250339586-Design-of-Hydraulic-Systems-for-Lift-Trucks.pdf

Ivan Gramatikov

Design of Hydraulic Systems for Lift Trucks

Second Edition

Preface to the Second Edition All information contained in the first edition has been retained. Some corrections and additions have been made to better serve the purpose of the book.

Design of Hydraulic Systems for Lift Trucks First Edition Published by Technical University- Sofia, Sofia 1000, Bulgaria ISBN: 978-954-438-730-3 Printed in Bulgaria

Second Edition Copyright 2011 by Ivan Gramatikov All rights reserved. No part of this book may be reproduced, stored in a retrieval system or transmitted in any form, or by any means, electronic, mechanical, photocopying, recording or otherwise, without the prior written permission of the author. For permissions e-mail: [email protected]

ISBN: 978-1-257-01500-9 Printed in the United States of America Front cover photos: Courtesy of Balkancar Record (http://www.balkancar-record.com)


i

CONTENTS Chapter 1:

Introduction

1

Preface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1

Definitions for design and system design . . . . . . . . . . . . .

2

Regulations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3

Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

4

Systems of units . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

4

Symbols used in formulae and hydraulic diagrams . . . . . .

5

Chapter 2:

Properties and parameters of the fluids

11

Properties Density . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

11

Specific weight . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

12

Specific gravity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

13

Viscosity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

13

Compressibility of fluids . . . . . . . . . . . . . . . . . . . . . .

16

Reynolds number and types of flow . . . . . . . . . . . . .

18

Parameters Pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

19

Flow and flow rate . . . . . . . . . . . . . . . . . . . . . . . . . .

20

Fluid velocity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

23

Work and Power . . . . . . . . . . . . . . . . . . . . . . . . . . . .

23

Drag and pressure loss . . . . . . . . . . . . . . . . . . . . . .

25

Hydraulic shock . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27

ii

Hydraulic Lock . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

27

Obliteration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

28

Stiction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29

Cavitation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

29

The Bernoulli Equation . . . . . . . . . . . . . . . . . . . . . . . .

30

The Torricelli Equation . . . . . . . . . . . . . . . . . . . . . . . .

31

Chapter 3:

Hydraulic system components 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18.

Flow Restrictors . . . . . . . . . . . . . . . . . . . . . . . . . . . . Pressure Relief Valves . . . . . . . . . . . . . . . . . . . . . . . Check Valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Reduction Valves . . . . . . . . . . . . . . . . . . . . . . . . . . . Pressure Compensated Flow Controls . . . . . . . . . . . Directional Control Valves . . . . . . . . . . . . . . . . . . . . . Hydraulic Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Motors . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Cylinders . . . . . . . . . . . . . . . . . . . . . . . . . . Pressure Sensors . . . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Accumulators . . . . . . . . . . . . . . . . . . . . . Hydraulic Filters . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Reservoirs . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Lines, Fittings and Couplings . . . . . . . . . . Manifold blocks . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hydraulic Fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Fluid Cleanliness . . . . . . . . . . . . . . . . . . . . . . . . . . . Electric Motors . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

33 34 36 37 39 40 42 48 59 60 64 66 70 77 83 88 90 95 98

Chapter 4:

Management and quality of hydraulic system design process 101 Brief history of quality . . . . . . . . . . . . . . . . . . . . . . . . . .

101

Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

103

Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

104


iii

Structuring the design process . . . . . . . . . . . . . . . . . . . .

106

Definitions of tools used . . . . . . . . . . . . . . . . . . . . . . . . .

108

Description of the design process steps . . . . . . . . . . . . .

110

Design guidelines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

116

Documenting the design activities . . . . . . . . . . . . . . . . . .

117

Project close-out criteria . . . . . . . . . . . . . . . . . . . . . . . . .

118

Failure and failure rate . . . . . . . . . . . . . . . . . . . . . . . . . . .

119

Patents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

120

Designing around an existing patent . . . . . . . . . . . . . . . .

122

Legal aspect of the design process . . . . . . . . . . . . . . . . .

123

Chapter 5:

Hydraulic systems for high lift trucks

125

Elevating system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

126

Hydraulic systems overview . . . . . . . . . . . . . . . . . . . . .

128

Design principles . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

129

Design requirements . . . . . . . . . . . . . . . . . . . . . . . . . . .

130

Hydraulic system with proportional manual directional valve 133 Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

146

Hydraulic system with electrically controlled proportional valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

153

Hydraulic system with emergency lowering . . . . . . . . . .

158

Energy recovery systems . . . . . . . . . . . . . . . . . . . . . . . .

160

Hydraulic steering system . . . . . . . . . . . . . . . . . . . . . . .

165

Electro-hydraulic steering system . . . . . . . . . . . . . . . . . .

171

Integrated hydraulic system . . . . . . . . . . . . . . . . . . . . . .

174

Smoothness of the lifting . . . . . . . . . . . . . . . . . . . . . . . . .

176

Chapter 6:

Hydraulic systems for low lift trucks

181

iv

Hydraulic system with independent power steering and lift circuits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

183

Integrated hydraulic systems for low lift trucks . . . . . . . .

185

Integrated hydraulic system with accumulator . . . . . . . .

189

Hydraulic system for pallet trucks with long fork attachments 194 Hydraulic power-assisted steering . . . . . . . . . . . . . . . . .

197

Integrated system with power-assisted steering . . . . . . .

199

Chapter 7:

Hydraulic systems for boom-type trucks 201 Hydraulic circuit for boom lift, extend and fork tilt . . . . . . .

202

Hydraulic lift & lower circuit for telescopic boom . . . . . . .

203

Hydraulic circuit with an automatic shut-off valve . . . . . .

207

High-speed extension of telescopic boom . . . . . . . . . . . .

208

Chapter 8:

Selected topics I.

211

Servicing the hydraulic systems . . . . . . . .

211

Troubleshooting principles . . . . . . . . . . . . . . . . . . . . . . . . System Life . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

212 212

Safety Rules . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

213

Servicing the fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

213

Servicing filters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

216

Servicing reservoirs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

216

Servicing rotary pumps and motors . . . . . . . . . . . . . . . . . .

217

Servicing hydraulic cylinders . . . . . . . . . . . . . . . . . . . . . . . .

218

Servicing valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

219

Servicing connectors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

220

Seals . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

221


v

II.

Components layout- general considerations

222

III.

Common problems . . . . . . . . . . . . . . . . . . . . .

223

IV.

Contamination of the hydraulic fluid . . . . . .

225

V.

The future of the hydraulics . . . . . . . . . . . . .

229

Appendixes

231

Appendix A

ITA classification

Appendix B

Physical properties of common fluids

Appendix C

Viscosity Classification of Industrial Lubrication Fluids

Appendix D

Coefficients of local resistance

Appendix E

Decision Matrix and QFD house

Appendix F

Hydraulic system calculation

vi


1

Chapter 1

Introduction Preface The purpose of this book is to illustrate design principles and methods for designing and calculating hydraulic systems for industrial lift trucks. Determining the main parameters of these systems is based on principles of hydraulics and mechanics. This book is to be used as a source of information for mechanical engineers involved in designing, manufacturing and servicing hydraulic systems for mobile lift trucks. This book can also be used by engineering students in Industrial Truck Programs. To combine these two purposes, there is an introductory chapter, “Properties and Parameters of Hydraulic Fluid”, and a chapter on “Hydraulic Components” describing the construction and the functions of components used in mobile hydraulic systems. This book will also be beneficial for engineers working in areas of design, fabrication and service of any other mobile off-highway equipment. In all universities, mechanical engineering students study the theoretical foundations of fluid mechanics, fluid dynamics, and thermodynamics. However few universities offer courses in hydraulics and pneumatics (also called: fluid power), which are the applications of these disciplines. That is why most design engineers learn the basics of the fluid power on the job. Fluid power learning time can be reduced significantly if some basic hydraulic principles are understood up front. This book will describe the hydraulic principles and operation of the main hydraulic arrangements which will give you the foundation for designing any system on your own. It is more difficult to design hydraulic systems for smaller lift trucks. That is because these systems must have the same performance as the bigger trucks but they have to be put into a smaller space envelope. The smaller design envelope is a major challenge to the design engineers. To meet this and all other challenges through the design process, engineers have to follow the principles of continuous improvement and design process quality. Quality of the design process depends on the proper execution of each step

2

Chapter 1: Introduction

of the process. The proper execution requires knowledge in engineering and management areas. The core necessary disciplines are: Mathematics, Mechanics of the Fluids, Hydraulic Circuits and Components, Management of Quality, Project Management, Design for Excellence and Professional Communication. Some of these courses, in most of the engineering programs, are not part of the engineering curriculum and therefore, engineers must take extra courses in order to acquire the right set of knowledge. Chapter 4, “Management and Quality of the Design Process”, describes the managerial aspect and the basic principles of the design process.

Definitions for design and system design • • •

“The best design is the simplest one that works” Albert Einstein Design is creative problem solving. System design is finding the balance in system performance that best satisfies the engineering requirements. This balance has to be achieved first at the conceptual level and then maintained throughout the whole design process.

Design of hydraulic systems is built on knowledge of several fundamental principles. Most fluid power engineers have them as background knowledge and do not even think about them. For people learning hydraulics, knowing the fundamental principles is the first step to designing energy and cost efficient systems. The milestones of the hydraulic principles are: • Knowledge of properties and parameters of the fluids • Velocity-pressure relationship (Bernoulli equation) • Knowledge of the hydraulic components Fluid properties, fluid parameters and the Bernoulli equation are described in Chapter 2. Chapter 3 describes the components used in the system. Good system designs would also require knowledge of: • • •

The engineering requirements (parameters) for the system Factors affecting system functionality and system life Constraints- cost, space, surrounding environment

When designing a system, the engineer must focus on four main aspects:


3

First: maximizing the system efficiency and the system life. In order to achieve this requirement, the design engineer has to select the components of the hydraulic system so that they will work together in a way leading to maximum system efficiency. Second: design for manufacturability and assembly Third: design for test and service Fourth: design a cost effective system These four aspects are described in chapters 4, 5, 6 and 7. In addition to designing the hydraulic system, the system engineer has to also consider how the system interacts with other systems (mechanical, electrical, control), type of vehicle (ICE or electric) and the ergonomic consequences of the design (the interaction of the system with the people). A definition of “system engineering” is given by the International Council of System Engineers (INCOSE) Systems engineering is an interdisciplinary approach and means to enable the realization of successful systems. It focuses on defining customer needs and required functionality early on in the development cycle, documenting requirements, and then proceeding with design synthesis and system validation while considering the complete problem. System engineering integrates all the disciplines and specialty groups into a team effort forming a structured development process that proceeds from concept to production to operation. System engineering considers both the business and the technical needs of all customers with the goal of providing a quality product that meets the user needs.

Regulations In some countries, such as Canada, the engineering profession is selfregulated through provincial organizations. The governing body is comprised of engineers chosen, through a voting process, by members of the engineering organization. In other countries, such as the USA, the state governments regulate the licensing, the practices of the profession and approve the governing body of the engineering organizations.

4


Professional organizations develop standards for minimum qualification, professional ethics and practices. They are also involved in the mediation of conflicts.

Calculations Clarity and accuracy of the technical calculations are an important part of a system design. All data, assumptions, mathematical and physical laws have to be specified clearly. Calculations are an intellectual asset for a company. Therefore any other engineer with the same background should be able to understand and use them. This reduces the development time of future projects and helps to bring new products to market in a shorter time. A good practice is to put all calculations on a server in HTML or PDF format. European countries (except the United Kingdom) use a comma as a decimal marker. The UK, the USA and English speaking provinces of Canada use a period as a decimal marker. In this book, since it is written in English, I am going to use a period.

Systems of Units International System (SI) of units This system was adopted in 1960 at the Eleventh General Conference on Weights and Measures as an international standard. SI is accepted by all countries in Europe and most countries in the world. In the future, it is expected to replace all other systems and to be used by all countries. In this book we will primarily use SI units. British Systems of Units • British Gravitational (BG) System In the past, the BG system was used in the English speaking countries. In the BG system the unit of length is foot (ft), the unit of force is pound (lb), the unit of mass is obscure (slug) and the unit of temperature is degree Fahrenheit (°F).


5

Fahrenheit (°F) = [Celsius (°C) x 9/5] + 32 Celsius (°C) = [Fahrenheit (°F) – 32] x 5/9 • English Engineering (EE) System The units in the EE system are similar to the units in the BG system. The unit of length is foot (ft), the unit of mass is pound mass (lbm), the unit of force is pound force (lbf) and the absolute temperature scale is degree Rankine (°R). The equation used to convert slugs to pounds is:

slug =

lbm gC

There are two gallons: British and US gallon 1 British gallon = 4.546 litters 1 US gallon = 3.785 litters

Symbols used in formulae and hydraulic diagrams Latin alphabet A

Area [m2]

D

Diameter [m]

dP

Pump displacement [cm3/rev]

dM

Hydraulic motor displacement [cm3/rev]

EV

Bulk Modulus of Elasticity (Bulk Modulus)

F

Force [N]

G

Gravity force [N]

GQ

Flow rate, weight [N/s]

h

Height, distance [m]

k

Ratio

L

Length or distance [m]

m

Mass (kg)

6


M

Mach number [-]

n

Rotational speed (frequency of rotation) [rev/min]

P

Power [Nm/s] and [W]

p

Pressure [N/ m2] and [Pa]

Q

Flow rate, volumetric [m3/s] and [L/min]

q

Flow rate, mass [kg/s]

RL

Lineal flow resistor

Re

Reynolds Number [-]

SG

Specific gravity [-]

t

Temperature [ºC]

T

Torque [Nm]

v

Velocity [m2/s]

V

Volume [m3] and [litter]

W

Work [Nm], [J]

Greek alphabet α

Angle [rad], [º]

β

Angle [rad], [º]

γ

Specific weight [N/m3]

δ

Deviation

ε

Angular acceleration [rad/s2]

η

Efficiency

ϕ

Angle [rad], [º]

µ

Dynamic (absolute) viscosity [Pa.s]

ν ν

ρ

Kinematic viscosity [m2/s], [St] Specific volume (m3/kg) Density [kg/m3]

ρ SG

Specific Gravity [-]

τ

Shear stress [N/m2] and [Pa]

ω

Angular velocity [rad/s]

θ

Angle [rad], [º]


Hydraulic symbols ________

Work line (suction, pressure and return)

--------

Pilot line

Flexible line

Crossing lines, junction

Crossing lines, not connected

Plugged line

Venting

Reservoir, open

Reservoir, pressurized

Filter

Accumulator

7

8


Pressure gage

Thermometer

Flow meter

Foot operated

Hand operated

Spring operated

Electrical control

Electrical control, proportional

Pump, constant volume, one direction of flow


9

Pump, variable volume

Pump, pressure compensated

Hydraulic motor, one direction of flow

Hydraulic motor, reversible flow

Pump- motor, reversible flow

Flow restrictor (orifice) fixed

Flow restrictor (orifice) variable

Flow control, pressure compensated, two-way

Flow control, pressure compensated, three-way

10


Pressure relief valve

Relief valve, proportional with indirect (pilot) control

Pressure reduction valve

Check valve

Pressure switch

Steering valve, type Orbitrol

Torque generator


11

Chapter 2

Properties and Parameters of the Fluids Fluid in general is any existing liquid or gas. In lift truck hydraulic, brake and steering systems, only liquids are used as working fluids. The science of Mechanics of Fluids consists of Hydrostatics and Hydrodynamics. Hydrostatics is based on Pascal's law, which states that a confined liquid that has a pressure placed on it will act with equal force on equal areas at right angles to the area. In Hydrostatic drives, the power is transmitted on the bases of applying pressure on the fluid or by the fluid’s potential energy. In Hydrodynamic drives, the power is transmitted by the kinetic energy of the fluid.

Properties Density Density of the fluid is defined as its mass per unit volume containing the mass.

ρ=

m ⎡ kg ⎤ V ⎢⎣ m 3 ⎥⎦

Where:

2.1

m

is mass of the fluid in a unit (kg)

V

is unit volume of the fluid (m3)

In SI system density has units of kg/m3). It is designated by the Greek letter ρ (rho). In BG system density is expressed in slug/ft3 where the mass is in slugs.

12

m=

Chapter 2: Properties and parameters of the fluids

WO [ slug ] , WO is the weight in pounds at sea level 32.174

A common reference for fluids is the density of water at 4°C temperature:

⎡ kg ⎤

ρ H 2O = 1000 ⎢ 3 ⎥ ⎣m ⎦ A common reference for non-liquids is the density of iron:

⎡ kg ⎤

⎡ t ⎤

ρ IRON = 7850 ⎢ 3 ⎥ or ρ IRON = 7.85 ⎢ 3 ⎥ ⎣m ⎦ ⎣m ⎦ Density can also be expressed as:

ρ=

1 ⎡ kg ⎤ v ⎢⎣ m 3 ⎥⎦

Where:

2.2

υ is specific volume (m3/kg)

Unlike gases, the density of the fluids depends little on pressure and temperature. Densities of different fluids are given in Appendix B.

Specific Weight Specific weight is a characteristic for bodies under the influence of the gravitational field. The gravitational field is not a force (because it is massless) but it produces a force when it interacts with mater. As a result, mater receives a gravitational acceleration which does not depend on the physical state of the mass. Specific weight of fluid is equal to the product of fluid density (ρ) and gravitational acceleration g = 9.806 m/s² (g = 32.174 ft/s²). It is defined as fluid weight per unit volume containing it.


13

⎡N⎤

γ = ρg ⎢ 3 ⎥ ⎣m ⎦

2.3

Specific weight is designated by the Greek letter γ (gamma). In the SI system it has units of N/m3 or kN/m3. In the BG system the units for specific weight are lb/ft³. The intensity of the gravitational field is stronger at sea level and diminishes farther away from earth which means that the gravitational acceleration changes. For engineering application the variation of the gravitation (g) is neglected therefore, only the variation in the fluid density causes variation in its specific weight. Specific weights of different fluids are given in Appendix B.

Specific Gravity Specific Gravity is the ratio of the density of the fluid to the density of the water at the same temperature.

ρ SG =

ρ 2.4

ρ H 2O

Specific Gravity is a dimensionless parameter and it has the same values in both SI and BG systems.

Viscosity Viscosity of the fluid is a measure of resistance against friction between fluid layers. It is related to the velocity gradient ( stress ( τ ) by the equation:

du

dy )

and the shear

14


τ =µ

du dy

[Pa.s]

2.5

Where, the constant of proportionality, µ (mu), is called dynamic (or absolute) viscosity of the fluid. Fluids, for which the velocity gradient is linearly related to shearing stress, are called Newtonian fluids (all common fluids). Graphically, the slope of shearing stress vs. velocity gradient is equal to the viscosity. The value of the viscosity depends on the fluid chemical content and temperature. In most fluid problems, viscosity is combined with the density in the equation:

µ ⎡ m2 ⎤ ν= ρ ⎢⎣ s ⎥⎦

2.6

Where, the Greek letter ν (nu) is called kinematic viscosity. The dimension of kinematic viscosity in SI units is m²/s. The units Stocks (St) and Centistokes (cSt) are also used.

1St = 1cm 2 / s = 10 −4 m 2 / s

1cSt = 1mm 2 / s = 10 −6 m 2 / s

The values of ν for different fluids are given in Appendix B. In the ISO classification system viscosity is related to ISO grade. There are 18 viscosity grades covering a range from 2 to 1650 centistokes. Viscosity of the ISO grades is measured at 40° C temperature. ISO system for viscosity measurement was adopted by The American Petroleum Institute and American Society for Testing and Materials (ASTM). Today all petroleum companies and manufacturers use this system as a standard for viscosity measurement. Prior to ISO adoption, viscosity of the ASTM grades was measured at 100° F (37.8° C) in SUS (Saybolt Universal Seconds) units. SUS unit range

To convert to cSt units

from 32 to 99

cSt = 0.2253 x SUS - (194.4 / SUS)

from 100 to 240

cSt = 0.2193 x SUS - (134.6 / SUS)

more than 240

cSt = SUS / 4.635


15

Because of the small temperature difference, ISO grades are a little more viscous than the corresponding ASTM grades in SUS units. Viscosity grade classification is given in Appendix C. Another characteristic given by fluid manufacturers is the Viscosity Index (V.I.). This index is a number that indicates changes of viscosity over change of temperature. High V.I. means that there is little change in viscosity with temperature change and vice versa. Fluid viscosity is a main factor that determines the amount of friction between the fluid layers, the boundary layers thickness along the inside walls and the friction between metal surfaces of the hydraulic components. Viscosity changes with the change of temperature, pressure and contamination. When the pressure on the fluid increases, the shear stress increases leading to viscosity increase. Also, when the fluid temperature increases its viscosity decreases. The effect of temperature on kinematic viscosity of some fluids is shown in Figure 2.1.

Fig. 2.1 Source: Webtec Products Ltd. (http://www.webtec.co.uk/)

16


Compressibility of fluids Compressibility of a fluid is a measure of how easy a fluid volume can be changed under pressure. Compressibility is characterized with the Bulk Modulus of Elasticity (Bulk Modulus or Modulus of Elasticity) EV. Modulus of Elasticity shows the resistance of the fluid to compression and is defined as:

Ev = −

dp ⎡ N ⎤ dV / V ⎢⎣ m 2 ⎥⎦

2.7

Where: dp is differential change in pressure needed to create a differential change in volume dV; V is the initial volume of the fluid; ∆V/V is specific volume. Because the specific volume is dimensionless, Modulus of Elasticity has the same units as pressure. The negative sign shows that an increase in pressure will cause a decrease in volume. In SI units Ev is given as N/m² (Pa). In BG (English) units it is given as lb/in² (psi). Some values of Ev are given in Appendix B. In the case of using hydraulic oil, the value of ∆V/V is very small (large Ev). For this reason, for the engineering applications we accept that fluids are incompressible and disregard the compressibility factor. Large values for the bulk modulus indicate that the fluid needs a great amount of pressure to make a small change in the volume. In other words, the bigger the number is the bigger resistance to compression the fluid has. Modulus of Elasticity can alternatively be expressed as

Ev = −

dp ρ / dρ

Where: dρ is differential change in density of the fluid; ρ is initial density of the fluid.

2.8


17

For most engineering applications we consider the fluids as incompressible. In doing so, we always have to keep in mind compressibility factor when designing or redesigning a system. In any hydraulic system, we have to look at not only rigidity of the fluid but also rigidity of the whole system. Bulk Modulus of the fluid is one of the main factors that determine the rigidity of the system. There are a number of cases when compressibility must be considered. • •

•

Compressing and decompressing large fluid volumes in hydraulic actuators such as piston cylinders. Presence of air in the fluid. Presence of air decreases fluid Bulk Modulus, which in turn increases compressibility of the whole system. Contents of 1% insoluble air can reduce Ev with 40%. Presence of air in the fluid usually is caused by improperly designed reservoir, incorrect selection of hydraulic components or damaged suction line. Use of an accumulator in the system.

For lift truck hydraulic systems compressibility is considered a negative characteristic because it reduces the rigidity of the system. Volume reduction as a result of compressibility of hydraulic oil is approximately 1% for every 15 MPa (2000 psi) pressure. Fig. 2.2 shows the relationship between Bulk Modulus E v and the temperature for two types of fluid.

Fig. 2.2

18


Reynolds Number and Types of Flow Fluid flow can be laminar, turbulent or a mixture of both. The factor that determines which type of flow is present is the ratio of inertia forces (vsρ) to viscous forces (µ/L) within the fluid. This ratio is expressed by the nondimensional Reynolds Number:

ρVL µ

Re =

2.9

Where: V is velocity characteristic L is lineal characteristic µ is the dynamic (absolute) fluid viscosity ρ is fluid density When the flow is in a pipe with a circular cross-section, the lineal characteristic L is equal to the pipe diameter D. Then the equation can be written as:

Re = ρ

VD

µ

We can also express the equation with the kinematic viscosity ν =

Re =

VD

ν

2.10

µ ρ 2.11

This number is named after Osborne Reynolds (1842-1912), who proposed it in 1883. Laminar flow is characterized with smooth flow and parallel layers. It occurs when the viscous forces are dominant (low Re number). Turbulent flow is characterized with turbulent behavior and whirlpools in flow and it occurs when the inertial forces are dominant (high Re number). For Reynolds Numbers up to 2000, the flow is laminar. Above Reynolds Number of 4000, the flow is completely turbulent. Between Re 2000 and


19

4000, the flow is transitional (between laminar and turbulent) and it has elements of both flow types. For flows within circular pipes the critical Reynolds number is generally accepted to be 2320.

Parameters Pressure Pressure is the normal force per unit area at a given point within the fluid. For most engineering problems we assume that the fluid moves as a rigid body (dealing with fluid at rest) therefore there is no shearing stress in it. So, the only forces acting on the fluid are pressure and weight. This allows us to obtain relatively simple solutions to most engineering problems. Pressure distribution (for incompressible fluids) is called hydrostatic distribution.

p1 = γh + p2

2.12

Where: h = z1 – z2 is the vertical distance from a point with pressure p1 to a point with pressure p2. This distance is called pressure head and it is interpreted as the height of a column of fluid of specific weight γ required to give a pressure difference (p1 - p2). If we have a surface exposed to the atmospheric pressure it is convenient to use a point on this surface as reference point 2. Thus, we let: p2=p0. In SI, unit pressure is expressed as Pa (Pascal), where: 1Pa=1N/m². In some cases we use the unit bar (1bar = 0.1 MPa). In BG, units are lb/ft² or lb/in² (psi). The relationship between the metric and the English systems is: 1 bar = 14.5 psi

20


In mobile truck hydraulic systems, positive displacement rotary pumps are used to create pressure. A disadvantage of using these type pumps is that they create pressure and flow pulsations in the discharge port. Pressure variation in a gear pump outlet is explained in Chapter 3, Hydraulic Pumps. Pressure measurement Pressure at a certain point measured relative to the local atmospheric pressure is called gage pressure. Absolute pressure, on the other hand, is measured relative to the perfect vacuum (absolute zero). Absolute pressure is always positive while the gage pressure can be either positive or negative. A negative gage pressure is also referred as a vacuum. Hydraulic systems used in the industrial trucks are classified according to the maximum pressure they are designed for: • • • •

Low pressure system- up to 5 MPa (< 50 bar) Medium pressure system- from 5 to 15 MPa (50 – 150 bar) Normal high pressure system- from 15 to 25 MPa (150 –250 bar) High pressure system- from 25 to 40 MPa (250 – 400 bar)

Flow and flow rate Flow is the motion of the fluid molecules from one point to another. Since the observation of all molecules is almost impossible, we are describing the flow as motion of part of the fluid, called small volume (or unit volume). Small volume contains numerous molecules. Flow is created when a new fluid is pushed into a fluid conductor (pre-filled pipe or hose). The molecules of the new volume push against fluid molecules already in the conductor and displace them. Displaced molecules move by pushing their neighbours and so on. So, the ejected fluid volume from the conductor at the opposite end will be the same as the one entered. The movement of fluid molecules causes a pressure wave traveling at the speed of sound (about 1400 m/s). The speed of sound in fluids is: c=

Ev

ρ

Where: Ev is the modulus of elasticity (Pa)

2.13


21

ρ is the density (kg/m3) For example, the speed of sound in hydraulic fluid (viscosity grade 32) is: c = (1.7x 109 / 870)1/2 = 1398 (m/s) The density values are given in Appendix B. When calculating the parameters of the hydraulic hydrostatic systems we assume that the velocity, v, at a given point in space does not vary with time dv/dt = 0. Such flow is called: steady flow. In a system with a steady flow, rapid closure or opening of a hydraulic component can cause unsteady effects, which have to be considered when a hydraulic system is designed. For example the “water hammer” affect, which results in loud banging of the pipes or tubers. There are three types of flow rate: • Volumetric flow rate, Q Volumetric flow rate is the unit volume flow per unit time passing through an observation cross section.

Unit _ volume V ⎡ m 3 ⎤ Q= = Unit _ time t ⎢⎣ s ⎥⎦

2.14

In SI units flow rate can be expressed either in cubic meters per minute [m3/min] or litters per minute [l/min]. In BG units the flow rate is expressed in gallons per minute [gpm]. In systems working with incompressible fluids we use volumetric flow rate in the calculations. In our further calculations, we are going to use exclusively this type flow rate. • Mass flow rate, q Mass flow rate is the unit mass per unit time

q=

Unit _ mass m ⎡ kg ⎤ = Unit _ time t ⎢⎣ s ⎥⎦

2.15

22


It can also be defined as:

q = ρQ, [kg / s ]

In BG units, mass flow rate is expressed as [slug/sec] or [slug/min]. • Weight flow rate, G Weight flow rate is the unit gravitational force per unit time

GQ =

Unit _ force F ⎡ N ⎤ = Unit _ time t ⎢⎣ s ⎥⎦

2.16

It can be defined as:

GQ = ρgQ, [ N / s ]

2.17

In BG, weight mass flow rate is expressed as [lb/sec] and [lb/min]. An example of flow rate distribution after the pump is shown at Fig. 2.3. The deviation in the flow rate is defined as:

δ = Where,

Fig. 2.3

Q − Qmin ∆Q × 100 = max × 100 _[%] Qm Qm

Qm

is the flow rate mean value.


23

Fluid Velocity Fluid velocity, in pipes is:

v=

Q ⎡m⎤ A ⎢⎣ s ⎥⎦

2.18

Where, Q [m³/s] is the volumetric flow rate passing through a cross section with area A [m2]. Designers must always consider velocity of flow through the pipes and hoses and maintain the flow velocity within recommended limits. Exceeding maximum flow rate limits may cause turbulence in the flow and reduce the efficiency of the system. The recommended flow velocities are shown in Chapter 3.14 (Hydraulic Connectors).

Work and Power Work, as we know from the course of Mechanics, is defined as force (F) acting through a distance (x).

W = Fx [Nm]

2.19

In hydraulics, we also apply force, F, to move a fluid volume at distance x. The force is equal to the pressure applied on a surface area.

F = pA [N ]

2.20

If we replace the force in the equation 2.19, work can be expressed as:

W = pAx [Nm]

2.21

24


If we further replace V = Ax [m³], we receive the formula most commonly used for solving fluid power problems.

W = pV [Nm]

2.22

Where, V is the fluid volume. In SI units work is expressed in Newton meters [Nm] or in Joules ( 1J = 1Nm ) Power is work per unit time

P=

W ⎡ Nm ⎤ t ⎢⎣ s ⎥⎦

2.23

In the SI units power is expressed in Watts [W], where: 1W = 1 Nm/s = 1 J/s. If we replace work with pressure multiplied by volume (equation 2.22), the equation (2.23) can be expressed as:

P=

pV = pQ [W ] t

2.24

Where:

Q=

V m3 [ ] is the flow rate; t s

p [Pa]

is the pressure.

The most convenient form of this formula for calculating the input power on the pump shaft is:

P=

pQ [kW ] 60η P

Where the units are:

2.25


⎡ l ⎤ Q ⎢ ⎣ min ⎥⎦

is flow rate in the pump outlet

p [MPa ]

is pressure at the pump outlet

25

η P = 0.8 − 0.9 is pump’s overall efficiency Power, in Hydrostatics, is transmitted on the bases of applying pressure on the fluid or by the fluid. First, the pump transmits energy to the fluid, and then the fluid transmits it to the actuators. Energy is the capacity to do work and it is expressed in the same units as work. We know that energy cannot be created or lost. In other words, we cannot get something without giving up something else. We can only transfer energy from one form to another and from one point to other. In mobile hydraulic systems, the fluid transfers energy from one location (the hydraulic pump) to another location (linear or rotary actuator). We put energy into the system and get energy out of the system, but there are always losses of energy due to friction, heat loss, etc. So, we can never get out more energy than we put in. Energy that we lose to friction is not lost to the universe; it is simply transformed to heat. Drag and pressure loss Drag is a force (in a direction opposite to the flow) due to the shear forces along the fluid layers. As we know, any fluid moving inside hydraulic lines (tubes or hoses) experience drag. Total drag is a function of the magnitude of the shear stress, τ, and the orientation of the surface on which it acts. Pressure loss is the energy that hydraulic fluid loses to overcome the friction between the moving fluid layers inside the hydraulic lines (pipes, tubes or hoses). The pressure loss is quantified as a pressure drop. Pressure drop is influenced by a number of factors such as: fluid velocity through the hydraulic components and connectors, fluid viscosity, hydraulic line inside wall roughness, etc.

26


Lineal pressure loss Lineal pressure loss is the pressure loss of laminar flow (with Re<2320) moving along the straight sections of the pipes. For laminated flow the pressure loss (pressure drop) due to friction is calculated with the D'ArcyWeisbach equation:

∆p L = λ

l ρ 2 (v ) d 2

2.26

Where:

l is the length of the pipe; d is the diameter of the pipe; v = Q/A is average flow velocity in the pipe; λ [lambda] is the coefficient of lineal flow resistance.

λ=n

64 , for round cross sections n=1 Re

2.27

In many cases, pressure drop (∆pL) for different lengths can be determined faster graphically by using nomograms. There are two type nomograms for determining: 1) in straight pipes and 2) in flexible hose. Local pressure loss Local pressure loss is a result of turbulence in the fluid when the flow changes its direction and velocity. This turbulence occurs inside hydraulic fittings. Local resistance occurs in the hydraulic fittings and it is a result of a change in the flow speed and direction. The pressure drop is calculated with the formula:

∆ pT = ζ

ρ 2

(v )2

2.28


27

Where: v = Q/A is the flow velocity at the outlet of the component; ζ [Zeta] is the coefficient of local flow resistance. Zeta depends on the geometrical shape, cross section and surface roughness of the local restrictor. Approximate values of Zeta are given in Table 2.1, Appendix D.

Hydraulic Shock A Hydraulic Shock is also called: “water hammer”. It is caused by quick closure of the hydraulic component causing pressure increases in the pressure side of the closing element. When the free flow is closed the kinetic energy of the moving fluid is transformed to potential energy, which in turn creates a pressure wave (shock wave). In order to absorb shock waves due to valve closure we use flexible hydraulic hoses as hydraulic lines. In the full power brake systems where hydraulic lines are metal tubing and a brake valve is used to redirect fluid to the wheel cylinders, the shock waves can be absorbed by an accumulator.

Hydraulic Lock One of the most common causes for failures in plunger type valves is excessive frictional force between the plunger and the housing. Frictional force (Fr) is due to uneven pressure distribution in valve clearances (fig. 2.4a). Different pressures on both sides of the plunger create a force perpendicular to the plunger axis. This force pushes the plunger off its center position against the housing increasing friction between internal surfaces. Friction force higher than the control force causes seizing of the plunger. This failure is called hydraulic lock. Valve designers add balancing grooves to equalize the pressure distribution around the plunger circumference (fig. 2.4b).

28


p1 Fr

pa

pb p2

[p]

[p]

pa

pa

pb

p2

pb

p1 x

Fig. 2.4

pm

p=p2-p1

y a)

distance

x

y b)

Obliteration It has been determined experimentally that flow rates through very small openings can gradually diminish and become zero. This phenomenon is called obliteration. It is caused by the adhesion forces between metal surface and the fluid which results in the buildup of layers of molecules on the surface. Adhesion force is an interaction at an atomic level and depends on the chemical composition of the fluid. Experiments show that obliteration exists in openings smaller than 0.01 mm and causes both surfaces to stick together plugging the opening. When the opening is plugged, the plunger is seized. This condition appears in plunger type hydraulic components with small internal clearances. To eliminate the stickiness and seizure of the valve, the plunger is subjected to vibrations with frequency higher than 30 Hz. The high frequency input to the valve is called dither signal.


29

Stiction The term stiction is created by combining the words stick and friction. Stiction occurs when the static friction force is higher than the moving force. It measures the spool resistance to initial motion.

Cavitation Cavitation in fluids is a process of formation and collapse of air or vapour bubbles. This leads to micro jets of oil pounding and eroding adjacent surfaces. Cavitation occurs when the absolute pressure of the fluid becomes close to zero. Cavitation also occurs when the pressure drop is enough that at a given temperature the air in the fluid starts to evaporate. In this case we say that the pressure becomes equal to the vapor tension of the fluid. When cavitation is formed at the suction of the pump, several things happen all at once. • • • •

The system experiences a loss in capacity The system can no longer build the same head (pressure) The efficiency drops The cavities or bubbles will collapse when they pass into the higher regions of pressure causing noise, vibration, and damage to many of the components.

The five basic reasons that form cavitation are: • • • • •

Vaporization Air ingestion Internal recirculation Flow turbulence Vane Passing Syndrome

Cavitation can have several root causes related to system and component design issues or related to service. 1. Tank design issues. Whirlpools in the tank churn the air into the oil or simply don't allow air to be released from the oil. This can be caused by turbulence in the returned fluid, low fluid level, reservoir that is not deep enough, lack of proper baffling, etc.

30


2. Suction-line leaks. Leaks between the tank and the pump can introduce air into the system. Often this is associated with the shaft seal at the pump that allows air to leak in. 3. Suction-line restriction. Sometimes suction lines are too long, too narrow or they are plugged (e.g., a plugged suction strainer). 4. Water vapor. When hot oils become contaminated with water, superheated seam will form vapor bubbles in the oil. 5. Insufficient head. Depending on oil viscosity and suction line conditions, the pump must be located at a sufficiently low elevation to enable oil to flow steadily from the tank to the inlet port of the pump. 6. Air release problems. As oils age and become contaminated, its air release properties become impaired. This means that once air bubbles are formed they stay locked into the oil and do not detrain out of the oil in the reservoir. Moisture contamination and oxidation are the main originators of this problem. ASTM D3427 is a test for air release properties. 7. High viscosity. When fluid temperature in the reservoir is too low, the viscosity may be too high to enable proper oil flow in the suction line and into the pump. Any other cause of high fluid viscosity can lead to the same problem.

The Bernoulli Equation The Bernoulli equation is a statement that the total pressure (pT) along a streamline remains constant (fig. 2.5). The assumption is that the fluid is incompressible and steady. Therefore, if the equation is applied for gases there will be an error built into it.

p+

1 ρϑ 2 + γz = pT = const. 2

2.29

First term p is the static pressure Second term

1 ρϑ 2 is 2

the dynamic pressure. The dynamic pressure is

the kinetic energy of the particle. Third term

γz = ρgz

is the weight of the fluid


31

The most popular engineering application of the above equation is when the equation is applied between two points on a steam line.

p1 +

1 1 ρϑ1 2 + γz1 = p 2 + ρϑ2 2 + γz 2 2 2

2.30

Fig. 2.5 The Bernoulli equation was formulated it in 1738 by the Dutch born mathematician and physicist Daniel Bernoulli (1700-1782).

The Torricelli Equation The Torricelli equation can be derived from the Bernoulli equation when the equation 2.29 is applied to a stream in a vessel with one free surface and one outlet nozzle (fig. 2.6)

Fig. 2.6

32

From,


p1 +

1 1 ρϑ1 2 + γz1 = p 2 + ρϑ 2 2 + γz 2 2 2

ϑ1

at the surface is very small therefore, ϑ1 becomes negligibly small and it can be ignored. Pressures p1 and p2 are equal to zero because they are equal to the atmospheric pressure. 2

Then, the equation can be simplified to

γ ( z1 − z 2 ) =

1 ρϑ2 2 2

When we replace the specific weight γ

= ρg , we receive

ϑ2 = 2 gh

2.31

This equation is called Torricelli's Theorem. It is named after the scientist and mathematician Evangelista Torricelli who in 1843 proved that the flow of liquid through an opening is proportional to the square root of the height of the liquid. Torricelli equation 2.26 can be used to find the flow rate

Q = A 2 gh

Q = ϑA 2.32


33

Chapter 3

Hydraulic Components Hydraulic components can be grouped according to their function: I. Valves Hydraulic valves are grouped into three general categories: pressure controls, flow controls and directional controls. Some valves can have multiple functions and can fall into more than one category. The most important valve characteristics are flow and pressure drop in the valve. Flow can be calculated based on the port diameter and the flow velocity. Pressure drop is more difficult to calculate accurately. That is why it is usually determined experimentally by the manufacturer. Based on the construction, the valves can be plunger or cartridge. Cartridge valves are a screw-in type, which offer the designers the potential of incorporating the valves into manifold blocks or the body of other hydraulic components, such as cylinders. II. Actuators This group consists of pumps, motors and cylinders. Actuators convert fluid energy into mechanical energy or vice versa. III. Accessories In this group are: Pressure and vacuum switches, accumulators, filters and connectors. IV. Hydraulic reservoirs There are two main types of reservoirs- open and closed. The hydraulic systems for industrial trucks use open type reservoirs. V. Hydraulic fluid The fluid is the single most important component of the hydraulic system. Its main function is to transmit energy.

34

Chapter 3: Hydraulic Components

In general, all hydraulic components introduce noise, vibration and losses into the system. In order to select them properly we have to understand their design, function and performance as a separate unit and as part of a system. When a component is used, there is always less energy out than energy in. In order to minimize the losses and the component cost, the components have to be sized per system requirements. Over-sizing will increase the component cost while under-sizing will increase the energy losses.

1.

Flow Restrictors (Orifices)

Flow restrictors (orifices) are local restrictions to the flow. It can be adjustable or non-adjustable. It is also called variable and fixed orifice.

Flow restrictor (orifice) fixed

Flow restrictor (orifice) variable Although, all hydraulic restrictors create some degree of turbulence, they can be lineal or turbulent depending on the type of flow passing though the component. Lineal is when L>d and the flow is predominantly laminar.

Turbulent is when L

35

In hydraulic schematics usually the lineal orifice symbol is used. The hydraulic restrictor is shown as turbulent in cases when we want to show that the pressure drop does not depend on the fluid viscosity. A disadvantage of the lineal restrictor is that its resistance (RL) changes with the temperature change of the fluid. The main function of orifices is to restrict flow and create a pressure drop in the system. They are used to control the actuator’s (motors, cylinders) speed. Although orifices’ main function is to create a pressure drop, they are also called: flow controls. It is important not to confuse them with the pressure compensated flow regulators which can also be called flow controls. Both flow controls have different symbols and the best practice to distinguish them is to look at the component’s symbol. Pressure drop, ∆p , in the orifice is proportional to the flow rate Q . Pressure drop is calculated with the formula 2.26 or 2.28 (Chapter 2) depending on the type of flow. When the relationship between the pressure losses and flow rate is nearly lineal, it can be expressed as:

∆p = RL Q

3.1

Where:

Q

is the flow rate through the restrictor

∆p = p1 − p2 is pressure drop across the restrictor

RL

is orifice resistance

In a system with an orifice, usually there is a varying pressure, p2, after the restrictor is determined by the variation in resistance of the actuator. The main flow restrictor characteristic is based on equations 2.26 and 2.28 and it is called: the flow-pressure drop characteristic. An experimental graph of such a characteristic is shown in Fig. 3.1

36


Orifice Flow-Pressure drop Characteristic 20 18

Pressure drop (MPa)

16 14 12 10 8 6 4 2 0 0

10

20

30

40

50

Flow (l/min)

Fig. 3.1 Orifice flow-pressure characteristics

2.

Pressure Relief Valves

Symbols: Pressure relief valves (also called: relief valves) are pressure control type valves. It is normally closed until it starts to operate. After the pressure is increased, the valve opens and the plunger (poppet or ball) finds a balance position. The balance is created between the pressure on one side and spring force on the other. The valve plunger can have infinite positions between closed and fully opened conditions. The relief valve’s main function is to protect the system against excessive pressure. It is usually


37

installed between the pressure line, after the pump, and the return line before the tank. Relief valves can be adjustable or non-adjustable. Valves are adjusted by changing the spring pre-compression. There are three types of balancing/closing elements: ball, poppet and spool (plunger). The balancing element used determines the type of the valve. Valves can be divided into two groups –1) Ball and Poppet and 2) Spool Valves. • Ball and Poppet Valves Ball and poppet valves are usually used for the construction of cartridge valves. Cartridge valves are less expensive and have higher flow rates than the same physical size spool valves. Ball and poppet valves are less susceptible to fluid contamination because when closed, the valve moving part is held tightly against a seat in the housing. On the other hand, they are more sensitive to flow and pressure irregularity. Their positioning is less balanced than spool valves, which leads to less accurate metering. • Spool Valves Spool valves are easier to control and can move at smaller steps because it is easier to proportionally control the stroke of the spool. On the other hand, they are more expensive and more susceptible to contamination. Spool valves have higher leakage rates than poppet valves.

3.

Check Valves

Symbol: Check valves are unidirectional control valves. They have two positions: ON or OFF. This valve has free-flow (open) and no-flow (close) direction. When the flow pushes the ball (or the poppet) away, the valve opens and permits free flow. Flow in the opposite direction pushes the ball against the seat. The built-up pressure keeps the passage sealed and the flow is blocked. These types of valves are designed to have a very small leakage rate when they are closed. Usually, valve seats are hardened steel which makes them more resistible to scoring from hard contaminants in the fluid.

38


The only difference between the check valve and the ball relief valve is the spring. Check valves have light springs which are used only to return the ball (poppet) to its seat when the flow stops. Because of the light spring, the pressure drop in the valve during operation is very small (about 0.05 to 0.1 MPa). There are three general check valve designs: plunger, poppet and ball design. The check valve has a relatively small effect on system noise, vibration and losses. When the check valve is built into another hydraulic component, the pressure loss from it is included in the total pressure loss of the main component. When we use an in-line check valve, it is acceptable to disregard the pressure loss in it. Therefore, it is very important not to undersize the valve. Undersizing it will increase the pressure drop, leading to inaccuracy in the calculated pressure demand. Check valves can have an internal or external pilot control. Fig. 3.2 shows two valves with pilot ports. outlet

outlet pilot

pilot inlet

inlet

Fig. 3.2 a) pilot-to-open

b) pilot-to-close

Pilot-to-open can be opened by external pressure. When there is no pressure in the pilot port, this valve allows flow only in one direction. When pressure is applied in the pilot, the flow can pass in both directions. The amount of pilot pressure required to open the check valve is:

p PILOT =

p OUTLET + p SPRING , r

Where:

pSPRING

is the pressure on the poppet due to spring force

r is the pilot ratio. It is the ratio of pilot piston area to poppet area.


39

Pilot-to-close also allows flow only in one direction in the absence of pilot pressure. When pilot pressure is applied from the pilot port, it overrides the free flow function and holds the valve closed. This feature is useful to control regenerative flow in a cylinder circuit or in a hydraulic logic circuit. Minimum pilot pressure required to close the valve is:

p PILOT =

p INLET − p SPRING , r

r is the pilot ratio

Two check valves can be combined together.

Fig. 3.3 Dual pilot check valve Dual pilot operated check valve is used for load holding applications or cylinder locking.

4.

Reduction Valves

Symbol: A reduction valve is a pressure control type of valve. Its function is to reduce pressure and maintain a pre-set lower pressure value in the outlet. The valve maintains a constant pressure in the outlet regardless of pressure and flow rate changes in the inlet. This valve is normally open.

40


Principle of operation Flow passes through an opening between a balanced plunger (spool) and housing. Pressure in the outlet is applied under the valve plunger through internal connection. A spring force, acting on the other side, balances the plunger. When the pressure in the outlet increases, the plunger is pushed up and the opening is reduced which, in turn, reduces the flow through the valve. Spring chamfer is connected to the reservoir therefore external drain to the reservoir is required for this valve.

5.

Pressure Compensated Flow Controls

The Pressure Compensating Flow Controls’ function is to regulate flow rate regardless of the system working pressure. These valves are also shortly called: Flow Controls or Flow Regulators. The flow rate is usually used to control an actuator’s speed. The valve can be placed before or after the actuator. Symbols:

Two-way flow control

Three-way flow control Construction of two-way flow controls This valve has two parts: a pressure balanced plunger and an orifice connected in series. It is called two-way because it has two ports. The balanced plunger (between point 1 and 2) controls the opening to maintain a constant pressure drop across the orifice. The flow through the valve is controlled by the orifice (between point 2 and 3) and the pressure drop ∆p = p2 – p3 (Fig. 3.4 a). The valve will maintain a constant flow rate from point 1 to point 2 within a specified pressure range. The valve regulates the flow rate only in one direction from point 1 (valve inlet) to point 3 (valve outlet).


41

Pressure in point 1 must be higher than the pressure in point 3. When the flow is reversed, from outlet to the inlet, the valve simply acts as a flow restrictor.

1 2 3

Fig. 3.4 Two-way flow controls: a) balanced valve before the orifice, b) balanced valve after the orifice. The flow equation through the orifice is:

Q = µA

2

ρ

∆p

3.2

Where: µ is the flow coefficient A is the area of the orifice opening ∆p is the pressure drop in the orifice The only variable in flow equation is the pressure drop (∆p). The purpose of the pressure balanced valve is to maintain a constant ∆p which ensures a constant flow rate (Q) through the orifice. The flow is as a function of the

pressure drop in the valve, Q = f ( ∆p ) which can be obtained experimentally. Flow rate vs. pressure drop relationship determines valve performance and it is called: Flow regulator static characteristic (shown in Fig. 3.5). Because of the nonlinear relationship between the flow and the pressure drop across the valve, the flow rate diverges slightly (the curve is never perfectly horizontal).

42


In the hydraulic systems, the nearly constant flow rate is used to achieve speeds of the hydraulic actuators which are independent from the pressure variations. In mobile forklifts it is mainly used to maintain a constant lowering speed of the load. A hydraulic system using this valve is described in Chapter 5 (Hydraulic Systems for High Lift Trucks). 50

Flow (l/min)

40

30

20

10

0 0

5

10

15

20

Pressure drop, delta p (MPa)

Fig. 3.5 Flow regulator static characteristic

6.

Directional Control Valves

Directional valves control the direction of the flow path. These valves can be classified on the basis of the number of possible ways the fluid can go. Most common types are 2-way, 3-way, 4-way or 5-way valves. Based on the neutral position of the valve plunger, there are three basic valve configurations shown in Fig 3.6 (a, b, c) •

Open center, a), is when the plunger is at a neutral position and the inlet flow is open to all ports.


43

•

Close center, b), is when the plunger is at a neutral position and the flow is blocked. • Tandem center, c), is when the plunger is at a neutral position and the flow is unloaded to the tank. Two other combinations of the first three, float center d) and open-to-three port e), are also shown in Fig 3.6

Fig. 3.6 Directional-control valves have two primary characteristics: 1) number of ports for the fluid and 2) number of positions for the controlling element. Valve ports are the passageway for fluid in or out of the valve. The numbers of positions refer to the number of distinct flow paths a valve can provide. There are three types of spool valve laps (fig. 3.7): zero, positive and negative. Valve lap is the distance the spool travels before valve opening. Valves with large overlaps have less leakage but they have less accurate flow metering.

A

P

B

a) zero lap

A

P

B

b) positive lap (overlap)

A

P

B

c) negative lap

Fig 3.7 Spool laps Port P (pump) is the valve inlet. Ports A and B are valve outlets. Valve is shown in closed centered position. In terms of plunger (spool) positioning, there are two major groups: 1) discrete/ finite positioning 2) infinite/ proportional positioning

44


Discrete positioning Finite positioning is when the plunger is shifted from one discrete position to another. For this reason these type valves are called discrete valves. Plunger shift occurs in an instant, causing the fluid to rapidly accelerate or decelerate. This causes fluid pulsations or in certain conditions it can cause fluid hammer.

Switching time for these valves depends on the size of the coil. Actuation time increases when the coil size and the valve size are increased. For example the switching times of directional valves size 6 (20 l/min nominal flow) with DC (direct current) magnet is about 40 milliseconds while the switching time of size 10 directional valve (80 l/min nominal flow) is about 80 milliseconds. Infinite positioning In these valves, the plunger is shifted proportionally to the input signal. The signal can be mechanical, electrical or hydraulic. The plunger can have infinitive intermediate positions, which makes these types of valves ideal for controlling speed and acceleration or deceleration of the actuators. Infinite positioning is illustrated by adding two extra parallel lines indicating that the plunger can slide inside the valve.

The infinite positioning directional valves can be further classified as: Proportional Valves, Servo Valves and Load Sensing Valves. Proportional Valves Proportional valves provide flow and pressure control proportional to the control input device. The control device can be either mechanical or electrical. When an electric signal is used to control the flow rate, the flow rate changes proportionally with the change of the signal to a solenoid. Inside a spool type valve there is a spool (plunger), which is the only moving part. Changing the electrical current, applied to the coil, changes the magnetic


45

field, which in turn creates a magnetic force on the armature and makes it move. The coil is placed in a metal housing which helps to retain the magnetic field. In most valves, a flat spring is used to resist plunger movement. The spring retains the plunger until the magnetic force on the armature exceeds the spring force. The main reason for the performance variation from one valve to another is the mechanical and geometrical tolerances that occur in the manufacturing process. Solenoid magnetic field can be adjusted so that it compensates for mechanical tolerances. Therefore, this valve can create a consistent relationship between the flow rate and the electrical current to the valve. Proportional valves can be an open-loop or a closed-loop construction. Open-loop valves do not have feedback between the solenoid input and the valve spool or valve output. They have a lower response time than the closed loop valves. Closed-loop valves have an outer loop for spool location feedback. An outer loop can be made by connecting a LVDT sensor to the spool. A LVDT sensor measures small changes (in the range of microns) of spool movement and converts them to electrical signals. Proportional valves can be spool or poppet type. Most of proportional valves are spool type designs because they have better control and metering capabilities. Poppet type proportional valves are less susceptible to fluid contamination. For this reason they are mainly used in systems subject to high contamination. To minimize the leakage from a section with high pressure to a section with low pressure, the plunger type valves are manufactured to have as a small gap as possible between the body and the plunger. The servo valves have 0.001 mm to 0.004 mm internal clearances while the discrete directional valves usually have 0.005 mm to 0.012 mm internal clearances. Servo valves Servo valves have a shorter response time than standard proportional valves. They are always closed-loop valves. There is a mechanical feedback link between the input command and the valve output. Servo valves usually consist of a two-stage spool. The spool position is controlled by two electromagnetic coils- one from each side. Manufacturing tolerances of these valves are in the range of micrometers. The tight tolerance requirements make them expensive to manufacture. Also, the reduced

46


clearances between the valve surfaces make this type of valve susceptible to fluid contamination which can jam the valve. Because of the high cost and the high fluid cleanliness requirements, servo valves are rarely used in mobile hydraulic systems. Servo valves are used in applications where short response time is critical. For this reason they are manufactured with zero laps or near-zero overlaps. Servo valves can have a response time as low as 0.0025 seconds (400 hertz). Where:

1Hz =

1 1sec

In contrast, standard proportional valves have a response in the range of 0.1-0.2 seconds (10 - 5 hertz). Proportional valve selection Proportional valves are selected per maximum flow that must go through the valve. Proportional and servo valves execute their control through a high-pressure drop. The valve's flow rating is usually based on a specific pressure drop. After we select valve size, it is recommended to measure the pressure drop across the valve. If (in our application) the pressure drop is significantly different than the rated pressure drop of the valve, we have probably selected the wrong valve. Usually, engineers end up with an oversized proportional valve. If an oversized valve is selected, the hydraulic actuators are unlikely to get the anticipated proportional performance. In most cases, the valve will open all the way before it is supposed to, providing a different resolution that we seek. When selecting the valve, pressure drop should be used to calculate the flow rate.

Q R = QOUT

∆p B ∆p A

Where, QR = valve's rated flow for our application QOUT = output flow needed for application ∆pB = rated pressure drop of proportional valve ∆p A = actual pressure drop needed for application

3.3


47

It is recommended that designers use this method to check the flow rate of their valve. In most cases, the flow rate they obtain through this method will differ from the flow rate in the catalogue.

Load Sensing Priority Valves Load sensing priority valves are simply called priority valves. They have infinitive positioning. There are two types of priority valves: static and dynamic.

Priority valve with static signal

Priority valve with dynamic signal

Load sensing priority valves are used to split the flow in open loop systems where one branch must have a guaranteed flow supply. This valve senses the flow requirements and provides metered priority flow to this port. The valve has one inlet and two outlets. One of the outlets is for the controlled fluid (CF) and the other one is for the excess fluid (EF). Dynamic load sensing valves have faster responses than static valves. They have a passage between CF and LS lines. This passage supplies a continuous pressurized flow to the LS line even when the line is not used which keeps the valve in a ready-to-respond position. Directional valves can have direct or indirect control. Direct control is applied directly to the valve control element. Indirect control (pilot operated design) is when the input signal controls a small pilot valve which in turn controls the main valve (fig. 3.8). Electrically controlled big valves require big and expensive solenoids. To reduce manufacture cost of these valves, they are controlled indirectly. Small solenoids are used to control the pilot valve which sends pressurized

48


fluid to control the main valve. Some proportional valves with indirect control have a course filter (screen) that protects the pilot stage. If a filter is used, the filter should be replaceable or washable.

a)

Symbol

P

T

b) Detailed symbol Fig. 3.8 Directional manual valve with indirect manual control

7.

Hydraulic Pumps

Symbols: Constant flow pump

Variable flow pump

Pumps are mechanical devices that convert mechanical energy into hydraulic energy. They draw fluid from a reservoir and send it to hydraulic actuators. There are two main types of pumps: positive displacement (vane, piston and gear pump) and non-positive displacement (centrifugal pumps). By definition, positive-displacement (PD) pumps displace a defined quantity of fluid with each revolution of the pumping elements. This is done by trapping fluid between the pumping elements and a stationary housing.


49

Pumping element designs include gears, lobes, rotary pistons, vanes, and screws. Positive displacement (PD) pumps can be either fixed or variable displacement. Fixed displacement pumps have a constant relationship between the flow rate and the drive shaft angular velocity. In variable displacement pumps, the displacement can be changed so that the flow rate can be independent from the drive shaft velocity. Gear pumps have fixed displacement while vane and the piston pumps can be either fixed or variable. Lift truck hydraulic systems use only PD type pumps such as: vane pump, piston (axial and radial) pump and gear pump. Systems with pressure up to 25 MPa usually have a gear or vane type pumps. While high pressure systems 25 to 40 MPa (3600 – 6000 psi) require using piston pumps.

Gear pumps Gear pumps can have external or internal gear meshing. External pumps have one or more sets of two spur gears while the internal have one or more sets of spur and ring gear. In fork lift application external gear pump is more popular because of the bigger selection and the lower cost. External gear pump construction A gear pump (Fig. 3.9) has a body in which there are two hardened steel gears intermeshing together. One of the gears is a drive gear and the other one is a driven gear. The drive gear is mounted on a shaft, which extends outside and is connected to a motor. Meshing gears create two chambersthe first is the inlet (suction port) the second is the outlet (pressure port). Rotating gears take fluid from the suction port, drive it around the gears and push it into the pressure outlet. The highest quality gear pumps have zero backlash gear meshing. Pumps with zero backlash meshing have high efficiency and low noise. Pump main parameters specified by the manufacturers are: flow rating (maximum and minimum shaft speeds), maximum pressure rating, and the type of mounting.

50


Fig. 3.9 Gear pump- external gear meshing Pump body is subject to cyclic loads due to pressurizing and depressurizing during pump operation. For this reason fatigue strength is a main requirement for the body design. A gear pump body can be made from die-cast aluminum alloy, aluminum alloy bar stock, cast steel or cast iron (ductile iron). Ductile iron is usually less expensive and has better noise and vibration dissipation than aluminum and steel however, it has the worst heat dissipation of the three. Ductile iron and cast steel have identical yield and tensile strengths. Ductile Irons have small volumetric changes and retain their strength at high temperatures due to the stability of the microstructure. For high temperature applications, ductile iron alloys with silicon and molybdenum are used. Silicon content of 4% to 6% provides the best combination of heat resistance and mechanical properties. Pump bodies from cast steel and ductile iron are designed to the yield point of the material. Aluminum pumps are designed for minimum deflection (high rigidity) because aluminum reaches its endurance limit sooner than steel and it has smaller plastic range and less tolerance to overload and deflection. Aluminum body pumps are good for low-temperature applications because at low temperatures (below 30°C), aluminum has a little change in properties (yield, tensile and impact strength). Cast iron pumps are preferred in wide temperature rage applications because the cast iron and the steel gears have similar expansion properties which, reduces the thermal distortion and the internal leakages.


51

Although a gear pump is tolerant to system contamination, the manufacturer must specify the acceptable contamination level. Direction of rotation of the shaft must be shown on the pump body. The pump can be bi-directional so that it delivers flow from either port. By using such pump, we can eliminate the directional valve in the system. Bi-directional pumps require a drive motor which is able to rotate in both directions. Flow is proportional to the shaft speed therefore the relationship between the shaft speed and the outlet flow is linear. A gear pump can include a built-in relief valve, check valve or both. The relief valve can be internal, fig. 3.10a, (the fluid is returned to pump inlet) or it can have external relief port, fig. 3.10b, (the fluid is returned to reservoir).

2

2 3

1 Fig. 3.10a

1 Fig. 3.10b

Internal reliefs can be used in systems in which the pump works on request. For steering systems in which the flow goes over relief 50% of the time, internal reliefs are not recommended. They heat up the oil and the pump and can cause leaks through pump inlet seals. When a system have an internal relief valve, it is important to keep the fluid temperature in its operating range. Overheating of the fluid can affect the relief valves performance. Pump delivery (flow rate in litters per minute)

QP =

dPn ⎡ L ⎤ ηV ⎢ 1000 ⎣ min ⎥⎦

3.4

52


Where: dP [cm3/rev] is the pump displacement. The displacement is a measure of the pump size and is given by the manufacturer; n [rev/min] is the shaft input rotational speed ηV is pump volumetric efficiency. In the BG units, flow rate is given in gallons per minute (gpm):

QP =

dPn ηV [gpm] 231

⎡ in 3 ⎤ dP ⎢ ⎥ ⎣ rev. ⎦

3.5

is the pump displacement (in cubic inches) per

revolution. Gear pump flow rate can be given at 1000 rpm by the manufacturer. In fixed-displacement pumps, the flow rate can be changed only by changing the drive shaft rotational speed. These pumps are used in open type systems in which the flow after each work cycle is returned to the reservoir. Variable displacement pumps are mainly used in closed systems (systems where the pump continues to operate at a stand-by in a neutral position) Torque on the pump shaft

T=

P

ω

η m [ Nm]

Where: P is the hydraulic power in Watts (Nm/s) ηm is pump mechanical efficiency w is the shaft angular velocity in (rad/s) In mechanical formulas, the shaft speed is expressed in radians per second w (rad/s) The angular velocity w (rad/s) can be converted to rotational speed, n (rev/s): w = 2πn


T=

P ηm 2πn

T=

pQ pd n η m = P ηm 2πn 2πn

T =

pd P ηm ( Nm) 2π

53

3.6

Slip Slip is a leakage of fluid from the pressure outlet back to the inlet. Slip increases with increasing pressure and wear. Increasing slip is referred to as a loss of efficiency. Slip can be reduced by constructing the pump for pressure and wear compensation. Pump Efficiency Overall efficiency is:

ηO =

Output _ Power = ηV η M Input _ Power

3.7

It is determined as the ratio between the hydraulic power at the pump outlet and mechanical power at the driving shaft at nominal pressure, rotational speed, and fluid viscosity (rated power). The overall efficiency has two components: volumetric and mechanical.

ηV =

Actual _ Flowrate Rated _ Flowrate

is the volumetric efficiency.

The actual flow rate is the flow at the pump output when the pump is working under load. It will vary at different pressures. Rated flow rate is the theoretical flow at the pump outlet without volumetric losses. Volumetric efficiency range is: ηv = 0.90 - 0.97. If volumetric efficiency is not known, for initial calculations we can take the average values for gear pumps: ηv = 0.90 (low speeds ≈ 1000 rev/min) and ηv = 0.97 (high speeds ≈ 3000 rev/min)

54


ηM =

Drive _ Saft _ Power is the mechanical efficiency. Input _ Shaft _ Power

It is a result of lost power due to friction in the bearings and between the meshing gears.

η M = 0.90 − 0.93

The gear pumps overall efficiency is in the range of 82% to 88% depending on the pressure and rotational speed. An example of pump overall efficiency at different pressures is shown in Fig 3.11. 100

Overall Efficiency [%]

90 80 70 60 50 40 30 20 10 0 0

50

100

150

200

250

300

Pressure [bar]

Fig. 3.11 A disadvantage of gear pumps is that they create pressure and flow ripples (pulses) in the discharge port. Pumps are one of the biggest sources of noise and vibrations in the hydraulic system. Every time the fluid between two teeth is pushed out of the pump, a peak in the pressure appears. A typical pressure distribution at a discharge port is shown in Fig. 3.12. Pulsations (δ) can be expressed as the ratio of pulsations amplitude (∆p) to an average value (p):

δP =

∆p

Where:

p1

× 100 [%] , ∆p is peak-to-peak amplitude.


55

p [MPa]

p

p3 p2 p1

t [s] Fig. 3.12 Where:

P1 is maximum continuous pressure P2 is maximum intermittent pressure P3 is maximum peak pressure

Pump pressure pulsations travel along the hydraulic lines at the speed of sound (about 1400 m/s in hydraulic fluid) until it is affected by a change in diameter or direction. Therefore, pulsation amplitude depends on the hydraulic lines (length and diameter) and fittings (type and size) in the system. Although the direction valve, after the pump, smoothes the flow and the pressure peaks, pulsations created by the gear pumps travel through the system to the hydraulic actuators. Internal gear pumps have smaller pressure pulses than external pumps because the spur and ring gear set have more teeth meshing than two external spur gears. Gear pumps are mainly used in systems with normalhigh pressures (from 15 to 25 MPa). For higher efficiency, they should be driven at speed close to their rated maximum because internal leakage is smaller at higher speeds. At low speeds, gear pumps have reduced lubrication between side plates and gears. Pump manufacturers always specify the minimum rotating speed. Intermittent pressure is used for selecting lift pumps that work intermittently. Continuous pressure is used for selecting steering pumps that have to run continuously.

56


Vane Pumps There are two types vane pumps: balanced and unbalanced. In the balanced design the rotor and the sliding ring surface are coaxial. In the unbalanced design they are not. Construction Similar to the gear pumps, a driving shaft coming from primary power source drives the vane pump. Inside the pump, the driving shaft is connected to a slotted rotor that is placed eccentrically from the center of the circular opening of a casting housing. Vanes placed in the rotor slots slide in and out. Centrifugal force causes them to slide out and the contour of the cavity pushes them back in. Tips of the vanes slide on the inside pump surface and seal the passage between the suction and the pressure ports. The vanes push fluid from the inlet to the outlet through the gap between the housing and the rotor. Vane pumps have higher efficiency than gear pumps because of less fluid leakage from the pressure outlet back to the inlet. They have less slip (smaller volumetric losses). Also, the efficiency remains constant over time. As the vane tips wear the slip remains the same because the centrifugal force always keeps the vanes in contact with the housing surface. Pump housing is made from the same materials as gear pumps. Mechanical efficiency is a result of the friction in the bearings and the friction between the cam contour and vane tips.

Fig. 3.13 Unbalanced vane pump


57

Disadvantages of vane pumps compared to gear pumps are: 1) A high efficiency into a narrow pressure range and 2) pumps are more susceptive to fluid contamination. When the contamination increases, their volumetric efficiency decreases.

Rotary Piston Pumps Rotary piston pumps have a rotational driving shaft. These pumps have some advantages over gear and vane pumps. They are used in systems with higher flow and pressure demands. Features: • • • •

High power-to-size ratio. We can get more hydraulic power out of a piston pump than we can from the same size gear pump. High pressures: some pumps can maintain pressure up to 70 MPa. Low power consumption at stand-by. High overall efficiency: for most pumps it is about 96%.

Construction There are two main types of rotary piston pumps: radial and axial. In the radial type, pistons are placed in a cylinder block. Pistons move radially in and out. The cylinder block (rotor) is located inside a fixed housing (stator) and is rotated by a drive shaft. The rotor centerline is offset from the stator centerline. The amount of offset determines piston stroke and pump displacement. In axial pumps (fig. 3.14), pistons move axially. They are placed into a cylinder block which is rotated by the drive shaft. The piston ends are depressed against a tilted disk (swash plate). The angle of the disk causes cylinders to move axially. If the disk is perpendicular to the axis of rotation (zero angle), pistons will not be compressed and there will be no flow through the pump. The disk can have different angles. When the disc is tilted to one side of the neutral, flow goes in one direction. When it is tilted to the other side, flow direction is reversed. When the disc angle is fixed the pump has fixed displacement. In pumps with variable displacement, the disk angle is controlled by a yoke. The yoke can have mechanical, electric or hydraulic control. At the released position of the yoke, the disk is

58


returned to neutral (zero) position and the pump stops delivering flow. When the yoke is hydraulically controlled, a pressure compensator maintains constant output pressure at different flow rates. Such pump is called: pressure compensated pump.

a

Fig. 3.14 Axial rotary piston pump Pressure compensated pumps are used mainly on internal combustion (IC) engine trucks because the engines speed is controlled by the truck’s speed requirements and not hydraulic system requirements. Piston pumps can have an integrated digital electronic control for pressure limiting, load sensing and anti-stall speed sensing. They can be constructed in a way to be able to operate either as a pump or as a motor. Design considerations when selecting a pump • • • • • • •

If a pump works at speeds outside the manufacturer specification, it has reduced lubrication and reduced life; When pump speed is higher than specified maximum, the inlet pressure is reduced (increased vacuum) which can cause cavitation. In this case, a floated pump design is recommended; The risk of cavitation in the pump inlet increases when: suction line is too long, pump is placed too high above reservoir or suction filter is undersized; Pressure above rated increases the torques on the pump drive shaft When a higher flow rate is required, it is recommended to use a larger pump instead of increasing the driveshaft rotational speed; At speeds higher than rated pumps have reduced pressure rating. Alignment tolerances of the pump and the motor shafts should be within limits specified by pump manufacturer


8.

59

Hydraulic Motors

Symbol:

Hydraulic motor, constant displacement

Similar to the PD pumps, there are three types of PD hydraulic motors: gear, vane and piston. They are powered by pressurized hydraulic fluid. Hydraulic motor’s main function is to convert hydraulic energy to mechanical energy and to transfer rotational kinetic energy to mechanical devices. There are two types of motors: • •

Fix displacement- constant flow at constant rotational speed Variable displacement- variable flow

Motors have the same pressure rating as pumps. Some of them are available with optional built-on holding (multi-disk) brake, dynamic (drum) brakes, flushing valves or speed sensors. Motor selection is based on two characteristics: motor shaft speed and shaft torque. The shaft torque is a product of motor volumetric displacement and pressure drop in the motor. Required shaft speed is determined by the flow rate and the motor displacement. Shaft rotational speed usually is determined by the performance of the other components in the truck and is given when calculating the hydraulic system. Motor displacement (dM) versus required output shaft torque (TM) is published by the manufacturer in charts or tables. For a required shaft torque, the motor displacement (dM) can be calculated with the formula:

dM

62.83TM = ( ∆p )η M

⎡ cm 3 ⎤ ,⎢ ⎥ ⎣ rev ⎦

3.8

Where: TM

is motor shaft torque (Nm)

∆p

is pressure drop across motor ports (bar)

ηM = 0.90 – 0.95

is motor mechanical efficiency

60


The flow rate, QM, is calculated from the formula:

QM =

d M nM 1000ηV

⎡ L ⎤ ,⎢ ⎣ min ⎥⎦

3.9

Where: dM

is motor displacement (cm3/rev)

nM

is motor shaft speed (rev/min)

ηV = 0.90 – 0.97 is motor volumetric efficiency Motor efficiency is identical to the pump efficiency.

9.

Hydraulic Cylinders

Cylinders convert fluid energy into mechanical energy in the form of linear motion and force. • Single-acting cylinders In elevating systems, either single-acting telescopic or single-acting ram type cylinders are used. Single-acting cylinders accept pressure fluid only on one side of the piston. Volume on the other side can either be vented to the atmosphere or connected to the tank. The return line to the tank is for collecting eventual leaks through the seal. These types of cylinders are called single-acting because the work done by the fluid is acting only in one direction- for lifting. Weight of the load and the mast does the work in the opposite direction. Ram type and telescopic type cylinders, shown in Fig. 3.15, are designed for applications where long strokes are required. Telescopic cylinders (fig. 3.15b) have two or more stages. The outside body is called: main cylinder and the smallest stage is called: plunger. When fully extended, their stroke exceeds the length of the cylinder at fully retracted condition. Collapsed length of a typical telescopic cylinder is about 20% to 40% of its extended length. Telescopic cylinders usually extend from the largest to the smallest stage. This means that the largest stage inside the main cylinder will start to extend first. There are telescopic cylinders that are designed to have all stages extend at the same time. This construction


61

provides constant speed and constant push force throughout the extracting and retracting.

INPUT

Fig. 3.15 a) Ram type, single acting

INPUT

b) Telescopic, single acting

Hydraulic cylinders have four main components: cylinder, piston, piston rod and seals. Cylinders are made from steel or ductile iron tube. Pistons are made from: alloy steel or high-tensile strength ductile iron. Piston rod is usually made from chrome plated and polished steel alloy. Inside surfaces of cylinders are coated and polished. Coating can be electrodeposited chrome or nitriding. Electro-deposition is the process of producing coating by putting a negative charge on the part and dipping it into a chemical solution that contains the coating metal. The thickness of the electrodeposited layer is determined by the duration of the process. Nitriding is a surface hardening process that introduces nitrogen into the surface. There are three nitriting technologies: ion, gas and salt. At the end of the telescopic cylinders there are at least one seal and a wiper. The seal is usually a U-cup lip seal. This seal relies on hydraulic pressure to press the seal lips against the rod and seal groove. It is better to use a seal with a pre-energized lip (by spring or O ring) in order to avoid leaks when the cylinder is unloaded and pressure is absent. The wiper prevents external contaminants entering the cylinder.

62


The linear force generated by a hydraulic cylinder is a product of input pressure and effective area.

F = ∆pAηCYL

3.10

Where:

∆p = p1 − p2

is the pressure (p2 = 0 for vented cylinders)

p1

is the input pressure

A [m 2 ]

is the piston area

ηCYL = 0.95 − 0.98 is

the mechanical efficiency of the lift

cylinder. When the cylinder diameter is increased, the natural frequency (the stiffness) of the system increases allowing the motion controller (the direction valve) to manage faster acceleration and deceleration, which in turn yields higher system performance. Roughly, the actuator (cylinder) natural frequency should be 3 - 4 times higher than the frequency of the motion controller. •

Double-acting, single rod cylinder A1

A2

p1 F p2

Fig. 3.16 Double-acting cylinders can have single piston rod (Fig. 3.16) or double piston rod. They are used for auxiliary operations such as tilting, side shift and reach/retract operations. These cylinders accept pressurized fluid on both sides of the piston. When the cylinder has a single-piston rod,


63

extension force is greater than retraction force (if equal pressures are used) because the area of the piston side (A1) is greater than the area of the face of the rod end side (A2). The force equation acting on the piston rod (shown in Fig. 3.16) is:

F = ( p1 A1 − p2 A2 )η CYL [ N ]

3.11

Where:

A1 = A2 =

πD 2 4

[m ] 2

π (D 2 − d 2 ) 4

is the larger area

[m ] 2

is the smaller area

At the same flow rate, the cylinder retracts faster than it extends. Faster retraction is a result of the smaller volume at the rod end side.

υE =

Q ⎡m⎤ , A1 ⎢⎣ s ⎥⎦

Extension speed

υR =

Q A2

⎡m⎤ ,⎢ ⎥ ⎣s⎦

Retraction speed

Q

is the flow rate

The relationships between the piston diameter (D) and the piston rod diameter (d) of double-acting cylinders are standardized and are given in table 3.1, where: φ = A1/ A2

64


D (mm)

d (mm) φ = 1.25

φ = 1.6

14 18 22 28 32 36 40 45 50 56 63 70 80 90

20 25 32 40 45 50 56 63 70 80 90 100 110 126

32 40 50 63 70 80 90 100 110 125 140 160 180 200 Table 3.1

10. Pressure Sensors

Symbol: There are two main types of sensors: 1) pressure switch and 2) electronic pressure sensor (pressure transducer). Pressure switch is an ON-OFF type switch which is controlled by pressure. Usually, it has a single-pole double-throw actuator. As the pressure increases, it pushes a piston against a retainer which compresses a spring and closes/opens the switch contact tips. The force on the piston


65

F = pAη m

counteracts the spring force FS = kx . Preloading of the spring is adjustable which allows the switch to be adjusted for different pressures. The manufacturer always specifies maximum recommended system pressure and the type of fluid with which the switch can work. Electronic pressure sensors provide output signal proportional to the pressure input. The output is an analog signal which can be voltage (0 to 10 V range) or current (4 to 20 mA range). Analog signal is a continuous electrical signal which varies analogously to a non-electric input signal (pressure variations).

Pressure sensor with voltage output signal

Pressure sensor with current output signal The accuracy of the electronic pressure sensors is about +/- 1% of the rated pressure and their life is about 10 million load cycles. Pressure spikes that exceed the nominal system pressure can cause failure of pressure sensors. Snubbers are installed before the sensors to protect them. Some sensors have built-in snubbers. A typical snubber consists of an orifice which suppresses and absorbs pressure spikes. Orifice diameter determines the level of dampening. The diameter selection is based on: • • •

System pressure: higher pressure requires smaller diameter (greater dampening) Fluid viscosity: higher viscosity requires smaller diameter Amplitude of pressure ripples: higher amplitude requires greater dampening. The amplitude of the spikes depends on the sensor location in the system.

Recommendations for pressure sensors selection: • • •

Select transducer with pressure range at least 20% above maximum working system pressure Avoid installing sensors after fast closing valves Install proper size snubber before transducer inlet

66


11. Hydraulic Accumulators

Symbol: Hydraulic accumulators are used as a source of power at the system’s request. They store non-compressible fluids under pressure. The pressure is created by an external force. There are three types of accumulators with regard to the external force: • • •

Weight loaded Spring loaded Gas loaded

Mobile equipment hydraulic systems usually use gas loaded accumulators. This type of accumulator uses the compressibility property of gases to store potential energy. When the fluid pressure in the system is higher than the gas pressure in the accumulator, fluid enters the accumulator compressing the gas. When the hydraulic pressure in the system drops, the gas trying to equalize the pressure expands and forces the fluid out of the accumulator. The most common type of accumulator employed in modern hydraulic systems is the nitrogen gas loaded type. An important characteristic of the nitrogen gas is that when the gas is compressed the its pressure increases exponentially. This results in storing more energy as the fluid pressure increases. Also, the nitrogen gas is safe for the user because it is nonflammable. Accumulators have several functions in hydraulic systems: • • • • • •

Storage and source of energy Pressure shock damper- at this application, the accumulator has to be installed as close as possible to the source of pulsation. Holding device Leakage compensator (used in full-power brake systems) Thermal expansion compensator Power source


67

Sizing Accumulators The following case is when the accumulator is used as a power source. The equation describing the gas in three different positions is (fig. 3.12): n

n

p1V1 = p 2V2 = p3V3

n

3.12

Where:

p1 = gas pressure in the pre-charged accumulator (initial) V1 = gas volume in the pre-charged accumulator (initial) p 2 = gas pressure in the charged accumulator (at maximum system pressure) V2 = gas volume in the charged accumulator p3 = gas pressure in the discharged accumulator (final) at minimum system pressure V3 = gas volume in the discharged accumulator (final) When charging and discharging take place slowly, there is enough time for the heat to dissipate. Then, we have an ISOTHERMAL process. In this case n=1. When charging and discharging occurs quickly without heat transfer we have an ADIABATIC process. Then n is equal to the ratio between the specific heat of the gas at a constant volume and its specific heat at a constant pressure. For nitrogen gas this ratio is equal to 1.4 (n=1.4). In bladder-type accumulators which are used in industrial truck hydraulic systems, we have adiabatic conditions.

68


Fig. 3.17 Determining the size of the accumulator Size is determined by the relationship between the initial gas volume, V1, and the needed fluid volume, VX: Equation 3.12 can be written as:

V1 ( p1 )1 / n = V2 ( p 2 )1 / n = V3 ( p3 )1 / n If we take the second two members we can express V2:

V2 ( p 2 )1 / n = V3 ( p3 )1 / n Also,

V3 = V2 + ∆V

Then

V2 ( p2 )1/ n = (V2 + V X )( p3 )1/ n = V2 ( p3 )1/ n + ∆V ( p3 )1/ n

3.13


69

V2 ( p2 )1/ n − V2 ( p3 )1 / n = ∆V ( p3 )1 / n V2 = ∆V

( p 3 )1 / n ( p 2 )1 / n − ( p3 )1 / n

From equation 3.13 we express the relationship between V1 and V2:

V1 ( p1 )1 / n = V2 ( p 2 )1 / n

( p 2 )1 / n V1 = V2 ( p1 )1 / n Replace V2:

( p 3 )1 / n ( p 2 )1 / n V1 = ∆V × ( p2 )1 / n − ( p3 )1 / n ( p1 )1 / n Divide both sides by ( p 2 )1 / n to determine the size of the accumulator.

V1 = ∆V

⎛ p3 ⎞ ⎜⎜ ⎟⎟ ⎝ p1 ⎠

1/ n

⎛p ⎞ 1 − ⎜⎜ 3 ⎟⎟ ⎝ p2 ⎠

1/ n

[m 3 ]

3.14

Pre-set accumulator pressure in the bladder is about 90% of the line pressure p3

p1 ≈ 0.9 × p3 It is recommended that the discharge pressure in the accumulator is: P2 < 3 P3

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12. Hydraulic filters Symbols:

a) filter (one direction flow)

b) filter for reverse flow

The devices for cleaning fluid can be classified as: filters, strainers and magnets. Filters are devices whose primary function is the retention of insoluble contaminants from the fluid. Strainers are course filters. Their filtration ranges from 50 to 300 microns (most strainers are 125 microns). Magnets function is to attract and remove iron from the fluid. They must be placed where they will attract most particles- between the return and the suction lines. Magnets can be installed inside the reservoir plug (magnetic plugs). Hydraulic filters are classified by pressure rating as: • • •

Low-pressure filters High-pressure filters Medium-pressure filters

There are five types of filters classified on the basis of their location and function in the system. • Suction filters These are low pressure filters located in the suction line before the pump. They are usually placed inside the reservoir. Suction or return filters can be combined with a breather when they are installed at the inlet or outlet of the reservoir. In order to determine what type of filter to use, we have to consider the requirements of each component and the hydraulic system as a system.


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• Pressure filter These types of filters that are placed in the pressure line either immediately after the pump or before a component with high fluid clearness requirements. They are either high or medium pressure. • Return filters Return filters are installed in the return line. They can be placed outside or inside the reservoir. It is recommended return filters have a bypass valve which protects the filter element during cold start and pressure spikes. Return filter which do not have a bypass valve mast be rated for pressures higher that the system pressure. They can be subject to high pressure when they are clogged. • Reverse flow filters Reverse flow filters are used when the flow direction is reversed in the lines. This filter has to retain contaminants in one direction and prevent returning the contaminants into the system when the flow is reversed. This is achieved by having two parallel lines and check valves in each line. In lift truck applications, reverse flow filters are used in hydrostatic transmissions and energy recovery systems with reversible pump/motor. • Air filters Air filters are placed on the reservoir and they are usually combined with a breather cap or a dip stick for fluid level indication. Breather cap air filters range from 2 to 40 microns filtration. Suction or return filters can be combined with a breather when they are installed at the inlet or outlet of the reservoir. In order to determine what type of filter to use, we have to consider the requirements of each component and the hydraulic system as a system. The most debated design to consider is the use of suction low-pressure filters. In general, they are recommended for systems working in areas with high air contamination. Suction filters main advantages and disadvantages are as follows. Advantages: • • •

Dissipate any degree of turbulence left over from the returned oil. Protect the whole system at the front and catch all contaminants entering the oil through the air breather. In order to minimize the pressure losses in the suction line, the suction filter must have a bigger filtering area which increases the

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cost of the filter. Schematically it can be presented as a couple of filters working in parallel. Disadvantages: •

•

•

Increase the resistance in the suction line, which may cause inefficient suction by the pump. For this reason they are mainly used in floated pump designs (the pump is located below fluid level of the reservoir). When the filter gets plugged, the flow bypasses the filter through the check valve and all benefits of the filtration are lost. It is recommended to change the filter during preventive maintenance after a number of hours and every time repair work is done on the system. To reduce change time, the filter should be located at easy to access location. Take extra space in the reservoir when the filter is places inside.

Filters can also be classified as: replaceable and permanent. Foreign particles collected in the permanent filter can be washed away and the filter can be reused. Filter efficiency Filter efficiency is based on three parameters: • Beta Rating • Dirt holding capacity • Pressure drop across the filter element In the past, there were two efficiency ratings: nominal and absolute. Currently there is no world standard describing nominal and absolute. If we know only the nominal or absolute rating of a filter in microns, this really does not quantify its retention efficiency at that size. If a filter is nominally rated, 6-micron, it does not describe how many particles, 6-micron and larger, are being retained by the filter. Also, there is a difference between absolute and nominal rating. Absolute rating has 98.7% to 99.5% retention efficiency, while nominal rating has a 50% to 98% efficiency. Most filter manufacturers specify absolute rating at a 99.5% efficiency. A better approach is to express the performance of filters with a Beta Rating Number (ISO 16889). A Beta rating (6 microns for example) is a measure of the number of particles greater than 6 microns upstream (before the filter) divided by the number of particles greater than 6 microns downstream (after the filter). Beta rating of 100 means that for every 100


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particles larger than six microns entering the filter, one passes through. The data given in the ISO standard or by the manufacturers are based only on specific test conditions prescribed by the standard. Actual field conditions may vary considerably. Another measure is filter retention efficiency (R). It can be calculated as: R = (1-1/BETA) x 100 [%] The relations between Beta and R can be given in table 3.2. Beta Number

Retention Efficiency

ß

R [%]

2

50.0

10 20

90.0 95.0

50 75

98.0 98.7

100 200

99.0 99.5

Table 3.2 Fine media filters can remove some additives from the fluid. Many types of lubricants have de-foaming additives. These are suspended semisolid particles in the 5-10 micron range and are filterable. Efficient 1-micron filters can potentially remove sulphur and phosphorus additives that are not dissolved as well as suspended solid anti-scuff additives. The best way to determine if the filter works properly is to do an upstream and downstream particle count. It is a good practice to combine the filter with a differential pressure switch. The switch signals the need for a filter change before the filter is fully clogged and the contamination passes through the check valve. Changing it on time prevents system component damage and premature replacement of partially used filter elements. Most filter manufacturers can determine how changes in the fluid viscosity at operating temperatures affect the flow rate

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Dirt holding capacity Dirt holding is a characteristic showing the retention capacity of the filter element until it is clogged. It is the weight (in grams) of the retained contamination at determined pressure drops across the filter. Pressure losses in the filter There is always a pressure drop across the filter. For this reason, in most lift truck hydraulic systems, filters are located in the return line. The pressure drop in the return line is desirable because it creates backpressure, which is necessary when lowering the load. Also, this location permits the use of low-pressure filters. Pressure drop depends on filter characteristics (area, particle block size) and system characteristics (fluid viscosity and flow rate). High pressure drop filters are not suitable for high viscosity oils. Filter manufacturers' publish PQ curves of their products to show the relationship between flow vs. pressure drop and viscosity vs. pressure drop. In some applications the cylinder return stroke is faster than the power stroke (lowering speed is faster than lifting speed). Therefore, the return flow rate is greater than the pump flow rate. If we have a filter in the return line, it has to be sized for the greater flow rate. If the return filter is sized for the pump flow, then the filter is undersized. Undersized filters will fail to properly clean the fluid and the system will build up heat. Filter Bypass Valves at Low Temps We want 100 percent flow to pass through the filter at all times. When the fluid temperature is low, its viscosity increases making it more difficult to pass through the filter element. The fluid takes the path of least resistance and goes through the bypass check valve or through ruptured sections in the element. The filter bypass condition happens usually for the first few minutes after a cold start till the fluid warms up. Loop filtration Off-line loop filtration is the most cost-effective method for cleaning the fluid. Filter efficiency and dirt-holding capacity are at a maximum because the fluid flow is steady. Off-line filtration can be installed permanently or it can be a portable unit. The main benefit is that loop filters have a lower cost per weight of captured dirt than in-line filters. Despite the benefits, loop


75

filtration is not used in the forklift truck hydraulic systems because it requires extra space and adds cost to the system. Start-up filtration Although the hydraulic fluid and components are cleaned to the recommended ISO levels, there is always contamination introduced during handling and assembly. For this reason, systems with high cleanliness requirements must be flushed after assembly and before start-up. During flushing, the fluid circulates through an outside filter (usually remote filter carts). The remote filter must have higher retention efficiency than the system filters. Also, the system has to be flushed with a flow rate higher than the maximum for the system. It is not recommended to flush the system with compressed air from common factory lines. Air will blow away bigger particles but it can bring in moisture and smaller contaminants. Air flushing can be used before fluid flushing. The flow Designers must always consider all system parameters when selecting the filter. When the fluid has cyclic behavior, the increase in flow can dramatically alter the efficiency of the element during the flow pitch. If a cylinder cycles at a fast rate, the filter is subjected to high flow and pressure fluctuations. This can cause leakage over the bypass check valve. The water Water absorption filters usually have an element that combines a particle removal media with water absorption material. Most water absorption media are sensitive to flow rate. Their efficiency is higher when flow velocity is low. It is common for these filters to be used on off-line circuits where the flow rate can be controlled. Filter media The most commonly used materials for filter media are: wire mesh, cellulose (wood pulp) and glass fiber. Wire mesh and cellulose are mainly used for filters with lower retaining efficiency. Most high efficient filters have high efficient glass fiber media.

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Filter electrostatic charging The friction between the flow and the filter element creates electrostatic charge. The charge generation depends on the type of fluid and filter media material. The generation increases at higher flow rate and fluid viscosity. This static electricity is collected in the filter housing and discharged to the truck frame. The electrostatic discharge has to be controlled because it increases the fluid thermal degradation and varnishing and it has negative effect on the performance of the electronic components. Filter failure monitoring There are two main of failure states: clogged and torn filter media. Filter condition is observed by measuring the pressure difference before and after the filter. Differential pressure switch, shown in Fig. 3.18, can be used to measure the pressure at both ends of the filter. The pressure drop across the filter should be within a certain limit. Pressure drop above the upper limit means that the filter is clogged. Pressure drop below the lower limit means that the filtering media is torn. When the pressure drop across the filter exceeds the spring force, the switch is turned on and it gives a reminding signal for filter change.

Fig. 3.18 (Filter with differential pressure switch) For suction and return filters, a pressure switch which is connected only to one point can also be used. The pressure drop in the filter can be obtained from sensing one side because pressure at the other side is approximately zero. When the filter has a bypass check valve, the pressure switch setting has to be below the setting of the bypass valve. Disadvantage of using a switch sensing only one side is that it works in narrow fluid viscosity range. If the fluid viscosity goes outside of this range the switch will give a false signal. Mechanical clogged filter indicators are also used. They are constructed as pop-up switches.


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Latest filter technologies have made possible filters with filtration as low as one micron with minimum pressure loss across the filter. In this design, the fluid is cleaned of ferrous particles by a magnet. This filter does not have a filtration barrier and it has a smaller pressure drop. In hydraulic systems for outdoor mobile equipment this filter is not used because most of the contamination comes from the environment where the contaminants are non-ferrous particles.

13. Hydraulic Reservoirs

Symbols:

a) Open

b) Pressurized

The reservoir has two main functions: to store the hydraulic fluid and to keep the fluid within defined working temperature limits. There are two types of reservoirs: open (vented) and pressurized (non-vented). In this section, we are going to discuss only the vented type since it is primary used in the mobile industrial equipment. A reservoir must be designed to meet the system requirements. These requirements are: • Proper size The reservoir has to have the smallest volume that holds the necessary fluid for the system. The most economical size has to be calculated based on the system’s requirements. The formulas used for determining the reservoir volume are based on calculation of the cooling surface and are given further in this section. • Good sealing and filtration against contamination of the fluid Air has to enter and exit the cylinder through a breather containing air filter. The capture efficiency of the air filter has to be the same as or higher than the capture efficiency of the main oil filter. Breathers without filters do not prevent contamination from the air. In dusty environments, the air entering

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the reservoir should be filtered. In moist environments, desiccant breathers are used to prevent the ingression of moisture. •

High degree of heat exchange between the fluid and surrounding air through the reservoir walls Reservoir walls have to ensure good heat exchange between the fluid inside and the surrounding air. Free air circulation around the reservoir must be guaranteed. Reservoirs without enough surface area to dissipate heat from natural circulation and reservoirs of systems with short work cycles should have a partition (baffle) which separates the reservoir into two parts. A baffle is a separation plate dividing the reservoir into two sections: return and suction. Baffles cause returned flow to circulate around the outer wall before it can get to the suction line. The benefit of this circulation is better heat exchange and turbulence dissipation of the return flow. Usually the lower corners of the baffle are cut off. The area of the cut offs must be larger than the cross-section area of the inlet line. •

Dissipation of air bubbles in the fluid along the path from the return to the suction pipe. Turbulent flow induces air bubbles in the fluid. These bubbles can enter the suction pipe and cause cavitation damage in the system. To minimize this risk, suction and return lines should be as far apart from each other as possible. If there is a baffle in the reservoir, the suction and return ports should be on opposite sides of the baffle. The suction line (or suction filter) has to be a minimum of 20 mm above the bottom of the reservoir in order to avoid taking contaminants along with the fluid. Also, return and suction lines have to be submerged a minimum of 30 mm below the lowest fluid level. Another way to dissipate the turbulence from the return line is to use a diffuser (spreader) at the return line so that the return flow passes through a spreader. Instead of diffuser, we can use two return filters connected in parallel or a strainer. The strainer will cause backpressure, which is desirable for some functions of the lift truck hydraulic systems. It is recommended to make reservoirs deep and narrow instead of shallow and wide in order to minimize the vortex effects in them. • Ability for fluid level observation. Maximum and minimum levels have to be shown on the fluid level indicator or marked on a deep stick. • Guaranteed atmospheric pressure. The maximum fluid volume has to be approximately 10% smaller than the reservoir volume. This will ensure a constant atmospheric pressure in the reservoir during level changes.


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• Easy serviceability. Main requirements include: accessible location of the hoses connected to the reservoir, ability for quick discharge and refill and easy fluid level check. • Flushing connection ports. Flushing is a procedure for cleaning the reservoir from contamination using turbulent flow. Flushing fluid must be compatible with the fluid used in the hydraulic system. The reservoir can be equipped with special flushing ports. There are three cases when having flushing ports is recommended: 1) The reservoir has to be flushed regularly. 2) Fluid oxidizes rapidly as a result of overheating. 3) There is a quick fluid contamination from outside.

Calculating the size of the reservoir Conventional and quick consideration for reservoir sizing is generally accepted as being three times the minute flow (l/min) of the pump. Reservoir volume (in litters) = 3 x Pump flow [litters per minute] However, using this criterion often results in selecting a reservoir that is oversized and larger than the available space. For high lift fork trucks, it is also necessary to consider the maximum fluid volume that the system needs. Maximum volume is the total volume of all cylinders when they are fully extended and the volume of all connectors in the system. The fluid volume in the reservoir is then increased to ensure that a minimum fluid level in the reservoir is maintained. Reservoir volume (in litters) = (1.1 to 1.5) x Maximum system volume It is a common practice to use one hydraulic system for different height elevating systems (masts). Every time the mast is replaced with another one with a higher fluid volume requirement, the reservoir volume has to be increased to meet the new requirements. When designing a reservoir we can select different shapes, materials, wall thickness and locations. In order to select the most economical design, we must calculate the reservoir surface required to dissipate the heat and limit

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the fluid temperature up to an acceptable maximum value TMAX. Maximum temperature is a design requirement which has to be determined before designing the reservoir. In hydraulic systems which use air cooling for the reservoir, the cooling surface is:

S=

QL 3600 PL = [m 2 ] k (∆TMAX ) k (TMAX − TO )

3.15

Where:

Q L = 3600PL is the amount of heat as a result of losses in the hydraulic system for a period of 1 hour

PL =

pQ (1 − η S ) is the lost power in the system 60η P

ηP

is the pump efficiency

ηS

is the system efficiency without the pump

∆TMAX = TMAX − TO is the temperature difference between the fluid maximum temperature and the surrounding air temperature

k=

1

α1

+

1 1

α2

+

δ λ

is a coefficient of heat transfer of the reservoir

walls

α1

is coefficient of heat transfer from the fluid to the walls

α2

is coefficient of heat transfer from the walls to surrounding air

δ

is the wall thickness

λ

is coefficient of heat conductivity of the wall

Coefficient of heat transfer, k, can be accepted as:


81

⎡ kJ ⎤ k = 25 ⎢ 2 ⎥ when the reservoir has poor air circulation around the ⎣ m hK ⎦ reservoir.

⎡ kJ ⎤ k = 45 ⎢ 2 ⎥ when the reservoir has good air circulation around the ⎣ m hK ⎦ walls.

⎡ kJ ⎤ k = 90 ⎢ 2 ⎥ when the reservoir walls are subject to air cooling by a ⎣ m hK ⎦ fan. Using the described requirements and method for calculating the optimum reservoir surface, the fluid in the reservoir can be reduced significantly by achieving the desired temperature and turbulence dissipations. Different materials have a different coefficient of heat transfer, k. For example, a stainless steel reservoir has half of the heat transfer capability of a carbon steel reservoir. Reservoir testing In order to guarantee that all design requirements for the reservoir are met, it is recommended that we measure some of the main parameters of the system. It is a good practice to build one prototype reservoir with plexiglas windows and a temperature gauge in it. This will allow monitoring of fluid turbulence as well as the change of temperature during the required duty cycle of operation. Achieving acceptable maximum oil temperature is the best measure of good reservoir design. It means that the reservoir can properly dissipate the heat generated by the system. Negative effect of moisture in the reservoirs Hydraulic oil can absorb water. In general, if the reservoir is designed wellthere is no turbulence- and the flow cycle is low, the water will settle out on the bottom. But, in lift truck applications, the hydraulic systems usually have short fluid cycles for lifting, reaching and side shifting. Therefore, in the reservoir design, we have to focus on the prevention of water entering the reservoir.

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The headspace of many tanks maintains a continuous moist fog. This can lead to a buildup of water in the oil, especially in cold storage applications where the truck has to go in and out of a freezer. The moisture condenses on the reservoir ceiling and walls and then drips into the fluid. Moisture signs can be found either by looking for oxidations (rust) on reservoir ceiling or by touching the inside ceiling with our fingers looking for moisture or rust. Materials •

• • • •

The most steel reservoirs are made from mild steel sheet metal. Reservoirs with volume less than 100 litters are usually made from 1.5-2 mm thick sheets. Reservoirs are painted only from the outside. The inside area above the oil can be a source of contamination. When carbon steel reservoirs are used for fluid storage, their internal walls must be coated against corrosion. Aluminized steel and stainless steel reservoir are designed to eliminate the contamination concern. Aluminized steel is mild steel coated with aluminum-silicon alloy. Stainless steel reservoirs are the most expensive to manufacture. For this reason they are mainly used for storing hydraulic fluids. Aluminum alloy reservoirs are usually die-casting. Plastic reservoirs are usually made from polypropylene, polyethylene or nylon. Resins used for reservoirs are usually rated for temperature ranges from - 40 ºC to +120 ºC. Design of plastic reservoir must consider coefficient of heat transfer (k) and the thermal deformation of the material. Polyethylene resins Crosslinked, low density Crosslinked, high density Linear, low density Linear, high density

Temperature

Change of material properties

115 °C (239 °F)

Becomes soft and deforms

Hydraulic fluid

170 °C (338 °F)

Becomes soft and deforms

Pressurized hot fluids

100 °C (212 °F)

Starts to melt

Hot water

130 °C (266 °F)

Starts to melt

Hydraulic fluid and chemicals

Table 3.3 Comparison between four polyethylene resins

Storage application


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The most widely used material for plastic reservoirs is polyethylene (lowdensity crosslinked or high-density linear). The name crosslinked shows that the polyethylene structure consists of bonded and linked together carbon chains at molecular level. Crosslinking of the molecules changes the polymer properties from thermo-plastic to thermo-elastic. Thermoelasticity makes the material more resistant to ruptures and cracks. At higher temperatures, thermo-elastic polyethylene will soften and become more flexible. This softness allows impact energy to be absorbed easier by the molecule chains. Crosslinking also improves the thermal properties of the polymer.

14. Hydraulic Lines, Fittings and Couplings Hydraulic lines used in pressurized hydraulic systems are: hoses, tubes and fittings. Hoses are rated by their inside diameter (ID) while tubes are rated by their outside diameter (OD) and wall thickness. Tube fitting sizes are based on the tubes outside diameter and thread size. Hose fittings are based on the hose size (hose ID) at one end and the thread size at the other. In general, there are four types of tubes. Hydraulic (fluid line) tubing is a metal tube. The most commonly used metals for hydraulic tubing are: low-carbon steel, stainless steel, alloy steel, copper, aluminum and copper-nickel alloy. Steel tubing is normalized to achieve softness necessary for ease of bending and flaring. Low-carbon steel tubing has one or more corrosion resistant coatings. Hydraulic tubing is produced to OD and wall thickness dimensions. Pneumatic tubing can be made from metal or non-metal materials: copper, aluminium, nylon, polyethylene or PVC. Nylon tubing is the most popular because of its flexibility and low cost. Mechanical tubing is steel tubing for structural applications. It can have round, rectangular or a square shape. Cylinders tubing is mechanical tubing with a finished inside diameter (ID) manufactured to be ready to use for hydraulic cylinders. The ID has tight

84


tolerances and surface finishing requirements. The cylinder tubing is produced to OD and ID dimensions. Mechanical and cylinder tubing will not be discussed in this book because they are not used for connectors. When the term “tubing” or “tube” are further used, they will refer only to hydraulic fluid line tubing. Hydraulic line selection is an important part of the design process. There are two important parameters- fluid velocity and maximum pressuretaken under account when selecting the size of the connectors. Few other parameters: temperature, construction and fluid type must also be taken into account. In addition, the effect on the system stiffness must be considered when selecting the type of connector. Elasticity of the flexible hoses reduces the dynamic stiffness of the system. Pressure All hydraulic fluid lines have rated pressure. Lines (for medium, normal and high pressure systems) are rated with a safety factor of four. That means that their safety (burst) pressure is four times the working pressure. Pressure rating varies in accordance with the different materials used. For example: copper tubing rating is about 10 MPa, steel tube rating is more than 25 MPa. Fluid velocity Fluid line size is selecting on the basis of fluid velocity inside the connector. The fluid velocity affects the pressure drop and the fluid Reynolds number. In general, pressure losses decrease when the diameter increases. In highpressure systems big diameters are not economical because of the increased cost of the connectors, fittings and all other components. Bigger pressure hoses are very stiff, require bigger bending radiuses and take more space. Also, it is more difficult to deal with larger fluid volume because it would require a larger reservoir. On the other hand, exceeding recommended maximum velocity of the fluid may cause turbulence in the flow. Small suction lines can increase the vacuum and create cavitation in the pump inlet. The cavitation damages the pump surfaces and causes noise and mechanical damages.


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The optimum recommended velocities in the connectors are: •

Suction line v= 0.5 to 1.5 m/s

•

Pressure lines Maximum Pressure

•

Recommended velocity

p < 50 bar (5MPa)

v = 4 m/s

p = 50 to 100 bar

v = 4 to 5 m/s

p = 100 to 200 bar

v = 5 to 6 m/s

p > 200 bar

v = 6 to 7 m/s

Return lines v= 2 to 3 m/s

It has been estimated that 80% of hose failures are caused by external physical damage to the hose. In order to increase durability of the hose, the hose manufactures use special hybrid compounds for the cover material instead of standard rubber covers. This more durable compound increases the service life, lowers the maintenance, and eliminates the need for costly hose protectors such as guards and sleeves. The pressure ratings for hoses are provided by the manufacturers. Design principles • • • • • • •

Maximum pick pressure in the system must be below maximum rated pressure for the hose. Hose burst pressure should not be used for hose selection. The burst pressure is only for safety purpose. Hoses used for the suction line must be able to withstand vacuum and pressure; Hose routing must ensure minimum length and number of bends, avoid twisting and avoid external heat sources; Hose sizes (inside diameters) are selected according to recommended velocity; Hoses have to be chemically compatible with the fluid in the system; Hoses must be protected from rubbing against metal edges or hard objects, snagging, cutting, pulling, bending and twisting; Use proper end fitting;

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• •

Hose cleanliness has to match the system’s cleanliness level. After a hose is cut to size and fittings are assembled, the hose assembly should be cleaned and plugged. It is recommended that mating connectors (tubing, end hose connectors and the fittings) are made from the same material.

Hose ratings and constructions Hose pressure ratings used in the industrial truck applications are: High-pressure hose is used for pressures of 20 bar to 400 bar. These type hose has three components: internal tube, steel mesh layers (four or six) and cover. The tube is made from synthetic rubber. Steel mesh is made from high-tensile still wire. The cover could be thermoplastic, synthetic rubber or fabric mesh. Medium-pressure hoses are used for pressures of 20 bar to 200 bar. This type hose has the same three components. The only difference is that the steel reinforcement consists of one layer still mesh. Low-pressure (suction or return) hoses are used for pressures from 0.6 bar to 20 bar. These hoses should be rigid enough to resist compression when the absolute pressures are below 1 bar (the hoses experience vacuum). This type has synthetic non-rigid reinforcement. High, medium and low-pressure hoses are rated for temperature range from -40 to 100 ºC. Connector restraining Restraining is achieved by clamping the connectors to rigid surfaces of the machine. Proper clamping increases the life of the tube and hose assemblies. Tube clamps can have non-metal (rubber or nylon) dampeners, which protects the tube from vibrations and mechanical shocks. Hose clamps are usually metal. The purpose is to protect the hose from twisting and rubbing against other surfaces or edges. Tubes and hoses can be attached to each other by a floating clamp. This clamp ensures that connectors are not rubbing against each other by creating a space between them. The clamp is called “floating” because it is not fastened to a rigid structure.


87

Fittings and Couplings The couplings in mobile forklifts have to meet the requirements for high flow and pressure, high and low temperatures and different fluids. Coupling connection has to be able to handle pressure pulsations and spikes higher than the rated pressure. These spikes appear when the work tool (the forks) hit a hard object or cylinder piston hits the hard stop at the end of the stroke. Tube fittings and adapters are cycle tested for endurance to 133% of the work pressure and with minimum torque. If the fittings are over tightened, their cycling endurance is reduced. Most commonly used types of couplings are: 37° flare (JIC) threaded couplings is one of the most commonly used. It has good performance and low cost. Both mating parts have coned surfaces which fit against each other to form a seal. Because the sealing is ensured by metal-to-metal contact, this type is torque sensitive. If they are under-tightened, there is leak path between the sealing surfaces. Overtightening can damage the treads and also causes leaks. It is recommended in extreme low or high temperature applications. O-ring coupling is a very popular threaded connection. There are O-ring face seal, O-ring ISO and O-ring SAE. The difference between ISO and SAE are: at machined ports ends, ISO uses metric thread, tube size and hut hex for metric wrenches while SAE has inch thread, tubes and hex. The male component has a rubber O-ring which is compressed against the machined seat in the female component. This rubber-to-metal seal is less torque sensitive. It is recommended for high-vibration and high-pressure applications. The O-ring coupling is not recommended for extreme temperature conditions because rubber ages and deteriorates at very low or high temperatures. Flanged connection (four-bolt split flange type) is designed to avoid the use of threaded connections. The fitting has two parts: a flanged head fitting and a flange clamp (one-piece or sprit version). Flange connection has high pressure capability because of the larger sealing surface and the lower torque requirements. Flange heads and clamps are made of plated or coated carbon steel. Flanged connections are designed to maximum inside diameter of the hole. The sealing is achieved by O-ring or by seal-plate which are compressed between the mating surfaces. This connection is

88


recommended for systems with high pressure and dynamic (pulsating) pressures. Pipe thread couplings are sealed by metal-to-metal contact. They have tapered male threads. The sealing is achieved by tightening both parts to certain torque. The thread surfaces deform and flatten against each other blocking the flow passage. For better sealing, a sealant material is added to the mating surfaces. This connection is used for low cycle pressure and high static (non-pulsating) pressure applications. The main limitations of it are: 1) shaped fittings (elbows) cannot be oriented in desired positions; 2) high risk for leaks at dynamic applications; 3) risk of thread cracking at extreme temperatures because of material expansion or contraction. Cutting face coupling is used for tube fittings. It has four components: body, cutting ring, collar and a nut. The cutting ring has two cutting edges. The ring slides on the tube. The ring geometry is design so that when the nut is tighten, the front edge cuts into tube surface first and the second edge follows. The collar ensures that all forces are equally distributed. Quick acting (connect/disconnect) is threadless coupling. Both components are pushed together by hand. A clip ring locks the male and the female parts. These couplings are usually used to ease the operator when additional attachments need to be attached to the truck. Flare tube end is used in tube connectors for low and medium (up to 200 bar) pressure systems. It has a flared tube which is clamped against a flared nose fitting. Clamping is achieved by a tube nut screwed to fitting.

15. Manifold blocks (Manifolds) Manifold is an aluminum, cast iron (ductile iron) or steel block which has ports for cartridge type valves and cavities providing interconnecting links between the hydraulic components. Manifolds reduce the number of external connections, which reduce the chances of leakage. Its design requires many considerations including available space, pressures, flows,


89

duty cycles, valves, port types, size, locations and materials. For systems with pressure up to 25 MPa, manifolds are made from steel or aluminium. For pressures above 25 MPa, it is more economical to have steel manifolds. After machining, the manifold blocks must be washed, deburred and surface finished against oxidation. Burrs and small particles are removed thermally. Thermal deburring is a process where the manifolds are put in a chamber filled with gas which is ignited. The temperature goes up to 3500 ºC for about 20 milliseconds. Manifolds have machined valve cavities specified by the valve manufacturers. Valves are screw-in cartridge type. Valves must be tightened to torque values provided by valve manufacturers. Solenoids, of the electrically controlled valves, must be spaced from each other so that their electromagnetic field does not interfere. The magnetic field can affect the performance of neighboring valves. Fittings are tightened according the manifold block material. Recommended torque values for fittings are given in table 3.4.

O-ring straight thread port- ISO 6149

Maximum Torque Values (Nm) Ductile Iron manifold

Aluminum manifold

Steel manifold

M8 x 1 M10 x 1 M12 x 1.5

8 10 15

8 12 25

10 18 30

M14 x 1.5 M16 x 1.5 M18 x 1.5

25 30 35

35 40 45

40 45 50

M20 x 1.5 M22 x 1.5 M27 x 2

45 55 75

55 65 100

70 120 145

100 120 150

120 150 180

180 200 250

M30 x 2 M33 x 2 M38 x 2 Table 3.4

90


Advantages of using manifolds are: • • • •

Design freedom in terms of placement Reduced space requirements Reduce the number of tubes, hoses, and fittings Reduce the assembly time

Limitations of manifold use are: • Manifolds increase the fluid temperature more than conventional plumbed systems because the valves are close to each other and flow paths are short. Therefore there is less heat dissipation. • They are sensitive to large contaminant particles. If the system does not have a suction filter, a manifold inlet filter rated at 20 to 25 µm is recommended.

16. Hydraulic Fluid Hydraulic fluid is the most important component of any hydraulic system. It affects the entire hydraulic system’s performance and the life of its components. As we have already discussed, filters, strainers and magnetic plugs are used to keep the fluid clean. The fluid has multiple functions in the system- it transmits energy from the energy source to the components, cools the system, cleans the system, lubricates components, reduces friction to minimize wear and absorbs pressure ripples. Fluids most important characteristics are: viscosity, seal compatibility, protection against component wear, deposit control, oxidative stability and water separability. Viscosity and effects of viscosity Viscosity is the single most important fluid property. It is the first thing to consider when selecting a hydraulic fluid. Before making a selection we have to determine the start-up viscosity and viscosity at its operating temperature. Another key characteristic in the selection process is the viscosity change relative to the temperature change. If the hydraulic reservoir is undersized, it can not dissipate all built-up heat and makes the fluid work at a higher


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temperature. This fact adds performance requirements to the fluid to maintain small viscosity changes in wider temperature ranges. When the fluid oxidizes, its viscosity increases. This has to be monitored because when the system works in a cold condition the higher viscosity can cause cavitation in the pump during a start. As we already defined viscosity as oil’s most important property, it makes sense to measure viscosity frequently, using on-site test equipment. There are two determinable parameters, absolute and kinematic viscosity. Kinematic viscosity measures the resistance of the fluid to flow and shear under gravity, such as oil flowing through a funnel. Absolute viscosity, on the other hand, determines oil's internal resistance to flow and shear. To visualize absolute viscosity, imagine the force needed to stir oil with a stick. Advantages of higher fluid viscosity are: reduces wear and leakage. Disadvantages of higher fluid viscosity are: increases filter pressure drops and possible filter bypass, reduced response to inputs and cold start sluggishness of the hydraulic system. Lift truck manufacturers should provide recommendations about the appropriate viscosity grade of the fluid for hydraulic systems for their equipment. It is especially important to provide recommendations if the system is subjected to extreme working conditions (advice from the fluid supplier can also be valuable). Viscosity grade classifications are given in Appendix C. Hydraulic systems for mobile forklift equipment use fluids with viscosity ν = (15 to 46) x 10-6 m2/s (viscosity grades 15, 22, 32 and 46). Seal compatibility Incompatibility between seals and fluid can cause seal failure and oil leaks. Seal manufacturers have done a lot of research because they serve a wide range of industries with a broad number of applications. Seal manufacturers can generally predict the chemical compatibility of a fluid and recommend a suitable seal material. Although following their recommendations is a good practice, the best verification is the system test at working conditions.

Component wear protection To achieve high efficiency and long system life, it is important not only to select fluid viscosity but also to select a fluid with anti-wear property. Antiwear property reduces friction between the metal surfaces in the hydraulic

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components. It depends on the fluid base and additives. Before the hydraulic fluids are approved by the equipment manufacturer, they are evaluated based on the system performance. Water Water in oil-based fluids leads to system damage and failure. The worst threat of water contamination is its reaction with additives and the damaging bi-products. Some anti-wear additives in presence of moisture can decompose and transform into a highly corrosive acid (sulfuric acid). When water is absorbed by hydraulic fluid, the fluid puffs up and creates white slime. This slime causes clogging in the filter. Water contamination is described in greater detail in Chapter 8-III (Common Problems). Specific gravity When users want to replace conventional hydraulic oil with another fluid, they have to take into account the specific gravity of the new fluid. If the specific gravity is more, hydraulic pumps cannot the new fluid as easily as the lighter one. For example: Conventional hydraulic oil has a specific gravity of about 0.85, while water glycol is around 1.0 and phosphate ester is 1.1. If a heavier fluid is put in the system and the design engineers don't make special accommodations for heavier fluids, cavitation in the pump will occur which will lead to excessive noise and failure of the pump. One way to solve the problems associated with using a heavier fluid is to put the reservoir higher than the pump. This arrangement is also called "flooded suction". The pump does not need to work as hard to move the heavier fluid. Determine the optimal fluid change period Designers should recommend a fluid change period. The best practice to determine the change period is to make a decision based on the statistical data of the contamination of the system. This data is collected through oil analyses, which include: measurement of the fluid viscosity, contamination particle count, water content and dissolved metal to determine how well the system is operating. We need to collect the data from the period of the system failure. Next, we can make the probability density distribution and determine L10 time for the system.


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Particle count The particle count is a “catch-all” type test. Almost anything that goes wrong in the machine will result, sooner or later, in an increased number of particles. If misalignment in the components, overloading, water contamination, viscosity breakdown, or bearing failure occurs, the particle count will rise. The challenge of this test is the correct interpretation of the test results. By definition, particle count is: the number of particles in the fluid greater than a particular micron size per unit volume of the fluid. It is often stated as particles larger than 10 microns per one milliliter (1mL). Usually an optical microscope is used to count the particles. Filter and hydraulic fluid manufacturers have recognized the importance of smaller contaminants. As a result, they include particle count tests and evaluations for particles smaller than 4µm. Analyzing fluid samples for the finest contaminants (4µm) is included in the latest ISO Cleanliness Code. In order to evaluate the fluid, we have to know the threshold for various performance characteristics. Test parameters for which threshold levels may be established include: • • • •

Dynamic leakage: external and internal Static leakage: external and internal Performance characteristics: pressure, flow rate, noise. Fatigue life of the hydraulic components. It is very difficult to determine whether a hydraulic component failure is caused by the fluid.

Synthetic fluids are a good choice for equipment that is used outside and is subject to temperature changes. Synthetic fluids can handle a wider range of temperature changes than petroleum-based fluid. Thermal effect The trend in the hydraulic systems design is to increase system power while using the same physical space. Power is increased by increasing fluid pressure and it results in increased operating temperature. Some systems now run with 100ºC fluid temperature instead of the recommended maximum of 80ºC. High temperature changes the viscosity of the fluid, therefore when selecting the fluid type designers should select the ones with greater resistance to change in viscosity as a result of temperature change.

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One of the most common causes of thermal failure in hydraulic fluids and some lubricating oils is aeration (presence of air bubbles). These bubbles can become rapidly compressed in hydraulic pumps and bearings. This results in extremely high localized temperatures (adiabatic compression) and leading to sudden formation of carbon fines. Hot surface carbonization is another form of thermal failure. When oil degrades as a result of high temperature, it creates problems associated with sludge, varnish, deposits, viscosity change and additive decomposition. Negative effect of air Air in the fluid can cause a great deal of damage to the hydraulic systems. Air properties that affect the system properties and performance are: Compressibility- it decreases the stiffness of the system, increases fluid temperature and causes thermal degradation. Compressed air bubbles can reach temperatures above 1000ºC and break down and darken the oil. Gaseous cavitation- causes wear of the hydraulic components and increases the noise in the system. Contains oxygen – promotes oxidation of the fluid. Oxidation also increases at high fluid pressures, temperature or both. Influence of additives Industrial hydraulic fluids normally contain anywhere between 99% to 99.5% base oil and about 0.5% to 1.0% by additives. These additives are important for the hydraulic system operation, particularly when the trend is to use longer-life fluids. There are many different additive combinations that can be used in fluids, and using the right additive combination is critical. One of the main ingredients, which represents approximately 60% to 70% of the additive package is a compound known as Zinc Dithiophosphate (ZDP). Fluids containing this additive reduce the wear rate of the hydraulic components. ZDP reacts with the metal and provides cushioning between sliding surfaces. It also improves the oxidation stability of the fluid. A fluid supplier can help identify what mix will provide the best performance for your equipment. Despite all benefits of the additives, if the fluid is not kept clean or it is overheated, the additives can be physically removed or chemically decompose in service. Removing and decomposing the additives is due to: oxidation, hydrolysis, thermal degradation or they can be removed by the filter.


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17. Fluid Cleanliness Most of hydraulic system failures are a result of fluid contamination from dirt, water or abrasive particles. We deal with the contamination level by setting up cleanliness level requirements to the fluid. Maintaining a low contamination level is even more important when the hydraulic system is new because valves have tighter clearances. Fluid cleanliness level Fluid cleanliness is a measure of the amount and the size of contaminating particles in the fluid. The size of the contaminants is given in microns. 1 micron (micrometer) = 1µm = 10-6 meters Fluid cleanliness level is defined in ISO 4406 standard. This standard uses a numbering system to define the contamination level. Cleanliness code Cleanliness code (per ISO 4406:1999) gives a number of particles greater than a specified particle size per defined volume of 1 mL (0.001 litters). The code format is: XX/YY/ZZ XX refer to the quantity of particles over 4 microns per 0.001 litters YY refer to the quantity of particles over 6 microns per 0.001 litters ZZ refer to the quantity of particles over 14 microns per 0.001 litters For example, hydraulic fluid cleanliness code 20/18/15 indicates that there are: 219 to 220 particles over 4 micrometers (µm) 217 to 218 particles over 6 micrometers (µm) 214 to 215 particles over 14 micrometers (µm) Fluid cleanliness requirements Fluid requirements depend on the pressure and the components in the system. Systems with higher pressure have higher cleanliness

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requirements for the fluid. This relationship is given in table 3.5. The cleanliness code is given only by the second and the third number (__/YY/ZZ format). Quantity of particles over 4 microns is not provided. Recommended fluid cleanliness requirements for some common components at different pressures are given in tables 3.6 (pumps), 3.7 (valves) and 3.8 (actuators). Component internal clearances between moving parts and system pressures are the main factors determining fluid cleanliness. The data is based only on the specific test conditions prescribed by the standard. For different field conditions, the actual fluid cleanliness requirements may vary. Recommended filtration (in microns)

System Maximum Pressure

Recommended cleanliness

p < 50 bar (5MPa)

19/16

15 to 25

p = 50 to 150 bar

18/15

12 to 15

p = 150 to 250 bar

16/13

10 to 12

p > 250 bar

15/12

5 to 10

for β (x) ≥ 75

Table 3.5 System components

Minimum recommended fluid cleanliness Less than

More than

(2000 psi)

14 to 20 MPa (2000 to 3000 psi)

Gear- constant flow

20/18/15

19/17/15

18/16/13

Vane- constant flow

20/18/15

19/17/14

18/16/13

Piston

19/17/15

18/16/14

17/15/13

Gear- variable flow

19/16/14

18/15/13

17/15/13

Vane- variable flow

18/16/14

17/15/13

16/14/12

Pumps

Table 3.6

14 MPa

20 MPa (3000 psi)


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Less than

More than

20 MPa

20 MPa

Directional (solenoid)

20/18/15

19/17/14

Check valves

20/18/15

20/18/15

Cartridge valves

20/18/15

19/17/14

Pilot operated check valves

20/18/15

19/17/14

Pressure control (modulating)

19/17/14

19/17/14

Flow control (standard)

19/17/14

19/17/14

Steering orbitrol (open center)

20/17/15

18/16/14

Steering orbitrol (load sensing)

19/16/14

17/15/13

Load-sensing directional

18/16/14

17/15/13

Hydraulic remote control

18/16/13

17/15/12

Proportional directional

18/16/13

17/15/12

Proportional pressure controls

18/16/13

17/15/12

Proportional cartridge

18/16/13

17/15/12

Proportional pressure relief

18/16/13

17/15/12

Servo valves

16/14/11

15/13/10

Valves

Table 3.7 Less than

More than

2000 psi

14 – 20 MPa 2000- 3000 psi

Gear motors

21/19/17

20/18/15

19/17/14

Hydraulic cylinders

20/18/15

20/18/15

20/18/15

Vane motors

20/18/15

19/17/14

18/16/13

Radial piston motors

20/18/14

19/17/13

18/16/13

Axial piston motors

19/17/14

18/16/13

17/15/12

Cam wave motors

18/16/14

17/15/13

16/14/12

Actuators

Table 3.8

14 MPa

20 MPa 3000 psi

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18. Electric Motors Although electric motors are not hydraulic components, they are included in this book because they are an integral part of the hydraulic systems. Most hydraulic systems use a constant displacement pump driven by an electric motor. There are four main types of electric motors used in mobile truck applications.

Series motor Armature and field (stator) windings are part of the same circuit. There are three configurations shown in fig. 3.19 a) Series Armature and stator winding are connected in series. b) Shunt Armature and stator winding are connected in parallel. c) Compound A compound motor has three windings: one armature and two fields. One of the field windings is connected in series and the other in parallel. These two windings create two magnetic fields. The current flowing through the rotor and the field is the same.

I = IK Fig. 3.19

I = IK + IB

I = IK + IB


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Separately excited motor Armature and field windings are separate. There is an independent control of the rotor voltage and field current (fig. 3.20)

IK

IB

Fig. 3.20

Permanent magnet motor Permanent magnet (PM) motors have a magnetic field created by permanent magnets in the stator. The rotor is similar to the series motor. Efficiency of the PM motors is within a wider range than DC motors with field windings. They can be more efficient than the other DC motors because of the free magnetic field. On the other hand, if the magnetic field distribution is not uniform, the motor is less efficient. Permanent magnet stepper (PMS) motors use permanent magnet rotors and are controlled by electrical pulses. These type motors are not popular choice for forklift truck applications.

AC induction motor Only three phase motors are used in lift trucks applications. AC motors require the use of an inverter which converts the DC to AC current. The power can flow through the inverter in both directions from the DC battery to the AC motor and vice versa. The inverter output voltage is always less than the input voltage. Knowing the characteristics of each motor, helps designers make better choices when they are selecting the pump type for the hydraulic system. Main advantages and disadvantages are listed in table 3.9

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Type motor

Series


Advantages Simple construction motor

Require brush maintenance

Low cost motor

Difficult to control high speed

Low cost controller

Separately excited

Permanent magnet

Disadvantages

Low system (motor Expensive & controller) cost motor Easy speed Require brush control. maintenance It can have speed feedback. Low cost motor for power under 1.5 kW No heat generation in the magnets

Require brush maintenance

Simple construction. Low cost AC induction

No brushes. Low maintenance Highest efficiency Good rotational speed control. It has speed feedback.

Table 3.9

Need DC to AC inverter More complex controller

Hydraulic system application Systems with constant pump speed

Systems with two & three pump speeds Good for energy recovery systems Good for integrated systems Low power systems Steering and small low lift truck systems Best for systems with variable pump speeds. AC motor/controller system can compensate for increased load or reduced battery voltage and maintain consistent speed. Best for energy recovery systems Best for integrated (lift and steering) systems


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Chapter 4

Management and Quality of Hydraulic System Design Process Brief history of quality At the end of the 19th and the beginning of the 20th century, a number of companies saw the connection between the quality of their products and their long-term success. Some of them started monitoring and controlling quality in order to achieve higher market success. In 1887, the president of the Procter and Gamble Company, William Procter, realizing the need for a quality product told his employees, “The first job we have is to turn out quality merchandise that consumers will buy and keep on buying.” At this time the managers and the engineers did the planning and supervising while the workers executed the production work. Later, in the 1940s, the supervision of the quality was transferred to quality inspectors. After World War II, Japanese companies used the tools of quality to improve their products and develop a culture of continuous improvement. The Union of Japanese Scientists and Engineers (JUSE) financed and encouraged companies to educate their employees on quality management practices. The Japanese recognized that inspecting a product does not add value to the product because it does not change its properties. Instead, they focused on improving the quality of the process which would lead to increased quality of the product. The Japanese called it Total Quality Control (TQC). For them, process control was equal to total quality. In the late 1960s, Professor Shigeru Mizuno and Yoji Akao developed a communication tool that converted customer needs to measurable design characteristics. They called it Quality Function Deployment (QFD). Later, the Japanese incorporated the value engineering principles in the QFD matrix. In the 1970’s, the quality of the products made by Japanese companies that applied the quality principles, started to exceed the quality of their competitor’s products in other parts of the world. This resulted in a penetration of the Japanese products into the Western markets.

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Chapter 4: Management and Quality of Hydraulic System Design Process

Toyota Production System (TPS) for example is a process toward reduction and elimination of waste and anything that does not add value to the product. In order to achieve this, Toyota involved the production workers in the process and established a culture of continuous improvement (called Kaizen in Japanese language). There are two main factors in developing the Kaizen culture in the company. The first factor is: give workers ownership of the process they do and the second is: guarantee employment to all full-time workers. In the 1980’s, American and European companies recognized the importance of quality improvement programs as a key factor for success and they started applying quality matrixes to measure their products and processes and to apply the continuous improvement principles. When Western companies recognized the value of Total Quality Control (TQC) and started applying it in North America and Europe they called it Total Quality Management (TQM). Western specialists used the word “management” instead of “control” because their understanding was that TQM was a system of managing the people. They hired quality control specialists and put them in charge of monitoring the quality of the supplies, the manufacturing process and the final product. European countries under communist rule also implemented the product quality concept. The communist parties took a leading role and determined the course of quality improvement in the country. They had a government quality improvement program and instructed the managers to implement this program. In some cases, people, who had little or no production experience, were chosen to lead and manage the quality process. As a result, they could not involve the production floor workers and did not achieve the quality level of their Japanese competitors. The fundamental difference between the Japanese and the Western companies in the manufacturing sector was the participation of the blue collar workers. The Japanese gave ownership of process quality to the people involved in the process while the Western companies hired specialists to manage the quality of the process. In North America and to some degree in Europe, when production was low or when a new process, eliminating labor is introduced, the extra workers are laid off. As a result, workers were not motivated to make suggestions on how to reduce the cycle time or improve the process because they might be laid off. Japanese companies also have ups and downs in the production volume but, in order to preserve worker’s continuous improvement thinking, they


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don’t layoff their full-time workers. Instead, they have part-time hourly employees. Part-time employees are given more or less hours depending on the production volume.

Introduction Quality of the design directly affects public welfare, health and property. The safety requirement must always be the number one requirement for any new product. To ensure an accurately designed product, in many countries, the firms that perform engineering design work are required to obtain a “Permit to Practice”. To obtain such a permit, the firm must satisfy two conditions: 1. Identify the individuals responsible for the engineering design and services. The responsible individuals are required to have a professional engineering license. 2. Carry liability insurance. In addition to the company insurance, some countries or states require all engineering personal to have secondary liability insurance. The United States was the first country to regulate engineering practice. There are three main factors that determine the quality of any product. First: quality of the final product is determined by the quality of the design process. Second: quality of the final product is determined by the quality of the manufacturing process. Third: quality of the design and manufacturing processes are achieved by defining and measuring each step of the process. In this chapter, a step-by-step methodology for a hydraulic system design process will be shown. In the past, quality strategies were focused on the manufacturing process. Now, the focus is on process management. For new system designs- the process management approach is a way to control each step in the design process. In addition, an integration of the organizational functions and the information flow through all steps of the design will be shown. The process approach, which includes engineering

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and management tools, will ensure a quality of the design that is built into the final product. During the process, the factual approach to measure the progress will be used. Factual approach is when each factor or parameter at each step is measured against a predefined value. ISO 9000:2000 states that the factual approach to decision making is "effective decisions that are based on the analysis of data and information." At the end of the chapter, a brief description of patents and their applications will be included. Novel ideas, whose outcome is obtaining patents, are the ultimate result of well managed and quality design process.

Factors Main factors that influence the quality of the design are: • Stick to the company’s goals and strategic objectives Strategic objective are a set of long-term directions that would allow the organization to achieve its long-term goals. • Effective professional communication Effective professional communication is a two-way interaction which includes listening, receiving and turning over information both verbally and nonverbally. The communication is considered effective when the inputs from this interaction are used to create desired results and solve problems. Effective communication is a main factor for the continuous improvement of the organization. • Leadership Leadership is an ability to positively influence and motivate people to achieve the team’s goals in an effective way. • Engineering knowledge Engineering knowledge is a base for any novel approach to new designs and improvements. In order to design a hydraulic system for lift trucks, an engineer has to know not only the hydraulic principles but also the equipment principles of operation such as:


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Load center of gravity- understanding the limits and maximum lifting capacity at the load center of gravity. o Mast tilt or reach and the effects on the load balance. o Kinematics of lift and steering mechanisms. o Elevating mechanism construction o Truck stability triangle. o Ramps and inclines. The engineer also has to know the operating rules and specific safety aspects for the equipment. o

• Innovations The generating and testing of new and novel ideas must be one of the main goals of any engineering organization. Innovating spirit must be encouraged and rewarded in order to become an organizational culture. o One technique to generate new ideas is using old ideas and adding something new to them. This strategy is called: knowledge brokering. o Another technique is to take an existing idea in one application and use it in another application. This technique has been used the most through-out the history of the technological development. Old ideas are the main source of new ideas therefore we need to educate ourselves and learn the existing practices in order to generate new ones. In this aspect, it is important that the organization has the means of collecting such data and making it available to all employees. • Motivation There are a number of factors that motivate people. It has been proven that the financial factor is not the main motivator because it has only short term effects. Recognizing the individual as a valuable asset to the organization is one of the biggest motivating factors for best performance and continuous improvement. • Focus on quality Quality of the design process is measured by the quality of the final product. The final product must meet customer expectations in term of performance, service, cost of operation and other factors identified by the customer. Time-to-market and greater value for the customer are the main driving forces for today’s companies success in the market.

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Structuring the design process The philosophy of having a structured design process is a key factor for implementing a total quality culture in the organization. The three main principles of the total quality are: 1. Process Infrastructure- that is the process management approach 2. Practices- principle of coordinating the design activities 3. Tools- these are all methods for: collecting data, analysis, calculations and approaches for improvement and problem solving The structured system development process explained in this book consists of 23 steps which are grouped in 7 stages. They are shown in Table 4.1. Structuring the process will help us to be not only more productive but also more predictive. Predicting the time for project completion allows the company to better manage their planning and budgeting. Stages

Design process steps

Tools

1. Project scope 2. Evaluate available people and financial resources Strategy 3. Form a design team 4. Design team goals and objective 5. Define customer requirements

Requirements 6. Define engineering definition requirement 7. Concepts generation Concept 8. Concept selection development

QFD, Functional decomposition QFD, DFX Go/No-Go Screening Decision matrix


9. System design 10. Benchmark 1 System design

11. Evaluation 12. Design review 1 13. Design documentation tune up

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Calculations QFD, DFX FMEA Taguchi Poke Yoke Problem solving

14. Prototype of the system 15. Test Design 16. Benchmark 2 evaluation 17. Design review 2

Calculations

18. Test at customer location 29. Validation of the design Design 20. Benchmark 3 validation 21. Decision to start production 22. Evaluate design process Process 23. Lessons learned improvement

Customer survey

Continuous improvement

Table 4.1 A hydraulic system is a product in which the cost of the components represents about 80% of the total manufacturing cost. Since the components are usually off-the-shelf purchased parts, their cost is easy to obtain. Therefore, our goal should be to predict the total cost of the system with no more than a +/- 5% error.

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Definitions of tools used All tools are well described in the literature. In this book, I will only give short descriptions of the tools. DFX stands for Design for Excellence. It is defined as a knowledge based approach whose goal is to design a product or system that maximizes the desirable characteristics (such as; quality, reliability, safety, time-to-market) and minimizes the undesirable characteristics (such as manufacturing cost). DFX includes Design for Manufacturability (DFM), Design for Assembly (DFA), Design for Producibility, Design for Environment, Quality Function Deployment, Taguchi’s Method, and Failure Mode and Effect Analysis. QFD (Quality Function Deployment) can be part of the DFX process or it can be used as a standalone tool. QFD is a systematic approach for improving product/ system quality by making sure that the final product meets the customer requirements. QFD tool is used to help us think through every aspect of what our customers want and how to deliver it. It does not guarantee that we will make a product exactly the way the customer wants but it does ensure that our product is as close to the customer’s requirements as possible. It also ensures that we design our products more efficiently. There are four key steps to QFD thinking: 1. Product Planning 2. Part Planning 3. Process Planning 4. Production Planning A basic QFD house shows the conversion process and the relationship between customer and engineering requirements. A QFD house for a forklift hydraulic system is shown in table 4.1 (Appendix E). Functional Decomposition is a technique of breaking down one problem to smaller and more easily managed sub-problems. This technique has two or more steps.


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Step one - find a single overall function that this system needs to accomplish. For the hydraulic system, described in Chapter 5 (Fig. 5.1), this function will be: Design an electric forklift truck hydraulic system which will allow the operator to manipulate the rated payload in all three degrees of freedom up to the maximum height of the mast. Step two - decompose the function into top level functions for the system. The top level functions will identify what the system is supposed to do. Step three - decompose further any of the top level functions into sub-function in order to refine the function as well as possible. The process will go on until all functions become measurable and simple to satisfy. Go/No-Go Screening is an evaluation of whether a proposed concept meets the corresponding engineering specification. If the answer is YES or MAYBE, the concept is GO. If the answer is No, then the concept is a NOGO. Decision matrix is used after a Go/No-Go screening. This tool provides a means of scoring each concept against pre-defined criteria for comparison. The matrix has weight columns in which each criteria is given a relative importance. A decision matrix for a hydraulic system is given in Table 4.2 (Appendix E). Steps for this method are: 1. Choose criteria for comparison- criteria can be either Customer Requirements or Engineering Requirements 2. Select concepts to be compared 3. Generate scores 4. Compare scores Taguchi’s Method is a technique to optimize the design process in order to minimize its cost. FMEA (Failure Mode and Effect Analysis) is a technique for improving the quality of a design and manufacturing process by first identifying and then eliminating or minimizing the potential quality problems.

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Description of the design process steps Step 1 - Project scope Defining the scope means breaking down the project deliverables into manageable tasks and establishing resources and milestones that can provide perspective on the project as a whole. Step 2 - Evaluate available people and financial resources Identify the people and material resources required to complete the project. Step 3 - Form a design team and obtain a project budget A design team is formed on the basis of people and resources from step 2. The best practice is to form a cross functional team to coordinate all phases of the product development, to reduce the time and enhance the crossfunctional communication. Step 4 - Design team goals and objectives • Define design team goals & objectives • Familiarize all team members with the team goals and objectives The design team objectives indicate what we want to achieve. Step 5 - Define customer requirements Defining customer requirements is a collaboration activity between engineering, marketing and the customers. The customer requirements are different for different hydraulic systems. The general requirements for lift truck application are: • • • • • • • • •

Safe operation Low acquisition cost Low total cost (purchase, maintenance, operation) Durability Ergonomics- ease to use controls and operator’s comfort Quiet operation Little or no down time during the warranty period (1-3 years) Ease of maintenance Ability to add attachments to the truck


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• Maximum capacity retention at certain height • Programmable lifting/lowering speeds • Ability to work at cold and hot surrounding temperatures • Smooth mast staging • Comply to safety standards In addition the general requirements, there are specific requirements. The specific requirements could be defined by one customer or by group of customers and target specific segment of the market. Step 6 - Define engineering requirements The engineering requirements (ER) are also called design specifications (DS). They tell how we want to achieve the customer requirements and list the limits in which the system has to work. The process of converting customer requirement (CR) to engineering requirement (ER) is called mapping. One of the most effective conversion tools is the Quality Function Deployment (QFD) technique which is also called house of quality. The basic QFD house shown in table 4.2 (Appendix E) has three main components: • List of customer requirements with assigned weights to them. • List of corresponding engineering requirements with measurable units. • List of competitors or benchmarks. After the structure of the house is constructed, we fill out the relationship between CR and ER. An example of a QFD for a hydraulic system is shown in table 4.3 (Appendix F). If for some ERs we cannot put a measurement unit, it means that they are too general. In this case, we use the functional decomposition technique to break the general requirements into smaller, simpler to satisfy and measurable requirements. Engineering requirements also include all safety and government requirements described in standards and practices. Step 7 - Concepts generation Our goal will be to generate as many concepts as possible for each of the lowest level functions identified in the decomposition process. Sources for concept ideas are: previous designs, customer suggestions, competitor designs, existing patents and technical literature. The most effective technique for concepts generation is brainstorming.

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Step 8 - Concept selection At this stage, the goal is to select one concept for each function and combine them into a conceptual design of the system. We can also combine a few design concepts into one that meets the customer’s needs most. There are usually two main steps. First step is to use the Go/No-Go screening technique. This step is usually used when we have more than 4 or 5 designs. The outcome will be reducing the number of designs to less than five. The second step is to use decision matrix tool. An example of such a matrix is shown in table 4.1. This technique relates and evaluates each concept against the customer or the engineering requirements. When the new concept is selected, we have to make sure that we are not infringing someone else’s intellectual property. It is recommended we obtain advice from a patent lawyer regarding the potential infringement. There are cases when we cannot come up with a better idea than our competitor’s product and we cannot use their idea because it is protected by a patent. The steps that need to be taken in this case are described in the Patents section of this chapter. Step 9 - System design System design includes the following activities: • • • • • •

circuit design calculation of the parameters component selection component layout methods of component control (sequence of commands) software development and tuning

System design will be our main focus in the following chapters of this book. Software development and tuning will not be described. This last step has to be done after assuring that all components are working properly. Step 10 - Benchmark 1 Benchmark 1 is a comparison of the calculated values against the engineering requirements, review manufacturing and assembly processes.


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Step 11 - Evaluation • • • • •

Safety and human factors (Ergonomics) Novelty of design- Is the design patentable? Innovation- consider emerging technologies that can improve the design Cost of the product- cost of components, assembly time & testing. Principle of operation

Step 12 - Design review 1 The design review is done together with manufacturing and service groups. They must confirm that the proposed design can be manufactured at a predetermined production volume and cost. Then, the manufacturer starts working on developing a manufacturing process and estimating cost per unit and cost of service. After they accept the design, these groups are committed to producing, maintaining and servicing the product. The ability to do these three activities in the most cost-effective way, determines the future profitability of the product and the level of customer satisfaction. Step 13 - Design documentation tune up At this stage, we tune up the engineering documentation. After this stage, we should have a set of parts and assembly drawings, components and system specifications that provide enough information for making a prototype. Step 14 - Prototype the system At this step we need to build a working model. Step 15 - Test (Two tests are required for a Hydraulic system) • Performance test • Reliability test The first goal of these tests is to verify the performance of the system. The second goal is to determine how the performance of the components and the whole system changes over a period of time. These tests are performed per duty cycle for the exact application. The third goal is to see the effects on other systems (electric, control) of the truck. Hydraulic component manufacturers usually provide graphs of the best performance of the new state of products at specific laboratory conditions.

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Our goal at this point is to verify the components and system parameters over a period of time. The best system performance is verified at 20 °C room temperature and 40 °C fluid temperature. The worst performance is tested at the extreme conditions defined in the project specification. All test results must be recorded and used in the next step. Another factor is the supplied power. When the hydraulic system is designed for an electric truck, a fully charged battery must be used. If the battery voltage goes down, the motor speed also goes down. As a result the truck performance is decreased and it goes outside the advertised operating range. For example: if the battery charge goes down to 50%, the lift motor speed can go down 10% or more during lifting. Step 16 - Benchmark 2 Benchmark 2 has three sub-steps: • • •

Benchmark for conformance- verification of quality of the design by benchmarking against engineering specifications. Competitive benchmarking- this comparison relates our future product performance to the performance of an existing product. Manufacturing evaluation- In order to reduce the risk of having manufacturing defects, improve productivity and reduce assembly time, at this step of the process, we also evaluate the design based on manufacturing capabilities.

There are uncontrollable factors influencing the system performance- such as weather and temperature- but our goal is to identify the limits in which the system operates. These limits are usually taken from the customer requirements. Step 17 - Design review 2 Review the design for safety, reliability, environmental risks and disposals. Step 18 - Test at customer location It is very beneficial for a company if they work together with their prospective customers. If a customer’s location is not available, the test is done at a test location where the proper conditions are established.


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Step 19 - Validation of design Validation of the quality of design is actually the validation of the final product performance. Be sure to always check the final product against the customer’s requirements. This step is necessary because we always lose information and give up something during the design process. At this step, we also collect customer feedback. Step 20 - Benchmark 3 Benchmark 3 compares the customer’s satisfactions against the initial customer’s requirements. Step 21 - Decision to start production Up to date results are presented to the senior management and the management makes the decision for launching the new product. Step 22 - Evaluate Design Process Several quantitative and qualitative measurements of the project’s activities must be taken and recorded during the process to determine whether the goals and objectives of the project have been met. The main measured activities are: o

Duration of time waiting for information- it should not exceed one business day

o

Information, libraries and database are available 99% of the time

o

Time schedule is met 90% of the time

o

Project budget is within 10% of forecasted cost

o

Each sub-task has an owner

o

All activities are recorded 100%

o

Funds are available at any time

o

On-going training is scheduled when it is necessary

o

Follow all technical standards

o

The total system cost is within +/- 5% of the projected cost

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o

Number of redesigns/ changes

o

Number of design iterations that needed to be retested

o

Customer satisfaction from the final product

Step 23 - Continuous improvement The main goal is to find a way to execute the design process more efficiently next time when we apply it. This goal is achieved through understanding the activities that influence the design time, quality and expenses. Continuous improvement has three steps. First, evaluate the process and identify where in the process the largest amount of money and time were wasted; secondly, document all findings and third, look for ways to improve this process. The factors that influence the design process can be divided into three groups: • •

•

Controllable by designers: people skills, motivation, knowledge, ability to work in a team Controllable by the company: Provide training, process, tools, literature, library, standards and conditions for collaboration between team members from different departments Uncontrollable: market and government regulations

Design guidelines 1. Address each function of the system separately before combining them in a system 2. Minimize the number of component 3. Use standard components 4. At each step of the process- consider the efficiency of the system 5. Lower potential assembly errors (mistake-proof design) • Easy to assemble • Parts cannot be assembled in a wrong way • Obvious when a part is missing


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• Assemble from one direction 6. Make the system easy to service • Special tools are not needed • Avoid special instructions 7. Consider liability Potential liability claims make documentation of quality assurance procedures a must for any company. By law, anyone who sells a product that is defective or unreasonably dangerous is subject to liability for any physical harm caused to the user. Quality of the design reduces the risk of product liability and provides support evidence in defense of the product. The company should record all evidences (tests, analysis, FMEA) that show that the design team made all necessary steps (did everything possible) to design a safe product including packaging and inspection. 8. Address environmental issues All designs have to comply with the environmental government regulations. Future regulations trend to have manufacturers be responsible for the full life cycle of the equipment. This means that hydraulic system designers have to think of the disposal cost at the beginning of the design process. Components that are easier to recycle and hydraulic lubricants containing environment friendly substances have to be preferred choices when making component selection. The main design requirements are: use of non-toxic substances and the use of recyclable and disposable materials. A low cost environment friendly product is achieved by reducing disposal time, regulatory cost and making it recyclable. There are few tools and guidelines available to optimize the complex designs. Two of most commonly used are: Design for Environment and Design for Disassembly (used to reduce disposal time).

Documenting the design activities Documenting the design activities facilitates long-term improvement. We cannot expect the same people to work on the same designs all the time. Some people get promoted; others go to other companies and some move to different positions. A main factor for making the design process better the next time is well documented steps of all design activities. Clarity and

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accuracy of the technical calculations are an important part of the practice of documenting. All given data, assumptions, mathematical and physical laws have to be specified clearly. Calculations are an intellectual asset for a company. Therefore, they should be documented in a way that allows any other engineer with the same background to be able to understand and use them. Good practice is to put all calculations on a server in HTML or PDF format. All design concepts, even if they are not used in production, are essential intellectual property for the company. Documenting all design concepts is necessary to prove the invention date if a company decides to seek a patent protection for the new design. In the last section, I am going to give a brief description of patents and the patenting process. General understanding of the process will enable the engineers to communicate more effectively with patent lawyers.

Project close-out criteria Any project must come to an end. The purpose of project closeout is to officially end the project and evaluate the level of success. The following criteria must be met to close the project: • • • • •

All issues and action items have been completed and signed off All required work products have been produced All deficiencies have been logged and signed off All quality assurance issues have been addressed A project termination or cancellation statement exists.


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Failure and failure rate Failure is a state at which a component is unable to perform a required function or performance. Failure rate is the number of failures per specified time period. One of the criteria used to determine the quality of a hydraulic system design is the failure rate of the system. When reliability is a customer’s requirement, this parameter can be part of the engineering requirements (ER). Further, the ER can include no failure period equal to or greater than the warranty period. When we evaluate a new design, first we estimate the reliability of each component separately and then, all components together as a system. After the system is in production, we look at the failure rates of the individual components. The term failure is referred to instantaneous event of malfunctioning of the system or the component. There are different types of failures such as: unusual noise, visual mechanical breakage, lost of power, deviation from system parameters, increased fluid temperature, etc. There are four failure rate distribution functions that are shown on Fig. 4.1. •

In Hyper exponential distribution the failure rate decreases in time. This distribution shows that there are built-in defects which show shortly after the system is put into service.

•

In negative exponential, the probability of failure is constant in time meaning that the likelihood of malfunctioning is the same at any period of time. Constant failure rate suggests that the malfunctioning is caused for random reasons.

•

Normal distribution is when failures increase in time due to wear and aging.

•

Weibull distribution is used because the results are easy to analize. If the failure rate decreases: ß < 1. If the failure rate is constant: ß = 1. If the failure rate increases: ß > 1.

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Fig. 4.1 (source: AKS Jardine, Maintenance, Replacement and Reliability 2002)

Patents A patent is an intellectual property, recognized to be a novel idea, granted by a Patent Office to the inventor (sole or joint). The legal purpose of the patent is to exclude others from making, using or selling the invention.


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There are three important parts which any patent must have: 1) a novel idea, 2) usefulness to society (diligence) and 3) not obvious. There are two types of industrial patents: • Design Patent Design type is a novel idea only in the appearance of an object. Its protection is 14 years from issuing the patent. • Utility Patent Utility type is a novel idea that is useful to society. These types are inventions of machines/mechanisms or processes (method of operation or manufacture); articles of manufacture (casting, molding). Its protection is 20 years from filing. The first step toward obtaining a patent is preparing and filing a patent application. The application contains three main parts: a specification (description of the invention and claims), drawings and an oath by the inventor(s). The claims define the invention. When more than one person works on a design, it is difficult to determine the inventors of the subject matter described in the claims. The rule is that anyone who made a significant contribution should be included as an inventor. When filing for a patent, the timing is very important. In the USA and Canada, the filing must be done within one year from the public disclosure of the invention. For this reason, it is very important to have confidentiality agreements with your business partners who have had any involvement during the design process. The US and the Canadian patent offices apply first-to-invent rule when they grant a patent, while the European offices use first-to-file rule. The government of the United States can stop any patent application from publishing if it is considered it to be a danger to the national security. In cases like this, they label the information as classified and instruct the inventor/s not to disclose or publish the invention. The government is not obligated to financially compensate the inventor/s. Patent protection extends only over the country that issues the patent. If we want to have protection in more countries, we have to file an application in each of these countries in their official language and according to their patents laws.

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Designing around an existing patent As mentioned earlier there are cases when we cannot come up with a better idea than an idea that is already protected by a patent. In cases like this there are two basic approaches. The first one is to use the patented idea. If they agree, we have to sign a License Agreement and in most cases pay a loyalty fee which is usually in the range of 1 to 5% of our product’s listed price. The second approach is to design our product around the patent. In this case, our focus will be to produce a similar product or design which does not infringe upon the existing patent. Designing around a patent is a widely used practice which usually results in more novel ideas or improvements. There are few steps that should be done to ensure we successfully get around the patent. 1. Read the existing patent. Claims are the most important part but they are also the most difficult to understand even by a patent lawyer. That why we should first concentrate on the “Description of the Invention” section in order to become familiar with the details of their idea. 2. Review other patents that are given as reference on the front page of the main patent or make your own prior art search to obtain all related patents. We have to find some of the other patents that are expired. Expired patents are free from infringement and they can be used by anyone. 3. Next we create a design based on a combination of ideas from one or more expired patents or a combination between expired, nonexpired patents and our own contribution. 4. If we make an improvement to an expired patent, we can file for patent protection for the improvement only. We can also make an improvement and obtain a patent for this improvement to a nonexpired competitor’s invention if we want to prevent our competitor from improving its design.


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Legal aspect of the design process When a product causes personal injuries, property damage or both, the harmed person will want to be compensated and he or she, in most cases, files a lawsuit against the product manufacturer. The party bringing the action or making the claim in the lawsuit is called a plaintiff. The plaintiff’s goal is to prove that the manufactured product has a flaw because of manufacturing or design error. The party against whom the claim is made, or the party defending the action, is called the defendant. Different states have different legal systems and standards. European countries and the Canadian province Quebec have civil laws where the legal principles are written by experienced legislation makers and approved by the government. The legal system in the USA and Canada (except Quebec) is based on the English common-law system. This system is also called the “judge-made law” system because court decisions establish the legal principles. In the USA, each state has its own legislation which can be different from the federal. For example, in California, even if the product meets all standards, the court can hold the manufacturer liable for any loss if the product does not perform according to the customer “reasonable expectations”. The main goal of management and quality of the design process is to ensure that each design meets specific conditions in order to guarantee error free design from legal prospective. In a number of cases, the court found that when the manufacturer builds a product to the engineering specifications, it has no legal duty to evaluate the safety and test the device. Therefore, when the design firm and the manufacturer are two separate entities, the design firm must assume full responsibility for the safe operation of the hydraulic system. Lawsuits dealing with design defects of hydraulic systems involve a complete review of the engineering calculations and standard practices. The plaintiff tries to prove that there were some of the following elements: 1.

2.

Negligence- Negligence is carelessly executed work or work below accepted standard. A manufacturer is liable because of negligence when: a) it fails to make an evaluation that would uncover a situation that may cause an injury, b) it failed to conduct a test that would uncover a defect. Incompetence- Incompetence is a lack of knowledge or skills of a person who carries out responsibilities of design engineers.

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3.

Lack of credentials- the engineering firm failed to obtain all necessary professional certifications.

The engineering design firm, on the other hand, has to prove that in their design process they: 1. 2. 3. 4.

Properly applied engineering principles. Anticipated modes of failure. Provided correct specifications for the user. Provided instructions for correct inspection and maintenance.

When the system design is completed, the design engineer should be able to predict the behavior of the system for a defined period of time and assume responsibility for the safe operation of the system.


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Chapter 5

Hydraulic Systems for High Lift Trucks In this chapter, hydraulic systems for high lift trucks- classes 1, 2, 4 and 5 will be described. The classifications that are used in this book are the ITA classifications shown in Appendix A. ITA (Industrial Truck Association) represents the manufacturers of lift trucks and their suppliers who do business in Canada, the United States or Mexico. ITA plays a major role in the development of the industrial truck standards and regulations. Class 1 is the most common class trucks. It includes counterbalanced sitdown and stand-up rider trucks. Trucks can have either electric motors or internal combustion (IC) engines. Class 2 contains the narrow-aisle lift trucks. These trucks don’t need counterweight because the load center is within the outline of the wheels. They have smaller turning radius and are mainly used at indoor warehouses with narrow aisles. Class 3 trucks are also called low-lift trucks or trucks for horizontal transportation. They can have walk-behind or ride-on operator. The hydraulic system of this class is described in Chapter 6. Classes 5 and 6 have the same construction as Class 1 but higher load carrying capacity. They can have solid rubber or pneumatic tires. Solid rubber types are made from softer core and harder outside layer and are preferred for indoor applications with smooth floors. Pneumatic tires are inflated with compressed air. Because of the better cushioning, they are preferred for outdoor work and uneven floors. Class 7 are rough-terrain lift trucks. They are used primarily for outdoor material handling. Their elevating system consists of a telescopic lifting arm. The hydraulic system of this class is described in Chapter 7.

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Chapter 5: Hydraulic Systems for High Lift Trucks

In fig. 5.1 is shown a sit-down counterbalanced lift truck. This type is the most common lift truck type. The weight the rear of the truck counterbalances the weight of the load. The truck has two parts: mobile (tractor) and elevating (mast). Tractor consists of chassis (containing the battery, the counterweight and the controls), steering and drive units. Mast frame is constructed from structural steel profiles. Hydraulic cylinders mounted on the mast are used to lift the loads. Mast can have two or more sections where the first section is always fixed while the others are movable.

Fig. 5.1 Sit-down counterbalanced lift truck

Elevating system A triple-mast elevating system is shown in fig. 5.2. This elevating system has three structural profiles (masts). One of them (1) is fixed while the other two (2 and 3) are movable vertically. There are two ram type lift cylinders. The free-lift cylinder (4) is mounted on the outside movable mast (3). It lifts


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the fork carriage without lifting the mast profiles (fig. 5.2b). The main lift cylinder (9) is mounted on the fixed carriage (1). It function is to elevate the movable mast sections 2 and 3 (fig. 5.2c). Vc = Vcarriage Vd = Vcyl Ve = 0 c

e

d

c

d

e

Vcarriage

Fig. 5.2 a) lowered position Components: 1. 2. 3. 4. 5. 6. 7. 8. 9.

Main mast Middle mast Free lift mast Free lift cylinder Pulley 1 (free lift) Chain 1 (free lift) Pulley 2 (main lift) Chain 2 (main lift) Main lift cylinder

b) free lift, maximum height

c) main lift

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Hydraulic Systems Overview The purpose of the hydraulic system is to control and manipulate the load functions such as: lifting, tilting, rotating, moving sideways, reaching and retracting. The hydraulic system should be able to provide enough power to accomplish some of these functions simultaneously when necessary. In addition to achieving the prime functions, the hydraulic controls have to provide smooth and fast motion. There are two main types of hydraulic systems: open and closed. Open systems are hydraulic circuits in which the pump draws fluid from a reservoir and at the end of the work cycle the fluid is returned back to the reservoir. These systems usually use non-compensated fixed displacement pumps as a primary source for flow and pressure. Flow rate is controlled by directional valves or by varying the speed of the pump motor. Maximum pressure in this system is set by a relief valve and the working pressure is controlled by resistance to the flow. When the system pressure exceeds the relief valve setting, the pump flow is bypassed through the valve to the tank. Open systems usually cost less and provide more freedom in the design. This type of system is used in most mobile applications for operating the payload. Closed systems are hydraulic circuits in which the fluid circulates between the pump and the actuator in a closed loop. Actuator rotational (or lineal) direction is reversed by using bi-directional pumps and reversing the flow. Pumps can be either fixed or variable displacement. Closed systems often use pressure compensated type pumps. The maximum pressure is controlled by a pump compensator, which cuts off the flow when the maximum pressure is reached. Supplied pressure continuously adjusts to a value slightly above pressure demand from the actuators. Sometimes, this system does not have a relief valve and there is minimum energy waste while the pump is working. Disadvantages of closed type systems are the high cost of the pump as well as energy waste in the form of generated heat when the pump strives to maintain the working pressure at all conditions. The hydraulic systems can also be classified as: 1) full load sensing systems or 2) systems with load sensing elements. The main focus of these two types is energy efficiency.


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A full load sensing system can also be closed or open. Closed systems use a pressure-flow compensated pump and at least one load sensing valve. The pressure-flow compensated pump is a variable volume piston pump, which senses the system flow and pressure requirements and delivers variable flow at variable pressures per these requirements. The pump stands by at a low pressure in order to save energy. In addition, this system has a load sensing directional valve which has a pressure feedback. The valve receives the pressure requirements from the actuator and sends a pressure signal to the pump. The advantage of this system is the higher power efficiency. A disadvantage of the system is its high cost. Because of the high cost, full load sensing systems currently are not used for industrial truck applications. Instead of full load sensing, the hydraulic system can be designed as an open type system which uses a non-compensated fixed-displacement pump and at least one load sensing valve. This combination appears to have the lowest cost-per-power ration. For this reason, these types are becoming more popular and all new designs have elements of pressure sensing feed back in order to improve efficiency. An example of systems with a load sensing valve will be discussed in this chapter (fig. 5.13) and next chapter (fig 6.4). In the 1990’s, lift truck manufactures started using electronically controlled AC and DC motors as power sources for the hydraulic systems. Availability of variable-speed motors to drive the hydraulic pumps provided more design options to the engineers. Mating variable-speed motors with fixeddisplacement hydraulic pumps allows designers to control pump flow rates by controlling the pump motor rotational speed which improves energy efficiency of the system. In this chapter, step 9 (System Design) of the design process, described in chapter 4, will be explained in detail. System Design includes the following activities: 1) circuit design; 2) calculation of the parameters; 3) component selection and 4) component layout.

Design Principles There are a number of design principles that need to be kept in mind during the systems design.

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1. 2.

Think of the system as part of a bigger system Address all safety concerns • accidental disconnection- it can cause hoses, fittings or fluid under pressure to fly into the air • over pressurization can cause mechanical failure • safe maintenance • safe hose routing and clamping of the hoses 3. Design the system where customization (modularity) is easy 4. Design the system for easy inspection and monitoring 5. Always look for an innovative approach 6. Use standard components where possible 7. Balance of performance and price (optimized price/performance ratio) 8. Integration with analog controllers 9. Consider using network communication. Network saves on wiring. It would also enable remote system monitoring, track lifecycle, identify failures. Improve future designs. 10. After the design is completed, take enough time to test the system and to evaluate the results. 11. Consider the cost of the system 12. Components integration (manifold design) where possible • cartridge valves take less space and cost less than line mounted valves • reduce number of fittings, hoses and tubing • reduce assembly time • occupy less space 13. Plan for operator friendly controls 14. Use flexible hoses when the hydraulic lines are subject to movements.

Design Requirements Before we start designing the system, we have to specify the system requirements. System requirements (design requirements) are the technical interpretation of the customers’ requirements. Two techniques- QFD and Functional Decomposition- are used for converting the customers’ requirements into manageable design requirements as described in Chapter 4. These design requirements become the performance and hydraulic system requirements. In addition, we have to include the


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applicable standards (ASME, ISO, Australian, etc). In this chapter the requirements listed in table 4.1 (Appendix E) will be used. • • • • • • • • • • • • • • • •

Maximum weight of the payload 3000 kg Maximum payload lifting height 3800 mm Lifting speed (empty) 30 +/- 3 cm/s Lifting speed (loaded) 30 +/- 3 cm/s Lowering speed (empty) 33 +/- 3.3 cm/s Lowering speed (loaded) 33 +/- 3.3 cm/s Maximum lowering speed 60 cm/s in case of failure in the load supporting the hydraulic circuit (required by ASME B56.1 Standard, article 7.25.8 ) Fork tilt Fork side shift Maximum work pressure of 25 MPa Ergonomic controls Controls arranged in sequence and direction of motion according ASME B56.1 Standard, article 7.25.6 Reliable system Maximum system noise level: 60 dB Temperature range: from -25°C to +80°C Fluid cleanliness: according ISO 4406

The next stage in the design process, as explained in Chapter 4, is the concept development. In this chapter, I am going to describe the most popular designs and options used by lift truck manufacturers. For each design, I will show a schematic diagram of a hydraulic system which satisfies the functional requirements such as lifting, lowering, tilting, etc. Then, the schematic diagram will be used for initial calculations and for selecting the parameters of the hydraulic system. When we know the system parameters, we can calculate the size of the system components based on the required pressure, flow and efficiency (losses). The next step, “creating the hydraulic component layout”, will not be described here because it is specific to and dependent upon truck chassis construction. Some mounting and layout rules will be described for specific components in Chapter 8 (Selected topics). When going through the design process steps, from chapter 4, the designer should keep in mind a few basic guidelines that may not be part of the design requirements but are important for creating a good design.

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• Minimize the number of components. The reason for this is not only that each component adds cost to the system but also because each component restricts the flow, creating pressure losses and heat. The first step to designing a high energy efficient system is to keep a minimum number of components. • Size the components properly. The factors that affect component selection were described in Chapter 3 (Hydraulic Components). Properly selected components provide quick response of the system functions and minimum drift after the command to the component is given. Also, the proper components increase the efficiency and reduce the cost of the system. • Think of the controls for the hydraulic system. Ergonomic and easy to use controls are becoming the number one user requirement. Operators spend usually 8 hours in a workplace consisting of a seat surrounded by controls. Therefore, the designers have to adapt the controls to the operator’s biological and physical size and ability. Also, there is a trend in the material handling industry to hire more female operators. As a result, more requirements such as: controls on an arm rest, adjustable controls, rotating controls with the operator, controls not restricting movement are added. Although most of these requirements are not put in writing, the design engineer should consider them in order to create a good design. For example, when selecting a directional valve, the engineer can specify either manual or electrically controlled. Any of them will do the job. The proper selection must be based on the solution that will make our system more user-friendly. One of the most important requirements usually overlooked is the noise level requirement. There is increasing health concerns about subjecting workers to a continuous noise level. In 2006, the European Union placed a mandatory noise protection directive (2003/10/EU) for all EU members. This directive set a maximum permitted uniform noise level of 80 dB for the duration of an 8 hours work shift. Considering that the truck operator is also exposed to noise from other sources, we lower the maximum noise level of the hydraulic system to 60 dB measured at the operator’s place of work.


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Hydraulic system with proportional manual directional valve Hydraulic systems for lift trucks have normal high pressures of 15 to 25 MPa. In order to create sufficient power for manipulation of the load, the systems design has to meet the pressure requirements at the worst operating conditions. Operating pressure is determined by the load. The operating pressure must be the same as or less than the rated pressure of the components in the system. Maximum system pressure is determined by the relief valve when the valve is at a fully opened position. Peak pressure is the peak before the relief valve starts to open and it is determined by the reaction time of the valve spring. It occurs for less than 100 ms.

Circuit design The hydraulic system should control different functions (lift/lower, side shift, tilt or others) independently or simultaneously when necessary. Each function is controlled by a directional valve. Valves could be independent components or combined in a valve stack which consisted of different sections. There are two main types of directional valve stacks: valves sections connected in parallel (Fig. 5.3) and valves sections connected in series (Fig. 5.4a and Fig. 5.4b). Valves have a bypass flow path which is closed when fluid is redirected to the actuators. The arrangement in figure 5.3 is the most popular directional valve stack used in the mobile hydraulic systems. For this reason, only this system will be used for further analysis and its operation will be described in details. While the system operation is described, we will also discus the different types of components and determine the ones that best fit our application. The size of the components will be determined after we calculate the systems pressure and flow requirements.

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Fig. 5.3 Schematic of a hydraulic system with an open-center directional valve sections connected in parallel. Components (Fig. 5.3): 1. Reservoir assembly 1a. Breather 1b. Filter 1c. Check valve 2. Flexible line 3. Electric motor 4. Hydraulic pump 5. Pressure relief valve

6. Directional control valve (four sections) 6a. Check valve 7. Flow restrictor (a) with check valve (b) 8. Lift cylinder with safety valve (main lift) 9. Lift cylinder (free lift) 10. Tilt cylinders 10a. Orifice 11. Side shift cylinder


Fig. 5.4a

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Fig. 5.4b

Directional valve with open-center sections connected in series. In the arrangement in figure 5.4a, the upstream sections (sections closer to the inlet) of the directional valve have priority to receive flow. Disadvantage is that when an upstream section is fully shifted, the bypass passage is closed and there is no flow to the downstream sections. An exception exists when the first spool is in lowering position (position 1). During lowering, the flow is available for other functions because lowering is caused by the weight of the load and it does not require pressurized fluid. In the arrangement in figure 5.4b, the return line from sections 2 and 3 is connected to the common line after the valve so that it supplies fluid to the next downstream section.

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Component selection and principle of operation (fig. 5.3) Hydraulic oil is stored in a reservoir (1). The reservoir is a vented type. In the lift truck design we are trying to use the weight of the oil in the reservoir for our benefit. For this reason, the reservoir is either located at the back to act as a counterweight (for counterweight trucks) or as low as possible to enhance the trucks stability. When the reservoir is made from non-metal (composite) material, the filter housing has to be ground to the truck frame in order to discharge the static electricity created in the filter. Breather (1a) is mounted on the top of the reservoir. The breather allows air to enter and exit the reservoir as the fluid level raises and falls. The breather can have a filter or not. Our goal is to design a hydraulic system that is able to work in a contaminated area. Knowing that the level of contamination entering the pump is a critical factor for the life of the whole system, we can install an air breather (1a) with a filter in it. There are advantages for the whole system when filtering the air that enters the reservoir. Using a filter with cellulose media is not popular in the hydraulic systems for lift truck applications. The main reason is that moisture in the air can clog the filter which means that it has to be replaced regularly. If the filter in the breather is clogged, it can create a vacuum in the reservoir and cause cavitation in the pump. Most breather filters are steel mesh type. Suction filter (1b) is mounted inside the reservoir (filter head and fitting ports are outside while the filter cartridge is inside). It is recommended to mount the filter vertical above the fluid level. Oil flows from the reservoir through the filter to the gear pump (4) in a flexible hose. A suction filter has to be equipped with a bypass check valve (1c). This valve protects the pump from cavitation by a providing parallel flow path and ensuring that the pressure in the pump inlet stays above a critical value. The valve starts to open when the filter begins to clog. It is fully open when the filter is fully plugged. The valve also opens when hydraulic fluid has increased viscosity because the truck has been parked in the cold or in sub-zero (freezing) temperatures. The negative side of filters with bypass valves is that the fluid will bypass the filters every time during cold start. In order to avoid this condition, the designer can use clogged filter sensor instead of bypass valve. The clogged filter sensor is a pressure sensor which gives signal to the pump motor (3) and disables the systems operation until the filter is changed. When a pressure sensor is used in addition to the bypass valve,


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its pressure setting should be about one (1) bar lower than the valve setting and its role is only to alert the operator. The suction line is the ideal location for a filter because the filter prevents contamination from entering the system. In some designs, suction filters are avoided because they create pressure drop and increase the risk of forming a vacuum in the pump inlet. A vacuum at the inlet can reduce up to 50% of the life of the gear and vane pumps. To avoid vacuum creation we can place the pump below the oil level in the reservoir. Pump (4) delivers the necessary flow for all consumers in the system. For systems with high pressure and high efficiency requirements, the first choice is an axial-piston pump. For a cost sensitive system, the first choice is a gear pump. In this system we are going to use a fixed-displacement gear pump. When we have an IC (internal combustion) lift truck, the hydraulic pump is mounted on the drive shaft of the IC engine and a dedicated electric motor for the pump is not required. In the electric trucks, the pump is connected to a dedicated electric motor (3). In this system, the pump always starts before the directional valve opens. The reason is that pump has to reach a minimum rotational speed and build up pressure first. Pressure relief valve (5) is built into the directional valve (6). Its function is to limit the maximum pressure in the system. We use an adjustable valve when different applications require different pressures. The spring tension is adjusted in order to provide the desired pressure level in the system. This can be done by the valve manufacturer or by the hydraulic system assemblers. The pressure level depends on the maximum load lifted by the cylinders. A relief valve has two positions that are controlled 1) when the valve starts to open and 2) when the valve is fully opened. At the first position, the valve begins to bypass part of the flow from the pressure line to the reservoir. The pressure at which this happens is called cracking pressure. The cracking pressure must always be more than the working pressure of the system. At fully open position, the valve allows all flow to go back to the reservoir. Usually the full relief pressure is 20% above the cracking pressure. When the same valve or the same hydraulic system is installed on different load- capacity trucks, assemblers on the production floor have to adjust the relief valve. There are two adjustment procedures used to ensure that the valve is fully closed during system operation. The first is when the relief valve is adjusted to the cracking pressure with a load on the forks 10% to 20% higher than the maximum rated load. The second is when the valve is

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adjusted with the maximum rated load on the forks. In the second case, pressure readings are used to set up the valve. First, the assemblers measure the maximum working pressure of fully loaded mast. Then, unloaded mast is extended until it hit the stops and the flow is bypassed through the relief valve. The valve spring is adjusted until pressure before the relief is 10% above the maximum pressure recorded earlier. Correctly adjusted relief valve does not affect the lifting times. Increased lifting times indicates that part of the fluid is bypassed to the reservoir. Lifting time within the design specification indicates that the valve setting is correct. Check valve (6a), prevents backflow from the system to the pump. It is placed before each section of the directional valve. The check valve can also be built into the pump or it can be an in-line type valve. Directional valve (6) used in this hydraulic system is a manually controlled proportional type. Proportional type valves allow smooth flow change and therefore smooth piston motion. The manual control is usually a lever which is moved by the operator. The valve has a stack of four sections built as one module. Each section has three positions- one neutral (middle position 2) and two operational (positions 1 and 3). In the neutral position there is no flow from the valve to the cylinders or from the cylinders to the valve. Sections are spring centered and manually controlled type. Each section has a separate valve spool with a separate control and can act independently from the others. In the neutral position, the pressure line is bypassed to the reservoir. When flow is required, the hydraulic pump starts to deliver flow first and then the plunger moves to redirect the flow through the valve. Shifting the spool opens the ports to the hydraulic actuators and at the same time closes the bypass line. The first section is for lifting and it has only one pressure outlet. Each of the other three sections has two outlets. The fourth section has two quick connect/disconnect connectors for utilization of additional power consumers. The manual control is acting only on one side of the spool by a lever. When the lever is pushed forward, it shifts the spool in one direction. When the lever is pulled back, it shifts the spool to the other direction. When the lever is released, the spring returns the plunger to its neutral position. This type of control allows us to have infinitive positioning of the spool. When we move the plunger slightly, it opens the pressure port partially and divides the pump flow sending part of it to the actuator and the rest of it back to the reservoir. Valve sections are connected in parallel. When all sections are fully open at the same time (require flow in all four


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outlets), the fluid will not go to all of them. The reason for this is that the system has different pressures in the different branches and when the operator shifts two or more sections at once, the fluid always takes the lowest resistance path and goes to the outlet with the lower pressure first. But, if the operator opens the ports, just a little, and creates back pressure from each spool, then two or more sections can operate at the same time. Orifice (7a) role is to restrict the flow from the lift cylinders to the reservoir during lowering of the load. When lifting, the fluid going to the cylinders passes through the check valve (7b). The check vale has very small pressure losses and it allows more energy efficient lifting. When the cylinder is lowered, the check valve (7b) closes and fluid is forced to pass through the orifice (7a). The orifice cross section diameter has to be calculated and properly selected in order to create enough back pressure to achieve the desirable lowering speed. Single-acting ram type cylinders (8 & 9) lift the payload to the designed preset height. Cylinders (8) are called main-lift cylinder and their function is lift the movable sections of the mast. They are mounted on the side or behind of the mast. Both cylinders (8) are connected via a rigid cross bar so that they always work together in tandem. Cylinder (9) is called free-lift cylinder. Its function is to lifts the fork carriage without lifting the sections of the mast. If there is only one free-lift cylinder, it is mounted in the middle of the mast. The piston area of the free-lift cylinder (9) is larger than the combined piston area of both main-lift cylinders (8). Therefore, when oil pressure is applied, the pressure creates higher force (F= p*A) in the middle cylinder and this cylinder starts to lift first. Free lift does not increase the overall truck height and allows the truck to transport the load through doors. This feature also allows faster lift speeds and requires less power. When the cylinder (9) finishes its stroke, the flow starts to lift cylinders (8) together. When the cylinders lower the load, their function is similar to a weighted (gravity) accumulator. The pressurized fluid holding the load accumulates potential energy which has to be released in a way that will absorb the shock waves and provide a smooth lowering piston movement. As we mentioned in chapter 2, when compressing and decompressing large fluid volumes in hydraulic cylinders, we have to consider the compressibility of the fluid factor. Compression of the hydraulic fluid accumulates potential energy similar to the potential energy accumulated in a compressed spring. Fluids that are less compressible absorb less energy. When the fluid is decompressed, it releases this energy creating noise. We have to control the decompression process otherwise the released energy

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can damage the systems components. Instantaneous decompression is controlled by installing a flow restrictor (orifice) (7a) after the hydraulic actuators (the lift cylinders). The restrictor absorbes some of the released energy and converts it into heat. For these reasons, hydraulic systems are designed with as little as possible circulating fluid. Therefore, the system needs to have as small as possible cylinder volumes and short hydraulic lines to the cylinders. Cylinder diameters are selected based on two requirements: 1) hydraulic- provide the desired lifting speeds and 2) mechanical strength- ability to support compression and bending loads without permanent deformation. Lift systems can have a lift speed limit switch that sends a signal to the electric motor (3) in order to reduce the rotational speed when the load is lifted above a certain pre-determined height. Reducing the motor speed in turn reduces the flow rate from the pump and the lifting speed. Safety valve (8b) is built into the inlet of the cylinder assembly (8). This valve is a flow control which limits flow rate out of the cylinder to a predetermined maximum rate in case of failure in the support hydraulics or disconnected hoses to the cylinder. The safety valve, shown in figure 5.2, has two positions. The first position (free flow) has a larger diameter ensuring free passage for the flow entering and exiting cylinder. The second position has two components connected in parallel: a small diameter orifice and a check valve which ensures flow only in one direction. The free flow position of the valve is connected to a valve inlet so that the pressure at the valve inlet acts on the plunger. In additions, there is a spring action on the same side of the plunger. The other side of the valve is connected to the cylinder input so that the cylinder pressure acts on the other side of the plunger. During lifting or controlled lowering, there is little pressure difference at both ends of the valve and the plunger is pushed to the free flow position by the spring. When failure in the system occurs, such as a broken hose or disconnected fitting, the pressure at the disconnected end (valve inlet) becomes almost zero. Then the pressure from the cylinder, acting on one side of the plunger, will create a force higher than the spring force and it will push the plunger in the second position closing the free flow passage. At this point, the fluid exiting the cylinder will be forced to pass through the orifice which will restrict the flow and slow down the cylinder speed to an acceptable safe level. The safety valves (8b) at the lift cylinder port can be replaced with low cost flow restricting (flow limiting) fittings. These fittings are placed at the end of


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the hoses attached to the lift cylinders. They are designed to be inserted into a hydraulic cylinder. The flow limiting fitting has a built-in valve which restricts the flow in the event of a hose failure. The hose failure is detected when the fluid flow exceeds a specified value. This fitting can control the flow in one direction- when the flow exits the cylinder during lowering. The flow is uncontrolled when it enters the cylinder. Two double-acting cylinders (10) are used to tilt the load by tilting either the mast or the carriage. The cylinder body is mounted to the chassis while the piston rod end is mounted to the pivoted mast. These cylinders can have built-in flow restrictors (10a) which limit the speed. Limiting forward tilt speed is very important because a high tilt speed leads to instability of the forklift truck. When the tilt cylinders don’t have built-in restrictors, an in-line orifice or flow control is used to control the tilt forward speed.

Selecting fluid lines and fittings Fluid lines and fittings are selected after choosing all other system components. First, we select the type of connecting line. For mobile system, it is best to use flexible hoses because they work better under vibration and have better dumping properties than metal tubing. Second, we select the inside diameter (ID) of the hoses based on their function, pressure rating and inside flow velocity. Recommended flow velocities are given in Chapter 3.14 (Hydraulic Connectors). Third, we select fittings based on the hose sizes, components port sizes and component’s layout. The fitting ID has to match the hose ID in order to minimize the local friction losses between the fluid and the internal walls. Forth, if there is no available component with the desired port size, a compromised based on minimum pressure loss has to be made. It will require a calculation of local pressure drops using different ID and type fittings to determine the best combination. Pressure drop in the lines is calculated with formula 2.26, or it is determined by using a selection chart. Pressure drop in the hydraulic fittings is calculated with formula 2.28. Often designers make a mistake of sizing the connectors by looking at the component port. This approach usually results in under sizing of the connectors which in turn causes larger pressure drops and more heat generation.

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Selecting hydraulic oil The hydraulic fluid is one of the main factors determining overall system efficiency. Pump efficiency largely depends on the fluid viscosity. When the fluid viscosity is low, pump mechanical efficiency is high but the volumetric efficiency is low because of the increased slip (internal leakage). When the viscosity is high, it reduces the mechanical and improves the volumetric efficiency. Thick oil must be avoided at low temperature conditions. During cold start, the hydraulic components have a reduced lubrication which causes increased contact surface wear. Thick cold oil can also cause cavitation in the pump inlet. The system engineer has to consider the temperature ranges in which the system will be working and select a fluid with the same or wider recommended temperature range. The mean temperature is the fluid temperature under normal operating conditions.

Filter selection The life of the whole system depends on the life of each individual component. The biggest contributor for component failures in the system is the number and the size of contaminants (dirt, dust, hard particles, etc) in the hydraulic fluid. The filter function is to limit the contaminants to an acceptable level. There are two steps of filter selection. The first step is selecting the location. Based on the location, the filter can be: suction, pressure or return. Selecting the location of the filter is based on the considerations described in chapter 3 (filters). The second step is sizing the filter which is done at components selection level.

Filter sizing Filter sizing is based on four main parameters: 1) Maximum flow rate through the filter. For suction filters it is the maximum pump flow rate. 2) Maximum pressure across the filter. Suction and return filters do not experience high pressures. For calculation purpose, we accept that the filter collapsed pressure is at least 1.5 times


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higher than the crack pressure of the bypass valve. Filters without bypass valves should have pressure switches which disable the system when the filter is plugged and yield a “replace filter” message. For a pressure line filter, designers have to select pressure ratings higher than the maximum pressure setting of the relief valve. This higher pressure rating is needed because of the pressure fluctuation in the fluid. 3) Fluid viscosity Fluid viscosity depends on temperature. Temperature range for the filter and the fluid should be the same. If a hydraulic system starts working at cold conditions, increased fluid viscosity puts extra pressure on the filter and can damage filtering element. If this risk exists, the filter has to be tested at cold conditions. 4) Filter retaining efficiency In the past, we worked with “absolute rating” and “nominal rating”. But these ratings did not describe the likelihood of particles larger than the rating to go through the filter. A better measure of filter retention efficiency is the Beta rating (Beta number) which shows the filter efficiency to capture specific size particles. Most manufacturers provide three beta ratings: 75, 100 and 200. Beta ratings (ISO 16889) are a measure of the number of particles greater than X microns entering the filter divided by the number of particles greater than X microns exiting the filter. Different Beta values are given in table 3.2, chapter 3. A procedure for filtration selection in order to maximize the life of a hydraulic system is described below. • Determine minimum component clearance We look at all components in the system and identify the components with the minimum clearances between their moving surfaces. These components require the highest fluid cleanliness. The first step is to collect information for all components, put it in a table and compare the data. There are two sources for this information: 1) component manufacturers and 2) standards. Clearances between moving surfaces is a mirror of quality of the product. It depends on the manufacturing technology, process

144


control and process quality. The component with the least clearance I am going to refer as: the most sensitive component. If we do not have data from the manufacturer, we will determine the most sensitive component based on ISO minimum cleanliness code. For the hydraulic system shown in Fig. 5.3 we will use the ISO guidelines. First we will use tables 3.6, 3.7 and 3.8 (chapter 3) and list all the components and the corresponding cleanliness code for 25 MPa (≈250 bar) pressure. The first number of the code is more relevant to precise servo systems and will be omitted. System components

Minimum Recommended Cleanliness for 25 MPa system pressure

Fixed gear pump

__/16/13

Check valves

__/18/15

Proportional Directional valve

__/15/12

Pressure relief valve

__/17/14

Flow control (Orifice)

__/17/14

Hydraulic cylinders

__/18/15

Table 5.1 From the table 5.1 we see that the most sensitive component to contamination is the directional valve. •

Determine minimum filtration o Select the retention efficiency and the corresponding β number from table 3.2 For example: Select 98.7 % or β (x) ≥ 75 o Compare the recommended clearance for 25MPa system pressure from tables 3.5 with the cleanliness requirements of the most sensitive components from table 5.1 and use the smaller numbers. For example: /15/12 < /16/13 Directional proportional valve cleanliness requirement code 15/12 is smaller and it will be used to determine filtration level. o Select corresponding level of filtration from Table 3.5, column 3. The recommended filtration is 5 µm to 10 µm. A proportional directional valve will require at least 10 micron filter at β (x) ≥ 75. Let’s look for a number smaller than 10 microns (at β (x) ≥ 75) in the


145

catalogue to find a filter that is already available on the market. For example, there is an off-the-shelf filter with ratings: Beta Number

β ≥ 75

β ≥ 100

β ≥ 200

Microns

8.7

9.6

12

Efficiency at β ≥ 200 is also referred to as absolute. Therefore the selected filter has 12 micron absolute efficiency. • Guaranteed system life We know that the time-to-failure is the time from start-up till the failure of the first component. The life of the components is given as the number of cycles to breakdown. So, the time-to-failure of the first failed component will determine the overall life of the system. With a filter having a 12 micrometer at β ≥ 200 we can guarantee that the most sensitive component (proportional directional valve) will be able to reach the “number of cycles to failure” specified by the manufacturer. L10 life of a component (for example L10 = 10,000 cycles) means that 90% of the components are expected to fail after 10,000 cycles of service and 10% of the components are expected to fail before they reach 10,000 cycles. The method for filter selection does not take into account system failures due to poor connections selection, improper assembly, system abuse or operation under extreme conditions.

System Efficiency The ultimate measure for system efficiency is the total pressure drop from the pump to the hydraulic actuators. A large pressure drop in the hydraulic components results in a requirement for higher input pressure. This increases the load on the hydraulic pump and requires more power from the drive motor. Even properly working pump loses its efficiency gradually so, the efficiency loss must be considered then the system is designed.

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Calculations Select hydraulic components Usually the first selected components are the pump and the actuators. From the initially selected maximum pressure and payload we determine the diameter of the lift cylinders. First we determine the diameter of the two main-lift cylinders and then we select the diameter of the free-lift cylinder so that the piston area of the free-lift cylinder is larger than the combined areas of both free-lift cylinders. The cylinder diameters are calculated from the formula:

LT

kASC =

5.1

p MAX η CYL

Where: ASC =

πd 2

[m2]

4

is the area of one cylinder.

k

is the number of lift cylinders

LT [N]

is the total load on the cylinders

p MAX [Pa]

is the maximum system pressure

ηCYL

is the efficiency of the lift cylinder. Ram (plunger) type cylinders have: ηCYL = 0.95 − 0.98

d=

2 LT

πp MAX η CYL

[m]

5.2

Select diameter that is larger than the calculated value. Calculate the work pressure in the lift cylinders. From equation 5.1:

pCYL =

LT

2 ASCη CYL

[Pa]

5.3


147

Calculate flow requirements After we select the diameters of the main-lift and the free-lift cylinders, we calculate the necessary flow rate, Q , for the given lifting and lowering speeds.

Q=

ASCϑ ⎡ m 3 ⎤ ⎢ ⎥ n ⎣ s ⎦

5.4

Where: n = 2 is called polyspast number (or mechanical advantage of pulley operated mast mechanism) and it is the ratio:

n=

V load _ lifting _ speed = C cylinder _ lifting _ speed Vd

(see fig. 5.2)

ϑ = 30 + / − 3 [m/s] lifting speed (given) Knowing the flow rate and the maximum pressure, next we select the hydraulic pump. Pump displacement (d) is calculated from equation 3.4

dP =

Q P ⎡ cm 3 ⎤ n ×ηV ⎢⎣ rev ⎥⎦

5.5

Where: QP [cm3/min] is the flow rate. n [rev/min] is the shaft input rotational speed ηV is pump volumetric efficiency. The next step is to determine the volume requirements of the system and to select the components per these requirements. The components that most affect the required fluid volume are the lift cylinders. As mentioned, for lifting we use single-acting ram type cylinders; for tilting we use differential (double-acting) cylinders; for a side shift we can use either differential or non-differential cylinders.

148


Select the length of the ram cylinder Length of the cylinder, S, depend on the maximum height of the lift mast. Lift mast height is derived from the payload lifting height and its type.

S CYL =

H −a+∆ [m] n

5.6

Where: H [m] is the height of the mast

a [m] is the fork thickness ∆

is the dimension tolerance compensation

n

is the polyspast number. This is a reduction number due to use of a pulley. In most cases, lift trucks have n = 2. Now, after these preliminary calculations and selections are complete, we can calculate all parameters of the system. If we find that some of the systems characteristics are unsuitable, we can go back and change the corresponding component and make the calculations again. In order to find the best balance between system performance, system cost and component availability, we might have to repeat all of the calculations a few times. Very often a new hydraulic system is designed around an existing one. In order to reduce the number of components, very often designers use the same components in different systems. So, finding the right balance that works best for a range of hydraulic systems will require a repetition of the same calculations for different parameters (pressure, flow and losses). The best practice is to use software, which will calculate all parameters automatically when changing any given value. Calculate the pressure losses in the system Pressure losses are calculated after the type of the fluid, and the hydraulic lines are selected. The given is: flow rate (Q), fluid density (ρ), fluid viscosity (ν), hose diameter and lengths and all fittings. There are two types of pressure loss (lineal and local) that were described in chapter 2. Lineal losses occur in the straight sections of the hydraulic tubes or pipes. Local losses are a result of change in the flow speed and direction. Pressure losses due to changes in elevation are insignificant and will be excluded from the total losses.


149

The procedure is: 1. Q V= A

Determine fluid velocity from equation 2.18. ⎡m⎤ ⎢⎣ s ⎥⎦

2. Determine the type of flow from equation 2.11. VD Re =

υ

If Reynolds Number is higher than 2000, it means that there is some turbulence in the flow. In this case we have to change some of the parameters (hose diameter, flow rate or type of fluid) in order to ensure a laminar flow. Having a laminar flow ensures minimum losses and higher system efficiency. 3. Calculate the coefficient of linear resistance for laminar flow, λ [Lambda ] 64 For round cross sections: λ = Re 4. Calculate linear pressure losses from formula 2.25 ∆p LIN

l ρ⎛q⎞ =λ ⎜ ⎟ d 2 ⎝S⎠

2

After we replace velocity V = ∆p LIN = λ

q S

l ρ 2 V d 2

5.7

5. Calculate the coefficient of local resistance for laminar flow, ζ [Zeta ] The values of Zeta are given in table 2.1 (Appendix D). 6. Calculate local pressure losses from formula 2.27. l Where, we replace ≈ 1 and receive: d

150


∆p LOC = ς

ρ 2

V2

5.8

7. Pressure loss of the hydraulic components This is the sum of the pressure loss, ∆pCOM, through all components that the fluid passes while going from the pump to the actuator. For the lifting circuit described in fig. 5.1, ∆pCOM = ∆p 6 + ∆p 7b

5.9

Where, ∆p6 are the losses in the directional valve (6) and ∆p7b are the losses in check valve (7b). Losses in the components at defined pressure and flow are given by the component manufacturers. 8. Total pressure loss from the pump to the lift cylinder is: ∆p SYS = ∆p LIN + ∆p LOC + ∆pCOM Calculate the pressure requirements at the pump discharge port

p PUMP = pCYL + ∆p SYS [Pa] Calculate system power requirements The electric motor power is (equation 2.25): PEM =

p PUMP Q 60η PUMP

[kW ]

Where:

η PUMP is the gear pump efficiency Q [L / min] is the flow rate in liters per minute

5.10


151

p PUMP [MPa] is the pressure at the pump discharge port in Mega Pascal. Calculate torque on the pump shaft To calculate the input from the drive motor torque, we use equation 3.6:

T=

pd P η m ( Nm ) 2π

Where:

p is pressure at the pump outlet dP [cm3/rev] is pump displacement ηm is pump mechanical efficiency.

Next, the electric motor is selected from the power and torque requirements of the system. A calculation of the parameters of the hydraulic system, described in fig. 5.3 and using the elevating section shown in fig. 5.2, is given in Appendix F.

Load lowering speed One of the requirements of the system is to maintain constant lowering speed (within +/- 10%) regardless of the load weight. The hydraulic restrictor (7a), fig. 5.3, creates backup pressure and reduces the lowering speed but it does not maintain the same lowering speed for different loads. Constant speed requirement can be met by connecting a flow regulator in the return line. The next example shows the difference between a flow regulator and an orifice when they are used to control the lowering speed of the lift cylinders (8) in fig. 5.3. Fig. 5.5 shows the pressure drop - flow (∆p – Q) characteristics of a flow regulator and an orifice during the lowering of the payload. Flow regulator characteristic has an initial steep flow increase due to the spring force in the counterbalanced valve (valve is described in chapter 3).

152


After the pressure drop ∆p reaches 2 MPa, the valve closes and maintains relatively constant flow. When the pressure drop across the valve increases, the valve has to compensate more which produces a tailing off effect. The tailing off effect tilts the graph up or down off its horizontal line. Since constant flow maintains almost constant lowering speed, when a constant lowering speed is required, we use a flow regulator. The flow regulator does not maintain a constant flow only in the first part of the graph for pressure drop where (∆p) < ∆pmin. The orifice characteristic is a quadratic relationship between the pressure drop (∆p) and the flow (Q). When the load is increased, ∆p increases as well. The valve has a high flow rate at high pressures (maximum load on the forks) and a low flow rate at low pressure (no load). Orifice valve characteristic provide high lowering speed when the lift truck is fully loaded and low lowering speed when it is empty. The reduced lowering speed reduces the overall truck productivity. The only advantage of using orifice is its simple construction and low price. For this reason, orifice is used to control lowering speed only in systems where the low cost of the system outweighs the increased productivity. 60

50

Flow (l/min)

40

30

20

10

0 0

Flow Regulator Orifice

5

10

15

Pressure drop B->A (MPa)

Fig. 5.5 Flow rate comparison

20

25


153

Hydraulic system with electrically controlled valves In the last decade, electrically controlled proportional valves are slowly replacing both discrete type valves and manually controlled proportional valves. This trend is a result of a number of factors such as ergonomic requirements, software development, new electronic components, new sensors and CAN bus technology. An example of a system with electrically controlled proportional directional valves (6a and 6b) is shown in Fig. 5.6.

Fig. 5.6

154


Components: 1. Reservoir assembly 2. Flexible line 3. Electric motor 4. Hydraulic pump 5. Pressure relief valve 6a. Proportional directional valve for lift and lower 6b. Proportional directional valve for tilt 6c. Proportional directional valve for side shift 7. Check valve 8. Flow control 9. Lift cylinders 10. Tilt cylinder 11. Side shift cylinder 12. Check valve

Principle of operation The main difference between this system and the first one (Fig. 5.3) is that the directional valves (6a, 6b, 6c) have electric controls. Lift and tilt valves (6a and 6b) are proportional type valves controlled by proportional electric signal while the side-shift valve (6c) has an ON/OFF control. Unlike the mechanical type where the control lever is mechanically linked to the valve, electrically controlled valves are connected to the controller with wires. Connecting components with wires allows more design freedom because the valves can be placed anywhere on the truck. Electrically operated valves can be equipped with a manual override. A manual override can be a screw or push-pull button. It is used to: 1) reset the plunger in its neutral position when the plunger is stuck as a result of contamination. Easy plunger movement shows that a plunger is not seized; 2) activate the valve and lower the load when power to the valve is lost. In the hydraulic system shown on Fig. 5.6, the lift/lower directional valve (6a) is equipped with manual push button.


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A control lever (joystick) that includes buttons and switches can be used to control the valves. Any movement of the joystick results in corresponding movement of the work mechanism. Using a joystick design allows designers to develop a more ergonomic operator’s compartment. Solenoids of the proportional valves can be controlled either directly by a potentiometer or by an electronic controller. The potentiometer is usually placed inside the joystick. Joysticks can have a mechanical potentiometer or solid state potentiometer (based on inductive technology) to translate mechanical motion into electric signal output. In the last few years the trend is to use joysticks that have a CAN bus interface. CAN bus technology allows electronic communication with a minimum number of wires and contactors. When a programmable controller is used, the potentiometer sends a signal to the controller which in turn sends a signal to the solenoid that controls the valve plunger. The position of the plunger determines the flow rate through the valve. The flow rate controls the position and movement of the work attachment. In this case, the joystick is configured only to deliver commands to the controller. Operators can activate two or more functions by pressing two or more buttons simultaneously. The controller can be programmed in such way that some commands are restricted when they are in conflict with commands already taking place. These restrictions are used for safety reasons. There are two basic types of valve controls: electric and electronic: An electric control is when the valve solenoid is controlled by changing the input voltage by continuous analog signal directly from the control handle (joystick). A potentiometer changes the value of the signal proportionally to the handle movement. The solenoid armature motion is proportional to the input voltage. Armature pushes the valve spool. Usually, proportional valves have a voltage or a current control. Discrete (ON-OFF) valves have a voltage control. Voltage is turned on and off by electric switch. An electronic control is when a programmable electronic device controls the input signal to the valve solenoids. Electronic control can be an analog or a digital signal. Digitized signals have definite number of steps. For example if the digital controller has eight-bit signal, there will be 256 (28 = 256) steps available. The resolution of eight-bit controller will be 0.4% (1/256 =

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0.0039). Modulated (PWM) digital signals are used to control to proportional solenoids. PWM signals are described in this chapter. ON/OFF signal are used to control solenoids for discrete valves. Voltage is a result of current and resistance. When the valve heats, the valve coil resistance changes which leads to a voltage change. Valves that need fine adjustments are controlled by changing the current because it does not depend on the temperature

Signal modulation Pulse Width Modulation (PWM) is a method of controlling the signal. The signal is controlled by a transistor. The transistor is an ON-OFF switch which controls the average value of the signal by controlling the width of the signal (the time signal is ON). Modulation time is the time (measured in per cent) when the signal is on. One hundred percent (100%) is practically a continuous signal and the voltage is equal to the maximum.

Signal [Volts]

Signal [V]

100% modulation

50% modulation

V = 24 V = 12 t [s]

t [s]

Switching frequency is the rate at which the signal is turned ON and OFF and it is called PWM frequency. PWM frequency must always be higher than the valve resolution. Then, the spool is too slow to respond to the instantaneous ON-OFF switching and it responds to the average value. Valve resolution is the smallest increment of input signal (input frequency) at which the valve spool can respond. The current to the valve solenoid will be: I = V/R (amp) Where: V is PWM average voltage and R is coil resistance


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Using an electronic control we can have pre-set (programmed) rates of spool travel, acceleration and deceleration. Advantages of electronically controlled systems are the flexibility of changing the system’s performance based on specific conditions. Different parameters correspond to different conditions. For example, when lifting and lowering speeds are controlled by software parameters, we can change them by simply changing these parameters. Also, we can adjust the speed according to outside conditions such as: surrounding temperature, truck speed, lift height and others. Valves can be manufactured as cartridge type valves and assembled into a common block (manifold). For the system shown in Fig. 5.4, we can assemble the three proportional valves (6), the relief valve (5) and the check valve (12) in one manifold.

Dither signal Dither signal is a high frequency input signal. Dither signal is added to the control signal in order to achieve smooth spool motion. The dither makes the valve spool vibrate at high frequency. This vibration makes the spool movement smoother by diminishing sticktion in the valve. It is important to choose a correct frequency and amplitude dither signal. Dithering reduces the valve hysteresis and makes valve performance more consistent. The minimum dither frequency is part of the valve specification and it is provided by the manufacturer. Excessive dithering can negatively affects the static and dynamic characteristics of the valve.

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Hydraulic System with Independent Emergency Lowering When the load is raised, the directional control valve can fail so that its plunger can not be switched into the lowering position. In this case the load has to be lowered by other means. In the previous examples, directional valves with a manual override were showed. The manual override is a manually operated control to switch the valve in lowering position. Another solution for lowering the load is adding a line between the lift cylinder and the reservoir. This additional connection is called emergency lowering circuit. Independent lowering circuit is used more often in hydraulic systems with electric spool type directional valves. Valves with electric controls are more susceptive to contamination than valves with manual controls. When the coil is energized, valve spool is magnetized and it attracts iron particles from the oil. For this reason, hydraulic systems with electric proportional valves have higher fluid cleanliness requirements. Spool valves with positive lap design (see fig. 3.7) are more likely to fail as result of iron contamination than poppet type valves. Two systems (5.7a and 5.7b) with emergency lowering will be described. In the first system (fig. 5.7a), the emergency lowering valve is independent from the lift/lower valve. Independent load holding is required for lift operator platforms or cabins. The lowering valve is always manually controlled. It can be a directional type or screw-in type needle valve. The screw type needle valve is more cost efficient. The needle valve has a small orifice with a tapered seat and a needle-like plunger which closes off the orifice when it is in the screwed-in position. This valve allows precise regulation of the flow rate because it has a fine-threaded screw and takes many turns to move the plunger from a fully open to a fully closed position. The needle valves can either be opened with a lever or allen wrench. Using a wrench is more inconvenient but it makes the valve fool-proof against unintentional actuation. When the valve can be activated by pushing/ pulling a button or lever, the valve must be located out of the operator’s reach during his/her normal


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work functions in order to eliminate the risk of activating it by mistake. Emergency lowering circuits are added when the main directional valves are not easily accessible.

Fig. 5.7a Hydraulic System with an emergency lowering valve connected in parallel to the main control valve In the second system (fig. 5.7b), the lowering valve has a manual override which is used for emergency lowering. The lowering valve is normally closed, ON-OFF type with electrical control. When the solenoid is deenergized, the flow can go only in one direction: from the priority valve to the lift cylinder. When the solenoid is energized, the valve is shifted and the flow goes to the return line.

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CF

EF

LS P

electrical lowering valve with manual override

Fig. 5.7b System with lowering directional valve which has a built-in check valve and a manual override for emergency lowering

Energy recovery systems Vehicle performance is measured by the length of time a truck maintains a normal handling operation. This length of time depends on two aspects: the amount of energy that the vehicle can carry and the efficiency of the hydraulic system. In the electric trucks the energy is stored in the vehicle battery while in the internal combustion (IC) engine truck the stored energy is in the fuel. In many truck configurations the battery (or fuel tank) size is limited to the available compartment space. That is why a goal for engineers is to design an energy efficient hydraulic system. An efficient


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system improves the fuel consumption in IC trucks or increases the battery life on electric trucks. Each hydraulic system consists of two groups: 1) a power source (motor and pump) that converts mechanical or electrical energy into hydraulic energy and 2) a system of components that transports the hydraulic energy and converts it back to mechanical power. When thinking about the efficiency of the power source, the first thing to consider is the type of motor that drives the hydraulic pump. There are two types of motors: 1) electric motors and 2) IC engines. Even a small increase in efficiency of the motor results in increased efficiency of the system and yields huge energy savings. When running continuously at or almost at a full load, the annual energy cost for running a motor can be up to 20 times its purchase price. In this regard, the increased price of a highefficient motor will be an insignificant factor in the motor selection decision. When energy efficiency of the drive motor is one of the design requirements, the following data has to be considered: • • • • • • •

Efficiency of an AC motor is about 96% Efficiency of a DC motor is about 90% Efficiency of an IC engine varies from 36% to 40% Lift motors operate less than 50% of the time Steering motors operate 100% of the time Electric motors for hydraulic systems are independent from the drive motors. Therefore, their speeds can be controlled per the hydraulic system flow requirements. IC engines run 100% of the time because they have to support all systems at the same time.

This chapter describes ways for increasing system efficiency by recovering hydraulic energy. Energy recovering systems use the energy of the returning fluid during the lowering process. The potential energy of the raised load is converted to hydraulic energy. Weight of the load is used to drive pressurized fluid through the hydraulic motor or to supply fluid to other hydraulic functions. Then the hydraulic motor drives a generator which produces electric energy and charges the battery. There are three main types of recovery systems (fig. 5.8).

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Fig. 5.8 a) unidirectional pump and flow


b) Reversible pump and flow

c) Separate pump and motor

1. Unidirectional pump and flow (fig. 5.8 a). The hydraulic pump is rotated only in one direction. The return fluid is connected to the pump inlet. Pump construction has to allow pressurized fluid at both ports. By connecting the return flow from the lift cylinder to the pump input the fluid can be supplied to other actuators while lowering the load. 2. Reversible pump and flow (fig. 5.8 b). Rotation of the pump is in one direction for lifting and is reversed during lowering. 3. Separate hydraulic pump and motor (fig. 5.8 c). The pump is used only for lifting the load while the motor is used only for energy recovery. Both pump and motor can be connected to the same motor/generator (shown on Fig. 5.8 c) or to separate electric machines (not shown). When they are connected to the same electric motor, there is a one-way clutch between the motor and the generator which allows torque transmission only in one direction: from the motor to the generator. In the next example, a system with one way pump and flow is described. The returning from the lowering cycle fluid is utilized in order to generate electric power. The returning flow is used for two functions: to supply other hydraulic functions and to drive the hydraulic pump/motor. A schematic of the hydraulic system is shown in Fig. 5.9.


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2

3

1

2

1

3

Fig. 5.9

Principle of operation Hydraulic pump/motor (4) is coupled with an electric machine (3) that can work as a motor or a generator. The electric machine (3) always works in the same direction. The relief valve (5) is placed between the pressure line and the return line. The return line has a return filter (2) with a check valve connected in parallel. This hydraulic system has three main branches that support three functions: lifting/lowering, tilt and side shift. Directional valves (6 and 7) have controls independent from each other which allows the operator to activate each of them separately or at the same time. Each three-position directional valve has a neutral middle position. The middle

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position has a bypass passage which ensures free flow from the pump to the reservoir. There are two ways to control the speed of the lift cylinder: flow resistance control and pump control. With flow resistance control, the flow rate to the cylinder is controlled by changing the opening of the proportional valve (6 or 7). When the proportional valve is partially open, it allows the excess flow to go for auxiliary functions. Pump control is when the flow rate is controlled by changing pump rotational speed. Pump controlled lifting is more energy efficient because there are less hydraulic losses. Disadvantage of it is that it has a longer response time. The unique feature of this system is the lowering branch configuration. One end of it is connected to the lift cylinder (11) inlet and the other end is connected to the suction line between the check valve (8) outlet and the pump inlet. This way, all returned fluid during lowering goes through the pump. Flow rate of the return fluid can be controlled either by the pump/motor (4) or by a flow control (9). In the case of motor controlled lowering, valve 9 and 10 can be replaced with one proportional directional two-position valve (not shown). The motor has a speed sensor (not shown) which sends actual rotational speed to the controller (not shown). The flow control (9) has a constant flow through the valve at different pressures in the lift cylinder. This constant flow provides lowering speed that is independent from the load. It also limits the rotational speed of the pump (4) and the electric machine (3) because pump rotational speed is proportional to pump volumetric flow. When the return fluid is not needed for another function, the pump/motor acts as a hydraulic motor and the electric machine works as a generator and recharges the battery. The fluid goes thought the bypass passages of the valves (7) back to the reservoir. When fluid is needed to move tilt (12) or side shift (13) cylinders, one or both valves (7) shift and redirect the flow to the auxiliary functions. Then, part of the flow goes to drive the auxiliary cylinders while the unused part is bypassed to the reservoir. If the tilt or side-shift actuators require more flow and pressure than supplied from the return line during lowering, the electric machine works as a motor and rotates the pump to create the necessary flow. In this case, in addition to the return fluid, the pump pulls out fluid from the reservoir.


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In the energy recovery modes the return fluid is used for other functions before going to the reservoir. If the fluid is too hot, it will have a negative impact on its life and on the efficiency of the whole system. For this reason, reservoir design that ensures proper cooling is essential for this design.

Hydraulic Steering System Steering system function is to provide power for manipulation of one or more wheels of the truck. Lift trucks’ power steering can be: electric, hydraulic and electro-hydraulic. This book will describe only hydraulic and electro-hydraulic systems. A hydraulic steering system can stand alone with its own power source (pump and motor) or it can be integrated with another hydraulic functions system (fig. 5.13). The stand alone system can have separate reservoir or it can share the reservoir used for other functions. A hydraulic steering system for lift trucks is usually a medium pressure system- from 5 to 15 MPa (50 – 150 bar) As we stated, before designing the steering hydraulics, we have to specify the system requirements. The system requirements (design requirements) for this system are: • • • • • •

Maximum pressure in the system = 8 MPa (the pressure requirement is calculated from the steering forces requirements) Revolutions of steering wheel that turn the truck wheels from full left to full right position: N = 3 to 5 revolutions Maximum revolutions per minute of the steering wheel = 90 [rev/min] Maximum system noise level: 50 dB Temperature range: from -25°C to +80°C Fluid cleanliness: according ISO 4406

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Steering system schematic All hydraulic steering systems must have a hydraulic pump, a steering valve connected to a steering wheel and an actuator which moves the steering mechanism. The actuator can be a cylinder which converts the hydraulic flow into linear motion (Fig. 5.10) or a motor which converts hydraulic flow to rotational motion (Fig. 5.11). The type of actuator depends on linkage design of the mechanical steering mechanism. The system usually has its own filter that is placed either in the suction line before the pump (Fig. 5.10) or in the return line (Fig. 5.11).

Fig. 5.10 An example of a stand alone steering system with a hydraulic cylinder as actuator for the steering mechanism Components: 1. 2. 3. 4. 5.

Reservoir Filter with check valve Hydraulic pump Steering valve, type: Orbitrol (with built-in relief and check valves) Steering wheel


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6. Double-acting cylinder (fig. 5.10) or Hydraulic motor (fig. 5.11) 7. Sprocket or gear (fig. 5.11)

Fig. 5.11 An example of a steering system with a hydraulic motor as actuator for the steering mechanism

Component selection and principle of operation (fig. 5.10) The first component that has to be considered in this system is the hydraulic cylinder (6) Fig. 5.10. It can be either single or double acting. A double-acting cylinder is preferred because it allows simplified geometry of the steering mechanism which reduces the number of mechanical links. The cylinder can be stationary mounted onto the truck chassis or the steering axle. Piston rods are connected to the wheels (not shown). Piston stroke is defined by the kinematics of the steering mechanism. Cylinder selection is determined from the piston rod strength requirements, the piston size and volume requirements.

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Steering valve (4) is selected on the basis of the steering wheel number of revolutions. Our goal, per design requirements, is four revolutions. The valve has to ensure that four full turns of the steering wheel will result in turning the truck wheels from full left to full right position. The steering valve has a built-in relief valve. When steering is performed, the relief valve is fully closed and all flow is circulated though the Orbitrol. When steering stops, the pressure increases and the valve opens. Then all flow from the pump goes through the valve to the reservoir. Some steering control valves have a return-to-center feature. The valve returns the steering control to neutral position when it is released by the operator. This returns the tires automatically to their center position. Steering wheel (5) is mechanically connected to the Steering valve (Orbitrol). Filter (2) is placed in the return line. It must have a check valve, connected in parallel, which opens when the filter is plugged. The check valve will ensure a continuous flow from the pump to the reservoir when the filter fails. Continuous flow is necessary to guarantee truck steering. Reservoir (1) is a vented type. It can be separated or combined with the reservoir of the hydraulic system for the lift/lower and the auxiliary functions. Pump (3) is a fixed displacement pump. Calculating the pressure requirement Let us look at the system in Fig. 5.10. Pressure in the system is a result of the steering force FS [N] acting on the piston in the cylinder. Steering force is determined by the moment which is necessary to turn the wheels. The pressure inside the cylinder is:

pCYL = Where,

FS 1 [ Pa] ACYL ηCYL

5.11


ACYL

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D2 − d 2 =π [m2] is the cylinder area on which pressure is acting. 4

D is the piston diameter d is the piston rod diameter

η CYL

is the cylinder efficiency

Then, the required pressure in the pump outlet is:

p PUMP = p CYL + ∆p H

5.12

Where,

∆p H

are the pressure losses in the hoses and the fittings from the pump to the cylinder. Calculating the flow requirement A.

Selecting Orbitrol displacement:

VCYL ⎡ cm3 ⎤ dS = N ⎢⎣ rev ⎥⎦

5.13

Where,

B.

VCYL

[cm3] is cylinder volume

N

is number of steering wheel revolutions

dS

[cm3/s] is steering valve displacement

Selecting steering pump flow rate:

QSP = Where,

N MAX (d S ) ⎡ l ⎤ ⎢⎣ min ⎥⎦ K

Q SP 5.14

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K = 1000

is a coefficient to convert cubic centimeters in liters.

NMAX

is the maximum number of revolutions per minute.

The maximum recommended speed is:

N MAX = (1.5 − 2) N A

5.15

NA is the average rotational speed of the steering wheel

Example: Determine the steering valve size and the required pump flow rate for a steering system with a double-acting cylinder with piston diameter D = 80 mm and rod diameter d = 56 mm. The mechanism kinematics require a cylinder stroke S = 140 mm. Design requirements specify from 3 to 5 steering wheel rotations from full left to full right turns.

Solution: Cylinder volume V in cubic centimeters is:

VCYL =

π (D2 − d 2 ) 4

S=

π (80 2 − 56 2 )10 −4 4

[ ]

(140)10 −2 = 358.9 cm3

⎡ cm3 ⎤ 358.9 = 89.7 ⎢ dS = ⎥ 4 ⎣ rev ⎦ We select Orbitrol. For example: Danfoss 80 which has a displacement of 80 [cm3/rev]. Using this valve the steering wheel rotations will be:

[ ]

VCYL 358.9 cm 3 N= = 4.5 [rev] = dS ⎡ cm3 ⎤ 80 ⎢ ⎥ ⎣ rev ⎦


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Revolutions = 4.5 is within out design specification. Knowing that an average operator turns the wheel with a speed of 60 revolutions per minute (NA = 60), the maximum rotational speed will be:

N MAX = (1.75) N A = 105 Then, the steering pump flow rate will be:

QSP =

N MAX (d S ) (105)(80) ⎡ l ⎤ = = 8. 4 ⎢ K 1000 ⎣ min ⎥⎦

Flow rate together with motor rotational speed are used to select the pump displacement. The displacement is calculated from equation 5.5. The process of pump selection is described in chapter 3 and in the example in Appendix F (Hydraulic System Calculation).

Electro-Hydraulic Steering System Microelectronics development in the 1980’s and the 1990’s allowed the electro-hydraulic steering systems to match the performance and manufacturing cost of traditional hydraulic steering. An example of an electro-hydraulic steering system is shown in Fig 5.12. In this system a proportional directional valve (4) replaces the Orbitrol type steering valve. The proportional valve (4) controls the flow to the hydraulic cylinder (5). Both ends of the cylinder piston rods are connected to steering mechanism (not shown) which rotates steered wheels (not shown). Steering mechanism converts the piston rod lineal movement to a turning motion of steered wheels.

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Steering control (7) rotation is monitored by an angular sensor (8). The sensor converts the rotational input parameter to a digital output signal (pulses) and it sends the signal to a controller (9). The controller controls the plunger movement of the proportional valve (4) by sending voltage (or current) signal to the solenoid. Proportional valve regulates the flow rate which in turn determines the speed of the cylinder (5) piston rod. There is a position sensor (6) which sends information about the actual position of the hydraulic actuator (the lineal piston rod speed of the rod). The actual value is fed back to the controller (9) in the form of voltage signal. This is a closed loop circuit because the controller receives required and actual parameters for the hydraulic cylinder’s movement and adjusts its output to a pre-calculated value. The controller has digital input and output features so that it can accept input signals from electronic sensors (6 and 8) and send output signals to the solenoid of the proportional valve (4). The most common angular sensor for this application is the rotary encoder. The encoder outputs information about steering wheel rotation such as: rotation angle, turning speed, total numbers of turns and angular acceleration. There are two types of rotary encoders: external and internal. An external encoder has better resolution but it requires an adapter housing which makes it a more expensive solution. An internal encoder is built inside a bearing. This design is more compact and provides more design freedom for the steering wheel location. In order to achieve smooth steering, it is important to minimize the positional hysteresis of the proportional valve (4). Positional hysteresis depends on friction inside the valve, magnetic hysteresis, temperature, fluid pressure and cleanliness. To minimize the negative influence of these factors, a high frequency noise (called dither signal) can be added to the control signal. The dither makes the valve spool vibrate which diminishes the valve stiction. A proportional directional valve has a high requirement for fluid cleanliness. When a return filter is used, for achieving maximum life of the system, it is recommended to have a 100 to 150 micron absolute mesh suction filter and a 10 to 15 micron absolute return filter.


Fig. 5.12 Electro-hydraulic steering system Components (fig. 5.12) 1. Reservoir 2. Wire-mesh filter (100 µm absolute) 3. Hydraulic pump 4. Proportional directional valve 5. Double-acting cylinder 6. Sensor 7. Steering wheel 8. Angular sensor (rotary encoder) 9. Programmable Controller 10. Filter (10 µm absolute) with bypass check valve

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Integrated Hydraulic System The hydraulic system, shown in Fig. 5.13, combines the hydraulic (lift and auxiliary) and the steering circuits in one. In this design, one common motor, pump and reservoir are used. A load sensing valve splits the flow giving priority to steering over load manipulating functions. The system can have either fixed or variable displacement pump. In electric industrial truck applications, fixed-displacement gear pumps combined with variable-speed motor is the most popular solution. In IC engine trucks, variabledisplacement, pressure-compensated pumps are preferred because the IC engine has a constant rotational speed. The distinguishing feature of the integrated systems is the use of a load-sensing priority valve to split the flow into two different circuits. The valve senses the flow requirements and sends the required flow to the steering branch and the rest of the flow to the other branch. When the steering valve is turned, it pressurizes the load sensing (LS) line to the priority valve. The pressure in the LS line shifts the plunger to provide flow for steering. The benefit of the priority valve is that it can maintain different working pressures in both branches. In an integrated system with priority valves, the excess flow (that is not needed for steering) is supplied to the second branch. When there are no flow requirements this flow has to be bypassed to the reservoir with minimum energy losses. There are two ways: first is to use a relief valve with external control which keeps the valve open when flow in the second branch is not needed (fig. 6.4 and fig.6.5) and second is to use a opencenter directional valve (fig. 5.3) where the pressure line is bypassed to the reservoir in its neutral position. It is a common practice for the high lift truck operator to perform two or more functions at the same time. In this case, both branches require flow at the same time. The sum of the required (in both branches) flow rate will be the maximum flow of the system. Therefore, we size the pump for the maximum flow. Although the pump is sized for the maximum flow, to save energy, we design the system to work continuously with the flow which is required only for steering. Maximum flow is supplied only when both functions are simultaneously required.


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Fig. 5.13 Integrated Hydraulic System When the integrated system is for electric trucks, designing an efficient system requires use of an electric motor and an electronic motor speed controller. By controlling the motor speed we can change the flow rate. There are two commonly used motor types for this application: 1) separate exited DC motors and 2) AC motors. The simplest motor controller for the DC motors has two different fixed speed settings. The low speed (1000 to 1500 rpm) is used continuously for steering. The high speed (2000 to 3000 rpm) is used when functions from the other branch requires flow.

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Smoothness of the Lifting One of the most important requirements for high lift trucks is the smoothness of the lifting process. Often, when lifting at a low speed, vibrations in the hydraulic cylinder cause a jerky motion of the whole elevating system. In applications when light and unstable loads have to be lifted high, any small thrust or vibration can make the load fall down. An example of such a load is a pallet with empty pop cans. Any excessive shaking of the elevation system would cause this light and unstable load to fall down from the pallet resulting in lost of production time. For carrying and lifting loads like this, smoothness of the lifting is crucial for the entire forklift performance. There are many factors: electrical, mechanical, hydraulic, operator skills, which can cause jerkiness in the elevating system. In this chapter we will describe only the hydraulic aspect of the problem and show few ways of achieving a smooth performance of the lift mast. High lift systems have either single-acting telescopic or single-acting ram type cylinders. The weight of the load and mast does the work in the opposite direction. Ram type or telescopic cylinders are designed for applications where long strokes are required. Both type cylinders are described in chapter 3.9. Performance requirements for these cylinders are: • • • •

Compact design High work pressure (usually 25 MPa) Operate in wide temperature range (-20 to +50º C) Maintain static and dynamic sealing

Factors, influencing the lifting process, are: A. B. C. D.

Driving (lifting) vs. resistance forces Pressure and flow pulsations in the cylinder inlet Compressibility of the fluid Dynamics of the directional valve

A. Driving vs. Resistance forces During lifting, the lifting force ( Fd ) that drives the cylinder piston (or plunger) acts against the resistance force ( Fr ) which consists of the friction force,


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load reaction and the inertia of the moving parts. Moving parts are the mast, cylinders, carriage, the load and all chains and hoses attached to the mast. The lifting force is:

Fd = pA η Where:

A is the piston area p is the pressure in the cylinder η is the efficiency of the cylinder

The resistance force contains three components.

Fr = FL + Fi + F f Where:

FL

is the Static Force of the load

Fi

is the Force of Inertia of the moving parts with mass (m) and acceleration of the moving parts (a).

Ff

is the Friction Force

Friction force itself has two components.

F f = Fcyl + Fmech Where:

Fcyl

is the friction inside the cylinder

Fmech is the friction between the moving parts of the elevating system One of the reasons for jerkiness in the system is the different rate at which the lifting force and friction forces are changing during the lifting process. Prof. Komitovsky (Components of Hydraulic and Pneumatic Systems) explains the origin of this jerkiness in the relationship between friction force and piston speed, Fig 5.14

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Fig. 5.14 In the first speed range -I-, the friction force (FT) has its highest value. In the second range -II-, FT is almost interchangeable and has its lowest value. Within this speed range the hydraulic cylinder has its best performance in terms of smoothness of motion and efficiency. When piston speed is in the third range, we observe a gradual increasing of FT but the slope of the curve is not as sharp as at the beginning. Let’s look in detail at the first section of this graph because the most vibrations appear in this speed range. After we open the directional valve, the flow starts moving the piston. When piston speed reaches a value v1, friction force is FT1. If a pulsation from the gear pump appears at this moment, additional kinetic energy will be added to the fluid entering the cylinder. This momentary kinetic energy increases the piston speed to v2, which in turn decreases the friction force to FT2. When the speed reaches v2, the momentary energy finishes and the friction force increases to FT1, reducing piston speed to v1. When a new pulsation appears in the cylinder, this process will be repeated and will make the piston speed vary between v1 and v2. This speed variation of the cylinder piston (or plunger) causes the jerky motion of the elevating system. There are two general ways to reduce the piston speed variation. The first way is to minimize flow variations (increase the “spring constant” of the system) by adding a hydraulic pulsation damper to the system. A damper is a hydraulic resistance between the direction valve and the lift


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cylinder. There are two types of damping: active and passive. Active uses a programmable controller which uses a cylinder position feedback signal to control the directional valve plunger’s movement. This method is expensive and is not used for forklift truck applications. Passive damping uses flow restriction to reduce pressure oscillations in the system. The most popular and cost effective damping element is the flow restrictor. On the other hand, its use is not energy efficient because it increases the hydraulic losses by converting hydraulic energy into heat. Another more efficient solution is using a pressure compensated flow control valve. The second way to minimize the piston speed variation is to reduce the mechanical friction between the piston and the internal cylinder surface. The most effective way to reduce friction is to use sliding non-metal rings. The use of non-metal sliding rings can reduce the friction force up to ten times, which comes from the difference in the coefficient of friction, f.

f teflon ≈ 0.01 ,

f metal ≈ 0.10

The friction force is:

Fcyl = nf teflonπDLp Where: n is the number of sliding rings, usually n = 2 D is the outside diameter of the ring L is the width of the ring p is the pressure in the cylinder Two main materials are used to manufacture sliding rings- teflon based and thermoplastic polyurethane materials. Optimal design is a combination of a seal, a scraper and a sliding ring. B. Pressure and flow pulsations entering the cylinder Flow and pressure variations in the pressure hoses are important to consider. In addition to pump induced flow pulsations, there could be shock waves after the directional control due to opening and closing of the valve. Pumps flow rate irregularity is:

δQ =

Q − Qmin ∆Q × 100 = max × 100 _[%] Qm Qm

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Where,

Qm is the mean value.

Pressure pulsations in the discharge port are created by the gear pump. Every time the fluid, situated between two teeth, is pushed out of the pump, a peak in the pressure appears. Pressure waves in the discharge port are described in Chapter 3.7 (Hydraulic Pumps). C. Compressibility of the fluid When calculating the hydraulic system we treat the fluid as incompressible but in reality it is not. Compressibility is characterized by Modulus of Elasticity (Bulk Modulus) EV. The higher EV is the stiffer and less compressible the fluid is. Fluids which have higher Bulk Modulus absorb less energy. Modulus of Elasticity depends strongly on the fluid temperature because when the temperature rises, the fluid expands creating additional pressure. Ev decreases when temperature increases. Experimental data are shown on Fig. 2.2, Chapter 2. Also, there could be small amounts of air (aeration) in the fluid. Aeration reduces the value of EV and makes the fluid more compressible. At high pressure, the air forms bubbles. When the pressure drops air bobbles are released causing cavitation, vibration and noise. Usually the first step in troubleshooting a jerky cylinder motion is to bleed-off the cylinders. For this reason, all high lift cylinders have bleed-off plugs or fittings. Some systems have automatic air bleed valves which allow air to escape from the hydraulic line without bringing the truck in for service. D. Dynamics of the valve Fluctuation of both flow speed and pressure decrease with the increase of the internal frequency of the directional valve. Higher frequency limits the amplitude of the relevant variables velocity and pressure. Speed fluctuations decrease with the increase of the switching frequency of the valve. Pressure fluctuations also decrease with the increase of the hydraulic capacity of the cylinder. Switching time of a normal solenoid valve is in the range of 30 to 50 ms. Servovalves have a switching frequency of 10 ms from fully open to fully closed position.


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Chapter 6

Hydraulic Systems for Low Lift Trucks There are three main configurations low-lift trucks: end rider, center rider and walk behind (fig. 6.1 and 6.2). According to the Industrial Truck Association (ITA), low lift trucks are class 2 (code 6 only) and class 3 trucks. ITA classification is shown in Appendix A. The main feature of the low-lift trucks is that the fork attachment (fork frame), which is designed to carry the load, is elevated only 150 mm (about 6 inches) above the floor. Therefore, these trucks do not require cylinders with long strokes. They are also called pallet trucks or trucks for horizontal transportation. In addition to the load, the fork attachment can carry the truck battery or mechanical lift mechanisms. Low-lift trucks are manual or electric powered because they are used inside warehouses and stores. There is no difference in the hydraulic lift system requirements between class 2 (code 6) and class 3 low-lift trucks. There are three types steering systems: • manual (mechanical) steering • power-assisted steering • power (hydraulic or electrical) steering When the trucks have manual or power-assist steering, the steering control is attached to the steered wheel (fig. 6.1 and 6.2). Trucks with power–assist steering use hydraulic torque generator (item 9, fig. 6.2). In trucks with hydraulic power steering, the control is detached from the wheel. These systems use a steering valve type “Orbitrol” (fig. 6.3a through 6.7). Power steering systems design and the orbitrol are described in chapter 5 (section: Hydraulic Steering Systems). Designs with torque generators are shown in fig. 6.11 and 6.12. The steering control can be a control arm (also called: tiller arm), steering wheel or joystick. Lighter pallet trucks have a tiller arm mechanically attached to one or more steered drive wheels (fig. 6.1). Heavier trucks have electric or hydraulic power steering (fig. 6.2).

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Chapter 6: Hydraulic Systems for Low Lift Trucks

The hydraulic steering system can be either a separate or integrated (with the lifting) circuit. Both types of hydraulic circuits will be described in this book. In the past only rider type trucks had power steering. In the last few years, driven by ergonomic reasons, manufacturers are adding power steering to walk behind trucks as well.

Fig. 6.1 Low-lift truck with manual steering and mechanical lift mechanism.

Fig. 6.2 Low-lift truck with power-assisted steering and hydraulic lift mechanism.


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Low-lift pallet trucks have four main applications: • loading/ unloading pallets from trailers • horizontal transportation of the load • low level orderpicking • stocking

Hydraulic system with independent power steering and lift circuits A two-circuit system that has independent power steering and lift hydraulic circuits is shown in fig. 6.3. Power steering is used on high-capacity trucks where large turning moment is required to turn the steered wheel. The steering circuit is independent and it has its own power supply. The steering valve is type Orbitrol. There is no mechanical connection between the steering control and steered wheels. Usually, the steering control is a wheel which makes 3 to 5 rotations from full left to a full right turn of the steered wheel. The steering circuit is the same design as described in Chapter 5. The calculation of the steering branch is also shown in Chapter 5. The lifting circuit generates power only for lifting. It does not have auxiliary functions such as side shift, tilt and attachments. The lifting circuit has few hydraulic components and it is relatively simple. For this reason it is constructed as a one power pack with four main components: a reservoir, an electric motor, pump and a valve block. The valve block consists of a manifold with hydraulic valves in it.

An advantage of systems with two circuits is that the lifting circuit can be an off-the-shelf power unit. Almost all hydraulic valve manufacturers offer such power units, which contain an electric motor, a hydraulic pump, a manifold with valves and a reservoir. When selecting a power unit, we look the combined performance of the pump and the electric motor from the manufacturer. Lift cylinders (12) usually are single acting. They are

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selected depending on the type of lifting mechanism and maximum load capacity of the truck. The lifting mechanism can have one or more cylinders (12).

Fig. 6.3 a) steering circuit; 1. 2. 3. 4. 5. 6.

Reservoir Filter Electric motor Hydraulic pump Check valve Relief valve

b) lift circuit 7. Steering valve, type Orbitrol 8. Hydraulic motor 9. Gear 10. Flow control 11. Directional valve 12. Lift cylinders

Disadvantages of two circuits system are: • •

Two circuits require two tanks, two electric motors and two hydraulic pumps which take up more space. Cost of the system is higher because more components are used.


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The steering circuit for low-lift trucks with a capacity below 3,000 kg is a low pressure system- up to 5 MPa (50 bars) and for trucks with capacity above 3,000 kg it is a medium pressure system- from 5 to 15 MPa (50 – 150 bars). The lifting circuit always has a normal high pressure from 15 to 25 MPa (150 – 250 bars).

Integrated Hydraulic Systems for Low Lift Trucks An Integrated System (Fig.6.4) is a system designed to support both steering and lifting functions. This design has two branches. Both branches have common motor, pump and reservoir. An integrated system which uses a load sensing priority valve (LSPV), item 4, as a flow divider is described below. LSPV can be: • • •

In-line type- the valve is built as a separate unit. It can be placed anywhere on the truck. Cartridge type- screwed into a manifold and combined with other hydraulic components into one power pack. Modular type (direct mounting)- assembled to the steering valve (Orbitrol) or the pump

A hydraulic system with in-line priority valve is shown in Fig. 6.4.

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Fig. 6.4 Integrated Hydraulic System Components: 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11.

Reservoir Suction filter with check valve Pump Load sensing priority valve (LSPV) Lift power pack Lift cylinders Motor Steering valve, type Orbitrol with LS feedback Steering control Hydraulic motor Gear


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Principle of operation (fig 6.4) A priority valve (4) is used in systems with two or more loops. One of the circuits needs to have a controlled flow rate regardless of load pressure changes. The valve has one inlet and two outlets. The outlet which supplies the controlled flow is marked as CF (controlled flow). The excess flow (EF) is delivered from the second outlet. The important characteristic of the priority valve (4) is that it can maintain different pressures at its outlets. Relief valves, installed in each circuit, limit the maximum pressure in the branch. In this system, the relief valve in the lift circuit is built into an assembly (5) and the steering relief valve is built into the steering valve (8). The relief valve in the lift circuit also has an external electric control which can keep the valve open when the solenoid is energized. The valve is kept open when there is no request for lifting. This way, it is used as a bypass valve. It allows the excess flow, not used for the steering, to go to the reservoir. When lifting is requested, the valve solenoid is de-energized and the valve closes. Then, it functions as relief valve to protect the system. One of the biggest restrictions for class 3 trucks is the small available space. The most space effective solution is when the priority valve is a cartridge type built into a hydraulic power unit (fig. 6.5). Such power unit contains: hydraulic manifold with cartridge valves, a reservoir, a pump and an electric motor.

Fig. 6.5 Priority valve built into a hydraulic power unit.

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Class 3 trucks usually have a fixed-displacement gear pump (3) and an electric pump motor (7). The motor (7) can be either constant-speed or variable-speed motor. Variable-speed motors are controlled by electronic controller. Systems with controllers are more energy efficient when variable flow rates are desired. The rotational motor speed matches the flow rate requirements of the system. When lifting is not required, the electric motor rotates at a low speed and supply a lower flow rate to support only the steering. When lifting is needed, the controller increases the motor speed, which results in a higher pump flow rate which is needed to support both: steering and lifting. System with constant speed motor (fig. 6.6)

Fig. 6.6 Hydraulic system with dual pumps and a priority valve


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In some cases the cost of the electronic controller is too high and it is more cost efficient to connect the electric motor directly to a battery. Then, the motor will have only a high speed and it will deliver the maximum flow at all times. When only steering is required, a lot of the power will be wasted. For this reason, when a one-speed motor and a fixed-displacement pump are used, energy efficiency can be achieved by using a dual pump (3) to split the flow. Both pumps work continuously but only the flow from one will circulate through the system to support steering. The flow from the second pump is used only when a lift is requested. If these is no lift request, the flow is bypassed through a valve (8) and returned to the reservoir (1). When more flow is needed for lifting, the directional valve (8) switches and redirects the flow to the priority valve (4).

Integrated hydraulic system with accumulator An integrated hydraulic system with accumulator is shown in fig. 6.7. In addition to the pump, this design uses an accumulator as a source of energy for the steering. The accumulator first stores energy then supplies it back to the system when needed. This circuit allows turning off the motor when the accumulator supplies the energy. Using accumulator makes the system more energy efficient. The accumulator sizing is shown in example 1 and the energy saving is shown in example 2. This design is feasible only for low-lift trucks because of the short lifting time (3 to 5 seconds) and the fact that the steering and lifting are rarely used simultaneously. Steering control is connected to a steering valve type “Orbitrol” (10). The valve and the control are detached from the steered wheel (not shown).

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Fig. 6.7 Integrated hydraulic system with accumulator and an intermittently working pump Principle of operation of the system in Fig. 6.7 The integrated system shown on Fig. 6.7 has different pressure requirements in both branches (for example: 14 MPa for steering and 24 MPa for lifting). A pressure switch, 8, set at 14 MPa, will limit the pressure in the steering branch. There is also a second relief valve (at 14.5 MPa crack pressure) which is built into the orbitrol, 10. The purpose of the second relief valve is to protect the steering branch in case of a pressure switch failure. A relief valve (5) (set at 25 MPa) limits the maximum pressure in the lifting branch.


191

When the operator steps on the dead-man pedal, the parking brake is released and the motor (7) is turned on. Then, the pump (3) starts to supply flow to the system. If the accumulator (9) is empty, it will take a few seconds to fill up. If there is a steering request before the filter is filled up, the flow from the pump will go to both the orbitrol (10) and the accumulator (9). As soon as the accumulator (9) is fully charged and the pressure reaches a pre-set value, the pressure switch (8) is activated and it turns off the electric motor. From this moment, the accumulator will supply the oil flow for the steering. When the accumulator is discharged to a pre-set minimum, the pressure will drop and the pressure switch (8) will turn on the electric motor. As mentioned above, steering and lift functions are rarely used simultaneously therefore this design does require a variable speed motor. Pump flow rate is selected to be a little more than the maximum required for steering. This way even during steering, there will be extra flow for charging the accumulator.

Example 1 Selecting an accumulator for the system shown in fig 6.7 Given (Engineering Specifications): • • • • • •

Steering valve (orbitrol) displacement = 80 cm3 per revolution Steering wheel turns = 4 revolutions for 180° turn from full left to full right position (2 rev. for 90°turn) Pump flow rate: Q = 6 (l/min) Maximum pressure (relief valve settings): pmax = 14 MPa Minimum pressure: pmin = 10 MPa Accumulator must be able to supply the steering valve for a period of time equal or more than the lift time (4 seconds)

Calculate accumulator volume If there is lifting and steering at the same time: for lifting time of 4 sec, we have 4 revolutions of the steering wheel (angular velocity n = 1 rev/s). So, the volume demanded by the orbitrol will be:

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∆V = (dS)(revolutions) ∆V = (80)(4) = 320 сm3 Where: ∆V = V3 – V2 is the difference in the volume of a fully charged accumulator V3 (maximum pressure) and uncharged accumulator V2 (minimum pressure). Then, the size of the accumulator, V1, is calculated by using formula 3.13

V1 = ∆V

⎛ p3 ⎞ ⎜⎜ ⎟⎟ ⎝ p1 ⎠

1/ n

⎛p ⎞ 1 − ⎜⎜ 3 ⎟⎟ ⎝ p2 ⎠

1/ n

1 / 1.4

⎛ 10 ⎞ ⎜ ⎟ 9 = 320 ⎝ ⎠ 1 / 1.4 = 1615 [cm3] ⎛ 10 ⎞ 1− ⎜ ⎟ ⎝ 14 ⎠

Where: p2 = 14 MPa is the maximum pressure in the steering circuit (fully loaded accumulator); p3 = 10 MPa is the minimum pressure in the system; p1 = 0.9 p3 = 9 MPa is the pressure in the bladder when the accumulator is no loaded. Based on the above calculation we are choosing a standard size of 2000 сm3

Example 2 Calculate the energy saving Calculate how much time the motor has to work in order to support the steering during one work shift. Assume that one work shift consists of 400 TWC (truck work cycles). Solution An example of TWC is shown in fig. 6.8. In one TWC, the truck goes from the load/unload area to the trailer (forks first), takes the load and drives back (tractor first). The truck will make six turns (three in each direction).


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Fig. 6.8 Steering duty cycle that requires three 90° turns in each direction It is known that one 90° turn requires two steering revolution. One (TWC) has six 90° turns = 12 steering wheel revolutions per truck work cycle or n1 =12 rev It is estimated that in an 8-hour work shift an average operator makes a maximum of 200 TWC (N = 200 cycles). Therefore, the total number of steering wheel revolutions per 8-hour work shift is:

nT = n × N = 200 × 12 = 2400 rev We are going to use 2000 cm3 accumulator (selected in example 1). The usable volume ∆V will be:

⎛ p.3 ⎞

1 1.4

1−⎜ ∆ V := V .1⋅

⎝ p.2 ⎠

⎛ p.3 ⎞ ⎜ ⎝ p.1 ⎠ ∆V = 396 cm3

1 1.4

The steering revolutions (nA) per one accumulator charge with the chosen 2000 сm3 accumulator will be: nA = ∆ V / ds = 2000/80 = 5

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where: ds = 80 cm3/rev is the orbitrol displacement The number of accumulator charges per 8-hour shift (200 TWC) will be: Nc = nT / nA =2400 / 5 = 480 charges Charging time (t) for one accumulator charge is: t = ∆ V/ QP = 4 sec,

where: QP = 6 l/min is the pump flow rate

So, the total charging time per 8-hour work shift will be: T = t x Nc = 4 x 480 = 1920 seconds = 32 minutes Therefore the electric motor will work only 32 minutes to support the steering. In systems without accumulator this time is about 7 hours because the electric motor runs continuously regardless whether steering is required or not.

Hydraulic system for pallet trucks with long fork attachments As described earlier, pallet trucks are used to lift and carry pallets. The fork attachment moves up and down relative to the tractor. Long forks are designed to transport three or four grocery pallets at the same time. Four pallets (in line) require forks that are 4880 mm (192 inches) long and have up to 4500 kg (10,000 lb) load capacity. Such forks must have high strength and little deflection. The strength calculations of such fork are the same as of a structural beam. The strength requirements are achieved by designing a beam with high section modulus. The section modulus is the ratio of the cross-section second moment of area to the distance of the surface to the neutral axis. The most effective way to increase the fork section modulus is to remove the mechanical lift links (push/pull rods) from the fork cavity and design the fork profile as a closed rectangular tube with reinforcing bars inside. In addition to having high strength, the fork attachment is requirement to move up and down parallel to the floor surface during lift and lower. Short attachments have a linkage mechanism for pivoting the load wheels downward when the forks are raised.


195

For long attachments such design is not efficient because of the increased cost and lack of space needed for the links. Also, the mechanical links have to be different for each fork length. To eliminate these downsides, pallet trucks with long forks have hydraulic cylinders pivoting the load wheel downward. Such hydraulic lift system is shown in fig. 6.2 (component layout) and in fig. 6.9 (schematics). The system shown in fig. 6.9 has two main cylinders (5a and 5b) mounted on the tractor and two slave cylinders (6a and 6b) mounted within the fork profile and attached to the load wheels.

Fig. 6.9

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Lifting (Fig. 6.2, 6.9 and 6.10) The main cylinders (5a and 5b) lift the fork assembly at one end while the slave cylinders (6a and 6b) lift the forks at the other end (see fig. 6.2). When the main cylinders extend, the upward movement of the fork is equal to the main cylinder stroke. Main cylinders are connected by hard mechanical link so that, they will always move together (they will have same strokes). Although both load wheels are mounted in separate forks, they are hydraulically linked to the main cylinders therefore they also work in tandem. When one of the load wheels hit a barrier, which stops the wheel movement, the other wheel will also stop. Cylinders (6a and 6b) are called “slave” because their stroke depends on the amount of flow coming from cylinders (5a and 5b). One end of each slave cylinder (the cylinder body) is pivoted to the fork frame while the other end (piston rod side) is attached to a link containing the load wheels (fig. 6.2 and 6.10). When a slave cylinder extends, the link rotates and the load wheel is pushed against the floor which in turn lifts the front end of the fork (fig. 6.10). Lowering The weight of a battery or/and the fork frame (when the battery in mounted on the tractor) is used to lower the forks. Directional valve (7) is energized to direct the flow to the reservoir. Lowering speed is controlled by a flow regulator (8). Because of the short strokes, non-proportional flow regulators are used.

A1

A2

Fig. 6.10 (only 5a and 6a cylinders are shown for clarity)


197

The relationship between cylinder areas and cylinder strokes is: A1 Y = A2 X

Where:

6.1 A1

is the rod area of the main cylinder

A2

is the piston area of the slave cylinder

X

is the main cylinder stroke

Y

is the slave cylinder stroke

Hydraulic Power-Assisted Steering Class 3 pallet trucks usually have one steered drive unit. The drive unit contains a drive motor, a gear box and a drive wheel. In most trucks, the steering control is a tiller arm. Power-assisted systems use a hydraulic actuator which links the steering control to the drive unit. Three common designs are shown in fig. 6.11. All three have hydraulic torque generator as an actuator. • • •

tiller arm control and drive unit with horizontal motor (6.11a) tiller arm control and drive unit with vertical motor (6.11b) steering control with steering ratio rs ≠ 1 (6.11c)

a) steering ratio rs = 1 Fig. 6.11

b) steering ratio rs = 1

c) steering ratio rs≠1

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Components: 1. 2. 3. 4. 5. 6.

Hydraulic torque generator Steering control (tiller arm or steering wheel) Steered drive unit Brake Drive motor Steered drive wheel

Steering ratio (rs) is defined as: rs = Nc / Nw,

6.2

Where: Nc is the number of revolution of the steering control Nw is the number of revolution of the steered wheel Steering ratio equal to one (rs = 1) means that the control (2) and the steered wheel (6) rotate at the same time and at the same angles. The hydraulic torque generator is activated mechanically by a steering control. The rotation from the control (2) to the steered drive unit (3) is transmitted through the input and output shafts of the torque generator (1). Both shafts are mechanically linked inside the generator. The steering control (2) is connected to the input shaft while the steered unit is connected the output shaft. Connection to the shafts can be direct (6.11a) or indirect by gear sets (6.11b and 6.11c). When both gear sets have the same gear ration (6.11b), the design has a steering ratio equal to one (rs = 1). By changing the gear ratio, different steering ratios can be achieved. The biggest advantage of using a hydraulic power steering over an electric steering is that the truck does not lose steering if a failure in the power system occurs. If there is no supply of pressurized fluid to port P, only manual steering will be available. Manual steering will be more difficult but will ensure control over the truck.


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Integrated Hydraulic System with PowerAssisted Steering An integrated system, shown in fig. 6.12, has two circuits: a power-assisted steering and a lift. This arrangement is used on class 3 pallet trucks where the steering control (6) and the steered drive unit (9) must have the same turning angles (rs = 1).

Fig. 6.12 Components: 1. 2. 3. 4. 5. 6. 7. 8.

Pump Motor, electric Priority valve with LS port Relief valve Gear Tiller arm Gear, ring Torque generator with LS port

9. Steered drive unit 10. Reservoir 11. Filter with check valve 12. Relief valve 13. Directional valve 14. Flow regulator 15. Lift cylinders

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Principle of operation (fig. 6.12) Class 3 trucks usually have a fixed-displacement gear pump (1). The electric motor (2) can be either a constant speed DC or a variable speed AC. Power steering is achieved by the torque generator (8) which converts hydraulic energy into mechanical torque on the output shaft. The generator has the input and output shafts and two hydraulic ports: inlet (P-port) which is connected to the pump, the outlet (T-port) which is connected to the reservoir and two mechanical shafts (input and output). Turning of the input shaft performs two functions: first, it transmits torque to the output shaft and second, it allows pressurized fluid from the pump to flow into "P" port. The pressurized fluid transmits additional torque to the output before going out through "T" port and returning to the tank. The torque generator has a load sensing (LS) port connected to a priority valve (3). When the torque generator is rotated, the LS port is pressurized. LS pressure acts on valve (3) and shift it to supply the steering branch with fluid proportional to the LS pressure. A load sensing priority valve (3) senses the flow requirements and supplies the required flow to the steering branch with priority. Priority valves are described in details in chapter 3.6 (Directional control valves). The directional valve (13) has three positions. The left and right positions are for lift and lower. The middle (neutral) position is open-to-tank. This way, when steering and lifting are not required, the directional valve (13) allows all flow to go to the reservoir. Thus, the system will not be pressurized when it is not needed. The flow regulator (14) has a parallel connected check valve. When lifting, fluid passes through the check valve. During lowering, the flow passes though the flow regulator.


201

Chapter 7

Hydraulic Systems for Boom-Type Trucks Boom-type reach vehicles are used mainly for off-road outdoor material handling on unfinished or uneven terrains. These trucks are class 7 according to ITA classification (see Appendix A). A typical truck arrangement is shown in Fig. 7.1. A class 7 truck has a telescopic lifting arm (2) one end of which is pivotally mounted to the truck (1) and is supported by a hydraulic cylinder (3). The other end is extendable and it has a hydraulically driven attachment. The telescopic arm (also called a boom) can have one or more extensions. Each extension is operated by a hydraulic actuator. There is an attachment mounted at the end of the inner extension. The attachment can either be a carriage with a pair of forks (6) or a lifting mast (not shown) for multi-directional handling. The fork carriage is pivotally mounted so that it can be tilted, extended or retracted relative to the boom arm. A tilt cylinder (5) is placed inside the inner boom. When the fork attachment is a mast type (not shown), the mast provides a vertical fork carriage movement without extending the boom. It can also have a side shift or a tilt function for the fork. 4

5 2 1 3

Fig. 7.1 Boom-type vehicle with a fork carriage at the end.

6

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Chapter 7: Hydraulic Systems for Boom-Type Trucks

The boom arm is pivoted at the rear side of the truck. It moves up or down using the hydraulic cylinder (3). One end of the cylinder is attached to the vehicle chassis and the other to the boom. Another hydraulic cylinder (4) extends and retracts the boom. Cylinder (4) can be placed inside or outside on the outer telescopic arm. Examples of hydraulic circuits for lift mast fork attachments were shown in chapter 5 (Hydraulic Systems for High Lift Trucks). In this chapter, we are going to describe only hydraulic circuits for boom manipulation.

Hydraulic circuit for boom lift, extend and fork tilt The main difference between a boom type truck and an in-door lift truck is the work duty cycle of the material handling. A truck with a boom is designed to move while carrying the load high in the air. When the truck moves, the load jumps on the forks and causes oscillation of the boom and the chassis. This oscillation can create an overloading condition which will have a negative effect on the truck stability. Also, the load oscillation interferes with the vehicle control, adds stress to the chassis and causes the truck body to deflect. Overloading is determined on the basis of 1) magnitude of the load and 2) length of the extension arm. Such conditions can be prevented by limiting engine power and acceleration. In addition to reducing engine performance, dumping the vibrations, created by the bouncing load, is achieved by connecting an accumulator to either the lifting or lowering side of the cylinder. Overloading can be monitored by sensing the pressure in the lift cylinder. Cylinder pressure will increase proportional to the load and arm length due to changes in the lever conditions. The sensing is obtained by pressure sensors. An example of a hydraulic system for a material handling vehicle with a telescopic arm is shown in Fig. 7.2.


203

Engineering Requirements for this system are: • • • • • •

Pivoted outer boom section At least one extendable inner section Each section to be operated by a dedicated hydraulic cylinder Boom cannot be lowered uncontrollably Boom to be protected against lowering in case of a broken hydraulic line. Vibration absorber when riding with the load up position.

Fig. 7.2

Hydraulic lift & lower circuit for telescopic boom arm The lift arm (boom) is almost horizontal at its low position. One end is pivoted to the chassis while the other end is lifted by a hydraulic cylinder. There are different circuit designs for boom lifting. Most of them use an accumulator as a vibrations dumper. Two examples of hydraulic circuits showing boom lifting and lowering are shown in Fig. 7.3 and Fig. 7.4.

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Fig. 7.3 Hydraulic circuit for lifting and lowering a boom-type arm 1. Reservoir assembly 2. Hydraulic pump 3. Motor/ IC engine 4. Flexible line 5. Pressure relief valve 6. Directional control valve 7. Directional valve with a check valve 8. Directional valve 9. Accumulator 10. Double-acting cylinder (item 6 in fig. 7.1) 11. Load


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Description of the components Reservoir assembly (1) consists of a reservoir with a breather and a suction filter, with bypass check valve, placed inside the reservoir. Most boom-type vehicles have IC engine. The reservoir of these vehicles has to be insulated from the IC engine because the engine is an undesirable heat source. Usually the IC engine and the reservoir are placed in separate compartments. Hydraulic pump (2) is usually a gear pump which is mounted on the drive shaft of the IC engine. IC Engine (3) can use diesel, petrol or propane fuel. Flexible lines (4) are an essential part of the system. They are required in order to allow boom pivoting movements and to transport fluid in the most efficient way. Hydraulic lines are sized and selected on the basis of the maximum flow velocity and pressure. Tube and hose selection is described in chapter 3.14. Pressure relief (5) valve is required to protect the hydraulic system from overloading. Directional control (6) valve is a three position two directional type valve. The first position is for lifting, the second position is neutral and the third position is for lowering. The directional valve can be operated manually, hydraulically or electrically. An electrically controlled valve is shown in this design. When there is no input signal to the solenoids, springs at both ends hold the valve in its neutral position. Directional valve (7) is a two-way, two-position discrete valve. It has a check valve in position 1 and an orifice in position 2. Directional valve (8) is a proportional two-way, three position valve. Accumulator (9) is connected to the lift cylinder immediately before the cylinder port. This way we eliminate the need for a flexible hose which in

206


turn minimizes the risk from breaking the connection. In most cases, metal tubing is used to connect the accumulator to the cylinder. Double-acting cylinder (10) is connected pivotally to the truck body on one side and to the boom on the other side. A single-acting cylinder can also be used. Load (11) consists of the weight of the boom, the boom attachments and the maximum payload. Description of the hydraulic circuit operation During lifting: the directional valve (6) is in position 1, valve (7) is deenergized and it has the check valve section connected to the line allowing oil flow to the cylinder. During lowering: the direction valve (6) is in position 3 and valve (7) is in position 2 forcing the fluid through an orifice in order to create a back pressure and smooth the lowering. During lifting or lowering, valve (8) is not energized and the spring force keeps the valve in closed position. Then, the accumulator (9) is disconnected from the system because the built-in check valve does not allowing flow to the accumulator. During transportation (with or without a load): the control valve (6) is switched to the neutral position, valve (7) is switched to position 1 preventing flow to the reservoir and valve (8) goes to position 3 connecting the accumulator to the lift arm support cylinder. Now, the pressure in the accumulator (9) will support the weight of the boom and will act as a vibrations dumper. When the load bounces, it forces the fluid from the hydraulic cylinder into and out of the accumulator. It is necessary for the accumulator to have the same pressure as the lifting side of the cylinder at the time it is connected to the system. To avoid pressure spikes, valve (8) is constructed as a proportional type. The valve solenoid is energized proportionally and it moves the plunger from position 1 to position 3 while going through the orifice (position 2). Switching at a slow rate allows the system to equalize the pressure before the valve is fully open. If the pressure in the accumulator is lower, a sudden lowering of the boom may occur. To avoid this effect, a small size accumulator is preferred for this application.


207

Hydraulic circuit for boom arm with an automatic shut-off valve

Fig. 7.4 Hydraulic circuit (with an automatic shut-off valve) for lifting and lowering a boom 1. Reservoir assembly 2. Hydraulic pump

7. Check valve

3. Motor

8. Directional valve, proportional

4. Relief valve- main

9. Accumulator

5. Relief valve

10. Cylinders

6. Directional valve

11. Load

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A hydraulic circuit with an accumulator and an automatic shut-off valve is shown in Fig. 7.4. The automatic shut-off valve consists of a pressure relief valve (5) and a check valve (7). The relief valve has two control lines connected to both sides of the cylinder. One line is connected to the lowering side and the other to the lifting side of the cylinder. During lifting (control valve 6 is in position 1), the spring maintains valve (5) in a closed position and the fluid goes through the check valve (7). During lowering (control valve 6 is in position 3), the check valve is closed. Pressure opens the relief valve (5) and the fluid passes through the valve to the reservoir. Relief valve (5) controls the lowering speed of the cylinder (10). When the pressure under the piston increases, it reduces valve (5) opening which slows down piston movement. During transportation with load, the control valve (6) is in neutral (position 2), valve (8) opens and connects the accumulator to the cylinder.

High-speed extension of telescopic boom Productivity of the truck increases when the time of a loading/unloading cycle is reduced. Faster extension of the boom is one way to increase truck productivity. A hydraulic circuit for high-speed boom extension is shown in fig. 7.5. The hydraulic actuator, extending the boom, is a double-acting cylinder (4) shown in fig. 7.1 and fig 7.5. The cylinder speed is load sensitive. Design Requirements: •

Fast extension, when the boom is unloaded.

•

Slow extension, when the boom is loaded.

Principle of operation Directional valve (1) has three positions. Position 2 is neutral. When the valve is in position 1, the cylinder extends and lifts the boom. When it is in position 3, the cylinder retracts. When the cylinder (4) is extending an unloaded boom, the pressure under the piston will be low and the relief valve (6) will be closed. Then, the return


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flow (Q1) will pass through check valves (5) and join the flow from the pump (Qp). Combined flow (Q1 + Qp) will go through the check valve (3) to the cylinder (4). When the boom is loaded, the pressure under the piston increases and the valve (6) opens. Then the return flow from the cylinder will go through the valve (6) to the reservoir (T). During cylinder retraction, the flow from the pump goes through the check valve (7) and enters the piston rod side of the cylinder. The return flow passes through the pressure-compensated flow regulator (2). The flow regulator (2) will maintain a constant lowering speed regardless of the load.

Fig. 7.5 The extension speed of an unloaded boom is:

υE =

QP + Q1 ⎡ m ⎤ ⎢⎣ s ⎥⎦ A1

7.1

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Where: Qp [m3/s]

is flow rate coming from the pump

⎡ m3 ⎤ Q1 = A2υ E ⎢ ⎥ ⎣ s ⎦

is flow rate coming from the cylinder

A1 = A2 =

πD 2 4

[m ] 2

π (D 2 − d 2 ) 4

is the larger piston area

[m ] 2

is the cylinder rod side piston area

After we replace Q1 in equation 7.1, the extension speed can be express as:

υE =

QP ⎡m⎤ A1 − A2 ⎢⎣ s ⎥⎦

7.2


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Chapter 8

Selected Topics I.

Servicing Hydraulic Systems

In the last few years, the service sector has grown rapidly as more services are offered by manufacturers and companies are focusing more on the quality of the service. High quality service leads to customer perception of a high quality product, which results in higher customer satisfaction and more orders. The American Management Association has estimated that companies lose as many as 25 percent of their customers each year because of poor customer service. Also, service is a very big revenue generator because it has higher turnover rate and higher profit margins than the original equipment. Studies show that having loyal, long-time customers can increase profitability by 100 percent even without increasing the market share. For these reasons, all organizations have service departments with trained professionals whose main objection is achieving maximum customer satisfaction. Service is an activity that does not create a new product. Its main function is to repair, maintain or increase the life of the existing equipment. For hydraulic systems, service means: 1) monitoring the system parameters within pre-defined limits; 2) repairing and replacing failed components and 3) preventive maintenance such as filter and fluid change. Three service requirements must be met in order to achieve a service excellence: Efficient - repairs done correctly the first time. Effective - repaired or replaced parts to last a satisfactory length of time. Economical - quick and cost efficient repairs. Service excellence can only be achieved on the basis of good designs for service and clear service procedures.

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Chapter 8: Selected Topics

Troubleshooting principles In order to perform effective troubleshooting, a number of hydraulic principles must be known and followed. These principles are: •

Hydraulic pumps create flow - not pressure

•

Resistance to flow creates pressure

•

Flow rate determines actuator speed

•

Pressure determines actuator force

•

Fluid under pressure takes the path of least resistance

•

Fluid movement from one place to another always results in pressure drop (pressure loss) and heat generation

System life As mentioned in the introduction of this book, one of the aspects of a good system design is: design for service and inspection. The main goal of this approach is to increase the overall life of the system and maintain a high efficiency throughout its life cycle. The expected life of individual components within a hydraulic system varies and is influenced by a number of factors. These factors include the type and construction of the component, circuit design, operating load and dutycycle. Forklift manufacturers determine the expected service life of components within a particular system by considering these variables in combination with historical data on achieved service life. Component life is normally available from the manufacturers upon request. This information is typically provided to long-term customers who have their own service departments. To minimize the chances of hydraulic components failing during service, the system manufacturers recommend expected service life. The service life is used for scheduling component replacements. All system components (valves, pumps, hydraulic lines) are flushed in order to have a certain cleanliness level. But regardless this fact, in the first hours after start up, the components continue to release small particles. In order to protect the system from the initial contamination, it is recommended the first filter change be done after 50 hour of operation.


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Safety Rules • • • •

Disconnect the power source (electrical and /or IC engine) before servicing any component. Lower or support moving parts before disconnecting any component. Put all controls in neutral position to release system pressure. Personal protection- ware safety glasses, safety shoes and proper work uniform.

Servicing the fluid Theoretically the new fluid is considered clean but in reality it is contaminated to some degree during transportation and installation. Most fluid contamination during service is from dirt, water or abrasive particles (metal or sand) entering the fluid. Metal particles not only increase the components wear but they are also activate catalysts in oxidation reactions. The most common particles are: iron, copper, tin, aluminum, zinc and lead. Dirt and the micro-glass filters are the main sources of silicon. It is very difficult to determine the optimal fluid change period. It is generally accepted that if the hydraulic system works well, the fluid can last up to two years. However, when the truck works at extremely cold or very humid environment this period can be as little as six months. In general, the fluid change is always done in two cases. • After every failure When a hydraulic component fails, metallic particles from this component are dumped into the hydraulic fluid. Therefore, the hydraulic fluid has to be changed at the same time as the failed component is changed. • Scheduled change Periodic fluid change is a preventive maintenance activity. It is a part of the hydraulic system long-term reliability recommendations. As described in Chapter 3, the change period should be determined on the basis of the statistical data of the contamination rate. This period can be different for the same systems that work in different environments. Regardless of how good the statistical data is, the fluid has to be inspected regularly for the presence of sludge. Sludge (mud) in hydraulic oil can significantly reduce the service life of the oil. One percent sludge in inhibited hydraulic oil reduces the service life by nearly 40%.

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Collecting fluid samples The best fluid samples are taken when the machine is running and the fluid is at operating temperature. It is recommended to take samples downstream of the hydraulic components and before the filter. If the samples are taken from the reservoir, take them from the mid level before the suction pipe. One oil sample shows the picture at a given moment of time. If we want to see the rate of change of a component’s wear, we have to monitor the oil over a period of time. It is known that all oil becomes darker over time because of oxidation. The important characteristic to record and understand is the rate of oxidation. Fluid tests • RPVOT test (per ASTM D2272) RPVOT (Rotating Pressure Vessel Oxidation Test) is a test that determines the oxidation stability of the oil. RPVOT measures the actual resistance to oil oxidation. Results from the test are compared to the test results of the new oil. The new oil base line can be used to convert RPVOT readings to remaining useful life (RUL) as a percentage of the new oil life (100%). Typically the caution limit is at 40% RUL and the critical limit is at 25% RUL, but this varies by application. This test is recommended for large fluid volumes that have long change periods and severe-duty applications. • FTIR (Fourier Transform Infrared Spectroscopy) test FTIR can detect chemicals and oxidants. Data is collected and converted from an interference pattern to a spectrum which allows this test to be computerized. • Acid number test This tests method is standardized in ASTM D 3339. It covers the determination of acidic constituents in petroleum-based products and lubricants. • Viscosity test The manufacturers usually specify the kinematic viscosity. When viscosity of the fluid is measured with a viscometer, two measurements are recommended: one at room temperature and one at operating temperature. Then, we compare the data to the manufacturer’s spec. The new fluid can vary within certain limits due to density change (see Appendix C). Fluid degradation, oxidation, contamination and other events cause the change in density. When fluid viscosity is reduced 20% or more, the fluid has to be replaced with new fluid.


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Abbreviations TAN (Total Acid Number) measures the level of acid and acid-products present in the oil. KV (Kinematic Viscosity) is a measure of oil thickness. Viscosity is considered abnormal when it has decreased by 10% or increased by 20% of the base line value. AF (Analytical Ferrography) is a method for visual observation of lubricant degradation. When a lubricant works beyond its carrying capacity, friction polymers are observed. When a lubricant is degraded, amorphous films are observed.

Consequences of high fluid temperature When the temperature of the fluid increases, its viscosity decreases. Certain levels of viscosity are required to lubricate the internal surfaces of the components by creating an oil film between them. If the viscosity is very low this oil film is too thin and friction between these surfaces may occur. Fluid temperature above 82°C (180°F) damages the seals and reduces the life of the fluid. A pressure relief valve should be set at higher pressure than the working pressure in the system. Using the relief valve to control the working pressure increases the pressure losses in the system and creates heat. When the hydraulic system starts to overheat, it has to be shut down. Then we troubleshoot to find the problem and fix it. Running a hydraulic system with fluid temperature above 82°C is not recommended because it damages the components. It is similar to operating an IC engine at overheating conditions. Storage Hydraulic fluids should not be stored above 60°C or below freezing temperature.

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Servicing filters Pressure drop in the filter When the filter pressure drop is measured, the fluid has to be at the operating temperature. At the operating temperature, hydraulic fluids have a lower viscosity than when they are cold. Pressure drop depends on the exact temperature and the viscosity index of the oil. In addition, most filters have bypass valves that are viscosity sensitive. Therefore, during a cold start or other cold operating temperature conditions, some of the flow passes through the filter through a bypass check valve. The fluid that goes through the valve is not filtered at this time. As the oil heats up, a higher percentage of the flow will pass through the filter. During bypass, a welldesigned filter will not permit particles from being pulled off the front side of the filter.

Servicing reservoirs The first requirement for having a clean reservoir is proper packaging and capping all ports to prevent contamination from entering the reservoir during storage. Regular reservoir service includes: fluid level check, moisture check and airflow around the reservoir check. The water in the system comes from humid air entering the reservoir through the breather. The temperature changes cause humidity to condensate into water droplets on the inside reservoir walls. Moisture in the reservoir forms rust on the inside metal walls. Vibrations, during operation of the system, knock the rust particles into the oil, where they are picked up by the pump and distributed throughout the system. New rust particles form where the old ones fell off which makes the contamination an endless process. Changing filters and off-line filtration are not solutions to the problem. Flushing or replacing the reservoir is sometimes the most costeffective solution. When the reservoir is flushed, a turbulent flow is required for more effective cleaning.


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Servicing rotary pumps and motors The pump is considered the heart of the hydraulic system. All rotary pumps can work as motors and vise versa. The main differences between pumps and motors are the seal design and construction. Each gear pump has a direction of rotation shown on the body. The rotational direction of a gear pump can be reversed. The basic steps to reverse a gear pump rotation are: 1) disassemble the pump, 2) flip over the wear plate, trust plate and the seal and 3) assemble the pump. When a pump is installed above fluid level, it is recommended its inlet port to be at the bottom. If the pump has a built-in air bleeder, it has to be connected to the reservoir below the minimum fluid level. Before the first start-up of the system there are a few procedures to be followed: 1. Pump mounting bolts have to be tightened per the manufacturer’s recommendation. 2. Inlet and outlet fittings have to be properly installed. 3. Never start up a dry pump. Pump case has to be filled out with fluid before the start. 4. After pump is installed, run the system for 2 to 4 minutes unloaded before pressurizing and then increase the pressure gradually. Pump is the first component to be inspected when the hydraulic system does not perform as designed. There are a few standard checks: 1. Visual inspection for leaks through the seals or through cracks in the pump body. 2. Inspect the connection (coupling or spline) between the pump and the motor. 3. Inspect for bad bearings (bushings). 4. Check pumps temperature. 5. Measure pumps volumetric efficiency. Use a flow-meter to measure the flow at the pump output when the pump is loaded. Volumetric efficiency is the ratio of the measured output flow divided into the theoretical output flow. If the theoretical flow is not known, it can be obtained (approximately) by measuring the pump output flow when the pump is not loaded. When a pump wears, the slip in the pump increases resulting in decrease of volumetric efficiency.

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More than 90% of hydraulic pump failures can be attributed to the three common causes listed below: • Mechanical: broken shaft, broken gear or cracked housing • Worn-out pump: worn-out pump has a high leakage path which results in pressure losses. • Wrong fluid type: the fluid must be selected per outside temperature conditions. Using a wrong type can cause premature pump failure.

Servicing hydraulic cylinders Cylinders have to be visually inspected every time when the truck is serviced. The visual inspection without removing the cylinder includes: check for leaks, bent piston rods, broken welds and check pivot shafts or supports. Cylinder failures: Leaking cylinders is the most common failure. Leaks can be internal or external. There are two main failures: seal damage or mechanical failure. Seal’s damage can be caused by a contamination (external or internal), mechanical damage to the piston rod (surface finish or straightness) or improper seal selection. New cylinder must be inspected for leaks. The leak test consists of five steps: • • • • •

Set up Pressurize to maximum pressure Hold the pressure for 1 to 2 minutes De-pressurize Check for leaks

When a cylinder leaks after being used, each part has to be checked in order to find the cause of the problem. There are a few basic steps in the troubleshooting sequence: 1. 2. 3. 4.

Remove cylinder from the truck Visual observation of the cylinder’s outside surfaces. Disassemble cylinder Visual observation of all components and inside surfaces


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• Piston and cylinder seals • Piston rod and cylinder polished surfaces • Check seal grooves for burrs and notches 5. Measure the piston rod straightness Bent cylinder rod is another common mechanical failure. This failure is caused by overloads (buckling effect) or eccentric loads as a result of improper cylinder mounting. Creeping Cylinder failure is as result of internal leakage: worn-out piston seals or inside cylinder surface is out of shape or scored.

Servicing valves The most common valve failures are: malfunction as a result of contamination, spring failure and O-ring failure. A typical contamination failure is a seizure between the spool and the valve bore. This seizure is called: silt lock. Silt lock occurs when the silt force exceeds the force available to actuate the valve. The most sensitive hydraulic components to seize are components with small internal clearances such as: priority valves and servo valves. In order to avoid locking the valves, we can install a filter in the pressure line before the valve. Most failures of the pilot control valves are due to contamination. Pilot (indirect) operated valves are less tolerant to contamination than directly operated valves. If a pilot stage of indirect control valve is plugged by contaminants, the failure can result in unintended fully open or fully closed position. If directly operated valve is contaminated, it is more likely to have an increased leakage or “sticky” plunger. For these reasons, every time a valve is serviced because of contamination, the hydraulic fluid has to be changed with new fluid. All valves have to be properly marked and stored lubricated in plastic bags. Sometimes the valves are performing normally but make excessive noise. Most common reasons for noise in the valves are:

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• • • • • • • •

Valve pressure setting (for relief valves) is too close to working pressure Spring out of adjustment Broken spring Sticking plunger Improper hydraulic fluid Contaminated or hot fluid Worn out internal surfaces Flow rate through the valve is more than recommended (undersized valve)

Servicing connectors Usually hose assemblies fail without warning. They will age and harden even under normal operating conditions. Therefore, they have to be inspected regularly for cracks, leaks wear and excessive corrosion of the fittings. Major causes for connectors’ failures are improper selection, use, routing and assembly. If hose failure occurs, the operator must immediately shut down the machine, move away from it and call a mechanic to de-pressurize the system. Then the hose is disassembled and examined for damages. A failed hose must not be repaired; it must always be replaced with a new one. Failures such as: high speed discharge of pressurized fluid, flying connector or wiping hose can endanger a person’s life or cause permanent injuries. Factors reducing hydraulic hose life are: 1. 2. 3. 4. 5. 6. 7.

Hose bend radius is less than the specified minimum radius Rubbing the hose against hard surfaces Twisting, pulling or cyclic bending of the hose Operating above or below the specified temperature Operating above maximum pressure Using non-compatible fluid Very high fluid velocity as result of undersized hoses


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Tubes are easier to service than hoses because they can be cut and flared in the field. Hoses have to be removed if welding or torch cutting is performed close to the hose. Cleanliness of the new hoses is very important. Replacement hoses must have both ends caped. Any hose contamination decreases the life of the other components.

Seals Failures and causes 1. Damage during installation. The causes for this failure are: seal cuts from sharp corners or threads, lack of lubrication and use of improper tools. 2. Ware-out as a result of rubbing against seal groove surfaces in dynamic sealing applications. The biggest contributor for this failure is the roughness of the groove surface. 3. Extrusion is changing the shape of the seal as a result of pushing it into the gap of the mating surfaces. This failure is common for seal rings in high pressure applications. 4. Swelling is a result of absorbing fluid by the seal. The reason for this failure is the use of incompatible fluid and seal material. 5. Loss of original shape. This is common for O-rings. The O-ring loses its original shape and develops two flat surfaces. This is mainly caused by excessive compression due to improper design or working for a long time under high pressure at high temperatures. 6. Contamination. Seals start to leak when hard particles enter seal contact surfaces.

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II. Components layout- general considerations Reservoirs • •

Avoid placing the reservoir next to a heat source with temperature above 80º C. If the reservoir is placed close to such heat source, reservoir walls may need air cooling by a fan. The reservoir level indicator or the dip stick should be visible and easy to access.

Filters • •

External in-line filters should be rigidly mounted to the truck body. Filters should be easily accessible for change.

Pumps • • • •

The pump should be easily accessible for maintenance and replacement Pump fitting should be replaceable without removing the pump from the truck Pumps and motors are noise and vibrations sources, so they have to be mounted to a rigid surface. Vibration dampers could be placed between the mounting surface and the pump motor. Pump body, pressure fitting and hose should not touch covers or any flexible parts.

Valves • • • • • • • •

In-line valves should be rigidly mounted and independent from connector mounting Use locknuts when valves are mounted in-line Easy to adjust, replace or repair Sufficient clearance for wrench manipulation around valve fittings Sufficient space for electrical connection to valve solenoids Avoid using connections with pipe threads which require the use of sealing compound Manual overrides should be easy to access and would not require removal of any valve component other than cover Valve solenoids must have enough spaces between them so that the magnetic field of one valve will not interfere with the magnetic field of


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the other valves. This requirement is especially important for proportional solenoids. Fluid Connectors (fluid lines) • • • • •

Hoses must be protected from rubbing against metal edges or hard objects, snagging, cutting, pulling, bending and twisting Hose routing must ensure minimum length and number of bends, avoid twisting and avoid external heat sources; The connectors should be fasten/clamped to a rigid surface All fluid lines that are closer than 100 cm from the operator must be guarded (EU directive) Connectors for serviceable components such as filters should not be placed above other components which may malfunction if a oil leak occurs

III.

Common Problems

Leaks There are two types of leaks: external and internal. External leaks are easy to see and repair. Internal leaks are caused by mechanical damage in the hydraulic components, damaged seals or pressure buildup. When fluid moves from an area of high pressure to an area of low pressure without performing useful work, there is a pressure loss which decreases the component’s efficiency and the component generates extra heat. This means that any component in the hydraulic system that has abnormal, internal leakage will increase the heat load on the system and can cause the system to overheat. This could be anything from a cylinder that is leaking pressurized fluid through the piston seal, to an incorrectly adjusted relief valve. Any heat-generating components need to be identified and changed. One way to quickly locate an internal leak is to measure the temperature of individual components. The hottest component in the system can lead us to the problem. First, locate the leak and determine whether it is through the housing, seal or thread. Second, look for things that may cause the leak. Leaks are often caused by pressure buildup. Look for plugged vents, overfilling and elevated heat levels.

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Overheating Any temperature above a specified limit is considered excessive heat. Excessive heat for a system with one type of fluid can be normal heat for a system with a different fluid. For this reason the first thing that a service person needs to know is the maximum acceptable temperature. Overheating causes fluid degradation and change in viscosity. Heat in the hydraulic system is caused by pressure losses in the fluid. Pressure losses are generated when fluid passes through the hydraulic components or any restriction in the hydraulic lines. Heat is equal to the power loss and is proportional to the pressure drop. Every component in the system, that creates a pressure drop, generates heat. When we design a system, first thing we have to calculate is the total power loss in the system (see chapter 5). Together with the heat generation, the system has heat dissipation. The amount of dissipation will determine the fluid temperature. If the temperature goes above the design value we say the system is overheating. To avoid overheating we should design a system with minimum pressure losses and size the reservoir according to system power loads. Design of the reservoir is described in Chapter 3, section 13.

Reduced pump flow delivery We know that a hydraulic pump produces flow, not pressure. Reduced flow rate can be a result of a number of factors: • Insufficient fluid in the pump intake If the pump does not have enough fluid in the intake, it cannot deliver the required flow rate. Three basic checks are required: 1) check if the reservoir is filled to the correct level; 2) the suction strainer or filter (if fitted) is not clogged, and 3) the pump intake line is unrestricted. • Pump internal ware and reduced efficiency • Reduced rotation speed of the drive motor as a result of reduced motor efficiency.


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IV. Contamination of the hydraulic fluid Contamination is the presence of a foreign substance in the hydraulic fluid such as: water, dirt, dust, hard particles (grit, metal particles), etc. A hydraulic system is a collection of hydraulic components. Specialists have estimated that as much as 75% of all hydraulic system failures are a result of fluid contamination. It has also been estimated that it is less expensive to control contamination than to remove it or deal with its negative consequences. The contamination control should be built into the system design and service procedure. Damage to the system is caused by hard particles flowing inside the hydraulic fluid. These particles accelerate the wear of the hydraulic components. The rate at which damage occurs depends on four main factors: 1) internal clearances of the components of the system, 2) size and quantity of the particles present in the fluid, 3) system pressure and 4) filtration. The first factor depends on the manufacturing process of the component manufacturer. It is usually considered as a given. The effect of the second factor depends on selection of a filter that would ensure an optimal cleanliness of the system, which in turn gives a predicted life of the system. The third factor- system pressure- is selected by the design engineer and it is based on maximum work pressure, pressure loses, lift cylinder size and hose diameters. The relationship between these factors and the calculation methods are described in the previous chapters. In order to specify a filtration requirement for a specific application, we need to know exactly what we are trying to remove. If the issue is air bubbles in the oil, then filtration will not improve the situation. Sources of contamination There are four major sources: built-in, environmental, generated and new fluid. Built-in is contamination in the new components that is left over from the manufacturing process. They are: weld spatter, chips, burrs, sand, dust, sealant, rubber, fiber. Environmental are contaminants from the surrounding environment entering the system through the reservoir air breather or through the cylinder wipers. They are: water, moisture, dirt, dust.

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Generated are contaminants as a result of system operation. Different processes causing this contamination are: •

Abrasion is the process of wearing away of a surface by constant scratching, usually due to the presence of hard particles such as dirt, grit, or metallic particles in the lubricant. It may also break down the surface finish of the material.

•

Adhesion is metal-to-metal contact between moving parts as a result of the loss of lubrication film.

•

Cavitation is a process of formation of air or vapor pockets (bubbles) due to pressure drop in a liquid (pressure drop is a sudden loss of pressure). The term cavitation comes from the word cavity and means formation of cavities. It can also occur in a hydraulic system as a result of low fluid levels that draw air into the system, producing tiny bubbles that expand explosively at the pump outlet, causing metal erosion and eventual pump destruction.

•

Corrosion is a chemical or electrochemical process by which metal is destroyed through reaction with the surrounding environment. When the metal is iron, the process is called rusting.

•

Erosion is a process of wearing down edges of the component’s surfaces due to the high pressure and high flow rate fluid in the system. This is an interesting problem because the fluid flow affects its boundaries through erosion and deposition, which in turn affects the fluid flow.

•

Fatigue is failure due to repeated stress. Initiation is caused by micro cracks from the hard abrasive particles.

•

Temperature- a 10° C rise over normal operating temperature can reduce the oil life from 4000 to 2000 hours.

New hydraulic fluids are manufactured under relatively clean conditions. But after traveling through many hoses and pipes to drums or tanks, the fluid is no longer clean. It picks up rubber and metal particles from lines, metal particles and rust from storage tanks. Funnels are also a source of contamination and should be avoided. When they are re-useable, funnels have to be cleaned and stored in zip-lock plastic bags.


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Hard particle contamination In general, contamination can cause abrasive scratching, corrosion, wear, buildup of deposits, or any combination of these. Hard particles are the biggest contributor for wearing the internal surfaces of the components. There are three sizes of hard particles relative to component clearance: •

Particles larger than the component’s internal clearances. They can pass through the flow or they can get stuck between the moving parts closing the clearance and seizing the component. This seizure condition is called a silt lock.

•

Particles that are the same size as the internal clearances and are passing through two surfaces. They cause scoring and heavy wear of sliding surfaces damage. These particles are the main cause for the abrasion wear.

•

Hard particles, smaller than the components’ internal clearances (usually smaller than 5 microns), can also be highly abrasive. The number of these particles in the system is the highest because most filters do not capture them. If they are present in large quantities, they act as a “sand rain” causing rapid wear of the component surfaces. Very often the only way to remove them is through an oil change.

A major clue showing that damage to a hydraulic component has been caused by abrasion, is the pattern of wear. When scoring is caused by abrasion, the wear is relatively evenly distributed across the entire surface. Evidence of mechanical problems due to wear are metal particles in the fluid. To find this evidence, the work fluid is examined. Samples of the fluid should be collected from a location that contains the highest concentration of 'evidence' of a problem. Sometimes samples are collected from the pressure line after the filter or from the return to reservoir line. This is generally because these locations are easier to access and allow low cost port installation and sample collections. However, these common locations are far from ideal because the 'evidence' may be filtered or settled out of the lubricant, leaving the program with little more than fluid properties information. Sample ports must be configured to enable the collection of lubricants in close proximity to the hydraulic components in question. Water contamination Water contamination is considered the second most serious contamination problem after hard particle contamination.

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There are three states of water when it enters a hydraulic fluid: dissolved, emulsified and free. Usually all three states are presented at the same time. Dissolved state is when a water molecule is captured by oil molecules and becomes part of the fluid. Dissolved water can only be removed from the oil chemically by using a Vacuum Dehydrator. A Vacuum Dehydrator machine can remove about 80% of dissolved water. Dissolved water contamination is the least harmful of the three states. Emulsified state is when water and oil are homogenously mixed. Emulsified water can be removed physically by using a moisture absorbing filter. Free water state is when water is in a free state. Because it is heavier, it settles to the bottom of the reservoir. Free water is the most damaging of the three states, because it can displace the oil and allow metal to metal contact of sliding surfaces and cause hydraulic component mechanical failure. Most of free water is settled on the bottom and can be removed simply by draining it. Water contamination accelerates the aging process resulting in oxidation, hydrolyses, additive depletion, reduced lubricant film strength, corrosion and damage to components. In addition it can cause cavitation. Hydraulic and lubrication fluids are best operated with a water content of 50% below the vapor tension. Contamination in the fluid increases the pressure losses and it is one of the most important factors negatively affecting the hydraulic system operation and reliability. One method of evaluating fluid cleanliness is to measure the motor current. The motor current draw will go up when the hydraulic system has more pressure losses (lower efficiency) due to resistance in the fluid. The environment contributes greatly to system contamination. Contamination enters the system via the fluid reservoir (air breathers and access covers), and any sealing pairs (cylinders seals, pump and motor seals). It is estimated that about 50-60% of contaminants enter via the cylinder seals and it can be expected that the amount of ingression will increase with seal wear. If we have a contamination problem, it is more cost effective to filter the oil than to do frequent oil changes. Research has shown that maintaining fluid cleanliness results in increasing the average time between system breakdowns. Particle contamination reduces the service life of hydraulic fluids by striping additives and promoting oxidation. When evaluating metal


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wear trend, the fluid sample has to be taken after the valve (downstream of the valve). A measure for contamination is the clearance code of the fluid. It is described in chapter 3, section 16. For servo systems and high quality proportional valves, it is recommended: •

10µm or 15 µm (β15>75) high pressure filter without bypass just before the servo valve or 3 to 5 µm (β3>75) low pressure filter in the return line • A breather filter as fine as the finest filter in the system (procedure is described in chapter 5) The aim is to limit the fluid contamination to the recommended level by ISO 4406 maximum limits 19/16/13 (18/15/12 for long life servo system). Inspection When draining oil from a reservoir, we have to look for the presence of sediment and sludge on the bottom of the reservoir. A good practice is to check the water content in the oil on a regular basis. We can use a device which can detect water in oil. Since water and oil have different dielectric properties, this device detects the water by sensing the change in the fluid dielectricity. One of the best tools for inspecting fluid contamination is analyzing the particles caught by the filter. Typical tests include ferrography and elemental analysis.

V.

The future of the hydraulics

In the first half of the 20th century industrial trucks had mechanical elevating systems. Then, within a period of 50 years the mechanical systems were replaced by hydraulic systems. The first hydraulic systems were mechanically controlled. Later, some mechanical controls were replaced by electrical. Today, we have computer controlled systems. The performances of mechanically and electrically controlled systems depend greatly upon the system design and selection of the right components. Every year manufacturers add more intelligence to the electronic controls to better control the performance of hydraulic systems. By using computers, we can change the performance of the system by simply changing the software.

230


In the past, a jerky motion of an actuator, due to pulsations in the fluid, required adding an orifice in the line. In the future, we are going to write an algorithm to accomplish the same goal. Controllers have the ability to control not only the movement but also the acceleration and deceleration rates of the plunger in proportional valves. Now, we design systems that work within a specific temperature range. If the oil gets outside this range, the system performs differently. We can solve the problem by using a pulse-width modulated (PWM) signal to control hydraulic valves. Using programmable devices to control the valve performance, allows the physics of fluid dynamics to be added into the equations. This way, we can control the systems’ performance in extreme conditions. The future hydraulic system will have electronics that monitor the machine’s performance and make adjustments “on the fly” in order to optimize the performance. Electronics will be able to perform self diagnostics and automatically find the most energy efficient system performance. The latest development of the hydraulic system is governed by the effort to improve the interaction between the machine and the operator. The future system will have built-in commands so that the operator does not have to know how to control it. The operator will simply tell the system what he (she) wants the system to do. Manufactures have already started making components that can share data with components from other manufacturers. New trucks have microcontrollers which control all functions of the truck. The on-board microcontroller is called a Vehicle Manager (VM). VM has built-in programmable parameters, which allows users to adjust the system’s performance. Hydraulic components with electronic controls are connected to the VM by CAN (Controlled Area Network) bus. The CAN bus transmits in and receives signals from micro-controllers that are built into the components of the hydraulic system. When the CAN communication is lost the VM gives an error code. Knowing the error code reduces the time to find the problem. Another area of development is remote diagnostics of the hydraulic system. The advantage is a reduced cost of service. Remote diagnostic is ability to receive information from the machine in the office. The information allows technicians to diagnose the problem and bring all necessary spare parts and tools when going to the customer to fix it. It also allows for better planning of the daily service activities.


Appendix A Source: Industrial Truck Association

Class 1

Class 1 Lift Code - 1 Counterbalanced Rider Type, Stand Up

Class 1 Lift Code - 4 Three Wheel Electric Trucks, Sit Down

Class 1 Lift Code - 5 Counterbalanced Rider, Cushion Tires, Sit Down

Class 1 Lift Code - 6 Counterbalanced Rider, Pneumatic or Either Type Tire, Sit Down

A1

A2


Class 2

Class 2 Lift Code - 1 High Lift Straddle

Class 2 Lift Code - 2 Order Picker

Class 2 Lift Code - 3 Reach Type Outrigger

Class 2 Lift Code - 4 Side Loaders, Turret Trucks, Swing Mast and Convertible Turret/Stock Pickers Class 2 Lift Code - 6 Low Lift Pallet and Platform (Rider)


Class 3 Class 3 Lift Code - 1 Low Lift Platform

Class 3 Lift Code - 2 Low Lift Walkie Pallet

Class 3 Lift Code - 3 Tow Tractors (Draw Bar Pull Under 999 lbs.)

Class 3 Lift Code - 4 Low Lift Center Control

Class 3 Lift Code - 5 Reach Type Outrigger Walk behind operator

A3

A4


Class 3 Lift Code - 6 High Lift Straddle Walk behind operator

Class 3 Lift Code - 7 High Lift Counterbalanced

Class 3 Lift Code - 8 Low Lift, Walk Behind (Walkie) or Rider Pallet Truck

Class 4 Class 4 Lift Code - 3 Fork, Counterbalanced (Cushion Tire) Load capacity above 8000 lb (3636 kg)


Class 5 Class 5 Lift Code - 4 Fork, Counterbalanced (Pneumatic Tire) Load capacity above 8000 lb (3636 kg)

Class 6 Class 6 Lift Code - 1 Sit-Down Rider (Draw Bar Pull Over 999 lbs.)

Class 7 Class 7 Lift Code - 1 Variable Reach Rough Terrain Fork Lift Truck

A5

A6


Appendix B Approximate physical properties of common fluids Fluids

Grade/ name

Temper.

Density

Specific Weight

Kinematic Viscosity

t

ρ

γ

υ

ºC

kg/m³

kN/m3

cSt -3

Air at 1 bar

20

1.21

11.8 x 10

15.1

Water

16

999

9.80

1.1

Gasoline

16

680

6.67

0.46

32 (Light)

40

870

8.53

160

46 (Medium)

40

876

8.59

227

68 (Med. Heavy)

40

Hydrocarbonbased hydraulic fluid

Synthetic hydraulic fluid

Hydrostatic Transmission fluid Automatic Transmission Fluid (ATF) Synthetic Automatic Transmission Fluid (ATF) Steering fluid

Hydraulic 882

8.65

340

100 (Heavy)

40

887

8.7

490

150 (Extra Heavy)

40

890

8.73

750

40

835

8.19

160

Mobile HSC-824

Applications

& Steering systems

Hydraulic

FIRLUBE 22

40

SAFETYTE X 216

40

1145

11.23

230

32/46

40

887

8.7

195

46/68

40

893

8.76

267

100

40

911

8.93

533

32/46

40

867

8.5

185

1110

10.88

230

& Steering systems

Hydrostatic transmission Automatic transmission Steering systems

AMSOIL

40

850

8.33

36.8

Automatic transmission

MAX

40

835

8.19

58.8

Steering systems

5W20

40

859

8.42

33.4

PENNZOIL

40

874

8.57

37.3

Steering system

130 at

Brake fluid

Castrol SRF

16

1058

10.3

-40ºC

Brake system

3.5 at 100ºC

Engine oil

SAE 30

16

912

8.95

420

Engine lubrication


A7

Appendix C VISCOSITY CLASSIFICATION OF INDUSTRIAL FLUIDS Kinematic Viscosity Limits

ISO Viscosity Grade

Mean Viscosity at 40o C

Minimum

Maximum

2

2,2

1,98

2,42

3

3,2

2,88

3,52

5

4,6

4,14

5,06

7

6,8

6,12

7,48

10

10

9,00

11,00

15

15

13,5

16,5

22

22

19,8

24,2

32

32

28,8

35,2

46

46

41,4

50,6

68

68

61,2

74,8

100

100

90,0

110

150

150

135

165

220

220

198

242

320

320

288

352

460

460

414

506

680

680

612

748

1000

1000

900

1100

1500

1500

1300

1650

(cSt)

A8


Appendix D Coefficients of Local Resistance

Table 2.1 (source: Komitovski M., Components of Hydraulic and Pneumatic Systems)

3000 3300

Our target

1

2 2 1

Benchmark values

1

3

30

30

3.6 min 3.7

cm/ s

1

1

2

m

1

2 2

2 2

3

33

33

cm/ s

1

1

2

3 2 2

Y

Y

Y/ N

1

55

60

cm/ s

1

1

2

Y

Y/ N

1

1

2 2 3

1

1

3

1 1

1

1

3

1 1 1

Y

Y/ N

1

3 1

1

min

min

kW

2

3

2 2

Y

Y/ N

1 1 1 2 1 3

900

900

hour

2 1 3

1

60

80

dB

3

1

1

0.6

0.5 min

m2

3

2

2 2 2

70

80

°C

3

1

2

1

3000 kg capacity

2 3

2

Lift height 3600 mm

3

3

Lifting speed 30 cm/s

1 1

4

Lowering speed 33 cm/s

1

5 Stop lifting in case of failure

1 1

6 Limit lowering speed in case of failure

1 1

7

Monitor load weight and position

2 2 2 2 2 2

8

Minimize vibration during lifting 2 2

9

Minimize vibration during lowering

2 2 1

10

Tilt and side shift option

2

11 Optimize system efficiency

2

12

Ergonomic controls

1 3 1 1 2 2

13

Time between failures

3 1 2 2 2 2 1 1 1

14

Max. system noise level

kg

5

10 10 10 10 10 10 10 8 8 7 7 7 6 5 5

16

15

Clearance between lift cylinders

Engineering requirements

Temperature range: -30°C to +80°C

Measurement unit

Ability to lift 3000 kg capacity Load /unload from 3 m high Load lifting Load lowering Safe lifting Safe lowering Safe truck travel with the load Smooth lifting Smooth lowering Easy to enter and exit pallets Low ownership cost Easy to use controls Reliable system Low noise level Good visibility Work in extreme cold and hot environment

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

16

Requirements

#

Customer

Relative importance( 1 to 10)

Table 4.1 QFD house with the relationships: 3 is strong, 2 is medium and 1 is weak

Appendix E A-9

Competitor 1

Concept selection

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

3000 kg capacity Lift/lower pallets at 3.6 m height Stop lifting in case of failure Limit the maximum lowering speed Monitor load position, speed and weight Minimize vibration during lifting Minimize vibration during lowering Lifting speed 30 cm/s +/- 10% Lowering speed 33 cm/s +/- 10% Tilt and side shift option Optimize system efficiency Ergonomic controls System reliability (time between service calls) Minimum system noise level Open area between lift cylinders

Engineering Requirements

Total + Total Overall total Weighted total

10 10 10 10 10 9 9 8 8 7 7 7 6 5 5

Weights I

II

III

Design Concepts Datum

In the first column we can list the Engineering Parameters. In the second column we list the importance/ weight of these parameters. In the last column we can list the points that our benchmark design has. For benchmark we have an existing design. It could be our design or a design of a competitor.

Table 4.2

A-10

Appendix E


A-11

Appendix F Calculate the hydraulic parameters and power requirement of a hydraulic system (hydraulic circuit fig. 5.3) for the elevating system shown in fig. 5.2

Given (Engineering parameters and requirements) Gmax := 3000⋅ kg v1 := 30⋅ v2 := 40⋅

cm s cm s

6 pmax := 26⋅ 10 Pa

Maximum payload Lift speed with maximum load (+/- 5 %) Lift speed empty (+/- 5 %) Maximum pressure in the system

Mast construction- one free lift cylinder and two main lift cylinders η cyl :=

0.97

η mast :=

0.98

Mast mechanical efficiency Mast polispast number

n := 2 g = 9.807

Lift cylinder efficiency

m 2 s

Gravitational acceleration

Weight of elevating system components Gcarrige := 115⋅ kg Gmast1 := 136⋅ kg

Gcyl.m := 42⋅ kg Gpiston := 26⋅ kg

Gfork := 63⋅ kg

Gm.chain := 4 ⋅ kg

Gmast2 := 158⋅ kg

Gex.chain := 6 ⋅ kg

A-12


Calculations Selecting cylinders 1. Main lift (side) cylinders Maximum load on cylinders is:

(

)

Lmax := n ⋅ Gmax + Gcarrige + 2 ⋅ Gfork + Gmast1 + Gcyl.m + 2 ⋅ Gm.chain + Gmast2 + 2 ⋅ Gpiston + 2 ⋅ Gex.chain 3 Lmax = 7.076× 10 kg

Calculate diameter of main lift cylinders using formula 5.2

2⋅

d1_min :=

Lmax⋅ g π ⋅ pmax⋅ η cyl ⋅ η mast

d1_min = 0.042m

We select standard size piston diameter bigger than the calculated minimum d1 := 45⋅ mm

The area of the main-lift piston is: A1 := π ⋅

2 d1

−3

A1 = 1.59× 10

4

m

2

2. Free lift (middle) cylinder We select the diameter of the free lift cylinder so that its area is bigger than the combined area of both main lift cylinders. A2_min := 2 ⋅ A1 d2_min := d1 ⋅

2

d2_min = 63.6⋅ mm

Select diameter size bigger than the calculated minimum d2 := 67⋅ mm


A-13

The area of the free-lift piston is: A2 := π ⋅

2 d2

4 −3

A2 = 3.526× 10

m

2

There are two main parameters which will be calculated first: follow and pressure. In order to have two lift speeds (one for empty lift and one for lift with maximum load), the system requires two flow rates. Mast construction has two stages (free lift and main lift) with different cylinder areas which produce different pressures. Therefore, the system has four main work points: Work point 1. Free lift without load (Maximum flow - minimum pressure) Work point 2. Free lift with maximum load Work point 3. Main lift without load Work point 4. Main lift with load (Minimum flow - maximum pressure) Calculating required flow rate for desired lift speed Work point 1 (Flow rate in free-lift cylinders, lift without load) Q1 :=

A2 ⋅ v2 n

Q1 = 42.3⋅

L min

Work point 2 (flow rate in free-lift cylinders, maximum load on the forks) Q2 :=

A2 ⋅ v1 n

Q2 = 31.7⋅

L min

A-14


Work point 3 (flow rate in main-lift cylinders, lift without load)

Q3 :=

2 ⋅ A1 ⋅ v2 n

Q3 = 38.2⋅

L min

Work point 4 (flow rate in main-lift cylinders, maximum load on the forks)

Q4 :=

2 ⋅ A1 ⋅ v1 n

Q4 = 28.6⋅

L min

Calculate pressures Work point 1 (pressure in free-lift cylinders, lift without load) G0 := 0 ⋅ kg

(

Zero payload on the forks

)

L1 := n ⋅ G0 + Gcarrige + 2 ⋅ Gfork + Gmast1 + Gcyl.m + 2 ⋅ Gm.chain L1 = 854kg p1 :=

L1 ⋅ g A2

6 p1 = 2.38× 10 Pa p1 = 23.8⋅ bar

Load on cylinders for work point 1


A-15

Work point 2 (pressure in free-lift cylinders, maximum load on the forks)

(

)

L2 := n ⋅ Gmax + Gcarrige + 2 ⋅ Gfork + Gmast1 + Gcyl.m + 2 ⋅ Gm.chain L2 = 6854kg

p2 :=


L2 ⋅ g A2

6 p2 = 19.1× 10 Pa p2 = 191⋅ bar

Work point 3 (pressure in main-lift cylinders, lift without load)

(

)

L3 := n ⋅ G0 + Gcarrige + 2 ⋅ Gfork + Gmast1 + Gcyl.m + 2 ⋅ Gm.chain + Gmast2 + 2Gpiston + 2 ⋅ Gex.chain L3 = 1076kg p3 :=


L3 ⋅ g

2 ⋅ A1

6 p3 = 3.3 × 10 Pa p3 = 33⋅ bar

Pressure without payload

Work point 4 (pressure in main-lift cylinders, maximum load on the forks) 3 Lmax = 7.076× 10 kg

p4 :=

Maximum load was calculated earlier

Lmax⋅ g

2 ⋅ A1

6 p4 = 21.8× 10 Pa p4 = 218⋅ bar

Pressure with maximum payload of 3000 kg

A-16


System Work Points- Summary 1 Flow (Q) Points Description l/min p.1 Free lift empty 42.3 p.2 Free lift with maximum load 31.7 p.3 Main lift empty 38.2 p.4 Main lift with maximum load 28.6

Pressure (p) bar 24 191 33 218

The system power requirements must be based on minimum two work points. Work points one and four are both extreems. Therefore, in this example only these two points will be considered.

Select components Pump displacement Pump displacement is function of pump flow delivery and shaft rotational speed. Gear pumps have best performance and reliability in the range of 1000 to 3000 rev/min. Electric motors have best performance and reliability in the range of 1500 to 5000 rev/min Based on this, we will target rotational speed of 2200 rev/min.

Given Q3 = 38.2⋅

Q2 = 31.7⋅ n := 2200⋅ η vol :=

L min L min

1 min

0.98

Maximum flow rate (lift empty)

Minimum flow rate (lift with maximum load) Rotational speed - target Pump volumetric efficiency


A-17

Calculate and select pump displacement (use formula 3.4) dmax :=

Q3 n ⋅ η vol

3 dmax = 17.7⋅ cm

dmin :=

Maximum pump displacement needed for empty lift at 2200 rev/min

Q2 n ⋅ η vol

3 dmin = 14.7⋅ cm

Minimum pump displacement needed for lift with maximum load at 2200 rev/min

Select standard pump displacement 3 dpump := 16⋅ cm η m :=

Pump mechanical efficiency at 25 MPa pressure & 2000rev/min

0.90

η vol :=

Select to use a gear pump

Pump volumetric efficiency at 25 MPa pressure & 2000rev/min

0.98

Calculate shaft rotational speed based on pump with 16 cm^3 displacement (use formula 3.4) ne :=

Q3 dpump⋅ η vol

ne = 2434⋅

nl :=

1 min

Maximum pump shaft rotational speed (empty lift)

Q2 dpump⋅ η vol

nl = 2024⋅

1 min

Minimum pump shaft rotational speed- during lift with maximum load

A-18


Select hydraulic line diameters (The diameters of the fluid lines are based on recomended fluid velocity, see Hydraulic Connectors, Chapter 3) Suction line vs := 1.5⋅ AS :=

m

Maximum recommended fluid velocity inside suction hose

s

Q1 vs

2 AS = 470.1⋅ mm

4⋅

dS :=

Area of inside cross section

AS π

dS = 24.5⋅ mm

Minimum suction diameter

ds := 25⋅ mm

Select 25 mm diameter for suction line

Pressure line vp := 6 ⋅ Ap :=

m

Recommended fluid velocity inside pressure hose

s

Q4 vp

2 AS = 470.1⋅ mm

dP :=

4⋅


Ap π

dP = 10.1⋅ mm dp := 10⋅ mm

Recommended diameter Select 10 mm diameter for pressure line


A-19

Return line

vr := 2.5⋅ Ar :=

m

Recommended fluid velocity inside pressure hose

s

Q1 vr

2 AS = 470.1⋅ mm

4⋅

dR :=


Ar π

dR = 19⋅ mm

Recommended diameter for the return lines

dr := 20⋅ mm

Select 20 mm diameter for return line

Hydraulic Losses Calculate pressure losses in two work points of the system (WP1 and WP4) Known p4 = 218⋅ bar Q4 = 28.6⋅

L min

ν :=

Flow rate during lift with maximum load Pressure

p1 = 24⋅ bar Q1 = 42.3⋅

Pressure during lift with maximum load

L min 2

32⋅ 10 ⋅ stokes

Maximum flow rate (lift empty) Fluid viscosity in viscosity grade 32

A-20


Losses in the hydraulic components Pressure losses are given by the manufacturer in a graph or table format

Losses in the directional control valve at temperature

6

∆p dc :=

0.08⋅ 10 ⋅ Pa

∆p fc1 :=

0.26⋅ 10 ⋅ Pa

∆p fc2 :=

1.5⋅ 10 ⋅ Pa

6

Pressure drop in the flow control when lifting.

6

∆p filter :=

6

Pressure drop in the flow control when lowering.

0.07⋅ 10 ⋅ Pa

Pressure drop in the suction filter (manufacturer range is from 0.05 to 0.10 MPa)

Losses in the hydraulic lines There are two types losses in the hydraulic lines which result in a pressure drop: lineal (due to friction along the walls) and local (due to change of direction of the flow). Lineal losses occur in straight tubes and hoses. Local losses occur in the fittings.

ρ :=

880⋅

kg m

3

Re := 1500 λ :=

Reynolds Number

64 Re

Lineal losses in the suction line, use formula 2.26 Ls := 300⋅ mm

AH := π ⋅

2 ds

4

Suction hose length Cross area of the hose


Lift empty (WP1)

Lift with maximum load (WP4)

2

⎛ Q1 ⎞ ρ Ls ∆p s1 := ⎜ ⋅λ ⋅ ⋅ 2 ds ⎝ AH ⎠ ∆p s1 =

0.005⋅ bar

2

⎛ Q4 ⎞ ρ Ls ∆p s4 := ⎜ ⋅λ ⋅ ⋅ 2 ds ⎝ AH ⎠ ∆p s4 =

0.002⋅ bar

Lineal losses in the presure line, use formula 2.26 Lp := 8000⋅ mm AP := π ⋅

dp

Hose length

2

4

Lift empty (WP1)


2

⎛ Q1 ⎞ ρ Lp ∆p p1 := ⎜ ⋅λ ⋅ ⋅ 2 dp ⎝ AP ⎠ ∆p p1 =

12.11⋅ bar

2

⎛ Q4 ⎞ ρ Lp ∆p p4 := ⎜ ⋅λ ⋅ ⋅ 2 dp ⎝ AP ⎠ ∆p p4 =

5.54⋅ bar

Lineal losses in the return line, use formula 2.26 Hose length

Lr := 900⋅ mm AR := π ⋅

dr

2

4

Lift empty (WP1) 2

⎛ Q1 ⎞ ρ Lr ∆p r1 := ⎜ ⋅λ ⋅ ⋅ 2 dr ⎝ AR ⎠ ∆p r1 =

0.04⋅ bar

Lift with maximum load (WP4) 2

⎛ Q4 ⎞ ρ Lr ∆p r4 := ⎜ ⋅λ ⋅ ⋅ 2 dr ⎝ AR ⎠ ∆p r4 =

0.02⋅ bar

A-21

A-22


Total lineal losses Lift empty (WP1)

Lift with maximum load (WP4) ∆p L4 := ∆p s4 + ∆p p4 + ∆p r4

∆p L1 := ∆p s1 + ∆p p1 + ∆p r1 ∆p L1 =

∆p L4 =

12.15⋅ bar

5.56⋅ bar

Local losses in the fittings (use formula 2.28) ∆p loc :=

6

2.1⋅ 10 ⋅ Pa

Total losses from the pump to the lift cylinder Lift empty (WP1)


∆p t1 := ∆p dc + ∆p fc1 + ∆p L1 + ∆p loc

∆p t4 := ∆p dc + ∆p fc1 + ∆p L4 + ∆p loc

∆p t1 =

6

3.66× 10 Pa

∆p t4 =

6

3 × 10 Pa

Pressure at pump outlet port Work poit 1 pp1 := p1 + ∆p t1 6 pp1 = 6.03× 10 Pa

Work poit 4 pp4 := p4 + ∆p t4 6 pp4 = 24.81× 10 Pa

System Work Points- Summary 2 Pressure in Pressure Flow (Q) Points Description lift cylinders losses l/min bar bar p.1 Free lift empty 42.3 23.8 36.6 p.4 Main lift with maximum load 28.6 218 30

Pressure in pump outlet bar 60.4 248

These two points (p.1 and p.4) will be used to determine the power requirements of the system


A-23

Power delivered by pump work point 1

work poit 4

lift without load Pemp :=

lift with MAX load

( pp1) ( Q3)

Pmax :=

η vol ⋅ η m

Pemp = 4.3⋅ kW

pp4 ⋅ Q4 η vol ⋅ η m

Pmax = 13.4⋅ kW

Pemp = 6 ⋅ hp

Pmax = 18⋅ hp

Motor torque (maximum load, main lift) η m :=

0.90

Pump mechanical efficiency at 25 MPa pressure & 1800 rev/min

work point 1

work point 4

lift without load

lift with MAX load

Te :=

Pemp

( ne) ⋅ 2 ⋅ π ⋅ η m

Tmax :=

Te = 19⋅ N·m

System Work Points Work Point p.1 p.4

Description Free lift empty Main lift with maximum load

Pmax

( nl) ⋅ 2 ⋅ π ⋅ η m

Tmax = 70.4⋅ N·m

Hydraulic parameters

Power requirements

l/min 42.3

Pressure in pump outlet bar 60.4

Pump speed rev/min 2434

Pump input power kW 4.3

Pump input torque Nm 19

28.6

248

2024

13.4

70.4

Flow (Q)

A-24


Notes


References 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17.

Beecroft G. Dennis, Management of Quality courseware Brezonick, Mike, New Flow Limiter, Velocity Fuse Target Improved Machine Safety Byrne Diane, Taguchi Shin, The Taguchi Approach to Parameter Design Casey Brendan, Hydraulic Supermarket Evans James, Lindsay William, The Management and Control of Quality Georgiev, George, Design of Lift Trucks Charles J. Murray, Fluid power lessons Clausing Don P., Total Quality Development Gramatikov Ivan, Hydraulic System for High Lift Truck, Master’s thesis, Technical University of Sofia Gramatikov Ivan, Filter Selection to Maximize Hydraulic System Life, University of Toronto Jeffrey K. Liker and David Meier, The Toyota Way Komitovski Michael, Components of Hydraulic and Pneumatic Systems Lazarov Stefan, Research and Improvements of the Hydraulic System for High-Lift Electric Trucks Moskov N., Lazarov C., Hydro- and Pneumatic Drives and Controls Munson, B., Yong, D., Okiishi, T., Fundamentals of Fluid Mechanics Stankov P., Antonov I., Mechanics of Fluids Lessons and Problems Vickers Mobile Hydraulics Manual

Fluid Power Journal Industrial Vehicle Technology, UKiP Media Events Hydraulics & Pneumatics, Penton publication Hydraulic Supermarket (www.hydraulicsupermarket.com) Machinery Lubrication, Noria Corporation


Notes

250339586-Design-of-Hydraulic-Systems-for-Lift-Trucks.pdf

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