Xe địa hình tốc độ cao hệ thống treo, đường ray, bánh xe và động lực học-High speed off-road vehicles suspensions, tracks, wheels and dynamics

Of course, the combination of these two areas into High Speed Off‐Road Vehicles is an amalgamation that is not only extremely exciting from an engineering perspective, but one that pre-s

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Guide to Load Analysis for Durability Johannesson November 2013

High Speed Off-Road Vehicles

Suspensions, Tracks, Wheels and Dynamics

Bruce Maclaurin

Retired ex Ministry of Defence

UK

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Library of Congress Cataloging-in-Publication Data

Names: Maclaurin, Bruce, author.

Title: High speed off-road vehicles : suspensions, tracks, wheels and

dynamics / Bruce Maclaurin.

Description: Hoboken, NJ : John Wiley & Sons, 2018 | Series: Automotive

series | Includes bibliographical references and index |

Identifiers: LCCN 2018007653 (print) | LCCN 2018012003 (ebook) | ISBN

9781119258803 (pdf) | ISBN 9781119258810 (epub) | ISBN 9781119258780

(cloth)

Subjects: LCSH: Off-road vehicles.

Classification: LCC TL235.6 (ebook) | LCC TL235.6 M25 2018 (print) | DDC

629.228/8–dc23

LC record available at https://lccn.loc.gov/2018007653

Cover Design: Wiley

Cover Images: Background: © solarseven/Shutterstock; Left: Range Rover Sport Offroad Presentation by

Wachauring is licensed under CC BY-SA; Middle: © Rockfinder/Gettyimages; Right: Reservists Train to be Challenger tank crew, WMID-2013-044-350 © Crown Copyright 2013

Set in 10/12pt WarnockPro by SPi Global, Chennai, India

10 9 8 7 6 5 4 3 2 1

humour are sorely missed.

To my wife Jacqueline, for her encouragement and tolerance of my many hours

communing with journals, papers, books and my computer.

Series Preface xiii

1.4.1 Challenger MBT Hydrogas Unit 8

1.4.2 Measured Characteristics of a Challenger Unit 9

2.1.3.2 Road Wheels 32

2.1.3.3 Track Tensioners 33

2.1.4 Track Loadings 33

2.1.4.1 Centrifugal Tension 33

2.1.4.2 Final‐Drive Torque Measurements 34

2.1.4.3 Lateral Horn Load 35

2.1.5 Rolling Resistance: Analytical Methods 35

2.1.5.1 On a Metal Wheel Path 35

2.1.5.2 On a Rubber Wheel Path 36

2.1.6 Rolling Resistance: Experimental Measurements 37

2.1.8 Approaches for Reducing Noise and Vibration 43

2.1.8.1 Finite Element Analysis and Experimental Sprockets 43

2.1.8.2 Fully Decoupled Running Gear 44

2.1.8.3 Flexible Rubber Tracks 44

2.1.9 Reducing Noise and Vibration 44

2.1.9.1 Stage (a): Establishing the Principal Noise Sources 45

2.1.9.2 Stage (b): Design and Production of the Resilient Mountings 46 2.1.9.3 Stage (c): Test Results with the Resilient Mountings 47

2.2 Flexible Tracks 48

2.2.1 Earlier Flexible Tracks 49

2.2.2 Contemporary Flexible Tracks 50

2.2.3 ‘Proof‐of‐Principle’ Flexible Tracks for a Spartan APC 51

3.1 Human Response to Whole‐Body Vibration (WBV) and Shock 59

3.2.2 DERA Suspension Performance Test Courses 65

3.2.3 Response of Multi‐Wheel Vehicles 66

3.2.4 Quarter‐Car Model 68

3.2.5 Computer Modelling 71

3.2.5.1 Parameter Specification 73

3.2.5.2 Assumptions 74

3.5.2.3 Examples of Use of the Model 74

3.5.2.4 Comparison with Trials Data 75

3.5.2.5 Upgrading the Suspension Performance of the Scorpion Family

4.3 Active Suspension Systems 91

4.4 DERA Active Suspension Test Vehicles 93

5.1.1 Leyland DAF DROPS 8×6 Logistic Load Carrier 103

5.1.2 MAN SX 8×8 High‐Mobility Load Carrier 105

5.1.3 Pinzgauer 4×4 and 6×6 Light Trucks 105

5.1.4 Range Rover 106

5.1.5 Alvis Stalwart 107

5.1.6 Caterpillar Mining/Dump Truck 108

5.1.7 Euclid (Later Hitachi) Mining/Dump Trucks 110

7.1.1 Skid Steering Mechanisms 133

7.1.2 Skid Steering Models 136

7.1.3 The Magic Formula 139

7.1.4 Deriving the Magic Formula Parameters for the Track 140

7.1.5 Steering Performance Model 144

7.1.6 Results from the Model 146

7.1.6.1 Driver Control Arrangements 146

7.1.6.2 Pivot Turn 146

7.1.6.3 Effect of Radius of Turn on Slewing Moment 147

7.1.6.4 Driving on a 15 m Radius Turn at Varying Speed to Show the Effects

of Track Tension and a Suspension System 148

7.1.6.5 Driving on a 15 m Radius Turn at Varying Speeds with New and Worn Pads

and on a Low‐Friction Surface 150

7.1.6.6 Driving at 15 m s–1 on Turns of Varying Radii 152

7.1.6.7 Effect of the Centre of Gravity (CG) Position 154

7.1.6.8 Model Validation 156

7.2 Comparing Skid and Ackermann Steered

Wheeled Vehicles 156

7.2.1 Tyre Force–Slip Data 157

7.2.2 Choice of Tyre Model 158

7.2.2.1 The Skid Steered Vehicle: Vehicle Model 159

7.2.3 Results from the Model 159

7.2.3.1 Neutral Turn 159

7.2.3.2 Variation of Slewing Moment with Radius of Turn 161

7.2.3.3 Cornering on 15 m and 30 m Radius Turns at Different Speeds 162

7.2.4 Ackermann Steered Vehicle Model 163

Appendix A: Equations of Motion 170

Appendix B: Equations of Motion 173

References 175

8.1 Basic Requirements 177

8.1.1 Soil 177

8.1.2 Basic Definitions 178

8.1.3 Soil–Vehicle Models 179

8.2 Models for Soft Cohesive Soils 180

8.2.1 Vehicle Cone Index (VCI) Model 180

8.2.1.1 Mobility Index for Tracked Vehicles 181

8.2.1.2 Mobility Index for Wheeled Vehicles 181

8.2.2 WES Mobility Number Model 182

8.2.3 Mean Maximum Pressure (MMP) 182

8.2.4 Vehicle Limiting Cone Index (VLCI) 183

8.2.4.1 Tyres 184

8.2.4.2 Tracks 187

8.3 Models for Dry Frictional Soils 189

8.3.1 WES Mobility Number for Wheeled Vehicles 189

8.3.2 DERA Trials 190

8.3.3 Tracked Vehicles 193

8.4 Space Efficiency of Running Gear Systems for Armoured Vehicles 194

8.5 Tractive Force–Slip Relationship for Tyres in Soft Cohesive Soils 197

8.5.1 Describing Force–Slip Characteristics 197

8.5.1.1 Rectangular Hyperbolae 197

8.5.1.2 Exponentials 197

8.5.2 The Magic Formula 198

8.5.3 Development of the Modified Magic Formula 199

9.1.2 Using the Braking System 204

9.1.3 Velocity‐Dependent Limited‐Slip Differentials 204

9.1.4 Frictional Limited‐Slip Differentials 205

9.2 Relationships for Frictional Limited‐Slip Differentials 206

9.3 Traction Performance 209

9.3.1 Traction Model 209

9.3.2 Model Results 210

9.3.2.1 Effect of Weight Transfer Across an Axle 210

9.3.2.2 Different Soil Strengths Under the Tyres 212

9.3.2.3 On a Split μ Surface 214

9.4 Steering Performance on a Road Surface 214

9.4.1 Steering Performance Model 214

9.4.2 Model Results 214

Reference 216

10.1 Articulated Tracked Vehicles 217

10.1.1 Traction Forces with Skid and Articulated Steering 221

11.2.2 Electronic Stability Programmes 230

11.2.3 Active Anti‐Roll Bars 230

11.2.4 Driver Training 230

11.3 Truck Rollover: A Case Study 230

11.3.1 Calculating the Rollover Angle 231

The automobile is a part of our society, and tightly linked to many aspects of our daily lives We see a wide variety of vehicles every day, passing us on the streets in our cities and on our motorways There is of course a vast plethora of different vehicles used for different aspects in our daily lives, and in more special applications Perhaps some of the more interesting and exciting applications are those that are far removed from our everyday lives Two of the more famous and popular vehicle applications in the automo-tive sector are high speed vehicles such as race cars and off‐road vehicles such as large earth‐moving equipment Being in the field for over 30 years, many of those years as a faculty member, I can testify to the fact that most people get very excited when they are inspecting a top‐rated race car or see these vehicles on the track The same is true when standing next to or watching large earth‐moving equipment in operation There is nothing quite like seeing an earth‐mover that is capable of effortlessly hauling the vol-ume of several large sedans up a steep grade out of a mining operation Of course, the combination of these two areas into High Speed Off‐Road Vehicles is an amalgamation that is not only extremely exciting from an engineering perspective, but one that pre-sents unique challenges to vehicle designers that are not faced in many other automo-tive sectors.

High Speed Off‐Road Vehicles is an excellent and in‐depth review of vehicle

perfor-mance in off‐road conditions with a focus on key elements of the running gear systems

of vehicles In particular, elements such as suspension systems, wheels, tyres and tracks are addressed in depth It is a well‐written text that provides a pragmatic discussion of off‐road vehicles from both a historical and analytical perspective Some of the unique topics addressed in this book include link and flexible tracks, ride performance of tracked vehicles, and active and semi‐active suspension systems for both armoured and unarmoured vehicles The book also provides spreadsheet‐based analytical approaches

to modelling these topic areas, providing insight into steering, handling and overall performance of both tracked and wheeled systems The author further extends these analyses to soft‐soil scenarios and thoroughly addresses rollover situations The text also provides some insight into more advanced articulated systems

It is quite clear that this text is a unique and valuable addition to the Automotive Series

whose primary goal is to publish practical and topical books for researchers and tioners in industry, and postgraduate/advanced undergraduates in automotive engi-neering The series addresses new and emerging technologies in automotive engineering, supporting the development of next‐generation transportation systems The series cov-ers a wide range of topics, including design, modelling and manufacturing, and provides

practi-Series Preface

High Speed Off‐Road Vehicles is written from a very pragmatic perspective, based on

the author’s extensive experience, and provides an excellent introduction to off‐road vehicles Simultaneously, it is a strong reference text for those practising design and analysis of such systems No other text covers the concepts and vehicle systems that are presented in this book It is an excellent read, very understandable and highly informa-tive The bottom line is that this book covers a very interesting topic area and is highly

unique in its content, making this book a welcome addition to the Automotive Series.

I would like to extend my gratitude to:

the UK Ministry of Defence, for sponsoring most of the work described in this book;the many colleagues at DERA who contributed, and in particular Robert Gray, Robin Warwick, Peter Cox, Narinder Dhillon and Matt Williams; and to my nephew Peter Maclaurin for producing some of the line drawings

the team at John Wiley and Sons, including Eric Willner and Anne Hunt (commissioning editors), Nithya Sechin and Blesy Regulas (project editors), P Sathishwaran (production editor) and Elaine Rowan (freelance copy editor)

Acknowledgements

To a large extent, this book reflects my time and experience working at the UK Ministry

of Defence Military Vehicle Establishment at Chertsey, Surrey During this time it was variously known as FVRDE (Fighting Vehicles Research and Development Establish­ment), MVEE (Military Vehicles Engineering Establishment), RARDE Chertsey (Royal Armaments Research and Development Establishment), DRA (Defence Research Agency) and DERA (Defence Engineering and Research Agency) before it was closed in

2002 and split between QinetiQ and DSTL (Defence Science and Technology Labo­ratory) For the purposes of the book, the establishment is generally called DERA The term ‘high speed’ in the title of the book is used somewhat loosely, but is meant to exclude mainly unsprung vehicles The vehicles described are largely military because they are the most common type of off‐road vehicles, although wheeled logistic vehicles spend much of their time on roads

The book mainly describes the running gear systems of vehicles, that is, the suspen­sion systems, tracks, wheels and tyres and their effects on vehicle performance in off‐road conditions The book does not review vehicle power trains, except for describing the mechanisms used for providing the differential track speeds required for steering tracked vehicles The engines used in wheeled vehicles, logistic and armoured, are typically uprated commercially available engines For main battle tanks (MBTs), more specialist units are required because of the need to combine high power (up to about

1100 kW) with very compact dimensions Apart from the Abrams tank, which uses a gas turbine engine, the diesel engine technology used is still fairly conventional The transmissions of tracked military vehicles tend to be specialist because of the need to combine the drive engagement and ratio‐changing functions with the system used for providing differential speeds to the tracks The drive engagement and ratio‐changing systems remain largely conventional The units also need to be mounted transversely

As described in Chapter 7, electric drive and steering systems are now being investi­gated for these functions

Much use has been made in the book of spreadsheet analysis using Microsoft Excel, and particularly the Solver routine for solving equations of motion Although the Excel/Solver approach can be somewhat laborious at times, an advantage is that the equations

of motion need to be written from first principles, particularly those for tracked vehi­cles, requiring careful analysis of and good insight into the systems being studied

The book is partly descriptive of past and present systems and partly analytical It is not an academic book, or intended to be so, but hopefully some of the methods shown will be of use to vehicle designers

Introduction

are examined in detail.

Chapter 2 describes vehicle track systems, both link tracks, and the flexible tracks that are increasingly being used Performance aspects considered include rolling resistance and the noise and vibration caused by link tracks

Chapter  3 examines the ride performance of tracked vehicles, including human response to vibration, terrain profiles, wheelbase filtering and computer modelling Pitch response to braking is also considered

Chapter 4 examines the potential advantages of active and semi‐active suspension systems and describes two DERA test vehicles and their ride performance

Chapter 5 describes the driveline and suspension systems of wheeled vehicles, both unarmoured and armoured Interconnected suspension systems are also described.Chapter 6 considers the suspension performance of wheeled vehicles including the use

of quarter‐car models and the effect of using the different ISO 2631 and BS 6841 HRV filters Also described are some ride performance measurements of a logistic vehicle.Chapter 7 examines the steering performance of tracked and wheeled vehicles The Magic Formula, widely used for describing the force–slip characteristics of pneumatic tyres, is used here to describe the force–slip properties of a track system in a skid steer­ing model Results are shown for steering response and also for the power flows through the double‐differential steering system Similar models are shown for comparing the steering performance of a skid steered and an Ackermann steered wheeled vehicle The effects of torque vectoring are also considered

Chapter 8 examines the soft‐soil performance of wheeled and tracked vehicles Most predictive methods are empirically based, particularly for wheeled vehicles, because of the difficulty of directly modelling the behaviour of a pneumatic tyre in a soft yielding soil The results of DERA field trials with single pneumatic tyres and a track rig are described together with the predictive models developed A tractive force–slip relation­ship for a tyre in a soft cohesive soil is also developed from the field trial results.Chapter 9 describes the effects of limited‐slip differentials on the traction and steer­ing performance of vehicles Relationships are developed to describe the effects of fric­tional limited‐slip differentials on tractive performance on different surfaces and soils Results are compared with those for free and locked differentials The effects on steer­ing performance on a road surface are also examined

Chapter 10 describes some previous, experimental and current articulated vehicles, both tracked and wheeled The traction forces required to steer skid and articulated tracked vehicles are compared Similarly, the traction forces required to steer articu­lated wheeled vehicles on hard and soft soils are compared with those for skid and Ackermann steered vehicles

Chapter 11 examines the basic relationships that determine the likelihood of a vehicle

to rollover Methods to reduce the likelihood of rollover are reviewed A study of a rollo­ver incident with a logistic vehicle is described A model is developed to predict the rollover angle of the vehicle and compare it with the measured vehicle tilt angle.The author has made every reasonable effort to trace copyright holders and to obtain permissions to reproduce images in the book Apologies are made if suitable permis­sions have not been obtained; copyright holders should then contact the publishers so that corrections can be made in any further print editions

High Speed Off-Road Vehicles: Suspensions, Tracks, Wheels and Dynamics, First Edition Bruce Maclaurin

© 2018 John Wiley & Sons Ltd Published 2018 by John Wiley & Sons Ltd.

1

The running gear systems used on high speed, mainly military, tracked vehicles provide four essential functions:

● the transmission of drive to a relatively large number of road wheels;

● the distribution of the weight of the vehicle over a relatively large area;

● a large suspension displacement to allow high speeds over rough terrains; and

● a particular requirement of military armoured vehicles, the running gear system should occupy the minimum space in the overall vehicle envelope in order to maxim-ise internal hull volume (as will be shown in Section 8.4, this is a particular attribute

of tracked vehicles compared to wheeled vehicles of similar soft‐soil performance)

In addition, the running gear must be of minimum weight, reliable, easy to maintain, and compared to some other vehicle components, relatively cheap to produce

1.1 General Arrangement

Figure 1.1 shows the running gear of the Warrior Infantry Fighting Vehicle (IFV) and

is typical of modern practice Trailing suspension arms carry rubber‐tyre road wheels and operate transverse torsion bars which run across the floor of the vehicle Rotary vane hydraulic dampers are incorporated into the pivots of the front, second and rear road wheel stations Link tracks run under the road wheels and around hull‐mounted drive sprockets and return idlers Track pretension is adjusted by means of oil‐filled rams reacting against the idlers, which are carried on short pivoting arms The drive sprockets are front‐mounted but could be at the rear of the vehicle, depending on the position of the power pack Small diameter rollers support the top run of the track The track link pivots are rubber‐bushed and the links are fitted with  replaceable rubber road pads to minimise road damage and reduce noise and vibration

Figure 1.2 shows the arrangement on the Leopard 2 Main Battle Tank (MBT) Rotary friction dampers are built into the front three and rear two axle arm pivots The vehicle

is fitted with rubber‐bushed double‐pin tracks (see Chapter 2)

Tracked Vehicle Running Gear and Suspension Systems

1.2 Transverse Torsion Bars

Modern high‐strength spring steels, used with suitable presetting, shot peening and corrosion prevention techniques, allow nominal shear stresses of up to 1250 mPa to be used with a reasonable fatigue life [1.1, p 226] Suspension torsion bars are only loaded

in one direction and so can be ‘preset’ To preset a torsion bar, it is wound up to induce

Figure 1.1 Warrior running gear layout Source: Courtesy of Ministry of Defence.

Feststellbremse Vorgeiege Tnebkranze

Loufrollen

Hydr endonschlag Stutzrolle

Spannvorrichtung

Tragarm mit lagerung

Figure 1.2 Leopard 2 running gear layout Source: Courtesy of ATZ.

partial yielding in the outer layers of the bar On release, the outer layers take on tive shear stresses and torques opposed by positive stresses and torques in the inner layers of the bar (Figure 1.3).

nega-The relationship between the various variables that affect the maximum shear stress

in the bar can be explored by setting up a suitable spreadsheet The vehicle will be considered as a notional MBT with a sprung mass of 600 kN and an effective torsion bar length of 2.13 m The variables that can be considered are the axle arm length (initially taken as 450 mm), the number of road wheels (initially taken as 12) and the stiffness of the bar The latter can deduced from the ratio of wheel loads at full bump and at static

FB/FS, initially taken as 3:1, and the required static to bump suspension displacement

ΔSB, taken as 350 mm This gives a heave natural frequency of about 1.2 Hz, which is

typical for an MBT The shear modulus C is set at 76 mPa [1.1, p 226] The diameter of

the bar is left open

This gives a maximum shear stress qmax of 1326 mPa, which can be considered too high for a good fatigue life Increasing the arm length to 500 mm increases maxi-

mum torque on the bar, but also reduces maximum wind‐up angle; qmax reduces to

1258 mPa This may be acceptable depending on the duty cycle Measurements show that the front wheels nearly always have the most severe duty, largely because

of the pitching motion of the vehicle; this can be controlled by an adequate measure

of damping

Softening the suspension to give a FB/FS value of 2.5 and with axle arm length R at

450 mm increases qmax to 1371 kPa With the stiffer suspension, increasing the

num-ber of wheels to 14 reduces the value of qmax to 1276 kPa With the 0.5 m wheel arms,

qmax reduces to 1211 mPa If the length of wheel arm can be further increased to

0.55 m without causing interference between the arms, then qmax further reduces to

1155 mPa

Another possibility is of course to simply reduce the static to bump displacement

to, say, 325 mm with 500 mm wheel arms, 14 wheels and the stiffer suspension; qmax

is then 1158 mPa Some of the different possibilities are summarised in the table overleaf

Direction of loading Shear

stress +

–

Figure 1.3 The principle of presetting a torsion bar.

n (m) travel (m) FS (kN) FB/FS bar (mm) (mPa) (kg)

nor-If it is not possible to obtain satisfactory values of shear stress with hull width torsion bars, then two strategies can be used to effectively lengthen the bars One is

to approximately double the length of the bar by ‘folding’ it back This arrangement was used on the Second World War (WW2) German Panther tank as shown in Figure 1.4 The vehicle used eight interleaved wheels per side, both to improve soft‐soil performance and to reduce loading on the rubber tyres of the wheels Apart from the extra complication, another disadvantage of this arrangement is the possi-bility of mud and stones becoming stuck between the wheels; at low temperatures this could freeze and immobilise the vehicle Maximum shear stresses in the torsion bars were limited to a mere 200 mPa because of the qualities of the available steel and the somewhat unrealistic – for a wartime tank – design life of 10 000 km Factors tending to increase stress levels were the very soft suspension (a pitch frequency of only 0.5 Hz) and the very short axle arms; the latter was a requirement of the inter-leaved wheels The static to bump displacement was only 200 mm, tending to reduce stress levels

A second strategy is to enclose the torsion bars in torsion tubes However, torsion tubes are intrinsically much stiffer than the torsion bars, and the diameter of the tubes

is increased as a result of the need to pass them over the torsion bar end fittings Some experimental work has been conducted on the bar and tube arrangement shown in Figure 1.5 The stiffness of the bar was measured at 0.204 kNm/degree and that of the

tube at 1.89 kNm/degree; that is, the tube is over 9 times stiffer than the bar The bined stiffness was 0.184 kN/degree.

com-The failure torque of the tube was measured at about 33 kNm and that of the bar at

14 kNm It is therefore tempting to reduce the wall thickness of the tube and hence its

Figure 1.4 Panther torsion bar arrangement.

Road wheel Pivot

bearing

Anchor block

Torsion tube Connector

Support

bearing

Torsion bar

Oil seal

Figure 1.5 Torsion tube over bar arrangement Source: Courtesy of Ministry of Defence.

If the requirement was to provide a bar with the combined stiffness of bar and tube, then its effective length would need to be approximately 11% (180 mm) longer In prac-tice, it would be preferable to use either longer wheel arms or stiffer suspension.

Various arrangements have been devised to use coil springs in tracked vehicle pensions Figure 1.6 shows the system used on the WW2 Cromwell MBT A bellcrank extension of the axle arm is pivoted to a cylindrical canister that contains the spring and operates one end of the spring The other end is reacted by a rod that passes through the spring and is pivoted to the hull side A compression spring is thus made to effectively act in tension The springs are of small diameter to minimise intrusion into the hull space However, this does give the springs a high Wahl stress correction factor Telescopic dampers are fitted to axles 1, 2, 4 and 5 Available suspension displacement

sus-8

Figure 1.6 Cromwell MBT suspension unit Source: Courtesy of Ministry of Defence.

was 226 mm bump and 190 mm rebound The suspension was soft with a heave natural frequency of about 1 Hz.

Figure 1.7 shows the arrangement used on the Centurion and Chieftain MBTs, ally called a Horstman bogie A coil spring pack reacts between leading and trailing wheel arms via bellcranks and ‘knife‐edge’ bearings so that virtually equal loads are applied to both wheels The wheels can also articulate without deflecting the springs The spring pack comprises three nested coil springs The innermost spring acts as a bump stop when both wheels move upwards, which limits the maximum average deflection of the two wheels to only 86 mm This can severely limit performance when the vehicle is pitching at or near resonance on longer wavelengths, especially likely because damping levels are low When the bogie articulates, the maximum bump dis-placement on one wheel can increase to 158 mm with the other wheel in the static position and the spring pack fully compressed This can be useful when traversing large short‐wavelength obstacles (e.g rocks and tree trunks) Maximum spring shear stresses are quoted at about 1000 mPa Telescopic dampers of fairly small force capac-ity are fitted The weight of a complete assembly is 777 kg, of which the coil spring pack is 137.4 kg Six units represent about 9% of vehicle mass, which is a high figure especially for a suspension of fairly limited performance In comparison the suspen-sion of Challenger 2 represents about 5.5% of vehicle mass for a suspension of far higher performance

usu-A larger improved bogie suspension was produced for the Khalid MBT Here the maximum average double bump deflection was increased to 180 mm and the

Figure 1.7 Chieftain bogie suspension unit Source: Courtesy of Ministry of Defence.

single‐wheel bump displacement increased to 241 mm The mass was increased to

898 kg with the spring pack at 162 kg

The Israeli Merkava Mk 4 MBT uses trailing arms operating individual coil spring units (see Figure 1.8) The first two and last two wheel stations have hydraulic rotary dampers Road wheel travel is quoted as 300 mm bump and 304 mm rebound The high static deflection implies a comparatively soft suspension with a low bump force The suspension is also fitted with long‐travel hydraulic bump stops similar to those on the Leopard 2

1.4 Hydrogas Suspensions

Hydrogas (otherwise oleopneumatic, hydropneumatic, gas‐over‐oil) suspensions use,

as their name implies, a gas volume as the spring medium actuated by a piston and an oil column The gas, usually nitrogen, is normally separated from the oil by a floating piston or rubber diaphragm Units have been built without a separator piston or dia-phragm between the gas and oil, a similar arrangement to that used on most aircraft undercarriage ‘oleos’ [1.2]

1.4.1 Challenger MBT Hydrogas Unit

The hydrogas suspension units fitted to the Challenger 1 and Challenger 2 tanks were designed and developed at the Military Vehicles and Engineering Establishment

Figure 1.8 Merkava 4 suspension unit Source: Courtesy of MANTAK.

(MVEE) Many hundreds of hours were spent testing units on 300 kN 0.5 m stroke hydraulic actuators in the MVEE test laboratories Actuator displacement duty cycles were based on real‐time inputs from test vehicles running on severe‐roughness cross‐country courses and from computer modelling studies A particular requirement was to develop a sealing system and cylinder bore finish that would allow the units to run without attention between servicing at approximately 2000 km intervals A 250 hour test was set as an objective to simulate this requirement Hydrogas suspension units have been fitted to Challenger MBTs since 1983.

Figure 1.9 shows a cutaway of the unit A one‐piece axle arm pivots on plain bearings and operates a pressure piston via a crank and connecting rod The connecting rod again uses a plain bearing at the crank end and a knife‐edge bearing at the piston end Oil and gas chambers are in line and separated by a light alloy floating piston A com-pact disc‐spring damping valve is fitted between the pressure and separator pistons The main pivot bearing housing is a steel casting with a screwed‐on forged‐steel oil cylinder The gas chamber is similarly a steel forging screwed onto the oil cylinder The gas pressure at static is about 12.8 mPa

Bump stops are not fitted to any of the wheel stations However, as will be discussed later (see Section 1.4.2.6), the suspension on the Challenger is heavily damped and this reduces maximum suspension displacement The suspension also has a strongly rising spring characteristic that results in a sort of built‐in bump stop The units are suffi-ciently robust to carry the resulting forces However, if the unit is overloaded (overpres-sured) it fails benignly by leakage around the screw thread between the cylinder and the gas chamber Each unit weighs 287 kg

1.4.2 Measured Characteristics of a Challenger Unit

1.4.2.1 Spring Characteristics

As part of a research project, an extensive series of laboratory tests were conducted to measure the spring and damper characteristics of a Challenger hydrogas unit The unit was subjected to sinusoidal inputs over a frequency range of 0.001–2.0 Hz and with amplitudes of ±175 mm and ±200 mm Tests were conducted with and without the damper unit in place

Figure 1.10 shows the measured force/displacement characteristic at 0.8 Hz from –50 mm to +350 mm The effective polytropic index is calculated at about 1.66

Figure 1.9 Challenger hydrogas suspension unit Source: Courtesy of Ministry of Defence.

This value is in good agreement with the values shown in Figure 1.11 [1.3] Figure 1.12 shows the polytropic indexes derived from the measured force/displacement character-istics over a frequency range of 0.001–1.0 Hz; this shows that the index is virtually

0 50 100

Figure 1.11 Ratio of specific heats for nitrogen at different pressures Source: Din, 1961 [1.3]

Reproduced with permission of Butterworths.

constant at a value of 1.66 down to a frequency of about 0.04 Hz The value then falls to 1.37 at 0.01 Hz Even at 0.001 Hz, that is, with a complete cycle time of almost 17 min, the process is not isothermal with an index of 1.15 This means that over the normal working frequency range the suspension operates at near‐adiabatic conditions with an index of 1.66 Pressures were also measured inside the gas chamber, yielding a slightly higher polytropic index of 1.69.

1.4.2.2 Damper Characteristic

The damper characteristics of a hydrogas unit can be derived from three measurements:1) the differential pressure across the damper valve;

2) the wheel force/displacement loop of the unit; and

3) by placing a damper assembly in a suitable flow rig

1.4.2.3 Differential Pressure Across the Damper Valve

Figure 1.13 shows the wheel load derived from the measured differential pressure across the damper for a total amplitude of 350 mm and at a frequency of 0.9 Hz The lag in the curves can be ascribed to compressibility of the hydraulic fluid, and friction and inertia of the separator piston The curves show a near‐linear rate with a value of

120 kN (m s–1)–1 at the wheel The unit starts to limit or ‘blow‐off’ at a force of mately 40 kN at the wheel, reaching 50 kN at a wheel speed of 1 m s–1 In the rebound direction the limiting force is shown to be approximately 30 kN It is important that rebound damping limits at a force less than the static wheel load so that the wheel does not ‘hang’, particularly so on the front wheel

approxi-1.4.2.4 Force/Displacement Loop

Damping forces derived from the force/displacement loop include the sliding friction component of the unit measured at about ±0.045 of wheel load The damping rate is in

1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7

fact too high to be derived from the loop Maximum force in the bump direction is

50 kN at a wheel speed of 1.0 m s–1, and –40 kN in the rebound direction

1.4.2.5 Flow Rig

A damper valve was mounted in a hydraulic flow rig with a controllable flow rate The differential pressure across the damper and the flow rate were measured; Figure 1.14

–40 –30 –20 –10 0 10 20 30 40

10 0 –10 –20 –30 –40 –50 –60

20 30 40 50 60

70 Towards bump

Force at wheel (kN)

–0.20 –0.40

–0.60

Figure 1.14 Challenger suspension unit damper characteristics from flow rig.

depicts the results Because the damper rate is controlled by a simple orifice it shows

an approximate square‐law characteristic The equivalent linear rate is approximately

220 kN (m s–1)–1, that is, appreciably more than that given by measuring differential pressure on the suspension unit actuator rig The unit starts to limit at approximately

50 kN in the bump direction and –40 kN in the rebound direction

Because the solid rubber tyres on tracked vehicles are comparatively stiff, wheel speeds of up to 10 m s–1 in the bump direction have been measured when a wheel comes into contact with large obstacles at high speeds It is therefore important that the damper valve can allow high flow rates without generating excessively high pressures The flow rig shows the wheel force only rises to about 80 kN at 10 m s–1

The overall conclusion is that the damper rate is between 120 and 220 kN (m s–1)–1

with bump limiting starting at around 50 kN and the rebound at around –40 kN

1.4.2.6 Suspension Damping of a Multi‐Wheeled Vehicle

Consider a six‐wheel‐per‐side vehicle with equal wheel spacing and a damping

coeffi-cient of Cw (kN (m s–1)–1) per wheel For the Challenger with a 1.0 m wheel spacing, the

pitch damping coefficient Cp (kN m (rad s–1)–1) is defined:

in this case) then it can be estimated by considering the vehicle as a uniform rectangle (length × height) and multiplying by a factor, usually taken as 1.15, to allow for the high masses at the ends of the vehicle (frontal armour and power pack) For the Challenger,

the pitch moment of inertia Ip is therefore:

The pitch natural frequency has been measured as approximately 1.0 Hz; we

there-fore have Cpc = 2 × 439 256 × 2π = 5520 × 103 kN m (rad s–1)–1 and individual wheel

criti-cal damping coefficient Cwc = (5520 × 103)/35 = 158 kN (m s–1)–1

This compares with the measured damping coefficient of between 120 and 220 kN (m s–1)–1, implying that the Challenger suspension is very heavily damped at or near

to critical damping This is confirmed by suspension performance measurements with the vehicle as described in Chapter 3

1.4.3 Temperature Effects

The ride height of hydrogas suspensions is sensitive to changes in ambient temperature and also temperature rises caused by damper heating; this affects the front suspension

units in particular, where displacements are greater The rear units may also be affected

by heat soak from the power pack To calculate the effects of temperature on the Challenger, the units will be assumed to be serviced at 20°C to give a nominal ride height with 100 mm of rebound displacement If the effects of track tension are ignored, ride height at 50°C increases by 59 mm and available rebound displacement reduces to

41 mm; at –30°C the ride height reduces by almost 100 mm In practice, the track has a considerable effect on changes in ride height

Figure 1.15 shows a simplified model of the track system All the suspension units are merged into one ‘super’ unit The suspension displacement is calculated by equating the vertical components of track tension with the suspension force These are dependent on the linear elasticity of the track, the approach and departure angles, and the deflection of the suspension units which are dependent on the isothermal gas laws The specific linear stiffness of the Challenger double pitch track has been measured as 17 280 (kN m–1)–1 The stiffness of a length of track is inversely propor-tional to its basic stiffness For a track length of 7.68 m and approach and departure angles of 30°, this gives an effective vertical stiffness of 562.5 kN m–1 at sprocket and idler

Figure 1.16 shows the changes in ride height for the Challenger with and without tracks With tracks the ride height reduces by 32 mm at –30°C compared to almost

100 mm without tracks At 50°C the ride increases by 18 mm with tracks, compared to

59 mm without tracks At –30°C the track tension reduces to 14 kN compared to the normal value of about 50 kN Damper heating will tend to warm the suspension units when the vehicle starts running, but it would still be desirable to retension the tracks to prevent sprocket jumping If it is required to restore the suspension to its normal ride height, one possibility is to alter the oil volume of the unit in a similar manner to that used on some Citroen road cars This system would not be suitable for the Challenger hydrogas unit, however If oil is bled from the unit to restore the normal ride height at high temperatures, then there is the possibility of the separator piston contacting the damper and depressurising the oil cylinder At –20°C the gas volume at static is reduced from 2.241 L to 1.935 L If oil is injected into the unit to restore the ride height, the pressure at 350 mm displacement will be increased to 106.5 MPa compared 56.8 MPa

If the vehicle was fitted with a compensating idler (see Figure 3.26) then the sion would tend to respond to temperature changes as if the vehicle was not fitted with

suspen-a trsuspen-ack

The conclusion is therefore that to restore the ride height to its normal condition, it would be necessary to restore the gas volume of the unit to the value at 20°C This would normally require the procedure to be carried out as a workshop operation Two meth-ods of reducing temperature sensitivity and increasing stiffness at the static position are either to use two‐stage units or counter‐spring units

1.4.3.1 Two‐Stage Units

As well as the thermal effects described above, a further disadvantage of ment single‐stage hydrogas suspension units is the comparatively low spring rate around the static position; this can result in comparatively large pitch changes when accelerating and braking or when operating on steep slopes These effects can be greatly reduced by using two‐stage units as shown diagrammatically in Figure 1.17 A smaller gas volume is used around the static position to increase stiffness As displacement increases, a second gas volume is engaged The alternative spring curves that can be produced can be analysed with the aid of a spreadsheet Figure 1.18 shows a possible load/deflection curve for a unit with a 50 kN static load This demonstrates another benefit of the arrangement, in that the second stage can have a fairly soft rate with a peak load at 350 mm deflection of 154 kN at a pressure of 39.6 mPa This compares to

large‐displace-a pelarge‐displace-ak lolarge‐displace-ad of 218 kN for the stlarge‐displace-andlarge‐displace-ard unit large‐displace-at large‐displace-a pressure of 56.3 mPlarge‐displace-a The stlarge‐displace-andlarge‐displace-ard unit has a stiffness of 144 kN m–1 at the static position compared to 380 kN m–1 for the two‐stage unit, that is, it is 2.64 times stiffer

80 100

At –30° the ride height would reduce by 32 mm and the static track tension would be halved to 25 kN At 50°C the ride height would increase by 18 mm and the track tension would increase to 64 kN.

Oil

Figure 1.17 Diagram of two‐stage hydrogas suspension unit.

300 200

Displacement (mm)

Wheel load (kN)

100 0

0 50 100 150 200 250

First stage

Second stage Standard unit

Figure 1.18 Two‐stage hydrogas suspension load/deflection characteristic.

1.4.3.2 Counter‐Spring Units

Another way of increasing the spring rate around the static position is to use a counter‐spring acting in opposition to the hydrogas spring The spring could be a small metal spring or a small hydrogas unit as shown diagrammatically in Figure 1.19 Again,

a simple spreadsheet analysis can be used to assess the different possibilities Figure 1.20 shows the spring characteristics for a possible hydrogas counter‐spring

Gas

Oil

Figure 1.19 Diagram of hydrogas suspension with counter‐spring.

200 150 100 50 0

–50

0 Combined

telescopic unit with the spring and counter‐spring either side of the piston.

1.4.4 Other Types of Hydrogas Suspension

1.4.4.1 Twin‐Cylinder Units

The Challenger is fitted with a tapered hull side to reduce the effects of mine damage However, many armoured vehicles have vertical hull sides to maximise internal hull volume This limits the space available for externally mounted suspension units Figure 1.21 shows the unit fitted to the Leclerc MBT This has two opposed cylinders that enable the unit to be narrower for a given piston area Other features of the unit are the use of rubber diaphragms to separate the oil and nitrogen, and heat pipes to help carry heat from the damping valves Each unit weighs about 250 kg

1.4.4.2 In‐Arm Units

Units can be made even narrower by adopting an in‐arm arrangement This is a matic inversion of the Challenger arrangement with a modified wheel arm attached to the hull and the axle and wheel attached to a modified oil/gas pressure cylinder as shown diagrammatically in Figure 1.22

kine-A US-produced in-arm unit is interesting on two counts: (1) it does not use a separator piston between the oil and the gas (an advantage of using a separator piston is that

servicing is generally easier because the correct quantities of oil and gas are more ily inserted into the unit); and (2) it uses a variable‐force friction damper The damper

read-is a hydraulically loaded multi‐plate friction brake built around the main unit pivot Hydraulic pressure is derived from suspension movement and generated by a small pis-ton actuated by a cam The oil feeds through an orifice to provide a velocity‐dependent damping characteristic, and a pressure relief valve is used for force limiting The posi-tion of the damper helps dissipate heat into the hull side It is not thought that this unit proceeded to production, however The South Korean K2 MBT is fitted with in‐arm hydrogas suspension units

An alternative to in‐arm units that are similarly narrow is to use telescopic hydrogas

or liquid spring units mounted above the axle arm as shown diagrammatically in Figure 1.23, similar to the arrangement shown in Figure 4.4 Liquid springs are con-sidered in Chapter 4

Gas Fixed pivot and crank

Oil Damper valve

Figure 1.22 Diagram of In‐arm suspension unit.

Figure 1.23 Suspension with an external hydrogas or liquid spring strut.

1.5.1 Hydraulic Dampers

Most tracked vehicles use some form of hydraulic damper, either: (1) telescopic; (2) lever‐operated opposed piston; (3) rotary vane; or (4) built in to a hydrogas suspen-sion unit

Telescopic dampers are fitted to the M 113 and Bradley IFV vehicles Although comparatively simple to manufacture and fit, dissipating heat on severe cross‐country terrain is difficult and usually requires the use of special high‐temperature hydraulic fluids

Lever‐operated opposed piston dampers are fitted to the Alvis Stormer vehicles Being hull‐mounted, they have good heat‐dissipation properties

Rotary vane dampers can be lever‐operated or built in to axle arm pivots, as on the Abrams MBT and Warrior IFV vehicles Figure 1.24 shows a cross‐section of the Horstman unit used on the Warrior vehicle Arm pivot mounting gives a neat installa-tion with good heat dissipation properties

1.5.2 Friction Dampers

The Leopard 2 MBT uses friction dampers supplemented by hydraulic bump stops as shown in Figure 1.25 A problem with friction dampers is that if comparatively high values of friction are used, for example a value equivalent to static wheel load, then

a lower value would need to be used in the rebound direction to prevent wheel ‘hang up’ Further, vibration from the track would be increased, especially noticeable when running on smoother surfaces As shown in Figure 1.26 the friction dampers on the Leopard 2 have a progressive action with force increasing with suspension displace-ment Friction force at the wheel in the static position is 6 kN, about 0.13 times that of

Relief

valve

Vane Pivot shaft