal combustion engine is being operated with a pure alcohol as fuel. The fuel molecule has 7 carbon atoms. A detailed analysis of the exhaust gas constituents on a wet molar basis (i.e., with the H2O present) in the exhaust manifold yielded the following: CO2: 13%, CO: 5.2%, H2O: 9.1%, H2:1.3%, and N2: 71.3% (a) Determine the elemental composition of the fuel molecule (b) Determine the relative air-fuel ratio. Is this lean stoichiometric or rich? (c) The engine is a 4-valve per cylinder naturally aspirated SI engine. It has a displaced volume of 3 liters, the airflow is 0.06 kg/sec and the engine speed is 3000 rpm. Estimate the engine load (i.e., the approximate ratio of the actual brake torque to the maximum brake torque at these conditions. The friction mean effective pressure is 200 kPa at this speed and is essentially independent of load at the loads considered here. The gross indicated fuel conversion efficiency is also independent of load. Hints: you may need to make appropriate assumptions concerning engine’s breathing characteristics and normalized torque, according to the given engine technology.
3.
Calculate the following parameters for a constant-volume combustion fuel-air cycle (Fig. 5-2a from text): (a) the pressures and temperatures at states 1, 2, 3, 4, 5 and 6 (b) the indicated fuel conversion efficiency (c) the imep (d) the residual gas fraction (e) the volumetric efficiency The engine is working at full load and the compression ratio is 8. Page131 1 of 2
Make appropriate assumptions for the inlet pressure and temperature, exhaust pressure and air-fuel ratio. Hint: start the calculations at point 1 using a residual gas mass fraction of 3% and a mixture temperature at start of compression of 343 K. One cycle calculation should suffice. 4.
Estimate from fuel-air cycle results the indicated fuel conversion efficiency, the imep, and the maximum indicated power at wide open throttle of two 4-stroke SI engines with the following specs: (a) six-cylinder, bore 9.2 cm, stroke 9 cm, compression ratio 7 and equivalence ratio 0.8. (b) six-cylinder, bore 8.3 cm, stroke 8 cm, compression ratio 10 and equivalence ratio 1.1. Assume that the actual indicated efficiency is 0.8 times the appropriate fuel-air cycle efficiency. The maximum permitted value of the mean piston speed is 15m/s. Explain also briefly why: - the efficiency of these two engines is approximately the same despite their different compression ratios. the maximum power of the smaller displacement engine is approximately the same as that of the larger displacement engine.
5.
Carbon dioxide (CO2) emissions from engines are a matter of increasing concern due to their contribution to the so-called “greenhouse effect”. Both the characteristics of the hydrocarbon fuel and engine factors determine an engine’s CO2 emissions levels. (a) Derive an expression for the mass of CO2 produced per unit brake work output from the engine in terms of the fuel H:C ratio (assume a fuel of composition (CHy)n), the fuel’s lower heating value QLHV, and the appropriate engine efficiencies. (b) Compare the CO2 producing potential of the following two fuel/engine combinations: - A methane (CH4) fueled spark-ignition engine with a compression ratio of 10 operating with a stoichiometric mixture. - A diesel (CH2)n fueled compression-ignition engine with a compression ratio of 16 operating with a fuel-air equivalence of 0.4. At the relevant load, the mechanical efficiency of the first engine is 0.6 and of the second 0.7. Assume the results of available fuel-air cycle calculations adequately describe the effects of compression ratio and equivalence ratio changes for both these fuel/engine combinations.
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2.61 Internal Combustion Engines Problem Set 4 Tuesday, March 2, 2004 Due: Thursday, March 11, 2004 1.
Hydrogen is a possible future fuel for SI engines, considering both emissions (only water vapor is present in the exhaust) and consumption (high heating value). A disadvantage of hydrogen fuel in SI engines is that the partial pressure of hydrogen in the mixture with air reduces engine’s volumetric efficiency, which is proportional to the partial pressure of air. (a) Find the partial pressure of air in the intake manifold downstream of the hydrogen fuel injection location at WOT (the engine is working with stoichiometric mixture). (b) Estimate the ratio of the fuel energy per unit time entering a hydrogen-fueled engine operating with a stoichiometric mixture to the fuel energy per unit time entering an identical gasoline-fueled operating at the same speed with a stoichiometric mixture. (c) How does the power output compare for both engines? What can you do to ensure that the output of each engine is the same? What are your limitations? (for reference see BMW paper: SAE 2003-01-3210, on 2.61 webpage)
2.
A conventional spark-ignition engine operating with gasoline will not run smoothly (due to incomplete combustion) with an equivalence ratio leaner than about I=0.8. It is desirable to extend the smooth operating limit of the engine to leaner equivalence ratios so that at part throttle operation (with intake pressure less than 1 atmosphere) the pumping work is reduced. Leaner than normal operation can be achieved by adding hydrogen gas (H2) to the mixture in the intake system. The addition of H2 makes the fuel-air mixture easier to burn. (a) The fuel composition with “mixed” fuel operation is H2 + C8H18 --- one mole of hydrogen to every mole of gasoline, which is assumed the same as isooctane. What is the stoichiometric air/fuel ratio for the “mixed” fuel? (b) The lower heating value of H2 is 120 MJ/kg and for isooctane is 44.4 MJ/kg. What is the heating value per kilogram of fuel mixture? (c) Engine operation with isooctane and the mixed (H2 + C8H18) fuel is compared in a particular engine at a part-load condition (brake mean effective pressure of 275 kPa and 1400 rev/min). You are given the following information about the engine operation: Fuel Equivalence ratio Gross indicated fuel conversion efficiency Mechanical rubbing friction mep Inlet manifold pressure Pumping mep
C8H18 0.8 0.35 138 kPa 46 kPa 55 kPa
H2 + C8H18 0.5 0.4 138 kPa ? ?
Estimate approximately the inlet manifold pressure and the pumping mean effective pressure with (H2 + C8H18) fuel. Explain your METHOD and ASSUMPTIONS clearly. Note that mechanical efficiency Km is defined as Km = bmep/imepg = bmep/(bmep + rfmep + pmep) Page133 1 of 3
3.
For four stroke engines, the inlet and exhaust valve opening and closing crank angles are typically: IVO: 15 CAD BTDC IVC: 50 CAD ABDC EVO: 55 CAD BBDC EVC: 10 CAD ATDC (a) Explain why these valve timings improve engine breathing relative to valve opening and closing at the beginnings and ends of the intake and exhaust strokes. (b) Mention at least one modification you would do to the above valve timings if the engine you are designing was - intended for a race car - turbo-charged - working at low speeds What other additional design issues are important?
4.
In a spark-ignition engine, a turbulent flame propagates through the uniform premixed fuel-air mixture within the cylinder and extinguishes at the combustion chamber walls. (a) Draw a carefully proportioned qualitative graph of cylinder pressure p and mass fraction burned xb as a function of crank angle T for –90q < T < 90q for a typical SI engine at wideopen throttle with the spark timing adjusted for maximum brake torque. Mark in the crank angles of spark discharge, and of the flame development period (0 to 10 percent burned) and end of combustion, on both p and xb versus T curves relative to the top-center crank position. (b) Estimate approximately the fraction of they cylinder volume occupied by burned gases when the mass fraction burned is 0.5 (i.e., halfway through the burning process). (c) A simple model for this turbulent flame is shown below in Fig. 1. The rate of burning of the charge dmb/dt is given by
dmb dt
(d)
Af Uu ST
Where Af is the area of the flame front (which can be approximated by a portion of a cylinder whose axis is at the spark plug position), Uu is the unburned mixture density, and ST is the turbulent flame speed (the speed at which the front moves relative to the unburned mixture ahead of it). The rate of mass burning is influenced therefore by combustion chamber geometry (through Af) as well as those factors that influence ST (turbulent intensity, fuel/air ratio, residual gas fraction, and EGR). Compare combustion chambers A and B shown in Fig. 1. Sketch the approximate location of the flame front when 50 percent of the mass has been burned. (A careful qualitative sketch is sufficient; however, provide a quantitative justification for your sketch.) Sketch the mass fraction burned versus crank angle curves for these two combustion chambers on the same graph, each with its spark timing set for maximum brake torque. You may assume the value of ST is the same for A and B. Compare combustion chambers A and C in the figure below, which have the same flame travel distance but have different chamber shapes. Which chamber has - the faster rate of mass burning during the first half of the combustion process; - the faster rate of mass burning during the second half of the combustion process; - the more advanced timing for maximum brake torque? Explain your answers Page134 2 of 3
Figure 1
5.
An approximate way to calculate the pressure in the end gas (the unburned gas ahead of the flame) just after knock occurs is to assume that all the end gas burns instantaneously at constant volume. We assume that the inertia of the burned gases prevents significant gas motion while the end gas autoignites. Use the data in Figure 9.5 from the text and assume that autoignition occurs at 10 CAD ATDC. You may also need to use the equilibrium charts of Chapter 4 from the text. (a) Determine the maximum pressure reached in the end gas after knock occurs (b) Estimate the volume occupied by the end gas as a fraction of the cylinder volume just before autoignition occurs
6.
Use the GM Vortec 4.2L DOHC engine, which appeared on problem set 2 to calculate the ratio of actual gross indicated engine performance to the equivalent fuel-air cycle predictions in Figures 59 and 5-10. (a) Maximum power, assume total friction mep = 200 kpa (b) Maximum torque, assume total friction mep = 150 kpa (c) Best brake specific fuel consumption: the best measured value is 250 g/kWh at part throttle and total friction mep = 140 kpa. At WOT the engine operates rich with fuel-air equivalence ratio of 1.2. The engine operates stoichiometric at other load points for emission control reasons. Explain briefly the reasons why the fuel-air cycle predicts better performance than measured
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2.615 Internal Combustion Engines Problem Set 5 Tuesday, March 16, 2004 Due: Tuesday, March 30, 2004 1. This problem requires you to use an Engine Simulation Program developed by General Motors Research Laboratories, which they have made available for free to academic institutions.** Unless otherwise stated, use the preset values of the variables, which define engine geometry and operating conditions. Assume burn duration of 50 CAD, unless otherwise stated. (a) At WOT, 2000 rpm and stoichiometric mixture conditions vary the spark timing in 10 CAD intervals from 50 to 10 CAD BTC x Plot the gross indicated mean effective pressure versus spark advance and determine the MBT spark timing. Explain briefly the shape of the curve. x Plot in the same graph T P max and T 50% burned versus spark advance. At MBT spark timing, what are the values of T P max and T 50% burned ? x Plot the indicated specific fuel consumption (isfc) versus spark advance. What spark advance gives a minimum isfc? How does it compare to MBT spark timing? Are there any other factors that you would take into account when setting the spark timing in an engine apart from efficiency and torque output? x Plot Pmax versus spark advance. Why is Pmax important? (b) At 2000 rpm and stoichiometric mixture conditions vary the inlet pressure in 0.25 bar intervals from 0.25 to 1 bar (recall that the inlet pressure approximately scales with the load). The burn durations are: Pi=0.25 bar, 70˚; Pi=0.5 bar, 50˚; Pi=0.75 bar, 44˚; Pi=1.0 bar, 38˚. Be sure to have the MBT spark timing for every point. x Plot the MBT spark timing versus the inlet pressure. Explain briefly the shape of the curve. x Plot in the same graph T P max and T 50% burned versus the inlet pressure. Do you see any variation of these values with the inlet pressure when the spark timing is always set to MBT? How do they compare to the values you got in part (a)? x Plot in the same graph the gross indicated, the net indicated and the brake mean effective pressures versus inlet pressure. Show on the graph the pumping and the mechanical friction losses. x Plot in the same graph the gross indicated, the net indicated and the brake fuel conversion efficiencies versus the inlet pressure. In the same graph include also the mechanical and volumetric efficiencies. Explain briefly the dependence of the mechanical efficiency on the inlet pressure. Does the gross indicated fuel conversion efficiency vary significantly with the inlet pressure? (c) At WOT and stoichiometric mixture conditions vary the engine speed in 1000 rpm intervals from 2000 to 5000 rpm. The MBT spark timing for each speed is 2000 rpm, 28˚ BTC; 3000 rpm, 32˚ BTC; 4000 rpm, 36˚ BTC; 5000 rpm, 40˚ BTC. Determine the 0100% burn duration at each of these speed conditions (that corresponds to MBT). Answer again part (b), having now the engine speed as a parameter.
**See GM Research Publication GMR-5758, "A User's Guide for the GM Engine-Simulation Program," March 4, 1987 Page136 1 of 2
(d)
At 2000 rpm and 0.5 bar inlet pressure, vary the fuel equivalence ratio in 0.1 intervals from 0.8 to 1.2. Be sure to have the MBT spark timing for every point. The burn durations are ĭ=0.8, 65˚; ĭ=0.9, 54˚; ĭ=1.0, 50˚; ĭ=1.1, 48˚; ĭ=1.2, 47˚; Answer again part (b), having now the fuel equivalence ratio as a parameter. (e) Compare the simulation and the fuel-air cycle results at the operating conditions defined in part (d). x Plot on the same graph the gross indicated mean effective pressure versus the fuel equivalence ratio for both cases. Plot on the same graph the gross indicated fuel conversion efficiency versus the fuel equivalence ratio for both cases. 2. In this problem you are asked again to use the GMR Engine Simulation Program. We are now interested in heat losses and emissions. Keep the combustion duration at 60 CAD for all cases. (a) Using MBT spark timings, sketch the following graphs: x absolute levels of total heat losses with respect to speed and load (intake pressure) for stoichiometric mixture. Use the same speed and load increments as in the last problem set and draw the total heat losses as a function of speed with contours of constant intake pressures. x relative levels of total heat losses (as a fraction of the fuel energy delivered into the engine) with respect to speed and load for stoichiometric mixture. (b) Answer part (a) having only the fuel equivalence ratio as a parameter. The operating conditions are: WOT, 2000 rpm and MBT spark timing for each point. (c) Explain briefly the dependence of total heat losses (both absolute and relative) on speed, load and mixture composition. (d) At WOT, 2000 rpm and MBT spark timing for each point, plot the molar concentration of CO2, CO, H2O, O2, H2 and NO in the exhaust stream as a function of fuel equivalence ratio. Plot also the exhaust gas temperature as a function of the equivalence ratio. Explain briefly the shape of the curves you obtain. (e) As already discussed in the lecture the Exhaust Gas Recirculation (EGR), is a widely used emission control technology for decreasing exhaust NO emissions. x Explain briefly why recycling exhaust gases reduces NO emissions. Is there any impact on the combustion process? At what loads is EGR used and why? x Plot NO emissions as a function of EGR mass fraction using the following operating conditions: 2000 rpm, stoichiometric mixture, spark timing set at 28 CAD BTDC and intake pressure such that each point has the same gross indicated mean effective pressure. Use as a baseline case intake pressure 50 kPa and EGR 0% and increase EGR until 30% with an increment of 10%. Plot also the intake pressure and the exhaust gas temperature as a function of EGR. x Do you think that the fraction of exhaust gases recycled must be higher or lower for diesel engines in order to have approximately the same relative impact on NO and why? x Suggest another way to affect the combustion process and NO emissions in the same way as EGR? Is there any way to control it? (f) It is also interesting to see during what part of combustion NO is formed. Therefore, plot NO emissions on a CAD basis from –120 to 120 CAD ATDC. Also plot the in-cylinder pressure and the mass fraction of fuel burned. Use the baseline operating conditions: WOT, 2000 rpm, stoichiometric mixture and MBT spark timing.
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3. Explain, using your understanding of the fundamentals of spark-ignition engine combustion, the causes of the observed variation in cylinder pressure cycle by cycle. What impacts do these cylinder pressure variations have on engine operation? Do you expect diesel engines to have higher or lower cycle by cycle variability?
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2.61 Internal Combustion Engines Problem Set 6 Tuesday, March 30, 2004 Due: Thursday, April 8, 2004 1.
Compare the three-valve two-plug combustion chamber shown, with a four-valve chamber with one spark plug on the cylinder axis. Make appropriate simple assumptions about the flame shape, as it grows outward from the spark(s). (a) For the four-valve, center plug, plot flame envelope area (flame outer boundary) A f 2Sr0 h against rf r0 through the combustion process. On the same graph, sketch carefully the equivalent curve for the three-valve two-plug design assuming both plugs fire at the same time: simple geometric calculations are required to short out the physics. (b) Plot the mass fraction burned curves versus CAD for these two combustion chambers on a qualitative but carefully proportioned graph from 45 CAD BTDC to 45 CAD ATDC. Show on the graph the spark discharge location, the TDC, the location of maximum incylinder pressure and the approximate end of combustion, with the spark timing for both chambers set at MBT timing. (c) With the two-plug chamber, the burn rate can be slowed down by firing one plug later than the other. Estimate approximately the mass fraction burned after which further delaying the second plug firing will have no impact on the burn rate.
r0
Intake
Intake
r0/2 1200
Spark plug
Exhaust
h
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2.
You are designing a high power density (high maximum power per unit displaced cylinder volume) port fuel injected naturally aspirated spark-ignition engine to have a maximum power of 150 kW for an automobile. Address the following questions: (a) Explain how you will achieve higher than “average” power density. (b) Explain how you will deal with knock. You may choose regular gasoline (Research octane no. 92, Motor octane no. 82) or premium gasoline (Research octane no. 98, Motor octane no. 88) which is $0.20 a gallon more expensive (c) Explain your emission control strategy and emission control system; you must meet strict standards for hydrocarbons, carbon monoxide, and oxides of nitrogen (d) Based on your choices in (a), (b) and (c) estimate the required displaced volume, bore, stroke, number of cylinders, and compression ratio. (e) Provide a description of your combustion system, explaining the objectives and logic behind your choices.
3.
In the disked-shaped combustion chamber of a spark-ignition engine, the flame propagates radially outward from the spark plug at the center (on the cylinder axis). The flame can be modeled as a thin smooth sheet at the location within the wrinkled turbulent flame, which on average separates the unburned and burned gases. (a) Derive an equation for the mass fraction burned x b m b m as a function of rf ro (rf : flame radius and ro : cylinder radius) and the density ratio rd U u U b . Plot this curve for an appropriate value of rd. (b) Draw a carefully proportioned graph of xb versus CAD for the engine operating at part load with MBT spark timing (30 CAD BTDC), indicating the approximate mass fraction burned at TDC and 10 CAD ATDC. Note that xb at T P max is about 0.75. Make an appropriate estimate of Pmax. The pressure in the cylinder at IVC (50 CAD ABDC) is the inlet pressure, i.e. 0.5 bar. The cylinder volume ratio, volume at IVC to volume at Pmax is 7.3. The mixture is stoichiometric and the compression ratio is 10. Make reasonable assumptions of any other quantities you may need.
4.
A spark-ignition engine driving a car uses, on average, 120 grams of gasoline per mile traveled. The average emissions from the engine (upstream of the catalyst) are 1.5, 2, and 20 grams per mile of NOx (as NO2), HC (interpret HC measurement as hydrocarbons with H/C ratio of 1.85), and CO, respectively. The engine operates with a stoichiometric gasoline-air mixture. (a) Find the average concentrations in parts per million of NOx, HC (as ppm C1), and CO in the engine exhaust. (b) Calculate the average combustion inefficiency associated with the given emissions levels. Include any hydrogen you estimate would be present in the exhaust stream. (c) Assuming the engine has a three-way catalytic converter with efficiency of 95%, find the rise in temperature in the catalyst, after the pollutants are removed. Neglect heat losses.
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141
142
143
144
145
146
147
148
149
150
151
Typical performance diagram SULZER RLB 90
2- STROKE ENGINE
(TURBOCHARGED)
Outputs, weights and dimensions
Number of cyls
RLB 90 Stroke 1900 mm Scre 900 mm
Rating
rev min
ERP 1
96
mep bar
cm m/s
13.42 5.21
A A B C D E F
18 550 21 200 23 850 25 500 25 200 23 800 32 400 36 000
31 800 43 200
bhp
11 760 14 700 17 640 20 580 21 520 25 460 29 400 16 000 20 000 24 000 28 000 32 000 36 000 40 000
35 280 48 000
bw bhp
10 000 12 500 15 000 13 600 17 000 20 400
17 500 23 800
kw
11 200 14 000 16 800 15 200 19 000 22 800
19 600 22 400 26 600 30 400 900 1 005
kw bhp
MCR 1
102
14.31 5.16
ERP 2
90
13.79 5.70
MCR 2
94
14.79 5.95
kw
bhp Net weight, without water and oil
10 780 12 500 14 220 15 780 17 480 19 200 20 320 24 360 11 270 12 550 14 710 4 000 4 000 4 000 4 000 4 000 4 000 4 000 4 000 1 800 1 800 1 800 1 800 1 800 1 800 1 800 1 800 9 250 9 250 9 250 9 250 9 250 9 250 9 250 9 250 1 185 1 185 1 520 1 520 1 520 1 520 1 520 1 520 13 150 13 150 13 150 13 150 13 150 13 150 13 150 13 150
tonnes
10 500 13 250 15 900 14 400 18 000 21 500
555
152
655
770
20 000 22 500 27 200 30 600
25 000 30 000 34 000 40 000
25 200 28 000 33 600 34 200 38 000 45 600 1 470 1 120 1 255
2.61 Internal Combustion Engines Design Project – Corrected Version Number 2 Wednesday, April 14, 2004 Due: Thursday, April 22, 2004 Heavy duty diesel engine with EGR and particulate trap For this project you need to design a heavy-duty truck diesel engine. The engine is 11 liter, 6 cylinder, with target maximum power of 360 kW and target bsfc of 185 g/kW-hr. You will also incorporate an emissions strategy to meet current emissions standards. 1. Base Engine Begin by sizing the base engine (no turbocharger), which will be 11 liter, 6 cylinder. Assume a maximum mean piston speed (Sp) of 10 m/s. Use data on Figure 13.7 on page 722 in the text for estimates of mechanical/rubbing plus auxiliary mep. Making any other reasonable assumptions necessary, determine the following parameters: (a) Bore and stroke (b) Compression ratio (c) Connecting rod length (d) Brake mean effective pressure, at maximum torque and maximum power (e) Maximum torque and maximum power (f) Maximum engine speed, at maximum power 2. Boost, turbomachinery, and intercooler Design the required turbomachinery and intercooler for the engine by addressing the following points: (a) Based on your calculations in part 1, calculate the amount of boost pressure required at maximum speed to produce the target power. (b) Use a turbocharger to produce this boost, and define the main operating parameters of the required turbomachinery. For both turbine and compressor, provide values for mass flow rate, pressure ratio, inlet and exhaust temperatures and pressures. Use typical values for isentropic efficiencies (Kt=0.85, Kc =0.80) and assume the exhaust temperature is 900K. (c) Include an intercooler to lower the temperature of the air coming out of the compressor. Provide the inlet and outlet temperatures of the air, as well as the coolant’s inlet and outlet temperatures and mass flowrate. Assume a counter-flow heat exchanger with effectiveness of 0.8. (Heat exchanger effectiveness is the ratio of the actual heat transfer to the heat transfer that would occur if the stream with the minimum capacity rate were heated (or cooled) from its inlet temperature to the inlet temperature of the other stream). (d) Draw a schematic of your system (Hint: You will need to include the effect of turbo-charging on pumping work)
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3. Brake efficiency Evaluate the brake fuel conversion efficiency of the design at half maximum speed and full load at that speed. Does it meet the target bsfc? If not, what changes in engine design would bring it closer? 4. Emissions – NOX The engine you are designing must comply with 2004 EPA NOx requirements which are set at 2.5 g/bhp-hr (assume for simplicity that the NOx requirement must be met at all operating points); EGR has been chosen as the technology to reduce NOx levels. Address the questions below using available data, and an appropriate safety factor: (a) Draw a schematic of your system, showing clearly how you will drive EGR from exhaust to intake. There are a few possibilities for doing this (the reference paper on EGR systems might be helpful). (b) Find the required amount of EGR to run at low load (25% max torque and 1600 rpm). (c) Find the required amount of EGR to run at maximum power. (d) Similar to part 2, recalculate the boost pressure at maximum power. Resize the turbomachinery and intercooler to reach the stated target or best case power output; use available data to determine the exhaust temperature. Also, include an EGR cooler to lower the re-circulated gas temperature before it enters the engine; using the same assumptions as part 2c, provide operating temperatures and flowrates. As a safety measure, it is common standard to reduce the amount of NOx, by an additional 20% to 40% of the required EPA standard. For this design please use a safety factor of 30% (i.e., reduce NOx to 1.75 g/bhp-hr, 30% below required standard). Assume that the equivalence ratio based on the mass of fresh fuel and fresh air must stay below the smoke limit of 0.7; for simplicity once you have selected the level of EGR, assume NOx levels remain constant, in spite of additional boosting (this is not the actual case). Also assume that beyond 8 CAD BTDC, for every additional CAD delay in injection you lose 0.25 percentage points in indicated fuel conversion efficiency. (Note: Watch the units in the emissions data)
5. Emissions – particulate The engine you are designing must also comply with 2004 EPA particulate requirements. Diesel Particulate Filter (DPF) has been selected as the technology to achieve the target PM levels of 0.05 g/bhp-hr. Assume that current DPF technology can reach 99% efficiency (i.e., 99% removal of particulates). Using a PM emissions safety factor of 50%, answer the following questions, all for maximum power conditions (a) What is the approximate level of PM coming out of the engine?
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(b) Size the trap and calculate the average pressure drop. Assume a space velocity in the range of 10,000 – 28,000 hr-1 that minimizes the physical size of the particulate trap (see SAE 2003-01-0047 for reference). (c) What impact does the particulate trap have on the performance of the turbocharged engine (be quantitative). What changes in boost pressure and turbomachinery operating conditions are needed to keep best case output? 6. 2007 Emissions requirements: Below is a table showing EPA Diesel engine emissions requirements for 2007. NOx (g/bhp-hr) 0.20
PM (g/bhp-hr) 0.01
(a) Based on engine out NOx levels of part 4, how efficient a NOx catalyst is needed to meet 2007 emissions levels? Assume that the catalyst is used in conjunction with EGR. (b) Is it possible to achieve these levels of PM with the trap described in the SAE 2003-01-0047 paper? (Note: Keep the same safety factors as in parts 4 and 5)
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Temperature (K)
950
1000
1050
800
850
900
156
5%
10%
0.5 CAD BTDC
15%
7.5 CAD BTDC
3.5 CAD BTDC
% EGR
20%
2.5 CAD ATDC
25%
6.5 CAD BTDC
Exhaust Temperature vs. EGR (For Different Start of Injection @ Maximum Power)
30%
35%
bsNOx [g/kW-hr]
2.500
3.000
3.500
4.000
4.500
5.000
0.000
0.500
1.000
1.500
2.000
157
-8
-6
-4
25% Max torque, 1600 RPM
32.7%
23.8%
8.6% Increasing EGR
-2
32.2%
24.9%
9.1%
2
Start of Main Injection [CA From TDC]
0
32.3%
24.2%
10.8%
bsNOx vs. Start Of Main Injection Low-Load Condition
4
32.1%
24.4%
11.0%
6
Reduced EGR
Stock EGR
Increased EGR
8
31.8%
23.6%
10.9%
10
bsNOx [g/kW-hr]
5.000
6.000
7.000
8.000
9.000
0.000
1.000
2.000
3.000
4.000
158
-10
Increasing
-8
29.5
20.0
7.6%
-6
0
30.9
20.4%
7.5%
Stock EGR
2
29.5
20.6%
9.4%
Reduced EGR
Start of Main Injection [CA From TDC]
-2
Increased EGR
-4
29.6%
20.2%
7.4
4
bsNOx vs. Start Of Main Injection (Maximum Power)
6
19.9%
8.5%
8
bsPM [g/kW-hr]
0.1000
0.1200
0.1400
0.1600
0.1800
0.0000
0.0200
0.0400
0.0600
0.0800
159
-8
Increasing G
-6
8.6%
23.8%
32.7%
-4
-2
0
10.8
22.9
32.3
2
11.0
22.2
4
6
8
Increased EGR Stock EGR Reduced EGR
25% Max torque, 1600 RPM
32.1
Start of Main Injection [CA From TDC]
9.1%
23.8
32.2
bsPM vs. Start Of Main Injection Low-Load Condition
10.9%
20.0 %
31.8
10
bsPM [g/kW-hr]
0.250
0.300
0.350
0.400
0.450
0.500
0.000
0.050
0.100
0.150
0.200
160
-10
Increasing
-8
7.6%
20.0%
29.5%
-6
-2
7.5
20.4
30.9
0
Stock EGR
2
20.6%
Reduced EGR
Start of Main Injection [CA From TDC]
Increased EGR
-4
7.4%
20.2%
29.6
9.4%
29.5%
bsPM vs. Start Of M ain Injection (M axim um Pow er)
4
6
8.5%
19.9%
8
2.61 Internal Combustion Engines Design Project Solution
Here is a possible solution for the design problem. 1. Base Engine Table 1 below summarizes the main parameters of the base engine Table 1 Base Engine Summary 11
Displacement (m3)
6
Cylinders Bore (m)
0.1326
Stroke (m)
0.1326 18
Compression Ratio Connecting Rod Length (m)
0.3315
Bmep @ max torque (kPa)
895
Bmep @ max power (kPa)
835
Maximum torque (N-m)
783
Maximum power (kW)
173
Maximum engine speed at maximum power(RPM)
2261
There are two possible methods to size the engine, and they should be consistent with each other: Method 1: P
K mK f ,iK v N Vd QHV U a ,oI F / A stoich
(1)
2
Method 2: Assume a bmep based on practical limits and fuel-air cycle charts, and solve for the power output: P
bmep Vd N
(2)
2000
161
Using the first method, we must determine K m , K f ,i , K v , U a ,o , I , N
Calculate Engine Speed (N) N
S p max find L 2L
(3)
Assume B ~ L, so
6
Vd
L
SB L 6SL 2
cylinders
4
3
(4)
4
1
1
§ 4Vd · 3 ¨ ¸ © 6S ¹
§ 4(0.011) · 3 ¨ ¸ © 6S ¹
0.1326m
so: N
10m/s 2(0.1326)
37.71revs / sec, or 2260 RPM
Determine I, and Kf,i Chose rc=18 (maybe a bit high), and I=0.7 (smoke limit, maximum possible fuel we can get in per mass of air). Using Fuel-air cycle results (Fig. 5-9, Heywood p. 182), then Kf,i= 0.575. Applying a correction factor of around 80%, actual Kf,i= 46%. The correction factor can be between 80% and 85%; For this case, I chose 80% so that Method 1, and 2, as explained above, are consistent with each other. Determine IMEP For phi=0.7, and rc=0.8, we get imep Pi
10.5 so imep
P1 10.5 ,
(5)
Note that Pi is not atmospheric pressure. At WOT, there is a pressure loss in the intake system, due to frictional losses that scale with speed. Pi will be less than atmospheric. Likewise, the exhaust pressure (Pe) is not atmospheric; a higher than atmospheric pressure is needed to pump the gases through the exhaust system. Once the gases leave the exhaust system and reach ambient conditions, they will expand to atmospheric pressure. Additionally, depending on the opening timing of the exhaust valves, the gases might exit at a higher pressure than what is required to overcome the pumping loss in the
162
exhaust system. To get an idea, of the value of Pi, look at Figure 13-13 in the text (Heywood P. 725). For a piston speed of 10 m/s, pmep
(Pe - Pi)
0.4 x ( S p) 2
0.4(10) 2
40
(6)
Now allocate this pumping loss between Pe and Pi. At high speeds around 18% of the loss is on the intake side, and the remaining 82% on the exhaust side. This will be consistent with volumetric efficiency as explained below. So: Pi= 101 kPa – 0.18(40 kPa) = 93.8 kPa Pe=101 kPa + 0.82(40 kPa) = 133.8 kPa We can now calculate an imep:
93.8 kPa 10.5
imep
984.9kPa
Determine Mechanical Efficiency Km
Km
§ imep tfmep · tfmep ¨¨ ¸¸ 1 ; imep imep © ¹
where tfmpe total friction mep
(7)
fmep (rubbing friction and auxiliary mep) pmep
From figure 13-7 (Heywood p 722), fmep for a fired engine at 2260 rpm | 140 kPa. So
tfmpe
140 40 kPa
Km
180 985
1
180kPa; and
81.7%
Determine Volumetric Efficiency and U a ,o Using figure 6-8 (Heywood p. 217), assume a volumetric efficiency of 90% for a piston speed of 10 m/s. Note that this volumetric efficiency measures the efficiency of the entire intake system. Also note that we have chosen the right pressure loss allocation for the intake system (as calculated in the imep section), consistent with volumetric efficiency. The air density U a ,o , is just calculated from ideal gas law, at ambient conditions. The value is 1.17 kg/m3
163
Fuel-to-Air Ratio & Heating Value From table D.4 in the text (Heywood p. 915) we get the stoichiometric Fuel-to-Air ratio of gasoline as 0.0697, and its heating value of 43.2 MJ/kg Power calculation With the estimates for each value, we can now calculate the power
P
(0.817)(0.46)(0.90)(37.7 m / sec)(0.011m ) 43.2e3kJ / kg 1.17 kg / m 3 (0.7)(0.0697) 2
P 173kW We also use method 2 to check for consistency. Rearranging equation 2.19b (Heywood p50), we get: P
bmep Vd N 805 kPa 11 dm3 37.7rev / sec 2000
2000
167 kW
the methods are close For low loads, follow the same procedure, with lower pumping loss, due to lower speed (see figure 13-13, Heywood), and lower rubbing and auxiliary friction (see figure 13-7 Heywood); additionally, the allocation of pressure losses is different, and must be consistent with volumetric efficiency.
2. Boost, Turbo-machinery and Intercooler Boost pressure: To find the boost pressure required, we use equation 1, and replace the volumetric efficiency for the entire inlet system with the volumetric efficiency for the valves only ( K v ~ 94%). We also replace the ambient air density with the air density right before the valves, U a ,i . This density can be determined from the ideal gas law, knowing the pressure (which is approximately cylinder pressure divided by volumetric efficiency), and the temperature (about the same as the cylinder temperature). Thus, we can vary the cylinder pressure until we get the required power level, as defined by equation 1. P
KmK f ,iKv N Vd QHV U a ,iI F / A stoich 2
Note that as we vary the cylinder pressure, and consequently the density, the mechanical efficiency (as defined by equation 7 above) will also change because the pumping loss will change.
164
Pmep= Pexhaust - Pintake Thus the solution to this problem is iterative, and can easily be done with a spreadsheet. After varying the cylinder pressure, determining the corresponding air density at the valves through the ideal gas law, and calculating the mechanical efficiency, we get the following target power:
P
(0.918)(0.46)(0.944)(37.7 m / sec)(0.011m ) 43.2e3kJ / kg 2.065kg / m 3 (0.7)(0.0697) 2
P
360kW
For this case the pressure that gives a density of 2.065 is 176 kPa, as dictated by the ideal gas law: Pcylinder
U a ,i (beforevalves ) RT(beforevalves ) *K v _ valves
.314kJ / kmoleK · ¸ 314 K (0.944) 2.065kg / m §¨¨ 828 .97kg / kmole ¸ 3
©
¹
which gives Pcylinder=176 kPa. The pressure that must come out of the compressor is approximately: Pcylinder
Pcomp
Kv
176kPa 0.944
186kPa
Thus the desired boost is 85 kPa. That is we have to compress 85 kPa above atmospheric. Note that to relate pressure before the valves, and after the valves, as a first approximation I have used the volumetric efficiency. Turbo-machinery Knowing the desired boost, the turbo-machinery can now be sized to generate the required pressure. This is done using the insentropic relationships for the compressor and turbine. First we must size the compressor by finding the work required to compress the gas to the desired pressure. Second, we must size the turbine to produce the work that drives the compressor. To determine the amount of work that is required to compress the gas we do an energy balance assuming an adiabatic compressor: Wc
m C p (T2 a T1 )
(8)
where, T2a= Actual compressor exit temperature T1= Compressor inlet temperature (300K)
165
We calculate T2a using the compressor efficiency, and isentropic relationships:
T2 a
T2 s T1
Kc
T1
(9)
and: T2 s
§ 1J · ¨ ¸ J ¸¹
§ P · ¨© T1 ¨¨ 1 ¸¸ © P2 ¹
(10)
T1, P1, P2, J, and Kc are all known, so T2a, and consequently the compressor work can be calculated. Knowing the compressor work, we now size the turbine using the following equations: Wc (11) Wt
Km where K m is the mechanical efficiency for the turbine and compressor system. 95% is reasonable estimate for this number W (12) T5 a T4 t Cp where: T5a=Actual turbine exhaust temperature T4 = Turbine inlet temperature (engine exhaust, given at 900K) To find the required turbine pressure ratio: P4 P5
§ T5 s ¨¨ © T4
Pr
· ¸¸ ¹
§ J ¨¨ © 1J
· ¸¸ ¹
(13)
where:
T5 s
T5 a T4
Kt
T4
(14)
Thus, enough equations for enough unknowns. Values for the temperatures, pressures, and compressor work, are show in table 2. Intercooler Adding an intercooler to lower the intake temperature, will increase the density of the gas, and consequently decrease the required boost, as reflected in table 2. For a given pressure rise we get a higher change in density (due to lower gas temperatures going into the engine). To size the intercooler you can select a coolant, and based on adequate
166
estimates for inlet and outlet coolant temperatures, you can determine the required mass flow-rate that is needed to achieve a certain temperature change in the air. You must use the definition for heat exchanger to determine the allowed change in air temperature: For the coolant I used water (Cp=4.2 kJ/kg K), and assumed that it goes in at 300 K, and I want it to leave at 380 K. To find the exit temperature of the air, and to determine the required flow rate of water, I use the definition for heat exchanger effectiveness in conjunction with an energy balance: For effectiveness we have:
H
m Cp coolant ,or air (Tin Tout ) coolant or air m Cp min (Tin Tout _ max )
and for the energy balance we have:
m Cp'T coolant
m Cp'T air
For this case, I chose the air and water to have about the same capacitance (mCp). Using the effectiveness equation I can solve for Tout air. Note that the capacitances will cancel out in the equation, and the maximum change in temperature occurs when Tout air= Twater in, thus:
H
(Tin Tout ) airr Tair _ out (Tin _ air Twater _ in )
Tair _ in H (Tair _ in Twater _ in )
Assuming, water temperature increases from 300 to 357, then Tair in is 314 K. Values for the heat exchanger temperatures and flow-rates are also shown in table 2.
Figure 1 Schematic of Turbocharged Engine with Intercooler
2
3 Heat Exchanger
Engine
7
6 4
compressor
turbine 167
1
5
Table 2 Turbocharged Engine with Intercooler: Operating Parameters No intercooling
With intercooling
State
Temperature (K)
Pressure (kPa)
Temperature (K)
Pressure (kPa)
1
300
101
300
101
2
401
233
372
186
3
401
233
314
186
4
900
223
900
188
5
801
127
833
127
6 7
N/A N/A
N/A N/A
101 101
300 380
Work Compressor
40 kW
Work Compressor
29 kW
Intercooler m_dot water m_dot air
0..097 kg/sec 0.404 kg/sec
3. Brake Efficiency At half maximum speed, and full load at that speed, I kept the same boost, but lowered the fmep, per figure 13-7 in the text. The BSFC came out to be 188 g/kw-hr. This number is actually quite good for industry standards. Other people perhaps got lower (around 175), however, as I previously explained, I was more conservative in my efficiency estimate from fuel air cycle tables, to be consistent with different ways of calculating power. My calculation is shown below:
bsfc
m f ( g / hr )
0.010078kg / sec(3600 sec/ hr )(1000 g / kg ) 193
Power (kW )
188 g / kW hr
Note that if we directly use break engine efficiency, we should get the same answer:
bsfc
1 K f ,b (QHV )
1 K mK f ,i (QHV )
1 0.96(0.460)(43.2)
188 g / kW hr
Ways to decrease bsfc include raising compression ratio, and reducing frictional losses. 4. Emissions – NOx
168
Figure 2 Schematic of Turbocharged Engine with Intercooler 7
2
Inter-
3
4 Venturi/ Mixer
cooler
InterEngine
cooler
5
compressor
turbine
1
6
The schematic shows how EGR will be driven from the engine. There are a few ways of doing this; one way is to use a Venturi system, as shown above. Another way is to optimize the system so that the pressures at the air and EGR intersection are about the same. It is necessary for these pressures to be equal, otherwise there will be backflow in the direction of lower pressure. Overall, however, the addition of EGR will impact the fuel economy of the engine. This is the price that we must pay to have lower emissions. The first step of this problem is to define the amount of EGR that is needed to meet EPA emissions levels. The emissions requirements along with their safety levels are shown in table 3 below. Table 3 NOx safety NOx Safegy NOx Standard Engine Mode target g/bkW- target g/bkWg/bhp-hr hr hr
EGR
Timing (CA from TDC)
Hit in Fuel economy (percentage points)
25% max torque, 1600 rpm
2.50
1.75
2.35
24%
1
2%
Maximum power
2.50
1.75
2.35
24%
0.5
2%
169
Using the figures provided, there are various possibilities for selecting EGR, depending on the hit on fuel economy. Figure 3 below is an example that shows that there is a range of timings and EGR levels that will give the proper amount of NOx
Figure 3 Acceptable operating area for low load
Must operate below dashed line
Once EGR has been calculated, the loss in engine efficiency can be assessed, as well as the required boost. Again this is an iterative process. There are many variables affecting engine power, and they are all related as well, thus at least a spreadsheet must be setup. For example, boost affects engine power, but it also affects mechanical efficiency, which in turn affects engine power, thus all these variables must be connected when solving the system. One important implication of adding EGR, is that the pressure in the cylinder chamber must increase if we are to maintain constant mass of fresh fuel and air; this is what we should desire if we are to maintain the same power output from the engine as the case without EGR. The total pressure is equal to the sum of the partial pressures of air and the EGR:
170
PT
Pair PEGR
Assuming the molecular weights of both Air and EGR are about the same, then the mole fraction is approximately equal to the mass fraction of each mixture, and PT can be expressed as: PT
Pair 1 EGR
Additionally, since there is a pressure loss of around 16 to 20 kPa associated with the venturi, a higher boost is still needed. To reach the target power output, a total boost of 194 kPa is required, for total pressure of 295 kPa. This is a high boost, higher than industry standard for this size engines. Perhaps a more practical boost is 150 kPa (PT =251 kPa), or less. However this limits the maximum power to 310 kW. If you recognized the practical limitations, this is a perfectly acceptable answer.
Table 4 With intercooling State
Temperature (K)
Pressure (kPa)
1
300
101
2
434
295
3
327
295
4
349
279
5
990
241
6 7
864 438
121 241
Table 5 Intercooler
Mass flowrate (kg/sec)
Tin (K)
Tout (K)
EGR
0.171
300
380
Compressor
0.131
300
380
171
5. Emissions Particulates (a) The particulate emissions corresponding to the chosen EGR level, can be obtained from the data provided (PM levels vs injection timing for various EGR fractions). At 24% EGR, the PM coming out of the engine is approximately 0.2 g/bkW-hr (b) To size the trap, use the space velocity that will minimize the volume of the trap, in this case 28,000 hr-1. This is evident from the relationship for space velocity:
V V
Space _ velocity
where V is the volume flow-rate of the gases going through the trap, and V is the volume of the trap. For a smaller trap volume we get a higher space velocity. Solving for the volume of the trap we get
V
V Space _ velocity
Using the ideal gas law to solve for V (m air m fuel m egr ) RTexhaust 1.10m 3 / sec V Pexhaust Solving for Volume
V
1.10m 3 / sec 28,000 / 3600 sec
0.141m 3 141 L
Values used are shown in table 6 below Table 6 Mdot_air&fuel (kg/sec)
Mdot_egr (kg/sec)
0.43
0.10
Texhaust (K) Pexhaust (kPa) 864.74
121.00
Space Velocity (1/sec)
Volume flowrate (m3/sec)
Trap Volume (L)
7.78
1.10
140.87
As shown in figure 4 of SAE 2003-01-0047, The maximum pressure loss through the trap, at a space velocity of 28,000/hr, is 6 kPa. This is a very small percentage of the total exhaust pressure (~2.5%), and the effect on turbo-machinery is small.
172
6. 2007 Emissions requirements Using the same safety factor as in part 6, the table below shows the new emissions that must be met: Table 7 NOx NOx safety Standard target g/bhp-hr g/bhp-hr 0.2000
0.1400
NOx Current Required Safegy Engine out efficiency target g/bkW-hr (catalytic) g/bkW-hr 0.1879
2.3490
PM Standard g/bhp-hr
0.9200
0.0100
PM safety PM Safegy Current Required target target Engine out Efficiency g/bhp-hr g/bkW-hr g/bkW-hr (trap) 0.0050
0.0067
0.2000
Where catalytic converter efficiency is defined as:
Kcat
1
m pollu tan t , out m pollu tan t ,in
As shown in table 7, a catalytic converter with 92% efficiency will be needed to meet 2007 NOx emissions requirements. The required particulate trap efficiency is fairly high (97%) but the trap presented in the Ford paper seems to have efficiencies of around 99%, so it should work fine for 2007 emissions requirements.
173
0.966
Bibliography on Internal Combustion Engines Obert, Edward F. Internal Combustion Engines and Air Pollution. New York: Intext Educational Publishers, 1973 edition. A good basic text on engines from the 1950s with modest updating in 1968; much excellent descriptive material. Taylor, C. Fayette, and Edward S. Taylor. The Internal Combustion Engine. International Textbook Company, 1961. A basic text now out of print and somewhat dated. Taylor, C. F. The Internal Combustion Engine in Theory and Practice. Vol. 1, and 2. Cambridge, MA: M.I.T. Press, 1966 and 1968. Reissued in paperback in 1977, and in 1985 as Second Edition with minor modifications. A much expanded version of reference 2; an advanced text with extensive material on engine design practice of the 1950s and 60s. Rogowski, A. R. Elements of Internal Combustion Engines. New York: McGraw-Hill, 1953. An elementary text used primarily for undergraduate teaching. Lichty, L. C. Combustion Engine Processes. 6th ed. New York, : McGraw-Hill Book Company, 1967. ISBN: 0070377200. A good basic text on all types of combustion engines, now somewhat dated. Khovakh, M., (general editor). Motor Vehicle Engines. English translation form Russian. Moscow: MIR Publishers, 1976. A Russian text with an excellent ordering of subject material. Patterson, D. J., and N. A. Henein. Emission from Combustion Engines and their Control. MI: Ann Arbor Science Publishers, Inc., 1972. A comprehensive text on engine emissions; now somewhat dated. Ayres, Robert U., and Richard P. McKenna. Alternatives to the Internal Combustion Engines. Baltimore: Johns Hopkins University Press, 1972. ISBN: 0801813697. A fundamental text on the alternative engines to the internal combustion engine. Nunney, M. J. The Automotive Engine. London: Newnes-Butterworths, 1974. A book which reviews modern automotive engine practice; contains descriptions of design and operation of engines and engine components. Yamamoto, Kenichi. Rotary Engine. Mazda: Toyo Kogyo Co., Ltd., 1969. Excellent text on the design and operation of Wankel engines.
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Ansdale, R. F. The Wankel RC Engine: Design and Performance. London: Iliffe Books, Ltd., 1968. Contains much technical and historical information on the Wankel engine. Springer, G. S., and D. J. Patterson, eds. Engine Emissions: Pollutant Formation and Measurement. New York, and London: Plenum Press, 1973. A set of contributed chapters on different emissions topics; some chapters are still useful. Sitkei, G. Heat Transfer and Thermal Loading in Internal Combustion Engines. Budapest: Akademiai Kaido, 1974. A monograph on heat transfer in spark-ignition and diesel engines and temperature distributions in engine components. Annand, W. J., and G. E. Roe. Gas Flow in the Internal Combustion Engine. Haessner Publishing, Inc., 1974. A review of selected topics related to gas flow in IC engine intake and exhaust systems. "Should We Have a New Engine?" An Automobile Power Systems Evaluation. Vol. 1. Summary, Jet Propulsion Laboratory, California Institute of Technology, JPL SP 43-17, August 1975. Popular summary of study which evaluates the internal combustion engine and its alternatives. Goodger, E. M. Hydrocarbon Fuels; Production, Properties and Performance of Liquids and Gases. London: Macmillan, 1975. Useful review of fuels, automotive and non-automotive. Cummins, Lyle. Internal Fire: The Internal Combustion Engine 1673 - 1900 Revised Edition. 2nd ed. Warrendale, PA: Society of Automotive Engineers, 1976. Excellent and readable history of the internal combustion engine by the son of the founder of the Cummins Engine Company. A History of the Automotive Internal Combustion Engine. Warrendale, PA: Society of Automotive Engineers special publication, SP-409, 1976. A set of four SAE papers reviewing the history of IC engine developments. Blackmore, D. R., and A. Thomas. Fuel Economy of the Gasoline Engine. New York, NY: John Wiley & Sons, 1977. A useful introduction to how fuel properties affect spark-ignition engine operation. Thomson, W. Fundamentals of Automotive Engine Balance. London: Mechanical Engineering Publications, Ltd., 1978. A short straightforward monograph on the balancing of various arrangement reciprocating engines.
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Benson, R. S., and N. D. Whitehouse. Internal Combustion Engines. Vol. 1, and 2. London: Pergamon Press, Inc. 1979. A modern text, limited in scope, with special emphasis on computer simulations of engine flow and combustion processes. Watson, N., and M. S. Janota. Turbocharging the Internal Combustion Engine. New York: John Wiley & Sons, 1982. An extensive and excellent professional reference text on turbochargers, and turbocharged engine performance. Benson, R. S. The Thermodynamics and Gas Dynamics of Internal Combustion Engines. Vol. 1. Edited by J. H. Horlock, and D. E. Winterbone. Oxford: Clarendon Press, 1982. Extensive and detailed monograph on unsteady engine intake and exhaust flow processes. Horlock, J. H., and D. E. Winterbone, eds. The Thermodynamics and Gas Dynamics of Internal Combustion Engines. Vol. 2. Oxford: Clarendon Press, 1986. Extensive and detailed monograph on in-cylinder engine processes and methods of analysis. Hilliard, J. C., and G. S. Springer, eds. Fuel Economy in Road Vehicles Powered by Spark Ignition Engines. New York, and London: Plenum Press, 1984. A set of contributed chapters on engine and vehicle factors which affect fuel economy; some are excellent. Stone, R. Introduction to Internal Combustion Engines. MacMillian Publishers, Ltd., 1985. 2nd ed. 1992. An introductory text appropriate to a survey undergraduate course on engines. Ferguson, C. R. Internal Combustion Engines--Applied Thermosciences. New York: John Wiley & Sons, 1986. A new text focusing primarily on Thermal/Fluids Science aspects of engine operation. Bosch. Automotive Handbook. 5th ed. Published by Robert Bosch GmbH. Warrendale, PA: Distributed by SAE, 2000. A concise and useful summary of technical data on engine and vehicle components and systems. Heywood, J. B. Internal Combustion Engine Fundamentals. London: McGraw-Hill, 1988. An extensive text and professional reference on the fundamentals behind engine operation and design. Bosch. Automotive Electric/Electronic Systems. Published by Robert Bosch GmbH. Warrendale, PA: Distributed by SAE, 1988. A practical guide to and description of automotive electrical systems.
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Arcoumanis, C., ed. Internal Combustion Engines. London; San Diego: Academic Press, 1988. A collection of contributed chapters on gasoline and diesel engines, turbocharged engines and automotive fuels; some are good. Blair, G. The Basic Design of Two-Stroke Engines. Warrendale, PA: Society of Automotive Engineers, 1990. A monograph with simple programs focused on two-stroke gasoline engine design issues and their underlying principles. Owen, K., and T. Coley. Automotive Fuels Handbook. Warrendale, PA: Society of Automotive Engineers, 1990. An extensive compilation of information on gasolines and diesel fuels and their effects on engine operation. Newton, K., W. Steeds, and T. K. Garrett. The Motor Vehicle. 11th ed. London; Boston: Butterworth, 1989. A useful source of practical information on engines, transmissions and vehicles. Lenz, H. P. Mixture Formation in Spark-Ignition Engines. New York, NY: SpringerVerlag, 1990. A resource for detailed information on gasolines, carburetors, fuel injection systems, and the mixture formation process. Ramos, J. I. Internal Combustion Engine Modeling. New York: Hemisphere Publishing Co., 1989. A review and useful introduction to the various models now available for engine processes. Heck, R. M., and R. J. Farranto. Catalytic Air Pollution Control. New York: Van Nostrand, Reinhold, 1995. A readily understandable review of catalyst fundamentals and application to vehicles. Blair, G. P. Design and Simulation of Two-Stroke Engines. Warrendale, PA: SAE, 1996. An update and extension of Blair’s earlier book; extensive information on small highperformance two-stroke spark-ignition engines. Sher, E., ed. Handbook of Air Pollution from Internal Combustion Engines: Pollutant Formation and Control. San Diego, CA: Academic Press, 1998. An extensive set of chapters, by different authors, on four-stroke and two-stroke cycle sparkignition and diesel engine operation and emissions, and fuel effects. Pulkrabek, W. W. Engineering Fundamentals of the Internal Combustion Engine. New York: Prentice-Hall, Inc., 1997. An introductory text on IC engine fundamentals.
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Borman, G. L., and K. W. Ragland. Combustion Engineering. WCB McGraw-Hill, 1998. A valuable reference volume on combustion processes in different practical systems, including IC engines, with extensive information on fuels. Heywood, J. B., and E. Sher. The Two-Stroke Cycle Engine: Its Development, Operation and Design. Warrendale, PA: SAE, Taylor & Francis, 1999. A comprehensive summary of the technical literature on two-stroke cycle engine processes which govern its operation and its design. Flagan, R. C., and John H. Seinfeld. Fundamentals of Air Pollution Engineering. Englewood Cliffs, CA: Prentice-Hall, Inc., 1988. A review of air pollutant formation processes and sources, and control approaches. Lenz, H. P., and C. Cozzarini. Emissions and Air Quality. Warrendale, PA: SAE, 1999. A concise handbook with data on transportation emissions, their impact, and ways to control their magnitude. Challen, B., and R. Baranescu, eds. Diesel Engine Reference Book. 2nd ed. Warrendale, PA: SAE, 1999. An extensive handbook on the theory, design, and applications of diesel engines. Bosch. Gasoline-Engine Management. 1st ed. Published by Robert Bosch GmbH. Warrendale, PA: Distributed by SAE, 1999. A handbook with extensive practical details on gasoline spark-ignition engines and their management and control. ———. Diesel-Engine Management. 2nd ed. Published by Robert Bosch GmbH. Warrendale, PA: Distributed by SAE, 1999. A handbook witih extensive practical details on diesel engines, their emissions, and their management and control. Stan, C., ed. Direct Injection Systems for Spark-Ignition and Compression-Ignition Engines. Published by Springer-Verlag, Berlin, Heidelberg. Warrendale, PA: Distributed by SAE, 1999. Multi-author volume on direct injection gasoline and diesel engines, focusing on the different practical approaches to direct injection of liquid fuel into the cylinder. Winterbone, D. E., and R. J. Pearson. Theory of Engine Manifold Design. Warrendale, PA: SAE, 2000. A text on the theory and methodology for analyzing unsteady gas flows in engine manifolds. ———. Design Techniques for Engine Manifolds. Warrendale, PA: SAE, 1999. A comparison text to #49, focusing on application of unsteady gas flow analysis tools to engine manifold design.
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Blair, G. P. Design and Simulation of Four-Stroke Engines. Warrendale, PA: SAE, 1999. A description of engine simulations, largely developed in the author’s laboratory, and their application to four-stroke engine performance prediction and design. Ferguson, C. R., and A. T. Kirkpatrick. Internal Combustion Engines Applied Thermosciences. 2nd ed. NY: John Wiley & Sons, Inc., 2001. A new edition of #27: An introductory text focusing on the thermal science processes important to internal combustion engine operations. Nuti, M. Emissions from Two-Stroke Engines. Warrendale, PA: SAE, 1998. A monograph on two-stroke cycle gasoline engines, the origins of their emissions and methods of control. Eastwood, P. Critical Topics in Exhaust Gas after-Treatment. Research Studies Press Ltd., 2000. A detailed monograph on engine exhaust gas treatment—catalysts, particulate filters—as well as exhaust treatment system issues. Makartchouk, A. Diesel Engine Engineering: Thermodynamics, Dynamics, Design, and Control. New York, and Basel: Marcel Dekker, Inc., 2002. Analysis based text, focused primarily on engine dynamics, structural design, and automated diesel engine control. Zhao, F., D. L. Harrington, and M-C. Lai. Automotive Gasoline Direct-Injection Engine. Warrendale, PA: SAE, 2002. An extensive review of the literature on GDI engine performance, combustion, efficiency, and emissions, and the state of GDI engine development.
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