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Advanced vehicle technology
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Advanced vehicle technology

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  • 1. Advanced Vehicle Technology
  • 2. To my long-suffering wife, who has provided sup-port and understanding throughout the preparationof this book.
  • 3. AdvancedVehicle TechnologySecond editionHeinz Heisler MSc., BSc., F.I.M.I., M.S.O.E., M.I.R.T.E., M.C.I.T., M.I.L.T.Formerly Principal Lecturer and Head of Transport Studies,College of North West London, Willesden Centre, London, UKOXFORD AMSTERDAM BOSTON LONDON NEW YORK PARISSAN DIEGO SAN FRANCISCO SINGAPORE SYDNEY TOKYO
  • 4. Butterworth-HeinemannAn imprint of Elsevier ScienceLinacre House, Jordan Hill, Oxford OX2 8DP225 Wildwood Avenue, Woburn, MA 01801-2041First published by Edward Arnold 1989Reprinted by Reed Educational and Professional Publishing Ltd 2001Second edition 2002Copyright # 1989, 2002 Heinz Heisler. All rights reservedThe right of Heinz Heisler to be identified as the author of this work has beenasserted in accordance with the Copyright, Designs and Patents Act 1988No part of this publication may be reproduced in any material form (includingphotocopying or storing in any medium by electronic means and whetheror not transiently or incidentally to some other use of this publication) withoutthe written permission of the copyright holder except in accordance with theprovisions of the Copyright, Designs and Patents Act 1988 or under the terms ofa license issued by the Copyright Licensing Agency Ltd, 90 Tottenham Court Road,London, England W1T 4LP. Applications for the copyright holders writtenpermission to reproduce any part of this publication should be addressedto the publishersWhilst the advice and information in this book are believed to be true andaccurate at the date of going to press, neither the authors nor the publishercan accept any legal responsibility or liability for anyerrors or omissions that may be made.Library of Congress Cataloguing in Publication DataA catalogue record for this book is available from the Library of CongressISBN 0 7506 5131 8 For information on all Butterworth-Heinemann publications visit our website at www.bh.comTypeset by Integra Software Services Pvt. Ltd, Pondicherry, Indiawww.integra-india.comPrinted and bound in Great Britain
  • 5. .......................................... 1 Vehicle structure................................................................ 1.1 Integral body construction 1.2 Engine, transmission and body structures .............................................................................................................................. 1.3 Fifth wheel coupling assembly ............................................................ 1.4 Trailer and caravan drawbar couplings.............................................................. 1.5 Semi-trailer landing gear .............................................................. system 1.6 Automatic chassis lubrication...................................... clutch 2 Friction........................................................ 2.1 Clutch fundamentals 2.2 Angular driven plate cushioning and torsional damping .............................................................................................. 2.3 Clutch friction materials 2.4 Clutch drive and driven member inspection ............................................................................................................ 2.5 Clutch misalignment................................................................. clutch 2.6 Pull type diaphragm ..................................................................... clutch 2.7 Multiplate diaphragm type ................................................................. clutch 2.8 Lipe rollway twin driven plate 2.9 Spicer twin driven plate angle spring pull type clutch ................................................................................................. brake 2.10 Clutch (upshift) 2.11 Multiplate hydraulically operated automatic transmission clutches ............................................................................ clutch 2.12 Semicentrifugal ................................................................... clutch 2.13 Fully automatic centrifugal .............................................................. 2.14 Clutch pedal actuating mechanisms 2.15 Composite flywheel and integral single plate diaphragm clutch .................... ............................................................. 3 Manual gearboxes and overdrives...................................................................gearbox 3.1 The necessity for a 3.2 Five speed and reverse synchromesh gearboxes ........................................... ........................................................... 3.3 Gear synchronization and engagement 3.4 Remote controlled gear selection and engagement m .................................... .............................................................. 3.5 Splitter and range change gearboxes................................................................... take-off 3.6 Transfer box power............................................................... 3.7 Overdrive considerations.................................................... ratios 3.8 Setting gear 4 Hydrokinetic fluid couplings and torque converters .................................................................................................. 4.1 Hydrokinetic fluid couplings
  • 6. 4.2 Hydrokinetic fluid coupling efficiency and torque capacity ........................................................................................coupling 4.3 Fluid friction 4.4 Hydrokinetic three element torque converter ................................................... 4.5 Torque converter performance terminology .......................................................................................................clutches 4.6 Overrun .................................................................. 4.7 Three stage hydrokinetic converter 4.8 Polyphase hydrokinetic torque converter ......................................................... 4.9 Torque converter with lock-up and gear change friction clutches .................... 5 Semi- and fully automatic transmission .................................................... ............................................................. 5.1 Automatic transmission consideration 5.2 Four speed and reverse longitudinally mounted automatic transmission..................................................... mechanical power flow 5.3 The fundamentals of a hydraulic control system .............................................. 5.4 Basic principle of a ....................................... hydraulically controlled gearshift 5.5....................................................... system Basic four speed hydraulic control 5.6 Three speed and reverse transaxle automatic transmission mechanical.................................. power flow 5.7 Hydraulic gear selection control components .................................................. 5.8 Hydraulic gear selection control operation ....................................................... 5.9 The continuously variable belt and pulley transmission ................................... 5.10 Five speed automatic transmission with electronic-hydraulic control ............ 5.11 Semi-automatic (manual gear change two pedal control) transmission............................ system 6 Transmission bearings and constant velocity joints ............................................................................................ 6.1 Rolling contact bearings ................................................................ joints 6.2 The need for constant velocity....................................................... 7 Final drive transmission 7.1 Crownwheel and pinion axle adjustments ......................................................................................................... locks 7.2 Differential................................................................ 7.3 Skid reducing differentials............................................................ axles 7.4 Double reduction.................................................. axles 7.5 Two speed..................................................................... 7.6 The third (central) differential ....................................................................... 7.7 Four wheel drive arrangements ............................................................. 7.8 Electro-hydralic limited slip differential
  • 7. 7.9 Tyre grip when braking and accelerating with good and poor road...............................surfaces.............................................................. system 7.10 Traction control.......................8 Tyres ........................................................... of tyres 8.1 Tractive and braking properties.............................................. 8.2 Tyre materials................................................... design 8.3 Tyre tread..................................................................... of tyres 8.4 Cornering properties .......................................................... stability 8.5 Vehicle steady state directional................................................................. 8.6 Tyre marking identification.................................................. 8.7 Wheel balancing............................ 9 Steering ............................................................. design 9.1 Steering gearbox fundamental .............................................................. steering 9.2 The need for power assisted ............................................................. joints 9.3 Steering linkage ball and socket .......................................................... 9.4 Steering geometry and wheel alignment...................................................................... pinion 9.5 Variable-ratio rack and 9.6 Speed sensitive rack and pinion power assisted steering ............................... 9.7 Rack and pinion electric power assisted steering ............................................................................... 10 Suspension............................................................ 10.1 Suspension geometry.............................................................. centres 10.2 Suspension roll..................................................................analysis 10.3 Body roll stability........................................................................stiffness 10.4 Anti-roll bars and roll .............................................................. stops 10.5 rubber spring bump or limiting............................................. location 10.6 Axle ..................................................................... 10.7 Rear suspension arrangements .................................................................. 10.8 Suspension design consideration............................................................ 10.9 Hydrogen suspension 10.10 Hydropneumatic automatic height correction suspension ........................... 10.11 Commercial vehicle axle beam location ....................................................... 10.12 Variable rate leaf suspension springs............................................................................................................................... bogies 10.13 Tandem and tri-axle..................................................................... 10.14 Rubber spring suspension 10.15 Air suspensions for commercial vehicles .....................................................
  • 8. 10.16 Lift axle tandem or tri-axle suspension ................................................................................................................ 10.17 Active suspension 10.18 Electronic controlled pneumatic (air) suspension for on and off road use .........................................system 11 Brake........................................... 11.1 Braking fun ................................................................ 11.2 Brake shoe and pad fundamentals ............................................................ 11.3 Brake shoe expanders and adjusters ........................................................... 11.4 Disc brake pad support arrangements .................................................................systems 11.5 Dual- or split-line braking........................................................ braking 11.6 Apportional ..................................................................... (ABS) 11.7 Antilocking brake system.............................................. servos 11.8 Brake 11.9 Pneumatic operated disk brakes (for trucks and trailers) ............................... 12 Air operated power brake equipment and ................. vehicle retarders ............................................................... brakes 12.1 Introductions to air powered ................................................................systems 12.2 Air operated power brake ............................................................. 12.3 Air operated power brake equipment.................................................... 12.4 Vehicle retarders...................................................................... brakes 12.5 Electronic-pneumatic.................................................. 13 Vehicle refrigeration....................................................... terms 13.1 Refrigeration 13.2 Principles of a vapour-compression cycle refrigeration system ..................... ................................................................. 13.3 Refrigeration system components 13.4 Vapour-compression cycle refrigeration system with reverse cycle................................. defrosting............................................................... 14 Vehicle body aerodynamics ....................................................................... 14.1 Viscous air flow fundamentals...................................................... 14.2 Aerodynamic drag................................................... 14.3 Aerodynamic lift................................................................ 14.4 Car body drag reduction.............................................................. control 14.5 Aerodynamic lift................................................. 14.6 Afterbody drag 14.7 Commercial ........................................... vehicle aeordynamic fundamentals 14.8 Commercial vehicle drag reducing devices ...................................................
  • 9. .................... Index
  • 10. 1 Vehicle Structure1.1 Integral body construction These box compartments are constructed in theThe integral or unitary body structure of a car can form of a framework of ties (tensile) and strutsbe considered to be made in the form of three box (compressive), pieces (Fig. 1.1(a & b)) made fromcompartments; the middle and largest compart- rolled sheet steel pressed into various shapes suchment stretching between the front and rear road as rectangular, triangular, trapezium, top-hat or awheel axles provides the passenger space, the combination of these to form closed box thin gaugeextended front box built over and ahead of the front sections. These sections are designed to resist directroad wheels enclosing the engine and transmission tensile and compressive or bending and torsionalunits and the rear box behind the back axle loads, depending upon the positioning of the mem-providing boot space for luggage. bers within the structure.Fig. 1.1 (a and b) Structural tensile and compressive loading of car body 1
  • 11. 1.1.1 Description and function of body Cantrails (Fig. 1.2(4)) Cantrails are the horizon-components (Fig. 1.2) tal members which interconnect the top ends of theThe major individual components comprising the vertical A and BC or BC and D door pillars (posts).body shell will now be described separately under These rails form the side members which make upthe following subheadings: the rectangular roof framework and as such are subjected to compressive loads. Therefore, they 1 Window and door pillars are formed in various box-sections which offer the 2 Windscreen and rear window rails greatest compressive resistance with the minimum 3 Cantrails of weight and blend in with the roofing. A drip rail 4 Roof structure (Fig. 1.2(4)) is positioned in between the overlap- 5 Upper quarter panel or window ping roof panel and the cantrails, the joins being 6 Floor seat and boot pans secured by spot welds. 7 Central tunnel 8 Sills 9 Bulkhead Roof structure (Fig. 1.2) The roof is constructed10 Scuttle basically from four channel sections which form11 Front longitudinals the outer rim of the slightly dished roof panel.12 Front valance The rectangular outer roof frame acts as the com-13 Rear valance pressive load bearing members. Torsional rigidity14 Toe board to resist twist is maximized by welding the four15 Heel board corners of the channel-sections together. The slight curvature of the roof panel stiffens it, thus prevent-Window and door pillars (Fig. 1.2(3, 5, 6, and 8)) ing winkling and the collapse of the unsupportedWindowscreen and door pillars are identified by a centre region of the roof panel. With large cars,letter coding; the front windscreen to door pillars additional cross-rail members may be used toare referred to as A post, the centre side door pillars provide more roof support and to prevent the roofas BC post and the rear door to quarter panel as crushing in should the car roll over.D post. These are illustrated in Fig. 1.2. These pillars form the part of the body structurewhich supports the roof. The short form A pillar and Upper quarter panel or window (Fig. 1.2(6)) Thisrear D pillar enclose the windscreen and quarter is the vertical side panel or window which occupieswindows and provide the glazing side channels, the space between the rear side door and the rearwhilst the centre BC pillar extends the full height of window. Originally the quarter panel formed anthe passenger compartment from roof to floor and important part of the roof support, but improvedsupports the rear side door hinges. The front and pillar design and the desire to maximize visibilityrear pillars act as struts (compressive members) has either replaced them with quarter windows orwhich transfer a proportion of the bending effect, reduced their width, and in some car models theydue to underbody sag of the wheelbase, to each end have been completely eliminated.of the cantrails which thereby become reactivestruts, opposing horizontal bending of the pas- Floor seat and boot pans (Fig. 1.3) These consti-senger compartment at floor level. The central BC tute the pressed rolled steel sheeting shape topillar however acts as ties (tensile members), trans- enclose the bottom of both the passenger and lug-ferring some degree of support from the mid-span of gage compartments. The horizontal spread-outthe cantrails to the floor structure. pressing between the bulkhead and the heel board is called the floor pan, whilst the raised platformWindscreen and rear window rails (Fig. 1.2(2)) over the rear suspension and wheel arches is knownThese box-section rails span the front window as the seat or arch pan. This in turn joins onto apillars and rear pillars or quarter panels depending lower steel pressing which supports luggage and isupon design, so that they contribute to the resist- referred to as the boot pan.ance opposing transverse sag between the wheel To increase the local stiffness of these platformtrack by acting as compressive members. The panels or pans and their resistance to transmittedother function is to support the front and rear vibrations such as drumming and droning, manyends of the roof panel. The undersides of the rails narrow channels are swaged (pressed) into the steelalso include the glazing channels. sheet, because a sectional end-view would show a 2
  • 12. Fig. 1.2 Load bearing body box-section memberssemi-corrugated profile (or ribs). These channels by the semicircular drawn out channel bottoms.provide rows of shallow walls which are both bent Provided these swages are designed to lay theand stretched perpendicular to the original flat correct way and are not too long, and the metal issheet. In turn they are spaced and held together not excessively stretched, they will raise the rigidity 3
  • 13. Fig. 1.3 (a±c) Platform chassis 4
  • 14. of these panels so that they are equivalent to a sheet spans between the rear end of the valance, where itwhich may be several times thicker. meets the bulkhead, and the door pillar and wing. The lower edge of the scuttle will merge with theCentral tunnel (Fig. 1.3(a and b)) This is the floor pan so that in some cases it may form part ofcurved or rectangular hump positioned longitudin- the toe board on the passenger compartment side.ally along the middle of the floor pan. Originally it Usually these panels form inclined sides to the bulk-was a necessary evil to provide transmission space head, and with the horizontal ledge which spans thefor the gearbox and propeller shaft for rear wheel full width of the bulkhead, brace the bulkhead walldrive, front-mounted engine cars, but since the so that it offers increased rigidity to the structure.chassis has been replaced by the integral box- The combined bulkhead dash panel and scuttle willsection shell, it has been retained with front wheel thereby have both upright and torsional rigidity.drive, front-mounted engines as it contributesconsiderably to the bending rigidity of the floorstructure. Its secondary function is now to house Front longitudinals (Figs 1.2(10) and 1.3(a and b))the exhaust pipe system and the hand brake cable These members are usually upswept box-sectionassembly. members, extending parallel and forward from the bulkhead at floor level. Their purpose is to with- stand the engine mount reaction and to support theSills (Figs 1.2(9) and 1.3(a, b and c)) These members front suspension or subframe. A common featureform the lower horizontal sides of the car body of these members is their ability to support verticalwhich spans between the front and rear road-wheel loads in conjunction with the valances. However, inwings or arches. To prevent body sag between the the event of a head-on collision, they are designedwheelbase of the car and lateral bending of the to collapse and crumble within the engine compart-structure, the outer edges of the floor pan are given ment so that the passenger shell is safeguarded andsupport by the side sills. These sills are made in the is not pushed rearwards by any great extent.form of either single or double box-sections(Fig. 1.2(9)). To resist the heavier vertical bendingloads they are of relatively deep section. Front valance (Figs 1.2 and 1.3(a and b)) These Open-top cars, such as convertibles, which do not panels project upwards from the front longitudinalreceive structural support from the roof members, members and at the rear join onto the wall of theusually have extra deep sills to compensate for the bulkhead. The purpose of these panels is to transferincreased burden imposed on the underframe. the upward reaction of the longitudinal members which support the front suspension to the bulkhead.Bulkhead (Figs 1.2(1) and 1.3(a and b)) This is the Simultaneously, the longitudinals are preventedupright partition separating the passenger and from bending sideways because the valance panelsengine compartments. Its upper half may form are shaped to slope up and outwards towards thepart of the dash panel which was originally used to top. The panelling is usually bent over near thedisplay the drivers instruments. Some body manu- edges to form a horizontal flanged upper, thusfacturers refer to the whole partition between engine presenting considerable lateral resistance. Further-and passenger compartments as the dash panel. If more, the valances are sometimes stepped andthere is a double partition, the panel next to the wrapped around towards the rear where they meetengine is generally known as the bulkhead and that and are joined to the bulkhead so that additionalon the passenger side the dash board or panel. The lengthwise and transverse stiffness is obtained.scuttle and valance on each side are usually joined If coil spring suspension is incorporated, theonto the box-section of the bulkhead. This braces valance forms part of a semi-circular tower whichthe vertical structure to withstand torsional distor- houses and provides the load reaction of the springtion and to provide platform bending resistance so that the merging of these shapes compounds thesupport. Sometimes a bulkhead is constructed rigidity for both horizontal lengthwise and lateralbetween the rear wheel arches or towers to reinforce bending of the forward engine and transmissionthe seat pan over the rear axle (Fig. 1.3(c)). compartment body structure. Where necessary, double layers of sheet are used in parts of the spring housing and at the rear of the valance where theyScuttle (Fig. 1.3(a and b)) This can be considered are attached to the bulkhead to relieve some of theas the panel formed under the front wings which concentrated loads. 5
  • 15. Rear valance (Fig. 1.2(7)) This is generally con- Torsional rigidity of the platform is usuallysidered as part of the box-section, forming the front derived at the front by the bulkhead, dash panhalf of the rear wheel arch frame and the panel and scuttle (Fig. 1.3(a and b)) at the rear by theimmediately behind which merges with the heel heel board, seat pan, wheel arches (Fig. 1.3(a, b andboard and seatpan panels. These side inner-side c)), and if independent rear suspension is adopted,panels position the edges of the seat pan to its by the coil spring towers (Fig. 1.3(a and c)).designed side profile and thus stiffen the underfloor Between the wheelbase, the floor pan is normallystructure above the rear axle and suspension. When provided with box-section cross-members to stiffenrear independent coil spring suspension is adopted, and prevent the platform sagging where thethe valance or wheel arch extends upwards to form passenger seats are positioned.a spring tower housing and, because it forms asemi-vertical structure, greatly contributes to the 1.1.3 Stiffening of platform chassisstiffness of the underbody shell between the floor (Figs 1.4 and 1.5)and boot pans. To appreciate the stresses imposed on and the resisting stiffness offered by sheet steel when it is subjected to bending, a small segment of a beamToe board The toe board is considered to form greatly magnified will now be considered (Fig.the lower regions of the scuttle and dash panel near 1.4(a)). As the beam deforms, the top fibres con-where they merge with the floor pan. It is this tract and the bottom fibres elongate. The neutralpanelling on the passenger compartment side plane or axis of the beam is defined as the planewhere occupants can place their feet when the car whose length remains unchanged during deforma-is rapidly retarded. tion and is normally situated in the centre of a uniform section (Fig. 1.4(a and b)). The stress distribution from top to bottom withinHeel board (Fig. 1.3(b and c)) The heel board is the beam varies from zero along the neutral axisthe upright, but normally shallow, panel spanning (NA), where there is no change in the length of thebeneath and across the front of the rear seats. Its fibres, to a maximum compressive stress on the outerpurpose is to provide leg height for the passengers top layer and a maximum tensile stress on the outerand to form a raised step for the seat pan so that bottom layer, the distortion of the fibres beingthe rear axle has sufficient relative movement greatest at their extremes as shown in Fig. 1.4(b).clearance. It has been found that bending resistance increases roughly with the cube of its distance from the neutral axis (Fig. 1.5(a)). Therefore, bend-1.1.2 Platform chassis (Fig. 1.3(a±c)) ing resistance of a given section can be greatlyMost modern car bodies are designed to obtain improved for a given weight of metal by takingtheir rigidity mainly from the platform chassis and metal away from the neutral axis where the metalto rely less on the upper framework of window fibres do not contribute very much to resistingand door pillars, quarter panels, windscreen rails distortion and placing it as far out as possibleand contrails which are becoming progressively where the distortion is greatest. Bending resistanceslender as the desire for better visibility is encouraged. may be improved by using longitudinal or cross- The majority of the lengthwise (wheelbase) bend- member deep box-sections (Fig. 1.5(b)) and tunneling stiffness to resist sagging is derived from both sections (Fig. 1.5(c)) to restrain the platform chas-the central tunnel and the side sill box-sections sis from buckling and to stiffen the flat horizontal(Fig. 1.3(a and b)). If further strengthening is floor seat and boot pans. So that vibration andnecessary, longitudinal box-section members may drumming may be reduced, many swaged ribs arebe positioned parallel to, but slightly inwards from, pressed into these sheets (Fig. 1.5(d)).the sills (Fig. 1.3(c)). These lengthwise membersmay span only part of the wheelbase, or the full 1.1.4 Body subframes (Fig. 1.6)length, which is greatly influenced by the design of Front or rear subframes may be provided to braceroad wheel suspension chosen for the car, the depth the longitudinal side members so that independentof both central tunnel and side sills, which are built suspension on each side of the car receives adequateinto the platform, and if there are subframes support for the lower transverse swing arms (wish-attached fore and aft of the wheelbase (Fig. 1.6 bone members). Subframes restrain the two halves(a and b)). of the suspension from splaying outwards or the 6
  • 16. Fig. 1.4 Stress and strain imposed on beam when subjected to bendinglongitudinal side members from lozenging as alter- the media of rubber mounts is that transmittednative road wheels experience impacts when travel- vibrations and noise originating from the tyresling over the irregularities of a normal road surface. and road are isolated from the main body shell It is usual to make the top side of the subframe and therefore do not damage the body structurethe cradle for the engine or engine and transmission and are not relayed to the occupants sittingmounting points so that the main body structure inside.itself does not have to be reinforced. This particu- Cars which have longitudinally positionedlarly applies where the engine, gearbox and final engines mounted in the front driven by the reardrive form an integral unit because any torque wheels commonly adopt beam cross-memberreaction at the mounting points will be transferred subframes at the front to stiffen and support theto the subframe and will multiply in proportion to hinged transverse suspension arms (Fig. 1.6(a)).the overall gear reduction. This may be approxi- Saloon cars employing independent rear suspen-mately four times as great as that for the front sion sometimes prefer to use a similar subframe atmounted engine with rear wheel drive and will the rear which provides the pivot points for thebecome prominent in the lower gears. semi-trailing arms because this type of suspension One advantage claimed by using separate sub- requires greater support than most other arrange-frames attached to the body underframe through ments (Fig. 1.6(a)). 7
  • 17. Fig. 1.5 Bending resistance for various sheet sections When the engine, gearbox and final drive are side members by utilising a horseshoe shaped framecombined into a single unit, as with the front longi- (Fig. 1.6(b)). This layout provides a platform fortudinally positioned engine driving the front wheels the entire mounting points for both the swing armwhere there is a large weight concentration, a sub- and anti-roll bar which between them make up theframe gives extra support to the body longitudinal lower part of the suspension. 8
  • 18. Fig. 1.6 (a±c) Body subframe and underfloor structure 9
  • 19. Front wheel drive transversely positioned modifies the magnitude of frequencies of theengines with their large mounting point reactions vibrations so that they are less audible to theoften use a rectangular subframe to spread out passengers.both the power and transmission units weight The installation of acoustic materials cannotand their dynamic reaction forces (Fig. 1.6(c)). completely eliminate boom, drumming, droningThis configuration provides substantial torsional and other noises caused by resonance, but merelyrigidity between both halves of the independent reduces the overall noise level.suspension without relying too much on the mainbody structure for support. Insulation Because engines are generally mountedSoundproofing the interior of the passenger close to the passenger compartment of cars or thecompartment (Fig. 1.7) cabs of trucks, effective insulation is important. InInterior noise originating outside the passenger this case, the function of the material is to reducecompartment can be greatly reduced by applying the magnitude of vibrations transmitted throughlayers of materials having suitable acoustic proper- the panel and floor walls. To reduce the transmis-ties over floor, seat and boot pans, central tunnel, sion of noise, a thin steel body panel should bebulkhead, dash panel, toeboard, side panels, inside combined with a flexible material of large mass,of doors, and the underside of both roof and based on PVC, bitumen or mineral wool. If thebonnet etc. (Fig. 1.7). insulation material is held some distance from the Acoustic materials are generally designed for one structural panel, the transmissibility at frequenciesof three functions: above 400 Hz is further reduced. For this type ofa) Insulation from noise Ð This may be created by application the loaded PVC material is bonded to a forming a non-conducting noise barrier spacing layer of polyurethane foam or felt, usually between the source of the noises (which may about 7 mm thick. At frequencies below 400 Hz, the come from the engine, transmission, suspension use of thicker spacing layers or heavier materials tyres etc.) and the passenger compartment. can also improve insulation.b) Absorption of vibrations Ð This is the transfer- ence of excited vibrations in the body shell to Absorption For absorption, urethane foam or a media which will dissipate their resultant lightweight bonded fibre materials can be used. energies and so eliminate or at least greatly In some cases a vinyl sheet is bonded to the foam reduce the noise. to form a roof lining. The required thickness of thec) Damping of vibrations Ð When certain vibra- absorbent material is determined by the frequencies tions cannot be eliminated, they may be exposed involved. The minimum useful thickness of to some form of material which in some way polyurethane foam is 13 mm which is effective with vibration frequencies above 1000 Hz. Damping To damp resonance, pads are bonded to certain panels of many cars and truck cabs. They are particularly suitable for external panels whose resonance cannot be eliminated by structural alterations. Bituminous sheets designed for this purpose are fused to the panels when the paint is baked on the car. Where extremely high damping or light weight is necessary, a PVC base material, which has three times the damping capacity of bituminous pads, can be used but this material is rather difficult to attach to the panelling. 1.1.5 Collision safety (Fig. 1.8) Car safety may broadly be divided into two kinds: Firstly the active safety, which is concerned with the cars road-holding stability while being driven,Fig. 1.7 Car body sound generation and its dissipation steered or braked and secondly the passive safety, 10
  • 20. collision, but overall alignment may also be neces- sary if the vehicles steering and ride characteristics do not respond to the expected standard of a simi- lar vehicle when being driven. Structural misalignment may be caused by all sorts of reasons, for example, if the vehicle has been continuously driven over rough ground at high speed, hitting an obstacle in the road, mount-Fig. 1.8 Collision body safety ing steep pavements or kerbs, sliding off the road into a ditch or receiving a glancing blow from somewhich depends upon body style and design struc- other vehicle or obstacle etc. Suspicion that some-ture to protect the occupants of the car from serious thing is wrong with the body or chassis alignment isinjury in the event of a collision. focused if there is excessively uneven or high tyre Car bodies can be considered to be made in three wear, the vehicle tends to wander or pull over toparts (Fig. 1.8); a central cell for the passengers one side and yet the track and suspension geometryof the welded bodywork integral with a rigid appears to be correct.platform, acting as a floor pan, and chassis with Alignment checks should be made on a level,various box-section cross- and side-members. This clear floor with the vehicles tyres correctly inflatedtype of structure provides a reinforced rigid crush- to normal pressure. A plumb bob is required in theproof construction to resist deformation on impact form of a stubby cylindrical bar conical shaped atand to give the interior a high degree of protection. one end, the other end being attached to a length ofThe extension of the engine and boot compart- thin cord. Datum reference points are chosen suchments at the front and rear of the central passenger as the centre of a spring eye on the chassis mount-cell are designed to form zones which collapse and ing point, transverse wishbone and trailing armcrumble progressively over the short duration of a pivot centres, which are attachment points to thecollision impact. Therefore, the kinetic energy due underframe or chassis, and body cross-member toto the cars initial speed will be absorbed fore and side-member attachment centres and subframeaft primarily by strain and plastic energy within the bolt-on points (Fig. 1.9).crumble zones with very little impact energy actu- Initially the cord with the plumb bob hangingally being dissipated by the central body cell. from its end is lowered from the centre of each reference point to the floor and the plumb bob con-1.1.6 Body and chassis alignment checks tact point with the ground is marked with a chalked(Fig. 1.9) cross. Transverse and diagonal lines between refer-Body and chassis alignment checks will be neces- ence points can be made by chalking the full lengthsary if the vehicle has been involved in a major of a piece of cord, holding it taut between reference centres on the floor and getting somebody to pluckTable 1.1 Summary of function and application of the centre of the line so that it rebounds and leavessoundproofing materials a chalked line on the floor. A reference longitudinal centre line may be madeFunction Acoustic materials Application with a strip of wood baton of length just greater than the width between adjacent reference marksInsulation Loaded PVC, Floor, bulkhead bitumen, with or dash panel on the floor. A nail is punched through one end without foam or and this is placed over one of the reference marks. fibres base, A piece of chalk is then held at the tip of the free mineral wool end and the whole wood strip is rotated about the nailed end. The chalk will then scribe an arcDamping Bitumen or Doors, side between adjacent reference points. This is repeated mineral panels, wool underside of roof from the other side. At the points where these two arcs intersect a straight line is made with a plucked,Absorption Polyurethane foam, Side panels, chalked cord running down the middle of the vehi- mineral wool, or underside of cle. This procedure should be followed at each end bonded fibres roof, engine of the vehicle as shown in Fig. 1.9. compartment, Once all the reference points and transverse and bonnet diagonal joining lines have been drawn on the 11
  • 21. Fig. 1.9 Body underframe alignment checksfloor, a rule or tape is used to measure the distances Both the variations of inertia and gas pressurebetween centres both transversely and diagonally. forces generate three kinds of vibrations which areThese values are then chalked along their respective transferred to the cylinder block:lines. Misalignment or error is observed when apair of transverse or diagonal dimensions differ 1 Vertical and/or horizontal shake and rockand further investigation will thus be necessary. 2 Fluctuating torque reaction Note that transverse and longitudinal dimen- 3 Torsional oscillation of the crankshaftsions are normally available from the manufac-turers manual and differences between paired 1.2.2 Reasons for flexible mountingsdiagonals indicates lozenging of the frameworkdue to some form of abnormal impact which has It is the objective of flexible mounting design to cope with the many requirements, some havingpreviously occurred. conflicting constraints on each other. A list of the duties of these mounts is as follows:1.2 Engine, transmission and body structure 1 To prevent the fatigue failure of the engine andmountings gearbox support points which would occur if they were rigidly attached to the chassis or1.2.1 Inherent engine vibrations body structure.The vibrations originating within the engine are 2 To reduce the amplitude of any engine vibrationcaused by both the cyclic acceleration of the reci- which is being transmitted to the body structure.procating components and the rapidly changing 3 To reduce noise amplification which would occurcylinder gas pressure which occurs throughout if engine vibration were allowed to be transferredeach cycle of operation. directly to the body structure. 12
  • 22. 4 To reduce human discomfort and fatigue by partially isolating the engine vibrations from the body by means of an elastic media.5 To accommodate engine block misalignment and to reduce residual stresses imposed on the engine block and mounting brackets due to chassis or body frame distortion.6 To prevent road wheel shocks when driving over rough ground imparting excessive rebound movement to the engine.7 To prevent large engine to body relative move- ment due to torque reaction forces, particularly in low gear, which would cause excessive mis- alignment and strain on such components as the exhaust pipe and silencer system.8 To restrict engine movement in the fore and aft direction of the vehicle due to the inertia of the engine acting in opposition to the accelerating and braking forces.1.2.3 Rubber flexible mountings (Figs 1.10, 1.11and 1.12)A rectangular block bonded between two metalplates may be loaded in compression by squeezing Fig. 1.10 (a and b) Modes of loading rubber blocksthe plates together or by applying parallel butopposing forces to each metal plate. On compres- When two rubber blocks are inclined to each othersion, the rubber tends to bulge out centrally from to form a `V mounting, see Fig. 1.11, the rubber willthe sides and in shear to form a parallelogram be loaded in both compression and shear shown by(Fig. 1.10(a)). the triangle of forces. The magnitude of compressive To increase the compressive stiffness of the force will be given by Wc and the much smaller shearrubber without greatly altering the shear stiffness, force by WS. This produces a resultant reaction forcean interleaf spacer plate may be bonded in between WR. The larger the wedge angle Â, the greater thethe top and bottom plate (Fig. 1.10(b)). This inter- proportion of compressive load relative to the shearleaf plate prevents the internal outward collapse of load the rubber block absorbs.the rubber, shown by the large bulge around the The distorted rubber provides support undersides of the block, when no support is provided, light vertical static loads approximately equal inwhereas with the interleaf a pair of much smaller both compression and shear modes, but withbulges are observed. heavier loads the proportion of compressive stiffnessFig. 1.11 `V rubber block mounting 13
  • 23. These modes of movement may be summarized as follows: Linear motions Rotational motions 1 Horizontal 4 Roll longitudinal 5 Pitch 2 Horizontal lateral 6 Yaw 3 Vertical 1.2.6 Positioning of engine and gearbox mountings (Fig. 1.15) If the mountings are placed underneath the com- bined engine and gearbox unit, the centre of gravity is well above the supports so that a lateral (side) force acting through its centre of gravity, such as experienced when driving round a corner, will cause the mass to roll (Fig. 1.15(a)). This condition is undesirable and can be avoided by placing the mounts on brackets so that they are in theFig. 1.12 Load±deflection curves for rubber block same plane as the centre of gravity (Fig. 1.15(b)). Thus the mounts provide flexible opposition toto that of shear stiffness increases at a much faster any side force which might exist without creating arate (Fig. 1.12). It should also be observed that the roll couple. This is known as a decoupled condition.combined compressive and shear loading of the An alternative method of making the naturalrubber increases in direct proportion to the static modes of oscillation independent or uncoupled isdeflection and hence produces a straight line graph. achieved by arranging the supports in an inclined `V position (Fig. 1.15(c)). Ideally the aim is to1.2.4 Axis of oscillation (Fig. 1.13) make the compressive axes of the mountings meetThe engine and gearbox must be suspended so that at the centre of gravity, but due to the weight of theit permits the greatest degree of freedom when power unit distorting the rubber springing theoscillating around an imaginary centre of rotation inter-section lines would meet slightly below thisknown as the principal axis. This principal axis point. Therefore, the mountings are tilted so thatproduces the least resistance to engine and gearbox the compressive axes converge at some focal pointsway due to their masses being uniformly distrib- above the centre of gravity so that the actual linesuted about this axis. The engine can be considered of action of the mountings, that is, the directionto oscillate around an axis which passes through of the resultant forces they exert, converge on thethe centre of gravity of both the engine and gearbox centre of gravity (Fig. 1.15(d)).(Figs 1.13(a, b and c)). This normally produces an The compressive stiffness of the inclined mountsaxis of oscillation inclined at about 10±20 to the can be increased by inserting interleafs betweencrankshaft axis. To obtain the greatest degree of the rubber blocks and, as can be seen infreedom, the mounts must be arranged so that they Fig. 1.15(e), the line of action of the mounts con-offer the least resistance to shear within the rubber verges at a lower point than mounts which do notmounting. have interleaf support. Engine and gearbox mounting supports are1.2.5 Six modes of freedom of a suspended body normally of the three or four point configuration.(Fig. 1.14) Petrol engines generally adopt the three pointIf the movement of a flexible mounted engine is support layout which has two forward mountscompletely unrestricted it may have six modes of (Fig. 1.13(a and c)), one inclined on either side ofvibration. Any motion may be resolved into three the engine so that their line of action converges onlinear movements parallel to the axes which pass the principal axis, while the rear mount is supportedthrough the centre of gravity of the engine but at centrally at the rear of the gearbox in approximatelyright angles to each other and three rotations about the same plane as the principal axis. Large dieselthese axes (Fig. 1.14). engines tend to prefer the four point support 14
  • 24. Fig. 1.13 Axis of oscillation and the positioning of the power unit flexible mountsarrangement where there are two mounts either side down at a uniform rate. The amplitude of this cyclicof the engine (Fig. 1.13(b)). The two front mounts movement will progressively decrease and the num-are inclined so that their lines of action pass through ber of oscillations per minute of the rubber mountingthe principal axis, but the rear mounts which are is known as its natural frequency of vibration.located either side of the clutch bell housing are not There is a relationship between the static deflec-inclined since they are already at principal axis level. tion imposed on the rubber mount springing by the suspended mass and the rubbers natural frequency1.2.7 Engine and transmission vibrations of vibration, which may be given by 30Natural frequency of vibration (Fig. 1.16) A sprung n0 ˆ pbody when deflected and released will bounce up and x 15
  • 25. Fig. 1.14 Six modes of freedom for a suspended blockwhere n0 = natural frequency of vibration the engine out of balance forces and the fluctuating (vib/min) cylinder gas pressure and the natural frequency of x = static deflection of the rubber (m) oscillation of the elastic rubber support mounting, i.e. resonance occurs when This relationship between static deflection and nnatural frequency may be seen in Fig. 1.16. ˆ1 n0Resonance Resonance is the unwanted synchron- where n = disturbing frequencyization of the disturbing force frequency imposed by n0 = natural frequency Transmissibility (Fig. 1.17) When the designer selects the type of flexible mounting the Theory of Transmissibility can be used to estimate critical resonance conditions so that they can be either prevented or at least avoided. Transmissibility (T) may be defined as the ratio of the transmitted force or amplitude which passes through the rubber mount to the chassis to that of the externally imposed force or amplitude generated by the engine: Ft 1 Tˆ ˆ 2 Fd n 1À n0 where Ft ˆ transmitted force or amplitude Fd ˆ imposed disturbing force or amplitude This relationship between transmissibility andFig. 1.16 Relationship of static deflection and natural the ratio of disturbing frequency and naturalfrequency frequency may be seen in Fig. 1.17. 16
  • 26. Fig. 1.15 (a±e) Coupled and uncoupled mounting points 17
  • 27. The transmissibility to frequency ratio graph rubber mountings is greater than 11¤2 and the trans-(Fig. 1.17) can be considered in three parts as follows: missibility is less than one. Under these conditions off-peak partial resonance vibrations passing to theRange (I) This is the resonance range and should be body structure will be minimized.avoided. It occurs when the disturbing frequencyis very near to the natural frequency. If steel mounts Range (III) This is known as the shock reductionare used, a critical vibration at resonance would go range and only occurs when the disturbingto infinity, but natural rubber limits the trans- frequency is lower than the natural frequency.missibility to around 10. If Butyl synthetic rubber is Generally it is only experienced with very softadopted, its damping properties reduce the peak rubber mounts and when the engine is initiallytransmissibility to about 21¤2. Unfortunately, high cranked for starting purposes and so quickly passesdamping rubber compounds such as Butyl rubber through this frequency ratio region.are temperature sensitive to both damping anddynamic stiffness so that during cold weather a Example An engine oscillates vertically on itsnoticeably harsher suspension of the engine results. flexible rubber mountings with a frequency of 800 Damping of the engine suspension mounting is vibrations per minute (vpm). With the informationnecessary to reduce the excessive movement of a provided answer the following questions:flexible mounting when passing through resonance,but at speeds above resonance more vibration is a) From the static deflection±frequency graph,transmitted to the chassis or body structure than Fig. 1.16, or by formula, determine the natural fre-would occur if no damping was provided. quency of vibration when the static deflection of the engine is 2 mm and then find the disturbing toRange (II) This is the recommended working natural frequency ratio. Comment on these results.range where the ratio of the disturbing frequency b) If the disturbing to natural frequency ratio isto that of the natural frequency of vibration of the increased to 2.5 determine the natural frequency Fig. 1.17 Relationship of transmissibility and the ratio of disturbing and natural frequencies for natural rubber, Butyl rubber and steel 18
  • 28. of vibration and the new static deflection of the 1.2.9 Subframe to body mountings engine. Comment of these conditions. (Figs 1.6 and 1.19) 30 30 One of many problems with integral body design isa) n0 ˆ p ˆ p the prevention of vibrations induced by the engine, x 0:002 transmission and road wheels from being transmitted 30 through the structure. Some manufacturers adopt a ˆ ˆ 670:84 vib/min subframe (Fig. 1.6(a, b and c)) attached by resilient 0:04472 mountings (Fig. 1.19(a and b)) to the body to which n 800 the suspension assemblies, and in some instances the ; ˆ ˆ 1:193 n0 670:84 engine and transmission, are attached. The mass of the subframes alone helps to damp vibrations. The ratio 1.193 is very near to the resonance It also simplifies production on the assembly line,condition and should be avoided by using softer and facilitates subsequent overhaul or repairs.mounts. In general, the mountings are positioned so that n 800 they allow strictly limited movement of theb) ˆ ˆ 2:5 n0 n0 subframe in some directions but provide greater freedom in others. For instance, too much lateral 800 freedom of a subframe for a front suspension ; n0 ˆ ˆ 320 vib/min 2:5 assembly would introduce a degree of instability 30 into the steering, whereas some freedom in vertical Now n0 ˆ p and longitudinal directions would improve the x quality of a ride. p 30 thus x ˆ n0 1.2.10 Types of rubber flexible mountings 2 30 30 2 A survey of typical rubber mountings used for ;xˆ ˆ power units, transmissions, cabs and subframes n0 320 are described and illustrated as follows: ˆ 0:008789 m or 8:789 mm A low natural frequency of 320 vib/min is well Double shear paired sandwich mounting (Fig.within the insulation range, therefore from either 1.18(a)) Rubber blocks are bonded between thethe deflection±frequency graph or by formula jaws of a `U shaped steel plate and a flat interleafthe corresponding rubber deflection necessary is plate so that a double shear elastic reaction takes8.789 mm when the engines static weight bears place when the mount is subjected to vertical load-down on the mounts. ing. This type of shear mounting provides a large degree of flexibility in the upright direction and1.2.8 Engine to body/chassis mountings thus rotational freedom for the engine unit aboutEngine mountings are normally arranged to its principal axis. It has been adopted for bothprovide a degree of flexibility in the horizontal engine and transmission suspension mountinglongitudinal, horizontal lateral and vertical axis of points for medium-sized diesel engines.rotation. At the same time they must have suffi-cient stiffness to provide stability under shockloads which may come from the vehicle travelling Double inclined wedge mounting (Fig. 1.18(b)) Theover rough roads. Rubber sprung mountings inclined wedge angle pushes the bonded rubbersuitably positioned fulfil the following functions: blocks downwards and outwards against the bent-up sides of the lower steel plate when loaded1 Rotational flexibility around the horizontal in the vertical plane. The rubber blocks are subjected longitudinal axis which is necessary to allow the to both shear and compressive loads and the propor- impulsive inertia and gas pressure components tion of compressive to shear load becomes greater of the engine torque to be absorbed by rolling of with vertical deflection. This form of mounting is the engine about the centre of gravity. suitable for single point gearbox supports.2 Rotational flexibility around both the horizontal lateral and the vertical axis to accommodate any horizontal and vertical shake and rock caused by Inclined interleaf rectangular sandwich mounting unbalanced reciprocating forces and couples. (Fig. 1.18(c)) These rectangular blocks are 19
  • 29. Fig. 1.18 (a±h) Types of rubber flexible mountings 20
  • 30. Fig. 1.18 contd 21
  • 31. Fig. 1.18 contddesigned to be used with convergent `V formation on either side of the power units bell housingengine suspension system where the blocks are at principal axis level may be used. Longitudinalinclined on either side of the engine. This configura- movement is restricted by the double `V formedtion enables the rubber to be loaded in both shear between the inner and two outer members seen inand compression with the majority of engine rota- a plan view. This `V and wedge configuration pro-tional flexibility being carried out in shear. Vertical vides a combined shear and compressive strain todeflection due to body pitch when accelerating or the rubber when there is a relative fore and aft move-braking is absorbed mostly in compression. Vertical ment between the engine and chassis, in addition toelastic stiffness may be increased without greatly that created by the vertical loading of the mount.effecting engine roll flexibility by having metal This mountings major application is for the rearspacer interleafs bonded into the rubber. mountings forming part of a four point suspension for heavy diesel engines.Double inclined wedge with longitudinal controlmounting (Fig. 1.18(d)) Where heavy vertical Metaxentric bush mounting (Fig. 1.18(e)) Whenloads and large rotational reactions are to be the bush is in the unloaded state, the steel innerabsorbed, double inclined wedge mounts positioned sleeve is eccentric relative to the outer one so that 22
  • 32. there is more rubber on one side of it than on the distortion within the rubber. Under small deflec-other. Precompression is applied to the rubber tion conditions the shear and compression isexpanding the inner sleeve. The bush is set so that almost equal, but as the load and thus deflectionthe greatest thickness of rubber is in compression increases, the proportion of compression over thein the laden condition. A slot is incorporated in shear loading predominates.the rubber on either side where the rubber is at its These mounts provide very good lateral stabilityminimum in such a position as to avoid stressing without impairing vertical deflection flexibility andany part of it in tension. progressive stiffness control. When used for road When installed, its stiffness in the fore and aft wheel axle suspension mountings, they offer gooddirection is greater than in the vertical direction, the insulation against road and other noises.ratio being about 2.5 : 1. This type of bush providesa large amount of vertical deflection with very littlefore and aft movement which makes it suitable for Flanged sleeve bobbin mounting with reboundrear gearbox mounts using three point power unit control (Fig. 1.19(a and b)) These mountingssuspension and leaf spring eye shackle pin bushes. have the rubber moulded partially around the outer flange sleeve and in between this sleeve and an inner tube. A central bolt attaches the inner tube to theMetacone sleeve mountings (Fig. 1.18(f and g)) body structure while the outer member is bolted onThese mounts are formed from male and female two sides to the subframe.conical sleeves, the inner male member being When loaded in the vertical downward direction,centrally positioned by rubber occupying the the rubber between the sleeve and tube walls will bespace between both surfaces (Fig. 1.18(f)). During in shear and the rubber on the outside of thevertical vibrational deflection, the rubber between flanged sleeve will be in compression.the sleeves is subjected to a combined shear and There is very little relative sideway movementcompression which progressively increases the stiff- between the flanged sleeve and inner tube due toness of the rubber as it moves towards full distor- rubber distortion. An overload plate limits the down-tion. The exposed rubber at either end overlaps the ward deflection and rebound is controlled by theflanged outer sleeve and there is an upper and lower plate and the amount and shape of rubberlower plate bolted rigidly to the ends of the inner trapped between it and the underside of the flangedsleeve. These plates act as both overload (bump) sleeve. A reduction of rubber between the flangedand rebound stops, so that when the inner member sleeve and lower plate (Fig. 1.19(a)) reduces thedeflects up or down towards the end of its move- rebound, but an increase in depth of rubber increasesment it rapidly stiffens due to the surplus rubber rebound (Fig. 1.19(b)). The load deflection charac-being squeezed in between. Mounts of this kind are teristics are given for both mounts in Fig. 1.19c.used where stiffness is needed in the horizontal These mountings are used extensively for body todirection with comparative freedom of movement subframe and cab to chassis mounting points.for vertical deflection. An alternative version of the Metacone mountuses a solid aluminium central cone with a flanged Hydroelastic engine mountings (Figs 1.20(a±c) andpedestal conical outer steel sleeve which can be 1.21) A flanged steel pressing houses and sup-bolted directly onto the chassis side member, see ports an upper and lower rubber spring diaphragm.Fig. 1.18(g). An overload plate is clamped between The space between both diaphragms is filled andthe inner cone and mount support arm, but no sealed with fluid and is divided in two by a separatorrebound plate is considered necessary. plate and small transfer holes interlink the fluid These mountings are used for suspension appli- occupying these chambers (Fig. 1.20(a and b)).cations such as engine to chassis, cab to chassis, Under vertical vibratory conditions the fluid willbus body and tanker tanks to chassis. be displaced from one chamber to the other through transfer holes. During downward deflec-Double inclined rectangular sandwich mounting tion (Fig. 1.20(b)), both rubber diaphragms are(Fig. 1.18(h)) A pair of rectangular sandwich subjected to a combined shear and compressiverubber blocks are supported on the slopes of a action and some of the fluid in the upper chambertriangular pedestal. A bridging plate merges the will be pushed into the lower and back again byresilience of the inclined rubber blocks so that way of the transfer holes when the rubber reboundsthey provide a combined shear and compressive (Fig. 1.20(a)). For low vertical vibratory frequencies, 23
  • 33. the movement of fluid between the chambers is unrestricted, but as the vibratory frequencies increase, the transfer holes offer increasing resist- ance to the flow of fluid and so slow down the up and down motion of the engine support arm. This damps and reduces the amplitude of mountings vertical vibratory movement over a number of cycles. A comparison of conventional rubber and hydroelastic damping resistance over the normal operating frequency range for engine mountings is shown in Fig. 1.20(c). Instead of adopting a combined rubber mount with integral hydraulic damping, separate diagon- ally mounted telescopic dampers may be used in conjunction with inclined rubber mounts to reduce both vertical and horizontal vibration (Fig. 1.21). 1.3 Fifth wheel coupling assembly (Fig. 1.22(a and b)) The fifth wheel coupling attaches the semi-trailer to the tractor unit. This coupling consists of a semi- circular table plate with a central hole and a vee section cut-out towards the rear (Fig. 1.22(b)). Attached underneath this plate are a pair of pivot- ing coupling jaws (Fig. 1.22(a)). The semi-trailer has an upper fifth wheel plate welded or bolted to the underside of its chassis at the front and in the centre of this plate is bolted a kingpin which faces downwards (Fig. 1.22(a)). When the trailer is coupled to the tractor unit, this upper plate rests and is supported on top of theFig. 1.19 (a±c) Flanged sleeve bobbin mounting with tractor fifth wheel table plate with the two halves ofrebound control the coupling jaws engaging the kingpin. To permit 24
  • 34. relative swivelling between the kingpin and jaws, the two interfaces of the tractor fifth wheel tables and trailer upper plate should be heavily greased. Thus, although the trailer articulates about the kingpin, its load is carried by the tractor table. Flexible articulation between the tractor and semi-trailer in the horizontal plane is achieved by permitting the fifth wheel table to pivot on hori- zontal trunnion bearings that lie in the same vertical plane as the kingpin, but with their axes at right angles to that of the tractors wheel base (Fig. 1.22(b)). Rubber trunnion rubber bushes normally provide longitudinal oscillations of about Æ10 . The fifth wheel table assembly is made from either a machined cast or forged steel sections, or from heavy section rolled steel fabrications, and the upper fifth wheel plate is generally hot rolled steel welded to the trailer chassis. The coupling locking system consisting of the jaws, pawl, pivot pins and kingpin is produced from forged high carbon man- ganese steels and the pressure areas of these com- ponents are induction hardened to withstand shock loading and wear. 1.3.1 Operation of twin jaw coupling (Fig. 1.23(a±d)) With the trailer kingpin uncoupled, the jaws will be in their closed position with the plunger withdrawn from the lock gap between the rear of the jaws, which are maintained in this position by the pawl contacting the hold-off stop (Fig. 1.23(a)). WhenFig. 1.20 (a±c) Hydroelastic engine mount coupling the tractor to the trailer, the jaws of the 25
  • 35. Fig. 1.21 Diagonally mounted hydraulic dampers suppress both vertical and horizontal vibrationsfifth wheel strike the kingpin of the trailer. The spring load notched pawl will then snap over thejaws are then forced open and the kingpin enters jaw projection to lock the kingpin in the couplingthe space between the jaws (Fig. 1.23(b)). The king- position (Fig. 1.24(c)). The securing pin shouldpin contacts the rear of the jaws which then then be inserted through the pull lever and tableautomatically pushes them together. At the same eye holes. When the tractor is driving forward, thetime, one of the coupler jaws causes the trip pin to reaction on the kingpin increases the lockingstrike the pawl. The pawl turns on its pivot against force between the jaw projection and the notchedthe force of the spring, releasing the plunger, allow- pawl.ing it to be forced into the jaws lock gap by its To disconnect the coupling, lift out the securingspring (Fig. 1.23(c)). When the tractor is moving, pin and pull the release hand lever fully outthe drag of the kingpin increases the lateral force of (Fig. 1.24(d)). With both the tractor and trailerthe jaws on the plunger. stationary, the majority of the locking force To disconnect the coupling, the release hand applied to notched pawl will be removed so thatlever is pulled fully back (Fig. 1.23(d)). This with very little effort, the pawl is able to swing cleardraws the plunger clear of the rear of the jaws of the jaw in readiness for uncoupling, that is, byand, at the same time, allows the pawl to swing just driving the tractor away from the trailer. Thusround so that it engages a projection hold-off stop the jaw will simply swivel allowing the kingpin tosituated at the upper end of the plunger, thus jam- pull out and away from the jaw.ming the plunger in the fully out position in readi-ness for uncoupling. 1.4 Trailer and caravan drawbar couplings 1.4.1 Eye and bolt drawbar coupling for heavy1.3.2 Operation of single jaw and pawl coupling goods trailers (Figs 1.25 and 1.26)(Fig. 1.24(a±d)) Drawbar trailers are normally hitched to the truckWith the trailer kingpin uncoupled, the jaw will be by means of an `A frame drawbar which is coupledheld open by the pawl in readiness for coupling by means of a towing eye formed on the end of the(Fig. 1.24(a)). When coupling the tractor to the drawbar (Fig. 1.25). When coupled, the towing eyetrailer, the jaw of the fifth wheel strikes the kingpin hole is aligned with the vertical holes in the upperof the trailer and swivels the jaw about its pivot pin and lower jaws of the truck coupling and an eyeagainst the return spring, slightly pushing out the bolt passes through both coupling jaws and draw-pawl (Fig. 1.24(b)). Further rearward movement of bar eye to complete the attachment (Fig. 1.26).the tractor towards the trailer will swing the jaw Lateral drawbar swing is permitted owing to theround until it traps and encloses the kingpin. The eye bolt pivoting action and the slots between the 26
  • 36. Fig. 1.22 (a and b) Fifth wheel coupling assemblyjaws on either side. Aligning the towing eye to the as a damping media between the towing vehicle andjaws is made easier by the converging upper and trailer. These rubber blocks also permit additionallower lips of the jaws which guide the towing eye as deflection of the coupling jaw shaft relative to thethe truck is reversed and the jaws approach the draw beam under rough abnormal operating con-drawbar. Isolating the coupling jaws from the ditions, thus preventing over-straining the drawbartruck draw beam are two rubber blocks which act and chassis system. 27
  • 37. Fig. 1.23 (a±d) Fifth wheel coupling with twin jaws plunger and pawl 28
  • 38. Fig. 1.24 (a±d) Fifth wheel coupling with single jaw and pawl 29
  • 39. 1.4.2 Ball and socket towing bar coupling for light caravan/trailers (Fig. 1.27) Light trailers or caravans are usually attached to the rear of the towing car by means of a ball and socket type coupling. The ball part of the attach- ment is bolted onto a bracing bracket fitted directly to the boot pan or the towing load may be shared out between two side brackets attached to the rear longitudinal box-section members of the body. A single channel section or pair of triangularly arranged angle-section arms may be used to form the towbar which both supports and draws the trailer. Attached to the end of the towbar is the socket housing with an internally formed spherical cavity. This fits over the ball member of the coupling so that it forms a pivot joint which can operate in both the horizontal and vertical plane (Fig. 1.27). To secure the socket over the ball, a lock device must be incorporated which enables the coupling toFig. 1.25 Drawbar trailer be readily connected or disconnected. This lock may take the form of a spring-loaded horizontally positioned wedge with a groove formed across its The coupling jaws, eye bolt and towing eye are top face which slips underneath and against thegenerally made from forged manganese steel with ball. The wedge is held in the closed engaged pos-induction hardened pressure areas to increase the ition by a spring-loaded vertical plunger which haswear resistance. a horizontal groove cut on one side. An uncoupling lever engages the plungers groove so that when the coupling is disconnected the lever is squeezed to liftOperation of the automatic drawbar coupling and release the plunger from the wedge. At the(Fig. 1.26) In the uncoupled position the eyebolt same time the whole towbar is raised by the handleis held in the open position ready for coupling to clear the socket and from the ball member.(Fig. 1.26(a)). When the truck is reversed, the jaws Coupling the tow bar to the car simply reversesof the coupling slip over the towing eye and in the the process, the uncoupling lever is again squeezedprocess strike the conical lower end of the eye bolt against the handle to withdraw the plunger and the(Fig. 1.26(b)). Subsequently, the eye bolt will lift. This socket housing is pushed down over the ball mem-trips the spring-loaded wedge lever which now rotates ber. The wedge moves outwards and allows the ballclockwise so that it bears down on the eye bolt. to enter the socket and immediately the wedgeFurther inward movement of the eye bolt between springs back into the engaged position. Releasingthe coupling jaws aligns the towing eye with the eye the lever and handle completes the coupling bybolt. The spring pressure now acts through the wedge permitting the plunger to enter the wedge locklever to push the eye bolt through the towing eye and groove.the lower coupling jaw (Fig. 1.26(c)). When the eye Sometimes a strong compression spring is inter-bolt stop-plate has been fully lowered by the spring posed between the socket housing member and thetension, the wedge lever will slot into its groove towing (draw) bar to cushion the shock load whenformed in the centre of the eye bolt so that it locks the car/trailer combination is initially driven awaythe eye bolt in the coupled position. from a standstill. To uncouple the drawbar, the handle is pulledupwards against the tension of the coil springmounted on the wedge level operating shaft 1.5 Semi-trailer landing gear (Fig. 1.28)(Fig. 1.26(d)). This unlocks the wedge, freeing the Landing legs are used to support the front of theeyebolt and then raises the eye bolt to the semi-trailer when the tractor unit is uncoupled.uncoupled position where the wedge lever jams it Extendable landing legs are bolted vertically toin the open position (Fig. 1.26(a)). each chassis side-member behind the rear wheels of 30
  • 40. Fig. 1.26 (a±e) Automatic drawbar coupling 31
  • 41. 1.6 Automatic chassis lubrication system 1.6.1 The need for automatic lubrication system (Fig. 1.29) Owing to the heavy loads they carry commercial vehicles still prefer to use metal to metal joints which are externally lubricated. Such joints are kingpins and bushes, shackle pins and bushes, steering ball joints, fifth wheel coupling, parking brake linkage etc. (Fig. 1.29). These joints require lubricating in proportion to the amount of relative movement and the loads exerted. If lubrication is to be effective in reducing wear between the moving parts, fresh oil must be pumped between the joints frequently. This can best be achieved by incorporating an automatic lubrication system which pumps oil to the bearingsFig. 1.27 Ball and socket caravan/trailer towing surfaces in accordance to the distance travelled byattachment the vehicle.the tractor unit, just sufficiently back to clear the 1.6.2 Description of airdromic automatic chassisrear tractor road wheels when the trailer is coupled lubrication system (Fig. 1.30)and the combination is being manoeuvred This lubrication system comprises four major com-(Fig. 1.28(a)). To provide additional support for ponents; a combined pump assembly, a power unit,the legs, bracing stays are attached between the legs an oil unloader valve and an air control unit.and from the legs diagonally to the chassis cross-member (Fig. 1.28(b)). The legs consist of inner and outer high tensile Pump assembly (Fig. 1.30) The pump assemblysteel tubes of square section. A jackscrew with a consists of a circular housing containing a ratchetbevel wheel attached at its top end supported by the operated drive (cam) shaft upon which areouter leg horizontal plate in a bronze bush bearing. mounted one, two or three single lobe cams (onlyThe jawscrew fits into a nut which is mounted at one cam shown). Each cam operates a row of 20the top of the inner leg and a taper roller bearing pumping units disposed radially around the pumprace is placed underneath the outer leg horizontal casing, the units being connected to the chassissupport plate and the upper part of the jackscrew bearings by nylon tubing.to minimize friction when the screw is rotated (Fig.1.28(b)). The bottom ends of the inner legs maysupport either twin wheels, which enable the trailerto be manoeuvred, or simply flat feet. The latter are Power unit (Fig. 1.30) This unit comprises aable to spread the load and so permit greater load cylinder and spring-loaded air operated pistoncapacity. which is mounted on the front face of the pump To extend or retract the inner legs, a winding assembly housing, the piston rod being connectedhandle is attached to either the low or high speed indirectly to the drive shaft ratchet wheel by way ofshaft protruding from the side of the gearbox. The a ratchet housing and pawl.upper high speed shaft supports a bevel pinionwhich meshes with a vertically mounted bevelwheel forming part of the jackscrew. Oil unloader valve (Fig. 1.30) This consists of a Rotating the upper shaft imparts motion directly shuttle valve mounted on the front of the pumpto the jackscrew through the bevel gears. If greater assembly housing. The oil unloader valve allows airleverage is required to raise or lower the front of the pressure to flow to the power unit for the powertrailer, the lower shaft is engaged and rotated. stroke. During the exhaust stroke, however, whenThis provides a gear reduction through a com- air flow is reversed and the shuttle valve is liftedpound gear train to the upper shaft which then from its seat, any oil in the line between the powerdrives the bevel pinion and wheel and hence the unit and the oil unloader valve is then discharged tojackscrew. atmosphere. 32
  • 42. Fig. 1.28 (a and b) Semi-trailer landing gear 33
  • 43. Fig. 1.29 Tractor unit automatic lubrication systemAir control unit (Fig. 1.30) This unit is mounted housing. Because the pawl meshes with one of theon the gearbox and is driven via the speedometer ratchet teeth and the ratchet wheel forms part oftake-off point. It consists of a worm and wheel drive the camshaft, air pressure in the power cylinder willwhich operates an air proportioning control partially rotate both the ratchet and pawl housingunit. This air proportioning unit is operated by a and the camshaft clockwise. The cam (or cams) aresingle lift face cam which actuates two poppet in contact with one or more pump unit, and so eachvalves, one controlling air supply to the power partial rotation contributes to a proportion of theunit, the other controlling the exhaust air from the jerk plunger and barrel pumping cycle of each unitpower unit. (Fig. 1.30). As the control unit face cam continues to rotate, the inlet poppet inlet valve is closed and the exhaust1.6.3 Operation of airdromic automatic chassis poppet valve opens. Compressed air in the air con-lubrication system (Fig. 1.30) trol unit and above the oil control shuttle valve willAir from the air brake auxiliary reservoir passes by now escape through the air control unit exhaustway of the safety valve to the air control (propor- port to the atmosphere. Consequently the com-tioning) unit inlet valve. Whilst the inlet valve is pressed air underneath the oil unloader shuttleheld open by the continuously rotating face cam valve will be able to lift it and any trapped air andlobe, air pressure is supplied via the oil unloader oil in the power cylinder will now be released viavalve to the power unit attached to the multipump the hole under the exhaust port. The power unitassembly housing. The power unit cylinder is sup- piston will be returned to its innermost position byported by a pivot to the pump assembly casing, the spring and in doing so will rotate the ratchetwhilst the piston is linked to the ratchet and pawl and pawl housing anti-clockwise. The pawl is thus 34
  • 44. Fig. 1.30 Airdromic automatic chassis lubrication system 35
  • 45. able to slip over one or more of the ratchet teeth to When the individual lubrication pump unitstake up a new position. The net result of the power primary plunger is in its outermost position, oilcylinder being charged and discharged with com- surrounding the barrel will enter the inlet port,pressed air is a slow but progressive rotation of the filling the space between the two plungers. As thecamshaft (Fig. 1.30). cam rotates and the lobe lifts the primary plunger, A typical worm drive shaft to distance travelled it cuts off the inlet port. Further plunger rise willrelationship is 500 revolutions per 1 km. For 900 partially push out the secondary plunger and soworm drive shaft revolutions the pumping cam open the check valve. Pressurised oil will thenrevolves once. Therefore, every chassis lubrication pass between the loose fitting secondary plungerpoint will receive one shot of lubricant in this and barrel to lubricate the chassis moving part itdistance. services (Fig. 1.30). 36
  • 46. 2 Friction clutch2.1 Clutch fundamentals 2.1.3 Multi-pairs of rubbing surfaces (Fig. 2.1)Clutches are designed to engage and disengage the An alternative approach to raising the transmittedtransmission system from the engine when a vehicle torque capacity of the clutch is to increase theis being driven away from a standstill and when the number of pairs of rubbing surfaces. Theoreticallygearbox gear changes are necessary. The gradual the torque capacity of a clutch is directly propor-increase in the transfer of engine torque to the tional to the number of pairs of surfaces for a giventransmission must be smooth. Once the vehicle is clamping load. Thus the conventional single drivenin motion, separation and take-up of the drive for plate has two pairs of friction faces so that a twingear selection must be carried out rapidly without or triple driven plate clutch for the same springany fierceness, snatch or shock. thrust would ideally have twice or three times the torque transmitting capacity respectively of that of2.1.1 Driven plate inertia the single driven plate unit (Fig. 2.1). However,To enable the clutch to be operated effectively, the because it is very difficult to dissipate the extradriven plate must be as light as possible so that heat generated in a clutch unit, a larger safety factorwhen the clutch is disengaged, it will have the mini- is necessary per driven plate so that the torquemum of spin, i.e. very little flywheel effect. Spin capacity is generally only of the order 80% per pairprevention is of the utmost importance if the vari- of surfaces relative to the single driven plate clutch.ous pairs of dog teeth of the gearbox gears, be theyconstant mesh or synchromesh, are to align in the 2.1.4 Driven plate wear (Fig. 2.1)shortest time without causing excessive pressure, Lining life is also improved by increasing thewear and noise between the initial chamfer of the number of pairs of rubbing surfaces because weardog teeth during the engagement phase. is directly related to the energy dissipation per unit Smoothness of clutch engagement may be area of contact surface. Ideally, by doubling theachieved by building into the driven plate some surface area as in a twin plate clutch, the energysort of cushioning device, which will be discussed input per unit lining area will be halved for a givenlater in the chapter, whilst rapid slowing down of slip time which would result in a 50% decrease inthe driven plate is obtained by keeping the diameter, facing wear. In practice, however, this rarely occurscentre of gravity and weight of the driven plate to (Fig. 2.1) as the wear rate is also greatly influencedthe minimum for a given torque carrying capacity. by the peak surface rubbing temperature and the intermediate plate of a twin plate clutch operates at2.1.2 Driven plate transmitted torque capacity a higher working temperature than either the fly-The torque capacity of a friction clutch can be wheel or pressure plate which can be more effect-raised by increasing the coefficient of friction of ively cooled. Thus in a twin plate clutch, half thethe rubbing materials, the diameter and/or the energy generated whilst slipping must be absorbedspring thrust sandwiching the driven plate. The by the intermediate plate and only a quarter eachfriction lining materials now available limit the by the flywheel and pressure plate. This is usuallycoefficient of friction to something of the order of borne out by the appearance of the intermediate0.35. There are materials which have higher coeffi- plate and its corresponding lining faces showingcient of friction values, but these tend to be evidence of high temperatures and increased wearunstable and to snatch during take-up. Increasing compared to the linings facing the flywheel andthe diameter of the driven plate unfortunately pressure plate. Nevertheless, multiplate clutchesraises its inertia, its tendency to continue spinning do have a life expectancy which is more or lesswhen the driven plate is freed while the clutch is in related to the number of pairs of friction faces forthe disengaged position, and there is also a limit to a given diameter of clutch.the clamping pressure to which the friction lining For heavy duty applications such as thosematerial may be subjected if it is to maintain its required for large trucks, twin driven plates arefriction properties over a long period of time. used, while for high performance cars where very 37
  • 47. Fig. 2.1 Relationship of torque capacity wear rate and pairs of rubbing faces for multiplate clutchrapid gear changes are necessary and largeamounts of power are to be developed, smalldiameter multiplate clutches are preferred.2.2 Angular driven plate cushioning and torsionaldamping (Figs 2.2±2.8)2.2.1 Axial driven plate friction lining cushioning(Figs 2.2, 2.3 and 2.4)In its simplest form the driven plate consists ofa central splined hub. Mounted on this hub is athin steel disc which in turn supports, by means ofa ring of rivets, both halves of the annular frictionlinings (Figs 2.2 and 2.3). Axial cushioning between the friction liningfaces may be achieved by forming a series of evenlyspaced `T slots around the outer rim of the disc.This then divides the rim into a number of seg-ments (Arcuate) (Fig. 2.4(a)). A horseshoe shapeis further punched out of each segment. The centralportion or blade of each horseshoe is given a per-manent set to one side and consecutive segmentshave opposite sets so that every second segment isriveted to the same friction lining. The alternativeset of these central blades formed by the horseshoepunch-out spreads the two half friction linings apart. An improved version uses separately attached, verythin spring steel segments (borglite) (Fig. 2.4(b)), pos-itioned end-on around a slightly thicker disc plate.These segments are provided with a wavy `set so asto distance the two half annular friction linings. Both forms of crimped spring steel segmentssituated between the friction linings provide Fig. 2.2 Clutch driven centre plate (pictorial view) 38
  • 48. Fig. 2.3 Clutch driven centre plate (sectional view) Fig. 2.4 (a and b) Driven plate cushion take-upprogressive take-up over a greater pedal travel and The spring take-up characteristics of the drivenprevent snatch. The separately attached spring plate are such that when the clutch is initiallysegments are thinner than the segments formed out engaged, the segments are progressively flattened soof the single piece driven plate, so that the squeeze that the rate of increase in clamping load is providedtake-up is generally softer and the spin inertia of the by the rate of reaction offered by the springthinner segments is noticeably reduced. segments (Fig. 2.5). This first low rate take-up A further benefit created by the spring segments period is followed by a second high rate engage-ensures satisfactory bedding of the facing material ment, caused by the effects of the pressure plateand a more even distribution of the work load. In springs exerting their clamping thrust as they areaddition, cooling between the friction linings occurs allowed to expand against the pressure plate andwhen the clutch is disengaged which helps to sta- so sandwich the friction lining between the flywheelbilise the frictional properties of the face material. and pressure plate faces. The advantages of axial cushioning of the facelinings provide the following: 2.2.2 Torsional damping of driven platea) Better clutch engagement control, allowing lower engine speeds to be used at take-up thus Crankshaft torsional vibration (Fig. 2.6) Engine prolonging the life of the friction faces. crankshafts are subjected to torsional wind-upb) Improved distribution of the friction work over and vibration at certain speeds due to the power the lining faces reduces peak operating tempera- impulses. Superimposed onto some steady mean tures and prevents lining fade, with the resulting rotational speed of the crankshaft will be additional reduction in coefficient of friction and subse- fluctuating torques which will accelerate and decel- quent clutch slip. erate the crankshaft, particularly at the front pulley 39
  • 49. the gear teeth, wear, and noise in the form of gear clatter. To overcome the effects of crankshaft torsional vibrations a torsion damping device is normally incorporated within the driven plate hub assembly which will now be described and explained. Construction and operation of torsional damper springs (Figs 2.2, 2.3 and 2.7) To transmit torque more smoothly and progressively during take-up of normal driving and to reduce torsional oscillations being transmitted from the crankshaft to the trans- mission, compressed springs are generally arranged circumferentially around the hub of the driven plate (Figs 2.2 and 2.3). These springs are inserted in elongated slots formed in both the flange of the splined hub and the side plates which enclose the hubs flange (Fig. 2.3). These side plates are riveted together by either three or six rivet posts which pass through the flanged hub limit slots. This thusFig. 2.5 Characteristics of driven plate axial clampingload to deflection take-up provides a degree of relative angular movement between hub and side plates. The ends of the helical coil springs bear against both central hub flange and the side plates. Engine torque is therefore transmitted from the friction face linings and side plates through the springs to the hub flange, so that any fluctuation of torque will cause the springs to compress and rebound accordingly. Multistage driven plate torsional spring dampers may be incorporated by using a range of different springs having various stiffnesses and spring loca- tion slots of different lengths to produce a variety of parabolic torsional load±deflection characteris- tics (Fig. 2.7) to suit specific vehicle applications. The amount of torsional deflection necessary varies for each particular application. For example, with a front mounted engine and rear wheel drive vehicle, a moderate driven plate angular movement is necessary, say six degrees, since the normal trans- mission elastic wind-up is almost adequate, but with an integral engine, gearbox and final drive arrange- ment, the short transmission drive length necessit-Fig. 2.6 Characteristics of crankshaft torsional ates considerably more relative angular deflection,vibrations undamped and damped say twelve degrees, within the driven plate hub assembly to produce the same quality of take-up.end and to a lesser extent the rear flywheel end(Fig. 2.6). If the flywheel end of the crankshaft Construction and operation of torsional damperwere allowed to twist in one direction and then the washers (Figs 2.2, 2.3 and 2.8) The torsionalother while rotating at certain critical speeds, the energy created by the oscillating crankshaft isoscillating angular movements would take up the partially absorbed and damped by the frictionbacklash between meshing gear teeth in the transmis- washer clutch situated on either side of the hubsion system. Consequently, the teeth of the driving flange (Figs 2.2 and 2.3). Axial damping load isgears would be moving between the drive (pressure achieved by a Belleville dished washer springside) and non-drive tooth profiles of the driven gears. mounted between one of the side plates and a fourThis would result in repeated shockloads imposed on lug thrust washer. 40
  • 50. Fig. 2.7 Characteristics of driven plate torsional spring Fig. 2.8 Characteristics of driven plate torsionaltorques to deflection take-up damper thrust spring The outer diameter of this dished spring pressesagainst the side plate and the inner diameter pushes i) a high degree of interface contamination toler-onto the lugged thrust washer. In its free state ance without affecting its friction take-up andthe Belleville spring is conical in shape but when grip characteristics.assembled it is compressed almost flat. As the fric-tion washers wear, the dished spring cone angle 2.3.1 Asbestos-based linings (Figs 2.2 and 2.3)increases. This exerts a greater axial thrust, but Generally, clutch driven plate asbestos-based lin-since the distance between the side plate and lugged ings are of the woven variety. These woven liningsthrust washer has increased, the resultant clamping are made from asbestos fibre spun around lengthsthrust remains almost constant (Fig. 2.8). of brass or zinc wire to make lengths of threads which are both heat resistant and strong. The woven cloth can be processed in one of two ways:2.3 Clutch friction materialsClutch friction linings or buttons are subjected to a) The fibre wire thread is woven into a cloth andsevere rubbing and generation of heat for relatively pressed out into discs of the required diameter,short periods. Therefore it is desirable that they followed by stitching several of these discshave a combination of these properties: together to obtain the desired thickness. The resultant disc is then dipped into resin to bonda) Relatively high coefficient of friction under the woven asbestos threads together. operating conditions, b) The asbestos fibre wire is woven in three dimen-b) capability of maintaining friction properties sions in the form of a disc to obtain in a single over its working life, stage the desired thickness. It is then pressedc) relatively high energy absorption capacity for into shape and bonded together by again dip- short periods, ping it into a resin solution. Finally, the rigidd) capability of withstanding high pressure plate lining is machined and drilled ready for riveting compressive loads, to the driven plate.e) capability of withstanding bursts of centrifugal force when gear changing, Development in weaving techniques has, inf) adequate shear strength to transmit engine certain cases, eliminated the use of wire coring so torque, that asbestos woven lining may be offered as eitherg) high level of cyclic working endurance without non- or semi-metallic to match a variety of working the deterioration in friction properties, conditions.h) good compatibility with cast iron facings over Asbestos is a condensate produced by the solidi- the normal operating temperature range, fication of rock masses which cool at differential 41
  • 51. rates. When the moisture content of one layer these fibres are difficult to inhale because of theiris transferred to another, fibres are produced on shape and size.solidification from which, as a result of high com-pression, these brittle, practically straight and 2.3.3 Metallic friction materialsexceptionally fine needle-like threads are made. Metallic and semi-metallic facings have been onlyDuring processing, these break down further with moderately successful. The metallic linings area diameter of less than 0.003 mm. They exhibit a normally made from either sintered iron or copper-length/thickness ratio of at least three to one. It is based sintered bronze and the semi-metallic facingsthese fine fibres which can readily be inhaled into from a mixture of organic and metallic materials.the lungs which are so dangerous to health. Metallic lining materials are made from a powder The normal highest working temperature below produced by crushing metal or alloy pieces intowhich these asbestos linings will operate satisfac- many small particles. They are then compressedtorily giving uniform coefficient of friction between and heated in moulds until sufficient adhesion and0.32 and 0.38 and a reasonable life span is about densification takes place. This process is referred to260 C. Most manufacturers of asbestos-based as sintering. The metallic rings are then ground flatlinings quote a maximum temperature (something and are then riveted back to back onto the drivenlike 360 C) beyond which the lining, if operated plate.continuously or very frequently, will suffer damage, Generally the metallic and semi-metallic liningswith consequent alteration to its friction charac- have a higher coefficient of friction, can operate atteristics and deterioration in wear resistance. higher working temperatures, have greater torque capacity and have extended life compared to that2.3.2 Asbestos substitute friction material of the organic asbestos based linings. The major(Figs 2.2 and 2.3) disadvantages of metallic materials are theirThe DuPont Company has developed a friction relatively high inertia, making it difficult to obtainmaterial derived from aromatic polyamide fibres rapid gear changes; high quality flywheel and pres-belonging to the nylon family of polymers and it sure plate. Cast iron must be used to match theirhas been given the trade name Kevlar aramid. friction characteristics and these facings are more The operating properties relative to asbestos expensive than organic materials.based linings are as follows: 1 High endurance performance over its normal 2.3.4 Cerametallic friction materials (Fig. 2.9) working pressure and temperature range. Cerametallic button friction facings are becoming 2 It is lighter in weight than asbestos material increasingly popular for heavy duty clutches. therefore a reduction in driven plate spin short- Instead of a full annular shaped lining, as with ens the time required for gear changing. organic (asbestos or substitute) friction materials, 3 Good take-up characteristics, particularly with four or six cerametallic trapezoidal-shaped buttons vehicles which were in the past prone to grab. are evenly spaced on both sides around the driven 4 Weight for weight Kevlar has five times the plate. The cerametallic material is made from a tensile strength of steel. powder consisting mainly of ceramic and copper. 5 Good centrifugal strength to withstand lining It is compressed into buttons and heated so that disintegration as a result of sudden acceleration the copper melts and flows around each particle of which may occur during the changing of gears. solid ceramic. After solidification, the copper 6 Stable rubbing properties at high operating forms a strong metal-ceramic interface bond. temperatures. It is not until a temperature of These buttons are then riveted to the clutch driven 425 C is reached that it begins to break down plate and then finally ground flat. and then it does not simply become soft and The inherent advantages of these cerametallic- melt, but steadily changes to carbon, the disin- lined driven plates are: tegration process being completed at about 1 A very low inertia (about 10% lower than the 500 C. organic disc and 45% lower than a comparable Kevlar exists in two states; firstly as a 0.12 mm sintered iron disc). Consequently it will result inthick endless longitudinal fibre, which has a cut quicker gear changes and, in the case of synchron-length varying between 6 and 100 mm, and secondly ized transmission, will increase synchronizer life.in the form of an amorphous structure of crushed 2 A relatively high and stable coefficient of friction,and ground fibre known as pulp. In either form providing an average value in the region of 42
  • 52. flywheel and pressure plate facings. A prolonged development programme has virtually elimin- ated this problem and has considerably extended the driven plate life span compared to driven plates using organic (asbestos-based) annular disc linings. 2.4 Clutch drive and driven member inspection This inspection entails the examination of both the driven plate linings and the flywheel and pressure plate facings and will now be considered. 2.4.1 Driven plate lining face inspection Driven plate friction facings should, after a short period of service, give a polished appearance due to the frequent interface rubbing effect. This smooth and polished condition will provide the greatest friction grip, but it must not be confused with a glazed surface created by the formation of films of grease or oil worked into the rubbing surfaces, heated and oxidized. A correctly bedded-in friction facing will appear highly polished through which the grain of the friction material can be clearly seen. When inFig. 2.9 Clutch driven plate with ceramic facings perfect condition, these polished facings are of a grey or mid-brown colour. A very small amount of lubricant on the facings will burn off due to the generated heat. This will only slightly darken the 0.4, which increases the torque capacity of facings, but providing polished facings remain so clutches using these driven plates. that the grain of the material can be clearly distin-3 The capability of operating at high working guished, it does not reduce its effectiveness. temperatures of up to 440 C for relatively long Large amounts of lubricant gaining access to the periods without showing signs of fade. friction surfaces may result in the following:4 Button type driven plates expose more than 50% of the flywheel and pressure plate surfaces to the a) The burning of the grease or oil may leave a carbon atmosphere during clutch engagement, so that deposit and a high glaze, this hides the grain of the heat transfer to the surrounding by convection material and is likely to cause clutch slip. may be improved by as much as 100%. b) If the grease or oil is only partially burnt and5 Button type friction pads do not suffer from leaves a resinous deposit on the facings it may warpage as do full ring metallic or organic linings result in a fierce clutch and may in addition and therefore are less prone to distort and cause produce clutch spin caused by the rubbing inter- clutch drag. faces sticking.6 Button type friction pads permit the dust worn c) If both carbon and resinous deposits are formed from the friction surfaces to be thrown clear of on the linings, clutch judder may develop during the clutch areas, thus preventing the possibility clutch take-up. of any trapped work-hardened particles from scoring the friction faces. 2.4.2 Flywheel and pressure plate facing inspection7 Cerametallic materials are not as sensitive to Cast iron flywheel or pressure plate faces should grease and oil contamination as organic asbestos have a smooth polished metallic appearance, but based linings. abnormal operating conditions or badly worn8 The early ceramic-metallic friction buttons had a driven plate linings may be responsible for the poor reputation as they tended to wear tracks in following defects: 43
  • 53. a) Overheated clutch friction faces can be identi- observe the reading. Acceptable end float values fied by colouring of the swept polished tracks. are normally between 0.08 and 0.30 mm. The actual surface temperatures reached can be distinguished broadly by the colours; straw, brown, purple and blue which relate to 240 C, 2.5.2 Crankshaft flywheel flange runout 260 C, 280 C and 320 C respectively. (Fig. 2.10(a))b) Severe overheating will create thermal stresses The crankshaft flange flywheel joint face must be within the cast iron mass of the flywheel and perpendicular to its axis of rotation with no permis- pressure plate, with the subsequent appearance sible runout. To check for any misalignment, keep of radial hairline cracks. the dial gauge assembly mounted as for the endc) Excessively worn driven plate linings with float check. Zero gauge the dial and rotate the exposed rivets and trapped work-hardened crankshaft by hand for one complete revolution dust particles will cause scoring of the rubbing whilst observing any dial movement. Investigate faces in the form of circular grooves. further if runout exists.2.5 Clutch misalignment (Fig. 2.10(a±d))Clutch faults can sometimes be traced to mis- 2.5.3 Flywheel friction face and rim face runoutalignment of the crankshaft to flywheel flange (Fig. 2.10(a and b))joint, flywheel housing and bell housing. Therefore, When the flywheel is centred by the crankshaft axis,if misalignment exists, the driven plate plane of it is essential that the flywheel friction face and rimrotation will always be slightly skewed to that of rotate perpendicularly to the crankshaft axis.the restrained hub which is made to rotate about Mount the dial gauge magnetic base to thethe spigot shafts axis. Misalignment is generally engine flywheel housing. First set the indicatorresponsible for the following faults: pointer against the friction face of the flywheel near the outer edge (Fig. 2.10(a and b)) and set1 Rapid wear on the splines of the clutch driven gauge to zero. Turn the flywheel one revolution plate hub, this being caused mainly by the tilted and observe the amount of variation. Secondly hub applying uneven pressure over the interface reset indicator pointer against the flywheel rim length of the splines. and repeat the test procedure (Fig. 2.10(b)). Max-2 The driven plate breaking away from the splined imum permissible runout in both tests is 0.02 mm hub due to the continuous cyclic flexing of the per 20 mm of flywheel radius. Thus with a 300 mm plate relative to its hub. diameter clutch fitted, maximum run-out would be3 Excessively worn pressure plate release mech- 0.15 mm. Repeat both tests 2 or 3 times and com- anism, causing rough and uneven clutch pare readings to eliminate test error. engagement.4 Fierce chattering or dragging clutch resulting in difficult gear changing. 2.5.4 Flywheel housing runout (Fig. 2.10(c)) If excessive clutch drag, backlash and poor When the gearbox bell housing is centred by thechanges are evident and the faults cannot be inside diameter and rear face of the engine flywheelcorrected, then the only remedy is to remove both housing, it is essential that the inside diameter andgearbox and clutch assembly so that the flywheel rear face of the housing should be concentric andhousing alignment can be assessed (Fig. 2.10). parallel respectively with the flywheel. Mount the dial gauge magnetic base to the fly-2.5.1 Crankshaft end float (Fig. 2.10(a)) wheel friction face and position. Set the indicatorBefore carrying out engine crankshaft, flywheel or pointer against the face of the housing. Make sureflywheel housing misalignment tests, ensure that that the pointer is not in the path of the fixing holesthe crankshaft end float is within limits. (Otherwise in the housing face or else incorrect readings mayinaccurate run-out readings may be observed.) result. Zero the indicator and observe the reading To measure the crankshaft end float, mount the whilst the crankshaft is rotated one complete revolu-magnetic dial gauge base to the back of the flywheel tion. Reset the indicator pointer against the intern-housing and position the indicator pointer against ally machined recess of the clutch housing and repeatthe crankshaft flanged end face. Zero the dial gauge the test procedure. Maximum permissible runout isand with the assistance of a suitable lever, force the 0.20 mm. Repeat both tests two or three times andcrankshaft back and forth and, at the same time, compare readings to eliminate errors. 44
  • 54. Fig. 2.10 (a±d) Crankshaft flywheel and clutch housing alignment 45
  • 55. 2.5.5 Detachable bell housing runout(Fig. 2.10(c and d))When the gearbox bell housing is located by dowelpins instead of the inside diameter of the engineflywheel housing (Fig. 2.10(c)) (shouldered bellhousing), it is advisable to separate the clutch bellhousing from the gearbox and mount it to theflywheel housing for a concentric check. Mount the dial gauge magnetic base onto theflywheel friction face and position the indicatorpointer against the internal recess of the bellhousing gearbox joint bore (Fig. 2.10(d)). Set thegauge to zero and turn the crankshaft by hand onecomplete revolution. At the same time, observethe dial gauge reading. Maximum permissible runout should not exceed0.25 mm.2.6 Pull type diaphragm clutch (Fig. 2.11)With this type of diaphragm clutch, the major com-ponents of the pressure plate assembly are a cast ironpressure plate, a spring steel diaphragm disc and a Fig. 2.11 Diaphragm single plate pull type clutchlow carbon steel cover pressing (Fig. 2.11). To actuatethe clutch release, the diaphragm is made to pivotbetween a pivot ring positioned inside the rear of thecover and a raised circumferential ridge formed on 2.7 Multiplate diaphragm type clutch (Fig. 2.12)the back of the pressure plate. The diaphragm disc is These clutches basically consist of drive and drivendivided into fingers caused by radial slits originating plate members. The drive plates are restrained fromfrom the central hole. These fingers act both as leaf rotating independently by interlocking lugs andsprings to provide the pressure plate thrust and as slots which permit axial movement, but not relativerelease levers to disengage the driven plate from the rotational spin, whilst the driven plates aredrive members. attached and supported by internally splined hubs When the driven and pressure plates are bolted to corresponding splines formed on the gearboxto the flywheel, the diaphragm is distorted into a spigot shaft, see Fig. 2.12.dished disc which therefore applies an axial thrust The diaphragm spring is in the form of a dishedbetween the pressure plate and the cover pressing. annular disc. The inner portion of the disc isThis clutch design reverses the normal method of radially slotted, the outer ends being enlargedoperation by pulling the diaphragm spring outwards with a circular hole to prevent stress concentrationto release the driven plate instead of pushing it. when the spring is distorted during disengagement. Owing to its configuration, the pull type clutch These radial slots divide the disc into a number ofallows a larger pressure plate and diaphragm release levers (fingers).spring to be used for a given diameter of clutch. The diaphragm spring is located in position withAdvantages of this design over a similar push type a shouldered pivot post which is riveted to theclutch include lower pedal loads, higher torque cover pressing. These rivets also hold a pair ofcapacity, improved take-up and increased dura- fulcrum rings in position which are situated eitherbility. This clutch layout allows the ratio of the side of the diaphragm.diaphragm finger release travel to pressure plate Whilst in service, the diaphragm cone angle willmovement to be reduced. It is therefore possible change continuously as wear occurs and as theto maintain the same pressure plate movement as clutch is engaged and disengaged during operation.that offered by a conventional push type clutch, To enable this to happen, the diaphragm pivotsand yet increase the ratio between clamp load and and rolls about the fulcrum rings. When the clutchpedal load from 4:1 to 5:1. is engaged the diaphragm bears against the outer 46
  • 56. Fig. 2.12 Multiplate diaphragm type clutchring, but when disengagement takes place the reac- bearings. There are two types of pressure platetion load is then taken by the inner ring. cover housings; one with a deep extended cover As the friction linings wear, the spring dia- rim which bolts onto a flat flywheel facing and thephragm will become more dished and subsequently shallow cover type in which the pressure platewill initially exert a larger axial clamping load. It is casting fits into a recessed flywheel.only when the linings are very worn, so that the The release mechanism is comprised of threedistance between the cover pressing and pressure lever fingers. The outer end of each lever pivotsplate become excessive, that the axial thrust will on a pin and needle race mounted inside eachbegin to decline. of the adjustable eye bolt supports, which are attached to the cover housing through an internally2.8 Lipe rollway twin driven plate clutch (Fig. 2.13) and externally threaded sleeve which is secured toThese clutches have two circular rows of helical coil the cover housing with a lock nut. Inwards fromsprings which act directly between the pressure the eye bolt, one-sixth of the release lever length, isplate and the cover housing, see Fig. 2.13. The a second pin which pivots on a pair of needle-release mechanism is of the pull type in which a bearing races situated inside the pressure platecentral release bearing assembly is made to with- lugs formed on either side of each layer.draw (pull out) three release levers to disengage theclutch. The clutch pressure plate assembly is bolted Release lever adjustmentto the flywheel and the driven plate friction linings Initially, setting up of the release levers is achievedare sandwiched between the flywheel, intermediate by slackening the locknuts and then rotating eachplate and pressure plate facings. The central hub of sleeve in turn with a two pronged fork adaptor toolthe driven plates is mounted on a splined gearbox which fits into corresponding slots machined out ofspigot shaft (input shaft). The splined end of the the adjustment sleeve end. Rotating the sleeve oneinput shaft is supported by a ball race bearing way or the other will screw the eye bolts in or outmounted inside the flywheel-crankshaft attachment until the correct dimension is obtained between theflange. The other end of this shaft is supported back of the release lever fingers and the outer coverinside the gearbox by either ball or taper roller rim edge. This setting dimension is provided by the 47
  • 57. Fig. 2.13 (a±b) Twin driven plate pull type clutch 48
  • 58. manufacturers for each clutch model and engine this instance one-sixth of the value if the springsapplication. Finally, tighten the locknuts of each were direct acting.eye bolt and re-check each lever dimension again. The operating characteristics of the clutch mechanism are described as follows:Release bearing adjustmentSlacken sleeve locknut with a `C shaped spanner. New engaged position (Fig. 2.14(a))Rotate the inner sleeve either way by means of the The spring thrust horizontal component of 2.2 kN,slotted adjusting nut until the recommended clear- multiplied by the lever ratio, provides a pressureance is obtained between the bearing housing cover plate clamping load of 13.2 kN (Fig. 2.14(a)). Theface and clutch brake. axial thrust horizontal component pushing on the pressure plate does not vary in direct proportioni.e. 9.5 mm for 355 mm Ð 1LP with the spring load exerted between its ends, but is 13 mm for 355 mm Ð 2LP a function of the angle through which the mounted 13 mm for 294 mm Ð 2LP springs operate relative to the splined input shaft. Finally tighten sleeve locknut and re-check clear-ance. Worn engagement position (Fig. 2.14(b)) When the driven plate facings wear, the release2.9 Spicer twin driven plate angle spring pull type bearing moves forward to the pressure plate soclutch (Fig. 2.14) that the springs elongate. The spring load exertedAn interesting clutch engagement and release pres- between the spring ends is thus reduced. Fortun-sure plate mechanism employs three pairs of coil ately, the inclined angle of spring axis to that ofsprings which are inclined to the axial direction of the thrust bearing axis is reduced so that as thethe driven plates. These springs are mounted spring load along its axis declines, the horizontalbetween the pressure plate cover housing, which thrust component remains essentially the same.takes the spring reaction, and the release lever cen- Therefore, the pressure plate clamping loadtral hub (Fig. 2.14). The axial clamping thrust is remains practically constant throughout the life ofconveyed by the springs to the six to one leverage the clutch (Fig. 2.14(b)).ratio release levers (six of them) spaced evenlyaround the release hub. These release levers span Release position (Fig. 2.14(c))between the release hub and a large annular shaped When the clutch is released, that is when the bear-adjustable pivot ring which is screwed inside the ing is pulled rearwards, the springs compress andpressure plate cover housing. Towards the pivot increase in load, but the spring angle relative to thepin end of the release levers a kink is formed so thrust bearing axis increases so that a greater pro-that it can bear against the pressure plate at one portion of the spring load will be acting radiallypoint. The pressure plate and intermediate plate are instead of axially. Consequently, the horizontalboth prevented from spinning with the driven component of axial release bearing load, causedplates by cast-in drive lugs which fit into slots by the spring thrust, gradually reduces to a valueformed into the cover housing. of about 1.7 kN as the bearing moves forwards, In the engaged position, the six springs expand which results in the reduced pedal effort. This isand push the release hub and, subsequently, the shown in Fig. 2.14(c).release levers towards the pressure plate so thatthe driven plates are squeezed together to transmit Internal manual adjustmentthe drive torque. Release bearing adjustment is made by unscrewing To release the clutch driven plates, the release the ring lock plate bolt and removing the plate. Thebearing assembly is pulled out from the cover hous- clutch pedal is then held down to relieve the releaseing. This compels the release lever hub to compress levers and adjusting ring load. The adjusting ring isand distort the thrust springs to a much greater then rotated to screw it in or out so that it alters theinclined angle relative to the input shaft axis and release lever hub axial position.so permits the pressure plate to be withdrawn by Turning the adjusting ring clockwise moves themeans of the retraction springs. release bearing towards the gearbox (increasing Because the spring thrust does not operate free pedal movement). Conversely, turning thedirectly against the pressure plate, but is relayed adjusting ring anticlockwise moves the releasethrough the release levers, the actual springs stiff- bearing towards the flywheel (decreasing freeness is reduced by a factor of the leverage ratio; in pedal movement). 49
  • 59. Fig. 2.14 (a±c) Twin driven plate angle spring pull type clutch 50
  • 60. The adjusting ring outer face is notched so that 2.10 Clutch (upshift) brake (Fig. 2.15)it can be levered round with a screwdriver when The clutch brake is designed primarily for use withadjustment is necessary. The distance between each unsynchronized (crash or constant mesh) gear-notch represents approximately 0.5 mm. Thus three boxes to permit shifting into first and reverse gearnotches moved means approximately 1.5 mm without severe dog teeth clash. In addition, therelease bearing movement. brake will facilitate making unshafts by slowing With the pedal released, there should be approxi- down the input shaft so that the next higher gearmately 13 mm clearance between the release bear- may be engaged without crunching of teeth.ing face and clutch brake. The brake disc assembly consists of a pair of Belleville spring washers which are driven by aInternal self-adjustment hub having internal lugs that engage machinedA clutch self-adjustment version has teeth cut on slots in the input shaft. These washers react againstthe inside of the adjusting ring and a small worm the clutch brake cover with facing material pos-and spring self-adjusting device replaces the lock itioned between each spring washer and outerplate. The worm meshes with the adjusting ring. cover (Fig. 2.15).One end of the spring is located in a hole formed When the clutch pedal is fully depressed, the discin the release lever hub whilst the other end is in will be squeezed between the clutch release bearingcontact with the worm. Each time the clutch is housing and the gearbox bearing housing, causingengaged and disengaged, the release lever move- the input spigot shaft to slow down or stop. Thement will actuate the spring. Consequently, once hub and spring washer combination will slip withthe driven plates have worn sufficiently, the respect to the cover if the applied torque loadincreased release lever movement will rotate the exceeds 34 Nm, thus preventing the disc brakeworm which in turn will partially screw round the being overloaded.adjusting ring to compensate and so reset the pos- In general, the clutch brake comes into engage-ition of the release levers. ment only during the last 25 mm of clutch pedalFig. 2.15 Clutch upshift brake (torque limiting) 51
  • 61. Fig. 2.16 Multiplate hydraulically actuated clutchestravel. Therefore, the pedal must be fully depressed up with the input and output splined drive membersto squeeze the clutch brake. The clutch pedal should respectively (Fig. 2.16). When these plates arenever be fully depressed before the gearbox is put squeezed together, torque will be transmitted frominto neutral. If the clutch brake is applied with the the input to the output members by way of thesegearbox still in gear, a reverse load will be put on the splines and grooves and the friction torque gener-gears making it difficult to get the gearbox out of ated between pairs of rubbing surfaces. These steelgear. At the same time it will have the effect of trying plates are faced with either resinated paper liningsto stop or decelerate the vehicle with the clutch brake or with sintered bronze linings, depending whetherand rapid wear of the friction disc will take place. moderate or large torques are to be transmitted.Never apply the clutch brake when making down Because the whole gear cluster assembly will beshifts, that is do not fully depress the clutch pedal submerged in fluid, these linings are designed towhen changing from a higher to a lower gear. operate wet (in fluid). These clutches are hydraul- ically operated by servo pistons either directly or indirectly through a lever disc spring to multiplate,2.11 Multiplate hydraulically operated automatic the clamping load which also acts as a piston returntransmission clutches (Fig. 2.16) spring. In this example of multiplate clutch utiliza-Automatic transmissions use multiplate clutches in tion hydraulic fluid is supplied under pressureaddition to band brakes extensively with epicyclic through radial and axial passages drilled in the out-compound gear trains to lock different stages of the put shaft. To transmit pressurized fluid from onegearing or gear carriers together, thereby providing member to another where there is relative angulara combination of gear ratios. movement between components, the output shaft These clutches are comprised of a pack of annular has machined grooves on either side of all the radialdiscs or plates, alternative plates being internally supply passages. Square sectioned nylon sealingand externally circumferentially grooved to match rings are then pressed into these grooves so that 52
  • 62. when the shaft is in position, these rings expand and lower speed range, the necessary extra clampingseal lengthwise portions of the shaft with their cor- thrust being supplemented by the centrifugal forceresponding bore formed in the outer members. at higher speeds. The release levers are made with offset bobFront clutch (FC) weights at their outer ends, so that they are centri-When pressurized, fluid is supplied to the front fugally out of balance (Fig. 2.17). The movementclutch piston chamber. The piston will move over due to the centrifugal force about the fixed pivotto the right and, through the leverage of the disc tends to force the pressure plate against the drivenspring, will clamp the plates together with consider- plate, thereby adding to the clamping load. Whileable thrust. The primary sun gear will now be the thrust due to the clamping springs is constant,locked to the input turbine shaft and permit torque the movement due to the centrifugal force varies asto be transmitted from the input turbine shaft to the square of the speed (Fig. 2.18). The reservethe central output shaft and primary sun gear. factor for the thrust spring can be reduced to 1.1 compared to 1.4±1.5 for a conventional helical coilRear clutch (RC) spring clutch unit. Conversely, this clutch designWhen pressurized, fluid is released from the front may be used for heavy duty applications whereclutch piston chamber, and is transferred to the greater torque loads are transmitted.rear clutch piston chamber. The servo piston willbe forced directly against the end plate of the rear 2.13 Fully automatic centrifugal clutchclutch multiplate pack. This compresses the release (Figs 2.19 and 2.20)spring and sandwiches the drive and driven plates Fully automatic centrifugal clutches separatetogether so that the secondary sun gear will now be the engine from the transmission system when thelocked to the input turbine shaft. Torque can now engine is stopped or idling and smoothly take upbe transmitted from the input turbine shaft to the the drive with a progressive reduction in slip withinsecondary sun gear. a narrow rising speed range until sufficient engine power is developed to propel the vehicle directly.2.12 Semicentrifugal clutch (Figs 2.17 and 2.18) Above this speed full clutch engagement isWith this design of clutch lighter pressure plate provided.springs are used for a given torque carrying capa- To facilitate gear changes whilst the vehiclecity, making it easier to engage the clutch in the is in motion, a conventional clutch releaseFig. 2.17 Semicentrifugal clutch 53
  • 63. Fig. 2.18 Semicentrifugal clutch characteristicslever arrangement is additionally provided. This bearing forwards. This movement pulls the reactormechanism enables the driver to disengage and plate rearwards by means of the knife-edge link andengage the clutch independently of the flyweight also withdraws the pressure plate through theaction so that the drive and driven gearbox member retractor springs so as to release the pressure platespeeds can be rapidly and smoothly unified during clamping load.the gear selection process. The automatic centrifugal mechanism consists of 2.14 Clutch pedal actuating mechanismsa reaction plate situated in between the pressure Some unusual ways of operating a clutch unit willplate and cover pressing. Mounted on this reaction now be described and explained.plate by pivot pins are four equally spaced bob-weights (Fig. 2.19). When the engines speed 2.14.1 Clutch pedal with over-centre springincreases, the bobweight will tend to fly outward. (Fig. 2.21)Since the centre of gravity of their masses is to one With this clutch pedal arrangement, the over-side of these pins, they will rotate about their pins. centre spring supplements the foot pressure appliedThis will be partially prevented by short struts when disengaging the clutch, right up until theoffset to the pivot pins which relay this movement diaphragm spring clutch is fully disengaged (Fig.and effort to the pressure plate. Simultaneously, 2.21). It also holds the clutch pedal in the `offthe reaction to this axial clamping thrust causes position. When the clutch pedal is pressed, thethe reaction plate to compress both the reaction master cylinder piston forces the brake fluid intoand pressure springs so that it moves backwards the slave cylinder. The slave piston moves the pushtowards the cover pressing. rod, which in turn disengages the clutch. After the The greater the centrifugal force which tends to pedal has been depressed approximately 25 mm ofrotate the bobweights, the more compressed the its travel, the over-centre spring change over pointsprings will become and their reaction thrust will has been passed. The over-centre spring tension isbe larger, which will increase the pressure plate then applied as an assistance to the foot pressure.clamping load. Adjustment of the clutch is carried out by adjust- To obtain the best pressure plate thrust to engine ing the pedal position on the master cylinder pushspeed characteristics (Fig. 2.20), adjustable reactor rod.springs are incorporated to counteract the maincompression spring reaction. The initial compres- 2.14.2 Clutch cable linkage with automaticsion length and therefore loading of these springs is adjuster (Fig. 2.22)set up by the adjusting nut after the whole unit has The release bearing is of the ball race type and isbeen assembled. Thus the resultant thrust of both kept in constant contact with the fingers of thelots of springs determine the actual take-up engine diaphragm spring by the action of the pedal self-speed of the clutch. adjustment mechanism. In consequence, there is Gear changes are made when the clutch is disen- no pedal free movement adjustment requiredgaged, which is achieved by moving the release (Fig. 2.22). 54
  • 64. Fig. 2.19 Fully automatic centrifugal clutch 55
  • 65. Fig. 2.20 Fully automatic centrifugal clutch characteristicsFig. 2.21 Hydraulic clutch operating system with over-centre spring When the pedal is released, the adjustment pawl whenever the driver depresses the clutch pedalis no longer engaged with the teeth on the pedal or maintains the pedal in a partially depressedquadrant. The cable, however, is tensioned by the position, as may be necessary under pull-awayspring which is located between the pedal and conditions. Movement of the clutch pedal is imme-quadrant. As the pedal is depressed, the pawl diately relayed by way of the servo to the clutch inengages in the nearest vee between the teeth. The proportion to the input pedal travel.particular tooth engagement position will gradu- As the clutchs driven plate wears, clutch actu-ally change as the components move to compensate ating linkage movement is automatically taken upfor wear in the clutch driven plate and stretch in the by the air piston moving further into the cylinder.cable. Thus the actual servo movement when the clutch is being engaged and disengaged remains approxi-2.14.3 Clutch air/hydraulic servo (Fig. 2.23) mately constant. In the event of any interruptionIn certain applications, to reduce the drivers foot of the air supply to the servo the clutch will stilleffort in operating the clutch pedal, a clutch air/ operate, but without any servo assistance.hydraulic servo may be incorporated into the actuat- Immediately the clutch pedal is pushed down,ing linkage. This unit provides power assistance the fluid from the master cylinder is displaced into 56
  • 66. Fig. 2.22 Clutch cable linkage with automatic adjusterthe servo hydraulic cylinder. The pressure created and so opens the exhaust valve. Compressed airwill act on both the hydraulic piston and the reac- in the air cylinder will now transfer to the reactiontion plunger. Subsequently, both the hydraulic plunger chamber. It then passes through thepiston and the reaction plunger move to the right exhaust valve and port where it escapes toand allow the exhaust valve to close and the inlet the atmosphere. The released compressed air fromvalve to open. Compressed air will now pass the cylinder allows the clutch linkage return springthrough the inlet valve port and the passage con- to move the air and hydraulic piston assembly backnecting the reaction plunger chamber to the com- to its original position in its cylinder and at thepressed air piston cylinder. It thereby applies same time this movement will be relayed to thepressure against the air piston. The combination clutch release bearing, whereby the clutch operat-of both hydraulic and air pressure on the hydraulic ing mechanism moves to the engaged drive positionand air piston assembly causes it to move over, this (Fig. 2.23(a)).movement being transferred to the clutch releasebearing which moves the clutch operating mechan- 2.15 Composite flywheel and integral single plateism to the disengaged position (Fig. 2.23(d)). diaphragm clutch (Fig. 2.24) When the clutch pedal is held partially This is a compact diaphragm clutch unit built asdepressed, the air acting on the right hand side of an integral part of the two piece flywheel. It isthe reaction plunger moves it slightly to the left designed for transaxle transmission applicationwhich now closes the inlet valve. In this situation, where space is at a premium and maximum torquefurther air is prevented from entering the air transmitting capacity is essential.cylinder. Therefore, since no air can move in or The flywheel and clutch drive pressing acts as aout of the servo air cylinder and both valves are support for the annular flywheel mass and func-in the lapped position (both seated), the push rod tions as the clutch pressure plate drive member.will not move unless the clutch pedal is again The advantage of having the flywheel as a twomoved (Fig. 2.23(c)). piece assembly is that its mass can be concentrated When the clutch pedal is released fluid returns more effectively at its outer periphery so that itsfrom the servo to the master cylinder. This permits overall mass can be reduced for the same cyclicthe reaction plunger to move completely to the left torque and speed fluction which it regulates. 57
  • 67. Fig. 2.23 (a±c) Clutch air/hydraulic servo 58
  • 68. Fig. 2.24 Integral single plate clutch and composite flywheel The diaphragm spring takes the shape of a assembled and therefore applies on axial thrustdished annular disc. The inner portion of the disc between the pressure plate and its adjacentis radially slotted, the outer ends being enlarged flywheel/clutch drive pressing. With this springwith a circular hole to prevent stress concentration leverage arrangement, a larger pressure plate andwhen the spring is deflected during disengagement. diaphragm spring can be utilised for a given overallThese radial slots divide the disc into many diameter of clutch assembly. This design thereforeinwardly pointing fingers which have two func- has the benefits of lower pedal effort, higher trans-tions, firstly to provide the pressure plate with mitting torque capacity, a highly progressivean evenly distributed multileaf spring type thrust, engagement take-up and increased clutch life com-and secondly to act as release levers to separate pared to the conventional push type diaphragmthe driven plate from the sandwiching flywheel clutch.and pressure plate friction faces. The engagement and release mechanism consists To actuate the clutch release, the diaphragm of a push rod which passes through the hollowspring is made to pivot between a pivot spring gearbox input shaft and is made to enter and con-positioned inside the flywheel/clutch drive pressing tact the blind end of a recess formed in the releasenear its outer periphery and a raised circumferen- plunger. The plunger is a sliding fit in the normaltial rim formed on the back of the pressure plate. spigot bearing hole made in the crankshaft endThe engagement and release action of the clutch is flange. It therefore guides the push rod and trans-similar to the pull type diaphragm clutch where the fers its thrust to the diaphragm spring fingers viadiaphragm is distorted into a dished disc when the release plate. 59
  • 69. 3 Manual gearboxes and overdrives3.1 The necessity for a gearboxPower from a petrol or diesel reciprocating enginetransfers its power in the form of torque and angularspeed to the propelling wheels of the vehicle toproduce motion. The object of the gearbox is toenable the engines turning effect and its rotationalspeed output to be adjusted by choosing a range ofunder- and overdrive gear ratios so that the vehicleresponds to the drivers requirements within thelimits of the various road conditions. An insightof the forces opposing vehicle motion and engineperformance characteristics which provide thebackground to the need for a wide range of gearboxdesigns used for different vehicle applications willnow be considered.3.1.1 Resistance to vehicle motionTo keep a vehicle moving, the engine has to developsufficient power to overcome the opposing roadresistance power, and to pull away from a standstillor to accelerate a reserve of power in addition to thatabsorbed by the road resistance must be availablewhen required. Road resistance is expressed as tractive resistance(kN). The propelling thrust at the tyre to road Fig. 3.1 Vehicle tractive resistance and effortinterface needed to overcome this resistance is performance chartknown as tractive effect (kN) (Fig. 3.1). For match-ing engine power output capacity to the opposingroad resistance it is sometimes more convenient to more energy as the wheel speed increases and there-express the opposing resistance to motion in terms fore the rolling resistance will also rise slightly asof road resistance power. shown in Fig. 3.1. Factors which influence the The road resistance opposing the motion of the magnitude of the rolling resistance are the ladenvehicle is made up of three components as follows: weight of the vehicle, type of road surface, and1 Rolling resistance the design, construction and materials used in the2 Air resistance manufacture of the tyre.3 Gradient resistance Air resistance (Fig. 3.1) Power is needed toRolling resistance (Fig. 3.1) Power has to be counteract the tractive resistance created by theexpended to overcome the restraining forces caused vehicle moving through the air. This is caused byby the deformation of tyres and road surfaces and air being pushed aside and the formation of turbu-the interaction of frictional scrub when tractive lence over the contour of the vehicles body. It haseffect is applied. Secondary causes of rolling resist- been found that the air resistance opposing forceance are wheel bearing, oil seal friction and the and air resistance power increase with the squarechurning of the oil in the transmission system. It and cube of the vehicles speed respectively. Thus athas been found that the flattening distortion of the very low vehicle speeds air resistance is insignifi-tyre casing at the road surface interface consumes cant, but it becomes predominant in the upper 60
  • 70. speed range. Influencing factors which determine gear with a small surplus of about 0.2% grade-the amount of air resistance are frontal area of ability.vehicle, vehicle speed, shape and streamlining of The two extreme operating conditions justbody and the wind speed and direction. described set the highest and lowest gear ratios. To fix these conditions, the ratio of road speed in highest gear to road speed in lowest gear at a givenGradient resistance (Fig. 3.1) Power is required engine speed should be known. This quantity isto propel a vehicle and its load not only along a referred to as the ratio span.level road but also up any gradient likely to beencountered. Therefore, a reserve of power must be Road speed in highest gearavailable when climbing to overcome the potential i.e. Ratio span ˆ Road speed in lowest gearenergy produced by the weight of the vehicle as itis progressively lifted. The gradient resistance (both road speeds being achieved at similar engineopposing motion, and therefore the tractive effect speed).or power needed to drive the vehicle forward, isdirectly proportional to the laden weight of the Car and light van gearboxes have ratio spans ofvehicle and the magnitude of gradient. Thus driving about 3.5:1 if top gear is direct, but with overdriveup a slope of 1 in 5 would require twice the reserve of this may be increased to about 4.5:1. Large com-power to that needed to propel the same vehicle up a mercial vehicles which have a low power to weightgradient of 1 in 10 at the same speed (Fig. 3.1). ratio, and therefore have very little surplus power when fully laden, require ratio spans of between 7.53.1.2 Power to weight ratio and 10:1, or even larger for special applications.When choosing the lowest and highest gearbox An example of the significance of ratio span isgear ratios, the most important factor to consider shown as follows:is not just the available engine power but also the Calculate the ratio span for both a car and heavyweight of the vehicle and any load it is expected to commercial vehicle from the data provided.propel. Consequently, the power developed perunit weight of laden vehicle has to be known. Thisis usually expressed as the power to weight ratio. Type of vehicle Gear Ratio km/h/1000 rev/mini.e. Power to weight Brake power developed ratio ˆ Car Top 0.7 39 Laden weight of vehicle First 2.9 9.75There is a vast difference between the power toweight ratio for cars and commercial vehicles Commercial Top 1.0 48which is shown in the following examples. vehicle (CV) First 6.35 6 Determine the power to weight ratio for thefollowing modes of transport: 39 Car ratio span ˆ ˆ 4:0X1 9:75a) A car fully laden with passengers and luggage 48 weighs 1.2 tonne and the maximum power pro- Commercial vehicle ratio span ˆ ˆ 8:0X1 6 duced by the engine amounts to 120 kW.b) A fully laden articulated truck weighs 38 tonne and a 290 kW engine is used to propel this load. 3.1.4 Engine torque rise and speed operating 120 range (Fig. 3.2)a) Power to weight ratio ˆ ˆ 100 kW/tonne Commercial vehicle engines used to pull large loads 1:2 are normally designed to have a positive torque 290 rise curve, that is from maximum speed to peakb) Power to weight ratio ˆ ˆ 7:6 kW/tonne. 38 torque with reducing engine speed the available torque increases (Fig. 3.2). The amount of engine3.1.3 Ratio span torque rise is normally expressed as a percentage ofAnother major consideration when selecting gear the peak torque from maximum speed (ratedratios is deciding upon the steepest gradient the power) back to peak torque.vehicle is expected to climb (this may normally betaken as 20%, that is 1 in 5) and the maximum level Maximum speed torque % torque rise ˆ Â 100road speed the vehicle is expected to reach in top Peak torque 61
  • 71. Fig. 3.2 Engine performance and gear split chart for an eight speed gearbox The torque rise can be shaped depending upon driver can either choose to operate the enginesengine design and taking into account such features speed in a range varying just below the maximumas naturally aspirated, resonant induction tuned, rated power to achieve maximum performance andturbocharged, turbocharged with intercooling and journey speed or, to improve fuel economy, wearso forth. Torque rises can vary from as little as 5 to and noise, within a speed range of between 200 toas high as 50%, but the most common values for 400 rev/min on the positive torque rise side of thetorque rise range from 15 to 30%. engine torque curve that is in a narrow speed band A large torque rise characteristic raises the just beyond peak torque. Fig. 3.2 shows that theengines operating ability to overcome increased economy speed range operates with the specific fuelloads if the engines speed is pulled down caused consumption at its minimum and that the engineby changes in the road conditions, such as climbing speed band is in the most effective pulling zone.steeper gradients, and so tends to restore the ori-ginal running conditions. If the torque rise is small 3.2 Five speed and reverse synchromesh gearboxesit cannot help as a buffer to supplement the high With even wider engine speed ranges (1000 to 6000torque demands and the engine speed will rapidly rev/min) higher car speeds (160 km/h and more)fade. Frequent gear changes therefore become and high speed motorways, it has become desirable,necessary compared to engines operating with and in some cases essential, to increase the numberhigh torque rise characteristics. Once the engine of traditional four speed ratios to five, where thespeed falls below peak torque, the torque rise fifth gear, and sometimes also the fourth gear, is anbecomes negative and the pulling ability of the overdrive ratio. The advantages of increasing theengine drops off very quickly. number of ratio steps are several; not only does Vehicle driving technique should be such that the extra gear provide better acceleration response,engines are continuously driven between the speed but it enables the maximum engine rotational speedrange of peak torque and governed speed. The to be reduced whilst in top gear cruising, fuel 62
  • 72. Table 3.1 Typical four and five speed gearbox gear Lubrication to the mainshaft gears is obtained byratios radial branch holes which feed the rubbing surfaces of both mainshaft and gears.Five speed box Four speed boxGear Ratio Gear Ratio 3.2.2 Five speed and reverse single stage synchromesh gearbox (Fig. 3.4)top 0.8 top 1.0 This two shaft gearbox has only one gear reduction4 1.0 3 1.3 stage formed between pairs of different sized con-3 1.4 2 2.12 2.0 1 3.4 stant mesh gear wheels to provide a range of gear1 3.5 R 3.5 ratios. Since only one pair of gears mesh, comparedR 3.5 to the two pairs necessary for the double stage gearbox, frictional losses are halved. Power delivered to the input primary shaft can follow five different flow paths to the secondaryconsumption is improved and engine noise and wear shaft via first, second, third, fourth and fifth gearare reduced. Typical gear ratios for both four and five wheel pairs, but only one pair is permitted to trans-speed gearboxes are as shown in Table 3.1. fer the drive from one shaft to another at any one The construction and operation of four speed time (Fig. 3.4).gearboxes was dealt with in Vehicle and Engine The conventional double stage gearbox isTechnology. The next section deals with five speed designed with an input and output drive at eithersynchromesh gearboxes utilized for longitudinal end of the box but a more convenient and compactand transverse mounted engines. arrangement with transaxle units where the final drive is integral to the gearbox is to have the input3.2.1 Five speed and reverse double stage and output power flow provided at one end only ofsynchromesh gearbox (Fig. 3.3) the gearbox.With this arrangement of a five speed double stage In the neutral position, first and second outputgearbox, the power input to the first motion shaft gear wheels will be driven by the correspondingpasses to the layshaft and gear cluster via the first gear wheels attached to the input primary shaft,stage pair of meshing gears. Rotary motion is but they will only be able to revolve about theirtherefore conveyed to all the second stage layshaft own axis relative to the output secondary shaft.and mainshaft gears (Fig. 3.3). Because each pair of Third, fourth and fifth gear wheel pairs are drivensecond stage gears has a different size combination, by the output second shaft and are free to revolvea whole range of gear ratios are provided. Each only relative to the input primary shaft becausemainshaft gear (whilst in neutral) revolves on the they are not attached to this shaft but use it onlymainshaft but at some relative speed to it. There- as a supporting axis.fore, to obtain output powerflow, the selected When selecting individual gear ratios, the appro-mainshaft gear has to be locked to the mainshaft. priate synchronizing sliding sleeve is pushedThis then completes the flow path from the first towards and over the dog teeth forming part ofmotion shaft, first stage gears, second stage gears the particular gear wheel required. Thus with firstand finally to the mainshaft. and second gear ratios, the power flow passes from In this example the fifth gear is an overdrive gear the input primary shaft and constant mesh pairs ofso that to speed up the mainshaft output relative to gears to the output secondary shaft via the first andthe input to the first motion shaft, a large layshaft second drive hub attached to this shaft. Gearfifth gear wheel is chosen to mesh with a much engagement is completed by the synchronizingsmaller mainshaft gear. sleeve locking the selected output gear wheel to For heavy duty operations, a forced feed lubrica- the output secondary shaft. Third, fourth andtion system is provided by an internal gear crescent fifth gear ratios are selected when the third andtype oil pump driven from the rear end of the fourth or fifth gear drive hub, fixed to the inputlayshaft (Fig. 3.3). This pump draws oil from the primary shaft, is locked to the respective gear wheelbase of the gearbox casing, pressurizes it and then dog clutch by sliding the synchronizing sleeve in toforces it through a passage to the mainshaft. The mesh with it. The power flow path is now trans-oil is then transferred to the axial hole along the ferred from the input primary shaft drive hub andcentre of the mainshaft by way of an annular selected pair of constant mesh gears to the outputpassage formed between two nylon oil seals. secondary shaft. 63
  • 73. Fig. 3.3 Five speed and reverse double stage synchromesh gearbox Transference of power from the gearbox output and accordingly power passes to both stub drivesecondary shaft to the differential left and right shafts.hand drive shafts is achieved via the final drivepinion and gear wheel which also provide a per-manent gear reduction (Fig. 3.4). Power then flows 3.3 Gear synchronization and engagementfrom the differential cage which supports the final The gearbox basically consists of an input shaftdrive gear wheel to the cross-pin and planet gears driven by the engine crankshaft by way of thewhere it then divides between the two side sun gears clutch and an output shaft coupled indirectly either 64
  • 74. Fig. 3.4 Five speed and reverse single stage synchromesh gearbox with integral final drive (transaxle unit)through the propellor shaft or intermediate gears to these shafts while the other is free to revolve on thethe final drive. Between these two shafts are pairs of second shaft at some speed determined by the exist-gear wheels of different size meshed together. ing speeds of the input and output drive shafts. If the gearbox is in neutral, only one of these To engage any gear ratio the input shaft has topairs of gears is actually attached rigidly to one of be disengaged from the engine crankshaft via the 65
  • 75. clutch to release the input shaft drive. It is then only ately the balls are pushed out of their groove, thethe angular momentum of the input shaft, clutch chamfered edges of the outer hubs internal teeth willdrive plate and gear wheels which keeps them revol- then be able to align with the corresponding teethving. The technique of good gear changing is to be spacing on the first motion gear. Both sets of teethable to judge the speeds at which the dog teeth of will now be able to mesh so that the outer hub can beboth the gear wheel selected and output shaft are moved into the fully engaged position (Fig. 3.5(c)).rotating at a uniform speed, at which point in time Note the bronze female cone insert frictional facethe dog clutch sleeve is pushed over so that both sets is not smooth, but consists of a series of tramlineof teeth engage and mesh gently without grating. grooves which assist in cutting away the oil film so Because it is difficult to know exactly when to that a much larger synchronizing torque will bemake the gear change a device known as the syn- generated to speed up the process.chromesh is utilized. Its function is to apply a fric-tion clutch braking action between the engaging 3.3.2 Positive baulk ring synchromesh unitgear and drive hub of the output shaft so that (Fig. 3.6(a, b and c))their speeds will be unified before permitting the The gearbox mainshaft rotates at propellor shaftdog teeth of both members to contact. speed and, with the clutch disengaged, the first Synchromesh devices use a multiplate clutch or a motion shaft gear, layshaft cluster gears, andconical clutch to equalise the input and output mainshaft gears rotate freely.rotating members of the gearbox when the process Drive torque will be transmitted when a gearof gear changing is taking place. Except for special wheel is positively locked to the mainshaft. This isapplications, such as in some splitter and range achieved by means of the outer synchromesh hubchange auxiliary gearboxes, the conical clutch internal teeth which slide over the inner synchro-method of synchronization is generally employed. mesh hub splines (Fig. 3.6(a)) until they engage With the conical clutch method of producing silent with dog teeth formed on the constant mesh geargear change, the male and female cone members wheel being selected.are brought together to produce a synchronizing When selecting and engaging a particular gearfrictional torque of sufficient magnitude so that one ratio, the gear stick slides the synchromesh outeror both of the input and output members rotational hub in the direction of the chosen gear (towardsspeed or speeds adjust automatically until they the left). Because the shift plate is held radiallyrevolve as one. Once this speed uniformity has been outwards by the two energizing ring type springsachieved, the end thrust applied to the dog clutch and the raised middle hump of the plate rests in thesleeve is permitted to nudge the chamfered dog teeth groove formed on the inside of the hub, the end ofof both members into alignment, thereby enabling the the shift plate contacts the baulking ring and pushestwo sets of teeth to slide quietly into engagement. it towards and over the conical surface, forming part of the constant mesh gear wheel (Fig. 3.6(b)).3.3.1 Non-positive constant load synchromesh The frictional grip between the male and femaleunit (Fig. 3.5(a, b and c)) conical members of the gear wheel and baulkingWhen the gear stick is in the neutral position the ring and the difference in speed will cause the baulk-spring loaded balls trapped between the inner and ing ring to be dragged around relative to the innerouter hub are seated in the circumferential groove hub and shift plate within the limits of the clearanceformed across the middle of the internal dog teeth between the shift plate width and that of the(Fig. 3.5(a)). As the driver begins to shift the gear recessed slot in the baulking ring. Owing to thestick into say top gear (towards the left), the outer designed width of the shift plate slot in the baulkingand inner synchromesh hubs move as one due to the ring, the teeth on the baulking ring are now out ofradial spring loading of the balls along the splines alignment with those on the outer hub by approxi-formed on the main shaft until the female cone of the mately half a tooth width, so that the chamferedouter hub contacts the male cone of the first motion faces of the teeth of the baulking ring and outer hubgear (Fig. 3.5(b)). When the pair of conical faces bear upon each other.contact, frictional torque will be generated due to As the baulking ring is in contact with the gearthe combination of the axial thrust and the differ- cone and the outer hub, the force exerted by theence in relative speed of both input and output shaft driver on the gear stick presses the baulking ringmembers. If sufficient axial thrust is applied to the female cone hard against the male cone of the gear.outer hub, the balls will be depressed inwards Frictional torque between the two surfaces willagainst the radial loading of the springs. Immedi- eventually cause these two members to equalize 66
  • 76. Fig. 3.5 Non-positive constant load synchromesh unit 67
  • 77. Fig. 3.6 (a±c) Positive baulk ring synchromesh unit 68
  • 78. their speeds. Until this takes place, full engagement centre, hold the bronze synchronizing cone ringsof the gear and outer hub dog teeth is prevented by apart. Alternating with the shouldered pins on thethe out of alignment position of the baulking ring same pitch circle are diametrically split pins, theteeth. When the gear wheel and main shaft have ends of which fit into blind bores machined inunified their speeds, the synchronizing torque will the synchronizing cone rings. The pin halves arehave fallen to zero so that the baulking ring is no sprung apart, so that a chamfered groove around thelonger dragged out of alignment. Therefore the middle of each half pin registers with a chamferedouter hub can now overcome the baulk and follow hole in the drive hub.through to make a positive engagement between If the gearbox is in the neutral position, both setshub and gear (Fig. 3.6(c)). It should be understood of shouldered and split pins are situated with theirthat the function of the shift plate and springs is to grooves aligned with the central drive hub (Fig.transmit just sufficient axial load to ensure a rapid 3.8(a and b)).bringing together of the mating cones so that the When an axial load is applied to the drive hub bybaulking ring dog teeth immediately misalign with the gear stick, it moves over (in this case to the left)their corresponding outer hub teeth. Once the cone until the synchronizing ring is forced against thefaces contact, they generate their own friction adjacent first motion gear cone. The friction (syn-torque which is sufficient to flick the baulking chronizing) torque generated between the rubbingring over, relative to the outer hub. Thus the cham- tapered surfaces drags the bronze synchronizingfers of both sets of teeth contact and oppose further ring relative to the mainshaft and drive hub untilouter hub axial movement towards the gear dog the grooves in the shouldered pins are wedged againstteeth. the chamfered edges of the drive hub (Fig. 3.8(c)) so that further axial movement is baulked.3.3.3 Positive baulk pin synchromesh unit Immediately the input and output shaft speeds(Fig. 3.7(a, b, c and d)) are similar, that is, synchronization has beenMovement of the selector fork synchronizing sleeve achieved, the springs in the split pins are able toto the left (Fig. 3.7(a and b)) forces the female expand and centralize the shouldered pins relative(internal) cone to move into contact with the male to the chamfered holes in the drive hub. The drive(external) cone on the drive gear. Frictional torque hub can now ride out of the grooves formed aroundwill then synchronize (unify) the input and output the split pins, thus permitting the drive hub to shiftspeeds. Until speed equalization is achieved, the col- further over until the internal and external doglars on the three thrust pins (only one shown) will be teeth of both gear wheel hub mesh and fully engagepressed hard into the enlarged position of the slots (Fig. 3.8(d)).(Fig. 3.5(c)) in the synchronizing sleeve owing to thefrictional drag when the speeds are dissimilar. Under 3.3.5 Split ring synchromesh unitthese conditions, unless extreme pressure is exerted, (Fig. 3.9(a, b, c and d))the dog teeth cannot be crushed by forcing the collars In the neutral position the sliding sleeve sits cen-into the narrow portion of the slots. However, when trally over the drive hub (Fig. 3.9(a)). This permitsthe speeds of the synchromesh hub and drive gear are the synchronizing ring expander band and thrustequal (synchronized) the collars tend to `float in block to float within the constraints of the recessthe enlarged portion of the slots, there is only the machine in the side of the gear facing the drive hubpressure of the spring loaded balls to be overcome. (Fig. 3.9(b)).The collars will then slide easily into the narrow For gear engagement to take place, the slidingportion of the slots (Fig. 3.5(d)) allowing the syn- sleeve is moved towards the gear wheel selected (tochronizer hub dog teeth to shift in to mesh with the the left) until the inside chamfer of the sliding sleevedog teeth on the driving gear. contacts the bevelled portion of the synchronizing ring. As a result, the synchronizing ring will be3.3.4 Split baulk pin synchromesh unit slightly compressed and the friction generated(Fig. 3.8(a, b, c and d)) between the two members then drags the synchron-The synchronizing assembly is composed of two izing ring round in the direction of whicheverthick bronze synchronizing rings with tapered member is rotating fastest, be it the gear or drivenfemale conical bores, and situated between them hub. At the same time, the thrust block is pulledis a hardened steel drive hub internally splined with round so that it applies a load to one end of theexternal dog teeth at each end (Fig. 3.8(a)). Three expander band, whilst the other end is restrainedshouldered pins, each with a groove around its from moving by the anchor block (Fig. 3.9(c)). 69
  • 79. Fig. 3.7 (a±d) Positive baulk pin synchromesh unit Whilst this is happening the expander is also mouth of the sliding sleeve. This energizing actionpushed radially outwards. Consequently, there attempting to expand the synchronizing ring pre-will be a tendency to expand the synchronizing vents the sliding sleeve from completely movingslit ring, but this will be opposed by the chamfered over and engaging the dog teeth of the selected 70
  • 80. Fig. 3.8 (a±c) Split baulk pin synchromesh unitgear wheel until both the drive hub and constant more friction torque because there is no relativemesh gear wheel are revolving at the same speed. speed to create the frictional drag. Therefore When both input and output members are uni- the expander band immediately stops exertingfied, that is, rotating as one, there cannot be any radial force on the inside of the synchronizing ring. 71
  • 81. Fig. 3.9 (a±d) Split baulk ring synchromesh unit 72
  • 82. The axial thrust applied by the gear stick to thesliding sleeve will now be sufficient to compress thesplit synchronizing ring and subsequently permitsthe sleeve to slide over the gear wheel dog teeth forfull engagement (Fig. 3.9(d)).3.4 Remote controlled gear selection andengagement mechanismsGear selection and engagement is achieved by twodistinct movements:1 The selection of the required gear shift gate and the positioning of the engagement gate lever.2 The shifting of the chosen selector gate rod into the engagement gear position. These two operations are generally performedthrough the media of the gear shift lever and theremote control tube/rod. Any transverse move-ment of the gear shift lever by the driver selectsthe gear shift gate and the engagement of the gate Fig. 3.10 Remote controlled double rod and bell crankis obtained by longitudinal movement of the gear lever gearshift mechanism suitable for both four andshift lever. five speed transversely mounted gearbox Movement of the gear shift lever is conveyed tothe selection mechanism via the remote controltube. Initially the tube is twisted to select the ment shaft. Consequently, this shifts the transversegate shift gate, followed by either a push or pull selector/engagement shaft so that it pushes themovement of the tube to engage the appropriate synchronizing sliding sleeve into engagement withgear. the selected gear dog teeth. For the gear shift control to be effective it musthave some sort of flexible linkage between the gear 3.4.2 Remote controlled bell cranked lever gearshift lever supported on the floor of the drivers shift mechanism for a four speed transversecompartment and the engine and transmission inte- mounted gearbox (Ford) (Fig. 3.11)gral unit which is suspended on rubber mountings. Gear selection and engagement movement isThis is essential to prevent engine and transmission conveyed from the gear shift lever pivot action tovibrations being transmitted back to the body and the remote control rod universal joint and to thefloor pan and subsequently causing discomfort to control shift and relay lever guide (Fig. 3.11).the driver and passengers. Rocking the gear shift lever transversely rotates the control shaft and relay guide. This tilts the selector relay lever and subsequently the selec-3.4.1 Remote controlled double rod and bell tion relay lever guide and shaft until the strikercranked lever gear shift mechanism, suitable for finger aligns with the chosen selector gate. A fur-both four and five speed transverse mounted ther push or pull movement to the gear shift levergearbox (Talbot) (Fig. 3.10) by the driver then transfers a forward orTwisting the remote control tube transfers move- backward motion via the remote control rod, con-ment to the first selector link rod. This motion is trol shaft and relay lever guide to the engagementthen redirected at right angles to the transverse relay lever. Movement is then redirected at rightgate selector/engagement shaft via the selector angles to the selector relay guide and shaft.relay lever (bell crank) to position the required Engagement of the gear required is finally obtainedgear gate (Fig. 3.10). A forward or backward by the selector/engagement shaft forcing the strik-movement of the remote control tube now conveys ing finger to shift the gate and selector fork alongmotion via the first engagement relay lever (bell the single selector rod so that the synchron-crank), engagement link rod and second relay izing sleeve meshes with the appropriate gearlever to rotate the transverse gate selector/engage- wheel dog clutch. 73
  • 83. Fig. 3.11 Remote controlled bell crank level gear shift mechanism for a four speed transversely mounted gearbox3.4.3 Remote controlled sliding ball joint gearshift mechanism suitable for both four and fivespeed longitudinal or transverse mounted gearbox(VW) (Fig. 3.12)Selection and engagement of the different gearratios is achieved with a swivel ball end pivot gearshift lever actuating through a sliding ball relaylever a single remote control rod (Fig. 3.12). Theremote control rod transfers both rotary and push-pull movement to the gate selector and engagementshaft. This rod is also restrained in bushes betweenthe gear shift lever mounting and the bulkhead.It thus permits the remote control rod to transferboth rotary (gate selection) and push-pull (select rodengagement shift) movement to the gate selector andengagement shaft. Relative movement between thesuspended engine and transmission unit and the car Fig. 3.12 Remote controlled sliding ball joint gear shiftbody is compensated by the second sliding ball mechanism suitable for both four and five speedrelay lever. As a result the gate engagement striking longitudinally or transversely mounted gearboxfinger is able to select and shift into engagement theappropriate selector rod fork. This single rod sliding ball remote control an additional relay lever mechanism (not shown) islinkage can be used with either longitudinally or needed to convey the two distinct movements oftransversely mounted gearboxes, but with the latter selection and engagement through a right angle. 74
  • 84. 3.4.4 Remote controlled double rod and hingedrelay joint gear shift mechanism suitable for bothfour and five speed longitudinal mounted gearbox(VW) (Fig. 3.13)With this layout the remote control is provided bya pair of remote control rods, one twists and selectsthe gear gate when the gear shift lever is given atransverse movement, while the other pushes orpulls when the gear shift lever is moved longitudin-ally (Fig. 3.13). Twisting movement is thus con-veyed to the engagement relay lever which makesthe engagement striking finger push the alignedselector gate and rod. Subsequently, the synchro-nizing sleeve splines mesh with the correspondingdog clutch teeth of the selected gear wheel. Relativemovement between the gear shift lever swivel sup-port and rubber mounted gearbox is absorbed bythe hinged relay joint and the ball joints at eitherend of the remote control engagement rod.3.4.5 Remote controlled single rod with selfaligning bearing gear shift mechanism suitable forboth five and six speed longitudinal mounted Fig. 3.14 Remote controlled single rod with self-aligninggearbox (Ford) (Fig. 3.14) bearing gear shift mechanism suitable for both fiveA simple and effective method of selecting and and six speed longitudinally mounted gearboxengaging the various gear ratios suitable forcommercial vehicles where the driver cab is for- Movement of the gear shift lever in the usualward of the gearbox is shown in Fig. 3.14. transverse and longitudinal directions provides both rotation and a push-pull action to the remote control tube. Twisting the remote control tube transversely tilts the relay gear shift lever about its ball joint so that the striking finger at its lower end matches up with the selected gear gate. Gear engagement is then completed by the driver tilting the gear shift lever away or towards himself. This permits the remote control tube to move axially through the mounted self-aligning bearing. As a result, a similar motion will be experienced by the relay gear shift lever, which then pushes the striking finger, selector gate and selector fork into the gear engaged position. It should be observed that the self-aligning bearing allows the remote control tube to slide to and fro. At the same time it permits the inner race member to tilt if any relative movement between the gearbox and chassis takes place. 3.4.6 Remote controlled single rod with swing arm support gear shift mechanism suitable for five and six speed longitudinally mounted gearbox (ZF) (Fig. 3.15)Fig. 3.13 Remote controlled double rod and hinged This arrangement which is used extensively onrelay joint gear shift mechanism suitable for both four and commercial vehicles employs a universal cross-five speed longitudinally mounted gearbox pin joint to transfer both the gear selection and 75
  • 85. pivots the suspended selector gate relay lever so that the transverse gate selector/engagement shift moves across the selector gates until it aligns with the selected gate. The gear shift lever is then given a to or fro movement. This causes the transverse selector/engagement shaft to rotate, thereby for- cing the striking finger to move the selector rod and fork. The synchronizing sleeve will now be able to engage the dog clutch of the appropriate gear wheel. Any misalignment between the gear shift lever support mounting and the gear shift mechanism forming part of the gearbox (caused by engine shake or rock) is thus compensated by the swing rod which provides a degree of float for the selector gate relay lever pivot point. 3.5 Splitter and range change gearboxes Ideally the tractive effect produced by an engine and transmission system developing a constant powerFig. 3.15 Remote controlled single rod with swing arm output from rest to its maximum road speed wouldsupport gear shift mechanism suitable for five and six vary inversely with its speed. This characteristic canspeed longitudinally mounted gearbox be shown as a smooth declining tractive effect curve with rising road speed (Fig. 3.16).engagement motion to the remote control tube In practice, the transmission has a fixed number(Fig. 3.15). Twisting this remote control tube by of gear ratios so that the ideal smooth tractivegiving the gear shift lever a transverse movement effect curve would be interrupted to allow for lossFig. 3.16 Ideal and actual tractive effort-speed characteristics of a vehicle 76
  • 86. of engine speed and power between each gearchange (see the thick lines of Fig. 3.16). For a vehicle such as a saloon car or light vanwhich only weighs about one tonne and has a largepower to weight ratio, a four or five speed gearboxis adequate to maintain tractive effect without toomuch loss in engine speed and vehicle performancebetween gear changes. Unfortunately, this is not the situation for heavygoods vehicles where large loads are being hauledso that the power to weight ratio is usually verylow. Under such operating conditions if the gearratio steps are too large the engine speed will dropto such an extent during gear changes that theengine torque recovery will be very sluggish(Fig. 3.17). Therefore, to minimize engine speedfall-off whilst changing gears, smaller gear ratiosteps are required, that is, more gear ratios are Fig. 3.18 Engine speed ratio chart for a vehiclenecessary to respond to the slightest change in employing either a ten speed range change or a splittervehicle load, road conditions and the drivers change gearboxrequirements. Figs 3.2 and 3.18 show that by dou-bling the number of gear ratios, the fall in enginespeed between gear shifts is much smaller. To cope applications, a three speed auxiliary gearbox can bewith moderate payloads, conventional double incorporated so that the gear ratios are trebled.stage compound gearboxes with up to six forward Usually one of these auxiliary gear ratios provides aspeeds are manufactured, but these boxes tend to be range of very low gear ratios known as crawlers orlarge and heavy. Therefore, if more gear ratios are deep gears. The auxiliary gearbox may be situatedessential for very heavy payloads, a far better way of either in front or to the rear of the conventionalextending the number of gear ratios is to utilize a two compound gearbox.speed auxiliary gearbox in series with a three, four, The combination of the auxiliary gearbox andfive or six speed conventional compound gearbox. the main gearbox can be designed to be used eitherThe function of this auxiliary box is to double the as a splitter gear change or as a range gear changenumber of gear ratios of the conventional gearbox. in the following way.With a three, four, five or six speed gearbox, thegear ratios are increased to six, eight, ten or twelve 3.5.1 Splitter gear change (Figs 3.19 and 3.20)respectively (Figs 3.2 and 3.18). For very special With the splitter arrangement, the main gearbox gear ratios are spread out wide between adjacent gears whilst the two speed auxiliary gearbox has one direct gear ratio and a second gear which is either a step up or down ratio (Fig. 3.19). The auxiliary second gear ratio is chosen so that it splits the main gearbox ratio steps in half, hence the name splitter gear change. The splitter indirect gear ratio nor- mally is set between 1.2 and 1.4:1. A typical ratio would be 1.25:1. A normal upchange sequence for an eight speed gearbox (Fig. 3.20), consisting of a main gearbox with four forward gear ratios and one reverse and a two speed auxiliary splitter stage, would be as follows: Auxiliary splitter low ratio and main gearbox first gear selected results in `first gear low (1L); auxiliaryFig. 3.17 Engine speed ratio chart for a vehicle splitter switched to high ratio but with the main gear-employing a five speed gearbox box still in first gear results in `first gear high (1H); 77
  • 87. Fig. 3.19 Eight speed constant mesh gearbox with two speed front mounted splitter changesplitter switched again to low ratio and main gear-box second gear selected results in 2L; splitterswitched to high ratio, main gearbox gear remainingin second gives 2H; splitter switched to low ratiomain gearbox third gear selected gives 3L; splitterswitched to high ratio main gearbox still in thirdgives 3 H. This procedure continues throughoutthe upshift from bottom to top gear ratio. Thus theoverall upshift gear ratio change pattern would be:Gear ratio 1 2 3 4 5 6 7 8 ReverseUpshift 1L 1H 2L 2H 3L 3H 4L 4H RL RHsequence It can therefore be predicted that alternatechanges involve a simultaneous upchange in the Fig. 3.20 Splitter change gear ratio±speed chart 78
  • 88. main gearbox and downchange in the splitter stage, Through the gear ratios from bottom to topor vice versa. `low gear range is initially selected, the main gear- Referring to the thick lines in Figs 3.2, 3.17 and box order of upchanges are first, second, third and3.18, these represent the recommended operating fourth gears. At this point the range change isspeed range for the engine for best fuel economy, moved to `high gear range and the sequence ofbut the broken lines in Fig. 3.17 represent the gear gear upchanges again becomes first, second, thirdshift technique if maximum road speed is the and fourth. Therefore the total number of gearcriteria and fuel consumption, engine wear and ratios is the sum of both low and high ranges,noise become secondary considerations. that is, eight. A tabulated summary of the upshift gear change pattern will be:3.5.2 Range gear change (Figs 3.21 and 3.22)In contrast to the splitter gear change, the rangegear change arrangement (Fig. 3.21) has the gear Gear ratio 1 2 3 4 5 6 7 8 Reverseratios between adjacent gear ratio steps set close Upshifttogether. The auxiliary two speed gearbox will have 1L 2L 3L 4L 1H 2H 3H 4H RL RH sequenceone ratio direct drive and the other one normallyequal to just over half the gear ratio spread frombottom to top. This is slightly larger than the maingearbox gear ratio spread. 3.5.3 Sixteen speed synchromesh gearbox with To change from one gear ratio to the next with, range change and integral splitter gearsfor example, an eight speed gearbox comprising (Fig. 3.23)four normal forward gears and one reverse and a This heavy duty commercial gearbox utilizes both atwo speed auxiliary range change, the pattern of two speed range change and a two speed splittergear change would be as shown in Fig. 3.22. gear change to enable the four speed gearbox toFig. 3.21 Eight speed constant mesh gearbox with two speed rear mounted range change 79
  • 89. When low or high splitter gears are engaged, the first motion shaft drive hub conveys power to the first or second pair of splitter gear wheels and hence to the layshaft gear cluster. Mid-four speed gearbox power flow (Fig. 3.23) Power from the first motion shaft at a reduced speed is transferred to the layshaft cluster of gears and subsequently provides the motion to all the other mainshaft gear wheels which are free to revolve on the mainshaft, but at relatively different speeds when in the neutral gear position. Engagement of one mid-gearbox gear ratio dog clutch locks the corresponding mainshaft drive hub to the chosen gear so that power is now able to pass from the layshaft to the mainshaft through theFig. 3.22 Range change gear ratio±speed chart selected pair of gear wheels. Reverse gear is provided via an idler gear which, when meshed between the layshaft and mainshaft,extend the gear ratio into eight steps and, when alters the direction of rotation of the mainshaft inrequired, to sixteen split (narrow) gear ratio the usual manner.intervals. The complete gearbox unit can be considered to Rear end range two speed gearbox power flowbe divided into three sections; the middle section (Fig. 3.23) When the range change is in the neu-(which is basically a conventional double stage four tral position, power passes from the mainshaft andspeed gearbox), and the first two pairs of gears at sun gear to the planet gears which then revolve onthe front end which make up the two speed splitter the output shafts carrier pin axes and in turn spingearbox. Mounted at the rear is an epicyclic gear round the annular gear and synchronizing drivetrain providing a two speed low and high range hub.change (Fig. 3.23). Engaging the low range gear locks the synchron- The epicyclic gear train at the rear doubles the izing drive hub to the gearbox casing. This forcesratios of the four speed gearbox permitting the the planet gears to revolve and walk round thedriver to initially select the low (L) gear range inside of the annular gear. Consequently, the carrierdriving through this range 1, 2, 3 and 4 then select- and output shafts which support the planet gearing the high (H) gear range. The gear change axes will also be made to rotate but at a speedsequence is again repeated but the gear ratios now lower than that of the input shaft.become 5, 6, 7 and 8. Changing to high range locks the annular gear If heavy loads are being carried, or if maximum and drive hub to the output shaft so that powertorque is needed when overtaking on hills, much flow from the planet gears is then divided betweencloser gear ratio intervals are desirable. This is the carrier and annular, but since they need toprovided by splitting the gear steps in half with rotate at differing speeds, the power flow forms athe two speed splitter gears; the gear shift pattern closed loop and jams the gearing. As a result, thereof 1st low, 1st high, 2nd low, 2nd high, 3rd low and is no gear reduction but just a straight throughso on is adopted. drive to the output shaft.Front end splitter two speed gearbox power flow 3.5.4 Twin counter shaft ten speed constant mesh(Fig. 3.23) Input power to the gearbox is supplied gearbox with synchromesh two speed rearto the first motion shaft. When the splitter synchro- mounted range change (Fig. 3.24)nizing sliding sleeve is in neutral, both the splitter With the quest for larger torque carrying capacity,low and high input gear wheels revolve on their closer steps between gear ratio changes, reducedneedle bearings independently of their supporting gearbox length and weight, a unique approachfirst motion shaft and mainshaft respectively. to fulfil these requirements has been developed 80
  • 90. Fig. 3.23 Sixteen speed synchromesh with range change and integral splitter gears 81
  • 91. Fig. 3.24 Twin countershaft ten speed constant mesh gearbox with synchromesh two speed range change 82
  • 92. adopting the two countershaft constant mesh gear- Power flow path Power flows into the main gear-box which incorporates a synchromesh two speed box through the input first motion shaft and gearrear mounted range change (Fig. 3.24). wheel. Here it is divided between the two first stage The main gearbox is in fact a double stage com- countershaft gears and is then conveyed via eachpound conventional gearbox using two counter- countershaft gear wheel to the corresponding sec-shafts (layshaft) instead of the normal single ond stage mainshaft gears. Each of these rotate atarrangement. relative speeds about the mainshaft. Torque is only transmitted to the mainshaft when the selected dog clutch drive hub is slid in to mesh with the desiredDesign and construction Referring to Fig. 3.24, gear dog teeth.there is a countershaft either side of the mainshaft The power flow can then pass directly to the out-and they are all in the same plane. What cannot be put shaft by engaging the synchromesh high rangeseen is that this single plane is inclined laterally at dog teeth. Conversely, a further gear reduction can19 to the horizontal to reduce the overall height of be made by engaging the low range synchromeshthe gearbox. dog teeth so that the power flow from the mainshaft The mainshaft is hollow and is allowed to float in auxiliary gear is split between the two auxiliarythe following manner: each end is counterbored, countershafts. The additional speed reduction isand into each counterbore is pressed a stabilizing then obtained when the split power path comesrod. The front end of this rod projects into the rear together through the second stage auxiliary outputof the input shaft which is also counterbored to gear. It should be observed that, unlike the main-house a supporting roller bearing for the stabilizer shaft, the auxiliary gear reduction output shaft hasrod. The rear projecting stabilizer rod has a spheri- no provision for radial float.cally shaped end which rests in a hole in the centre Reverse gear is obtained by incorporating an idleof a steel disc mounted inside the auxiliary drive gear between the second stage countershaft reversegear immediately behind the mainshaft. This gear gears and the mainshaft reverse gear so that theitself is carried by a ball bearing mounted in the mainshaft reverse gear is made to rotate in thegearbox housing. When torque is transmitted opposite direction to all the other forward drivethrough the gearbox, the centrally waisted 11 mm mainshaft gears.diameter section of both stabilizers deflects untilradial loads applied by the two countershaft gears 3.6 Transfer box power take-off (PTO) (Fig. 3.25)to the mainshaft gear are equalized. By these A power take-off (PTO) provides some shaft drivemeans, the input torque is divided equally between and coupling to power specialized auxiliary equip-the two countershafts and two diametrically oppo- ment at a specified speed and power output. Powersite teeth on the mainshaft gear at any one time. take-offs (PTOs) can be driven directly from theTherefore, the face width of the gear teeth can be engines timing gears, but it is more usual andreduced by about 40% compared to gearboxes using practical to take the drive from some point offsingle countershafts. Another feature of having a the gearbox. Typical power take-off applicationsmainshaft which is relatively free to float in all radial are drives for hydraulic pumps, compressors, gen-directions is that it greatly reduces the dynamic erators, hoists, derricks, capstain or cable winchloads on the gear teeth caused by small errors of platform elevators, extended ladders, hose reels,tooth profile during manufacture. A maximum drain cleaning vehicles, tippers, road sweepers,radial mainshaft float of about 0.75 mm has proved snow plough blade and throwing operations andto be sufficient to permit the shaft to centralize and any other mechanical mechanism that needs adistribute the input torque equally between the two separate source of power drive output.countershafts. To minimize end thrust, all the gears The power take-off can be driven either by one ofhave straight spur teeth which run acceptably the layshaft cluster gears, so that it is known as aquietly due to the balanced loading of the gears. side mounted transfer box, or it may be driven from Each of the five forward speeds and reverse are the back end of the layshaft, in which case it isengaged by dog teeth clutches machined on both known as a rear mounted transfer box (Fig. 3.25).ends of the drive hubs. The ends of the external Transfer boxes can either be single or two speedteeth on the drive hubs and the internal teeth in the arrangements depending upon the intended appli-mainshaft gears are chamfered at about 35 to cation. The gear ratios of the transfer box are soprovide some self-synchronizing action before chosen that output rotational speeds may be any-engagement. thing from 50 to 150% of the layshaft input speed. 83
  • 93. Fig. 3.25 Six speed constant mesh gearbox illustrating different power take-off point arrangement3.6.1 Side mounted single speed transfer box ated. With this gear train layout, the drive is con-(Fig. 3.25) veyed to the intermediate shaft by a gear wheelWith the single speed side mounted transfer box, which is in constant mesh with both the layshaftthe drive is conveyed to the output gear and shaft gear wheel and the high speed output gearby means of an intermediate gear mounted on a (Fig. 3.25). The output shaft supports the highsplined idler shaft which is itself supported by two speed output gear which is free to revolve relativespaced out ball bearings (Fig. 3.25). Engagement of to it when the transfer drive is in neutral or low gearthe transfer output shaft is obtained by sliding the is engaged. Also attached to this shaft on splines isintermediate straight toothed gear into mesh with the low speed output gear.both layshaft gear and output shaft gear by a selec- High transfer gear ratio engagement is obtainedtor fork mounted on a gear shift not shown. by sliding the low speed output gear towards the high speed output gear until its internal splines3.6.2 Side mounted two speed transfer box mesh with the dog teeth on the side of the gear.(Fig. 3.25) This then transfers the drive from the layshaft toIf a more versatile transfer power take-off is the output shaft and coupling through a simplerequired, a two speed transfer box can be incorpor- single stage gear reduction. 84
  • 94. Low transfer gear ratio engagement occurs when By selecting a 20% overdrive top gear, say, thethe low speed output gear is slid into mesh with transmission gear ratios can be so chosen that thethe smaller intermediate shaft gear. The power engine and road resistance power curves coincide atflow then takes place through a double stage peak engine power (Fig. 3.26). The undergearing(compound) gear reduction. has thus permitted the whole of the engine power curve to be shifted nearer the opposing road resist-Rear mounted two speed transfer box (Fig. 3.25) ance power curve so that slightly more engineIn some gearbox designs, or where the auxiliary power is being utilized when the two curves inter-equipment requires it, a rear mounted transfer sect. As a result, a marginally higher maximumbox may be more convenient. This transfer drive vehicle speed is achieved. In other words, thearrangement uses either an extended monolayshaft engine will be worked at a lower speed but at aor a short extension shaft attached by splines to the higher load factor whilst in this overdrive top gear.layshaft so that it protrudes out from the rear of the If the amount of overdrive for top gear isgearbox (Fig. 3.25). The extended layshaft sup- increased to 40%, the engine power curve will beports a pair of high and low speed gears which are shifted so far over that it intersects the road resist-in permanent mesh with corresponding gears ance power curve before peak engine power hasmounted on the output shaft. been obtained (Fig. 3.26) and therefore the max- When the transfer box is in neutral, the gears on imum possible vehicle speed cannot be reached.the extended layshaft are free to revolve indepen- Contrasting the direct drive 20% and 40%dently on this shaft. Engagement of either high or overdrive with direct drive top gear power curveslow gear ratios is achieved by sliding the output with respect to the road resistance power curve atdrive hub sleeve in to mesh with one or other sets of 70 km/h, as an example, it can be seen (Fig. 3.26)adjacent dog teeth forming part of the transfer box that the reserve of power is 59%, 47% and 38%layshaft constant mesh gears. Thus high gear ratio respectively. This surplus of engine power over thepower flow passes from the layshaft to the constant power absorbed by road resistance is a measure ofmesh high range gears to the output shaft and the relative acceleration ability for a particularcoupling. Conversely, low gear ratio power trans- transmission overall gear ratio setting.mission goes from the layshaft through the low A comparison of the three engine power curvesrange gears to the output drive. shows that with direct drive top gear the area in the loop made between the developed and opposing power curves is the largest and therefore the engine3.7 Overdrive considerations would respond to the changing driving conditionsPower is essential to propel a vehicle because it is a with the greatest flexibility.measure of the rate of doing work, that is, the If top gear is overdriven by 20%, as shown inamount of work being developed by the engine in Fig. 3.26, the maximum engine power would beunit time. With increased vehicle speed, more work developed at maximum vehicle speed. This thenhas to be done by the engine in a shorter time. provides the highest possible theoretical speed, The characteristic power curve over a speed but the amount of reserve power over the roadrange for a petrol engine initially increases linearly resistance power is less, so that accelerationand fairly rapidly. Towards mid-speed the steep- response will not be as rapid as if a direct driveness of the power rise decreases until the curve top gear is used. Operating under these conditions,reaches a peak. It then bends over and declines the engine speed would never exceed the peakwith further speed increase due to the difficulties power speed and so the engine could not `over-rev,experienced in breathing at very high engine speeds and as a result engine wear and noise would be(Fig. 3.26). reduced. Benefits are also gained in fuel consump- A petrol engined car is usually geared so that in its tion as shown in Fig. 3.26. The lowest specific fuelnormal direct top gear on a level road the engine consumption is shifted to a higher cruising speedspeed exceeds the peak power speed by about 10 to which is desirable on motorway journeys.20% of this speed. Consequently, the falling power Indulging in an excessive 40% overdrive top gearcurve will intersect the road resistance power curve. prevents the engine ever reaching peak power so thatThe point where both the engine and road resistance not only would maximum vehicle speed be reducedpower curves coincide fixes the road speed at which compared to the 20% overdrive gearing, but theall the surplus power has been absorbed. Therefore much smaller difference in power developed toit sets the maximum possible vehicle speed. power dissipated shown on the power curves would 85
  • 95. Fig. 3.26 Effect of over and undergearing on vehicle performanceseverely reduce the flexibility of driving in this gear. prolonged engine life, reduced noise, better fuelIt therefore becomes essential for more frequent economy and less driver fatigue will be achieved.down changes with the slightest fall-off in road Another major consideration is the unladen andspeed. A further disadvantage with excessive over- laden operation of the vehicle, particularly if it isdrive is that the minimum specific fuel consumption to haul heavy loads. Therefore most top gear over-would be shifted theoretically to the engines upper drive ratios are arrived at as a compromise.speed range which in practice could not be reached. An analysis of matching an engines performance 3.7.1 Epicyclic overdrive gearingto suit the driving requirements of a vehicle shows Epicyclic gear train overdrives are so arranged thatthat with a good choice of undergearing in top the input shaft drives the pinion carrier while thegear for motorway cruising conditions, benefits of output shaft is driven by the annular gear ring 86
  • 96. (Figs 3.27 and 3.28). The gear train may be either of The amount of overdrive (undergearing) usedsimple (single stage) or compound (double stage) for cars, vans, coaches and commercial vehiclesdesign and the derived formula for each arrange- varies from as little as 15% to as much as 45%.ment is as follows: This corresponds to undergearing ratios of between 0.87:1 and 0.69:1 respectively. Typical overdriveSimple gear train (Fig. 3.27) ratios which have been frequently used are 0.82:1 A (22%), 0.78:1 (28%) and 0.75:1 (37%).Overdrive gear ratio ˆ S‡S Example 1 An overdrive simple epicyclic gear also A ˆ S ‡ 2P train has sun and annulus gears with 21 and 75 where A ˆ number of annulus ring teeth respectively. If the input speed from the engine gear teeth drives the planet carrier at 3000 rev/min, determine S ˆ number of sun gear teeth a) the overdrive gear ratio, P ˆ number of planet b) the number of planet gear teeth, gear teeth c) the annulus ring and output shaft speed, d) the percentage of overdrive.Compound gear train (Fig. 3.28) A 75 PL (PL ‡ PS ‡ S) a) Overdrive gear ratio ˆ ˆOverdrive gear ratio ˆ A ‡ S 75 ‡ 21 PL (PL ‡ PS ‡ S) ‡ PS S 75 also A ˆ PL ‡ PS ‡ S ˆ ˆ 0:78125 96 where A ˆ number of annulus ring gear teeth b) A ˆ S ‡ 2P S ˆ number of sun AÀS Therefore Pˆ gear teeth 2 PS ˆ number of small planet 75 À 21 gear teeth ˆ 2 PL ˆ number of large planet 54 gear teeth ˆ ˆ 27 teeth 2Fig. 3.27 Simple epicycle overdrive gear train 87
  • 97. 3000c) Output speed ˆ ˆ 3840 rev/min b) A ˆ PL ‡ PS ‡ S 0:78125 ˆ 21 ‡ 24 ‡ 15 3840 À 3000 ˆ 60 teethd) Percentage of overdrive ˆ 100 3000 4000 840 Â 100 c) Output speed ˆ ˆ 4878 rev/min ˆ ˆ 28% 0:82 3000Example 2 A compound epicyclic gear train over- 4878 À 4000 d) Percentage of overdrive ˆ Â 100drive has sun, small planet and large planet gears 4000with 21, 15 and 24 teeth respectively. Determine 878 Â 100the following if the engine drives the input planet ˆcarrier at 4000 rev/min. 4000a) The overdrive gear ratio, ˆ 21:95%b) the number of annulus ring gear teeth,c) the annulus ring and output shaft speed,d) the percentage of overdrive. 3.7.2 Simple epicyclic overdrive gear train (Fig. 3.27) PL (PL ‡ PS ‡ S) If the sun gear is prevented from rotating anda) Overdrive the input shaft and planet carrier are rotated, the ˆ gear ratio PL (PL ‡ PS ‡ S) ‡ PS S pinion gears will be forced to revolve around the 24 (24 ‡ 15 ‡ 21) fixed sun gear and these pinions will revolve simul- ˆ taneously on their own axes provided by the carrier 24 (24 ‡ 15 ‡ 21) ‡ (15 Â 21) pins. 24 Â 60 1400 As a result, motion will be transferred from the ˆ ˆ carrier and pinion gears to the annulus ring gear (24 Â 60) ‡ 315 1755 due to the separate rotary movement of both the ˆ 0:82 planet carrier and the revolving planet gears, thusFig. 3.28 Compound epicycle gear train 88
  • 98. the annulus and therefore the output shaft will be the engine, the output shaft will try to run fastercompelled to revolve at a slightly faster speed. than the input shaft and so tend to release the unidirectional clutch rollers, but this is prevented3.7.3 Compound epicyclic overdrive gear train by the inner cone clutch locking the sun gear to the(Fig. 3.28) annulus, thereby jamming the sun, planet andFor only small degrees of overdrive (undergearing), annular epicyclic gear train so that they cannotfor example 0.82:1 (22%), the simple epicyclic revolve relative to each other.gearing would need a relatively large diameter Engagement of the inner cone clutch to the exter-annulus ring gear; about 175 mm if the dimension nal cone surface of the annulus gear is provided byof the gear teeth are to provide adequate strength. four stationary thrust springs (only one shown)A way of reducing the diameter of the annulus ring which are free to exert their axial load against agear for a similar degree of overdrive is to utilize a thrust plate. This in turn transfers thrust by way ofcompound epicyclic gear train which uses double a ball bearing to the rotating cone clutch supportpinion gears on each carrier pin instead of one size member splined to the sun gear sleeve. This over-of pinion. By this method, the annulus diameter is run and reverse torque will be transmitted betweenreduced to about 100 mm and there are only 60 the engine and transmission in direct drive.teeth compared to the 96 teeth annulus used with Owing to the helical cut teeth of the gear wheels,the simple epicyclic gear train. an end thrust exists between the planet gears and To transmit power, the sun gear is held still the sun gear during overrun and reverse whichwhilst the input shaft and planet carrier are rotated. tends to push the latter rearwards. Therefore, addi-This compels the large planet gear to roll around tional clamping load between the cone clutch facesthe stationary sun gear and at the same time forces is necessary.each pair of combined pinion gears to revolveabout their carrier pin axes. Consequently the small pinion gear will impart Overdrive (Fig. 3.29) When overdrive is engaged,both the pinion carrier orbiting motion and the the cone clutch, which is supported on the splinedspinning pinion gear motion to the annulus ring sleeve of the sun gear, is moved over so that itsgear so that the output shaft will be driven at a outer friction facing is in contact with the internalhigher speed to that of the input shaft. cone brake attached to the casing. Consequently the sun gear is held stationary. With the sun gear held still and the input shaft and planet carrier3.7.4 Laycock simple gear train overdrive rotating, the planet gears are forced to rotate about their own axes and at the same time rollDescription (Fig. 3.29) The overdrive unit is around the fixed sun gear, with the result that theattached to the rear of the gearbox and it consists annulus gear is driven at a faster speed than theof a constant mesh helical toothed epicyclic gear input shaft. This causes the unidirectional clutchtrain which has a central sunwheel meshing with outer member (annular carrier) to overrun thethree planet gears which also mesh with an intern- inner member (planet carrier) so that the wedgedally toothed annulus gear. The planet gears are rollers on their ramps are released. Pulling the conesupported on a carrier driven by the input shaft clutch away from the annulus cone and into fric-whilst the annulus is attached to the output shaft tional contact with the brake casing cone againstvia a carrier forming an integral part of both the axial load of the six thrust springs is achievedmembers. A double cone clutch selects the different by means of hydraulic oil pressure. This pressureratios; when engaged one side of the clutch pro- acts upon two slave pistons (only one shown) whenvides direct drive and when the other side is used, a valve is opened by operating the driver controlledoverdrive. selector switch. The outward movement of the slave pistons,Direct drive (Fig. 3.29) Direct drive is obtained due to the hydraulic pressure, draws the station-when the inner cone clutch engages with the outer ary thrust plate, ball bearing and rotating clutchcone of the annular gear. Power will then be con- member away from the annular cone and intoveyed via the unidirectional clutch to the output engagement with the outer brake cone, therebyshaft by means of the rollers which are driven up locking the sun gear to the casing. Of the helixinclined ramps and wedged between the inner and angle of the gear teeth, the torque reaction tendsouter clutch members. When the vehicle overruns to push the sun gear forward so that extra end 89
  • 99. Fig. 3.29 Laycock single epicycle overdrivethrust is necessary to maintain sufficient clamping cone clutch frees the sun gear and removes the loadthrust between the frictional faces of the cones in from the engine. The engine speed increases imme-the brake position. diately until it catches up with the output shaft, at which point the unidirectional clutch rollers climb up their respective ramps and jam. The input shaftsDirect and overdrive controlled gear change action power coming from the engine is now permitted to(Fig. 3.29) When direct drive is selected, hydrau- drive the output shaft, which in turn transmits drivelic pressure is steadily increased and this gradually to the propellor shaft. At the same time the double-releases the double-sided cone clutch member from sided cone clutch completes its movement andthe cone brake fixed to the casing. The release of the engages the annular ring cone. 90
  • 100. Overdrive is engaged when the double-sided cylinders (only one shown) and also to the solenoidcone clutch moves away from the annulus gear controlled valve and dashpot regulator relief valve.cone and makes contact with the stationary cone The dashpot pressure regulator ensures a smoothbrake, thus bringing the sliding cone clutch mem- overdrive engagement and disengagement under dif-ber and sun gear to rest. As a result of the sun gear fering operating conditions. When in direct drive thebeing held stationary, the gears now operate as an pump to slave cylinders line pressure is determinedepicyclic step up gear ratio transmission. During by the regulator relief valve spring tension whichthe time the double-sided cone clutch member controls the blow-off pressure of the oil escaping tomoves from the annulus cone to the brake cone the lubrication system. This line residual pressure inthe clutch will slip. This now permits the unidirec- direct drive is normally maintained at about 2.8 bar,tional roller clutch to transmit the drive. Whilst the but when engaging overdrive it is considerably raisedinput ramp member rotates as fast as the output by the action of the dashpot to about 20±40 bar.ramp member the roller clutch drives. However, asthe annular ring gear speed rises above that of theinput shaft, the rollers will disengage themselves Overdrive engagement Energizing the solenoidfrom their respective ramps thereby diverting the draws down the armature, thereby opening thedrive to the epicyclic gear train. inlet valve and closing the outlet valve. Oil at resi- dual line pressure will now pass through the control orifice to the base of the dashpot regulator relief valve causing the dashpot to rise and compressElectrical system (Fig. 3.29) Overdrive or direct both the dashpot spring and relief valve spring.drive gear ratio selection is controlled by an elec- Consequently, the pump to slave cylinder pressuretrical circuit which includes an overdrive on/off circuit will gradually build up as the dashpot springswitch, inhibitor switch and a relay switch. An shortens and increases in stiffness until the dashpotinhibitor switch is incorporated in the circuit to piston has reached its stop, at which point theprevent the engagement of overdrive in reverse operating pressure will be at a maximum. It is thisand some or all of the indirect gears. A relay switch gradual increase in line pressure which provides theis also included in the circuit so that the overdrive progressive compression of the clutch thrust springson/off switch current rating may be small com- and the engagement of the cone clutch with the fixedpared to the current draw requirements of the cone brake.control solenoid. The overdrive may be designedto operate only in top gear, but sometimes theoverdrive is permitted to be used in third or even Direct drive engagement De-energizing the sole-second gear. Selection and engagement of over- noid closes the inlet valve and opens the outletdrive by the driver is obtained by a steering column valve. This prevents fresh oil entering the dashpotor fascia panel switch. When the driver selects cylinder and allows the existing oil under the dash-overdrive in top gear or one of the permitted indir- pot to exhaust by way of the control orifice and theect gears, say third, the on/off switch is closed and outlet valve back to the sump. The control orificethe selected gearbox gear ratio selector rod will restricts the flow of escaping oil so that the pressurehave pushed the inhibitor switch button into the drop is progressive. This enables the clutch thrustclosed switch position. Current will now flow from springs to shift the cone clutch very gradually intothe battery to the relay switch, magnetizing the contact with the annulus cone.relay winding so that as the relay contacts close, alarger current will immediately energize the sole-noid and open the control valve so that overdrive 3.7.5 Laycock compound gear train overdrivewill be engaged. (Fig. 3.30) Overdrive When overdrive is selected, the double-Hydraulic system (Fig. 3.29) A plunger type sided cone clutch contacts the brake cone whichpump driven by an eccentric formed on the input forms part of the casing. This brings the sun gearshaft supplies the hydraulic pressure to actuate the which is attached to the sliding clutch member to aslave pistons and thereby operates the clutch. The standstill.pump draws oil from the sump through a filter (not The input drive passes from the pinion carrier toshown). It is then pressurized by the plunger and the annulus ring and hence to the output shaftdelivered through a non-return valve to both slave through the small planet gear. At the same time, the 91
  • 101. Fig. 3.30 Laycock double epicycle overdrivelarge planet gear absorbs the driving torque reaction planet gears, drive forward the annular ring gear at aand in the process is made to revolve around the faster speed relative to that of the input.braked sun gear. The overdrive condition is created The overall gear ratio step up is achieved byby the large planet gears being forced to roll `walk having two stages of meshing gear teeth; oneabout the sun gear, while at the same time revolving between the large pinion and sun gear and theon their own axes. As a result, the small planet gears, other between the small pinion and annulus ringalso revolving on the same carrier pins as the large gear. By using this compound epicyclic gear train, a 92
  • 102. relatively large step up gear ratio can be obtained the low pressure ball valve will open and relieve thefor a given diameter of annulus ring gear compared excess pressure. Under these conditions the axialto a single stage epicyclic gear train. load exerted by the clutch thrust springs will clamp the double-sided floating conical clutch member to the external conical shaped annulus ring gear.Direct drive (Fig. 3.30) Direct drive is attained byreleasing the double-sided cone clutch member from Overdrive engagement To select overdrive thethe stationary conical brake and shifting it over so solenoid is energized. This closes the solenoid ballthat it contacts and engages the conical frictional valve, preventing oil escaping via the lubricationsurface of the annulus ring gear. The power flow system back to the sump. Oil pressure will nowfrom the input shaft and planet carrier now divides build up to about 26±30 bar, depending on vehicleinto two paths Ð the small planet gear to annulus application, until sufficient thrust acts on bothring gear route and the large planet gear, sun gear slave pistons to compress the clutch thrust springs,and double-sided clutch member route, again finish- thereby permitting the double-sided clutch membering up at the annulus ring gear. With such a closed to shift over and engage the conical surface of theloop power flow arrangement, where the gears can- stationary brake. To enable the engagement actionnot revolve independently to each other, the gears to overdrive to progress smoothly and to limit thejam so that the whole gear train combination rotates maximum hydraulic pressure, a high pressure valveas one about the input to output shaft axes. It jumper is made to be pushed back and progres-thereby provides a straight through direct drive. It sively open. This controls and relieves the pressureshould be observed that the action of the unidirec- rise which would otherwise cause a rough, andtional roller clutch is similar to that described for the possibly sudden, clutch engagement.single stage epicyclic overdrive.Clutch operating (Fig. 3.30) Engagement of 3.8 Setting gear ratiosdirect drive and overdrive is achieved in a similar Matching the engines performance characteristicsmanner to that explained under single stage epicyc- to suit a vehicles operating requirements is pro-lic overdrive unit. vided by choosing a final drive gear reduction and Direct drive is provided by four powerful springs then selecting a range of gear ratios for maximumholding the double-sided conical clutch member in performance in terms of the ability to climb gradi-frictional contact with the annulus ring gear. Con- ents, achievement of good acceleration through theversely, overdrive is obtained by a pair of hydraulic gears and ability to reach some predeterminedslave pistons which overcome and compress the maximum speed on a level road.clutch thrust springs, pulling the floating conicalclutch member away from the annulus and into 3.8.1 Setting top gearengagement with the stationary conical brake. To determine the maximum vehicle speed, the engine brake power curve is superimposed onto the powerHydraulic system (Fig. 3.30) Pressure supplied by requirement curve which can be plotted from thethe hydraulic plunger type pump draws oil from the sum of both the rolling (Rr ) and air (Ra ) resistancesump and forces it past the non-return ball valve to covering the entire vehicles speed range (Fig. 3.31).both the slave cylinders and to the solenoid valve The total resistance R opposing motion at anyand the relief valve. speed is given by: R ˆ Rr ‡ RaDirect drive engagement When direct drive is ˆ 10Cr W ‡ CD AV 2engaged, the solenoid valve opens due to the sole-noid being de-energized. Oil therefore flows not where Cr ˆ coefficient of rolling resistanceonly to the slave cylinders but also through the W ˆ gross vehicle weight (kg)solenoid ball valve to the overdrive lubrication sys- CD ˆ coefficient of aerodynamic resist-tem where it then spills and returns to the sump. Arelatively low residual pressure will now be main- ance (drag)tained within the hydraulic system. Should the oil A ˆ projected frontal area of vehicle (m2 )pressure rise due to high engine speed or blockage, V ˆ speed of vehicle (km/h) 93
  • 103. Fig. 3.31 Forces opposing vehicle motion over its speed Fig. 3.32 Relationship of power developed and roadrange power required over the vehicles speed range 3.8.2 Setting bottom gear The top gear ratio is chosen so that the maxi- The maximum payload and gradient the vehicle ismum road speed corresponds to the engine speed at expected to haul and climb determines the necessarywhich maximum brake power is obtained (or just tractive effort, and hence the required overall gearbeyond) (Fig. 3.32). ratio. The greatest gradient that is likely to be Gearing is necessary to ensure that the vehicle encountered is decided by the terrain the vehicle isspeed is at a maximum when the engine is develop- to operate over. This normally means a maximuming approximately peak power. gradient of 5 to 1 and in the extreme 4 to 1. TheThus Linear wheel speed ˆ Linear road speed minimum tractive effort necessary to propel a vehicle up the steepest slope may be assumed to be approxi- dN 1000 ˆ V (m/min) mately equivalent to the sum of both the rolling and GF 60 gradient resistances opposing motion (Fig. 3.31). 60 dN The rolling resistance opposing motion may be; Final drive gear ratio GF ˆ 100 V determined by the formula: dN Rr ˆ 10Cr W ˆ 0:06 Vwhere GF ˆ final drive gear ratio where Rr ˆ rolling resistance (N) N ˆ engine speed (rev/min) Cr ˆ coefficient of rolling resistance d ˆ effective wheel diameter (m) W ˆ gross vehicle weight (kg) V ˆ road speed at which peak power is Average values for the coefficient of rolling developed (km/h) resistance for different types of vehicles travelling Example A vehicle is to have a maximum road at very slow speed over various surfaces have beenspeed of 150 km/h. If the engine develops its peak determined and are shown in Table 3.2.power at 6000 rev/min and the effective road wheel Likewise, the gradient resistance (Fig. 3.33)diameter is 0.54 m, determine the final drive gear ratio. opposing motion may be determined by the for- mula: 0:06 dNGF ˆ 10W V Rg ˆ or 10W sin G 0:06 Â 3:142 Â 0:54 Â 6000 where Rg ˆ gradient resistance (N) ˆ 150 W ˆ gross vehicle weight (10W kg ˆ WN) ˆ 4:07 X 1 G ˆ gradient (1 in x) ˆ sin 94
  • 104. Table 3.2 Average values of coefficient of rollingresistance where GF ˆ final drive gear ratio GB ˆ bottom gear ratio Coefficient of rolling resistance (Cr) M ˆ mechanical efficiencyVehicle type Concrete Medium hard soil Sand E ˆ tractive effort (N) T ˆ maximum engine torque (Nm)Passenger Car 0.015 0.08 0.30 R ˆ effective road wheel radius (m)Trucks 0.012 0.06 0.25Tractors 0.02 0.04 0.20 Example A vehicle weighing 1500 kg has a coefficient of rolling resistance of 0.015. The trans-Note The coefficient of rolling resistance is the ratio of therolling resistance to the normal load on the tyre. mission has a final drive ratio 4.07:1 and an overall R mechanical efficiency of 85%.i:e: Cr ˆ Wr If the engine develops a maximum torque of 100 Nm (Fig. 3.34) and the effective road wheel radius is 0.27 m, determine the gearbox bottom gear ratio. Assume the steepest gradient to be encountered is a one in four. Rr ˆ 10Cr W ˆ 10 Â 0:015 Â 150 ˆ 225N 10W Rg ˆ G 10 Â 1500 ˆ ˆ 3750N 4 E ˆ Rr ‡ Rg ˆ 3750 ‡ 225 ˆ 3975N eR GB ˆ TGF MFig. 3.33 Gradient resistance to motion 3975 Â 0:27 ˆ ˆ 3:1X1 100 Â 4:07 Â 0:85Tractive effort ˆ Resisting forces opposing motion EˆR ˆ Rr ‡ Rg (N)where E ˆ tractive effort (N) R ˆ resisting forces (N) Once the minimum tractive effort has been cal-culated, the bottom gear ratio can be derived in thefollowing way:Driving torque ˆ Available torque ER ˆ T GB GF M ER; Bottom gear ratio GB ˆ TGF M Fig. 3.34 Engine torque to speed characteristics 95
  • 105. 3.8.3 Setting intermediate gear ratiosRatios between top and bottom gears should be NH NL Also ˆspaced in such a way that they will provide the G2 G3tractive effort±speed characteristics as close to the NLideal as possible. Intermediate ratios can be best ; G3 ˆ G2 NHselected as a first approximation by using a geo-metric progression. This method of obtaining the NH NLgear ratios requires the engine to operate within the and ˆ G3 G4same speed range in each gear, which is normally NLselected to provide the best fuel economy. ; G4 ˆ G3 Consider the engine to vehicle speed character- NHistics for each gear ratio as shown (Fig. 3.35). When NH NL ˆchanging gear the engine speed will drop from the G4 G5highest NH to the lowest NL without any change in NLroad speed, i.e. V1 , V2 , V3 etc. ; G5 ˆ G4 NHLet G1 ˆ 1st overall gear ratio NL G2 ˆ 2nd overall gear ratio The ratio is known as the minimum to max- NH G3 ˆ 3rd overall gear ratio imum speed range ratio K for a given engine. G4 ˆ 4th overall gear ratio NL G5 ˆ 5th overall gear ratio Now, gear G2 ˆ G1 ˆ G1 K, NH NLwhere Overall ˆ Engine speed (rev/min) since ˆ k (a constant) gear ratio Road wheel speed (rev/min) NHWheel speed when engine is on the high limit NH in NL gear G3 ˆ G2 ˆ G2 K ˆ (G1 K)K NH NHfirst gear G1 ˆ (rev/min) G1 ˆ G 1 K2Wheel speed when engine is on the low limit NL in NL NL gear G4 ˆ G3 ˆ G3 K ˆ (G1 K 2 )Ksecond gear G2 ˆ (rev/min) NH G2These wheel speeds must be equal for true rolling ˆ G 1 K3 NH NL NLHence ˆ gear G5 ˆ G4 ˆ G4 K ˆ (G1 K 3 )K G1 G2 NH NL ˆ G 1 K4 : ; G2 ˆ G1 NH Hence the ratios form a geometric progression.Fig. 3.35 Gear ratios selected on geometric progression 96
  • 106. The following relationship will also apply for a Example A transmission system for a vehiclefive speed gearbox: is to have an overall bottom and top gear ratioG2 G3 G4 G5 NL of 20:1 and 4.8 respectively. If the minimum to maxi- ˆ ˆ ˆ ˆ ˆK mum speeds at each gear changes are 2100 andG1 G2 G3 G4 NH 3000 rev/min respectively, determine the following: G5 a) the intermediate overall gear ratiosand G5 ˆ G1 K 4 or K4 ˆ b) the intermediate gearbox and top gear ratios. G1 1 r NL G5 4 Kˆ 4 G5 NHHence K ˆ or G1 G1 2100 ˆ ˆ 0:7 In general, if the ratio of the highest gear (GT ) 3000and that of the lowest gear (GB ) have been deter-mined, and the number of speeds (gear ratios) of a) 1st gear ratio G1 ˆ 20:0:1the gearbox nG is known, the constant K can be 2nd gear ratio G2 ˆ G1 K ˆ 20  0:7 ˆ 14:0X1determined by: 3rd gear ratio G3 ˆ G1 K 2 ˆ 20  0:72 ˆ 9:8X1 4th gear ratio G4 ˆ G1 K 3 ˆ 20  0:73 ˆ 6:86X1 1À1 5th gear ratio G5 ˆ G1 K 4 ˆ 20  0:74 ˆ 4:8X1 GT nG Kˆ GB 20:0 GT ‡1 b) G1 ˆ ˆ 4:166X1So ˆ K nG 4:8 GB 14:0 À1 G2 ˆ ˆ 2:916X1; GT ˆ GB K nG 4:8 For commercial vehicles, the gear ratios in 9:8the gearbox are often arranged in geometric G3 ˆ ˆ 2:042X1 4:8progression. For passenger cars, to suit the chan-ging traffic conditions, the step between the ratios 6:86 G4 ˆ ˆ 1:429X1of the upper two gears is often closer than that 4:8based on geometric progression. As a result, thiswill affect the selection of the lower gears to some 4:8 Top gear G5 ˆ ˆ 1:0X1extent. 4:8 97
  • 107. 4 Hydrokinetic fluid couplings and torque convertersA fluid drive uses hydrokinetic energy as a means 4.1 Hydrokinetic fluid couplingsof transferring power from the engine to the trans- (Figs 4.1 and 4.2)mission in such a way as to automatically match The hydrokinetic coupling, sometimes referred tothe vehicles speed, load and acceleration require- as a fluid flywheel, consists of two saucer-shapedment characteristics. These drives may be of a discs, an input impeller (pump) and an outputsimple two element type which takes up the drive turbine (runner) which are cast with a number ofsmoothly without providing increased torque or flat radial vanes (blades) for directing the flow paththey may be of a three or more element unit of the fluid (Fig. 4.1).which not only conveys the power as required Owing to the inherent principle of the hydro-from the engine to the transmission, but also multi- kinetic coupling, there must be relative slip betweenplies the output torque in the process. the input and output member cells exposed to eachFig. 4.1 Fluid coupling action 98
  • 108. other, and the vortex flow path created by pairs of The two half-members are put together so thatadjacent cells will be continuously aligned and the fluid can rotate as a vortex. Originally it wasmisaligned with different cells. common practice to insert at the centre of rotation a With equal numbers of cells in the two half hollow core or guide ring (sometimes referred to asmembers, the relative cell alignment of all the cells the torus) within both half-members to assist inoccurs together. Consequently, this would cause a establishing fluid circulation at the earliest momentjerky transfer of torque from the input to the output of relative rotation of the members. These couplingsdrive. By having differing numbers of cells within had the disadvantage that they produced consider-the impeller and turbine, the alignment of each pair able drag torque whilst idling, this being due mainlyof cells at any one instant will be slightly different to the effectiveness of the core guide in circulatingso that the impingement of fluid from one member fluid at low speeds. As coupling development pro-to the other will take place in various stages of gressed, it was found that turbine drag was reducedcirculation, with the result that the coupling torque at low speeds by using only a core guide on thetransfer will be progressive and relatively smooth. impeller member (Fig. 4.2). With the latest designFig. 4.2 Fluid coupling 99
  • 109. these cores are eliminated altogether as this also by the impeller around its axis and secondly itreduces fluid interference in the higher speed range circulates round the cells in a vortex motion.and consequently reduces the degree of slip for a This circulation of fluid only continues as long asgiven amount of transmitted torque (Fig. 4.6). there is a difference in the angular speeds of the impeller and turbine, because only then is the cen-4.1.1 Hydrokinetic fluid coupling principle of trifugal force experienced by the fluid in the fasteroperation (Figs 4.1 and 4.3) moving impeller greater than the counter centri-When the engine is started, the rotation of the fugal force acting on the fluid in the slower movingimpeller (pump) causes the working fluid trapped turbine member. The velocity of the fluid aroundin its cells to rotate with it. Accordingly, the fluid is the couplings axis of rotation increases while itsubjected to centrifugal force and is pressurized so flows radially outwards in the impeller cells due tothat it flows radially outwards. the increased distance it has moved from the centre To understand the principle of the hydrokinetic of rotation. Conversely, the fluid velocity decreasescoupling it is best to consider a small particle of when it flows inwards in the turbine cells. It there-fluid circulating between one set of impeller and fore follows that the fluid is given kinetic energy byturbine vanes at various points A, B, C and D as the impeller and gives up its kinetic energy to theshown in Figs 4.1 and 4.3. turbine. Hence there is a transference of energy Initially a particle of fluid at point A, when the from the input impeller to the output turbine, butengine is started and the impeller is rotated, will there is no torque multiplication in the process.experience a centrifugal force due to its mass andradius of rotation, r. It will also have acquired some 4.1.2 Hydrokinetic fluid coupling velocitykinetic energy. This particle of fluid will be forced diagrams (Fig. 4.3)to move outwards to point B, and in the process The resultant magnitude of direction of the fluidof increasing its radius of rotation from r to R, will leaving the impeller vane cells, VR, is dependentnow be subjected to considerably more centrifugal upon the exit velocity, VE, this being a measure offorce and it will also possess a greater amount of the vortex circulation flow rate and the relativekinetic energy. The magnitude of the kinetic energy linear velocity between the impeller and turbine, VL.at this outermost position forces it to be ejected The working principle of the fluid couplingfrom the mouth of the impeller cell, its flow path may be explained for various operating conditionsmaking it enter one of the outer turbine cells at assuming a constant circulation flow rate by meanspoint C. In doing so it reacts against one side of the of velocity vector diagrams (Fig. 4.3).turbine vanes and so imparts some of its kinetic When the vehicle is about to pull away, the engineenergy to the turbine wheel. The repetition of fluid drives the impeller with the turbine held stationary.particles being flung across the junction between the Because the stalled turbine has no motion, the rela-impeller and turbine cells will force the first fluid tive forward (linear) velocity VL between the twoparticle in the slower moving turbine member members will be large and consequently so will the(having reduced centrifugal force) to move inwards resultant entry velocity VR. The direction of fluidto point D. Hence in the process of moving inwards flow from the impeller exit to turbine entrance willfrom R to r, the fluid particle gives up most of its make a small angle Â1 , relative to the forward direc-kinetic energy to the turbine wheel and subsequently tion of motion, which therefore produces consider-this is converted into propelling effort and motion. able drive thrust to the turbine vanes. The creation and conversion of the kinetic As the turbine begins to rotate and catch up toenergy of fluid into driving torque can be visualized the impeller speed the relative linear speed isin the following manner: when the vehicle is at rest reduced. This changes the resultant fluid flowthe turbine is stationary and there is no centrifugal direction to Â2 and decreases its velocity. The netforce acting on the fluid in its cells. However, when output thrust, and hence torque carrying capacity,the engine rotates the impeller, the working fluid will be less, but with the vehicle gaining speed therein its cells flows radially outwards and enters the is a rapid decline in driving torque requirements.turbine at the outer edges of its cells. It therefore At high turbine speeds, that is, when the outputcauses a displacement of fluid from the inner edges to input speed ratio is approaching unity, there willof the turbine cells into the inner edges of the be only a small relative linear velocity and resultantimpeller cells, thus a circulation of the fluid will entrance velocity, but the angle Â3 will be large.be established between the two half cell members. This implies that the magnitude of the fluid thrustThe fluid has two motions; firstly it is circulated will be very small and its direction ineffective in 100
  • 110. Fig. 4.4 Relationship of torque capacity efficiency and speed ratio for fluid couplings Fig. 4.5 Relationship of engine speed, torque and slipFig. 4.3 Principle of the fluid coupling for a fluid coupling 101
  • 111. rotating the turbine. Thus the output member will in impeller speed, considerably raises the couplingslip until sufficient circulating fluid flow imparts torque carrying capacity. A further controlling fac-enough energy to the turbine again. tor which affects the torque transmitted is the It can be seen that at high rotational speeds the quantity of fluid circulating between the impellercycle of events is a continuous process of output and turbine. Raising or lowering the fluid level inspeed almost, but never quite, catching up to input the coupling increases or decreases the torquespeed, the exception being when the drive changes which can be transmitted to the turbine (Fig. 4.4).from engine driven to overrun transmission drivenwhen the operating conditions will be reversed. 4.3 Fluid friction coupling (Figs 4.6 and 4.7) A fluid coupling has the take-up characteristics4.2 Hydrokinetic fluid coupling efficiency and which particularly suit the motor vehicle but ittorque capacity (Figs 4.4 and 4.5) suffers from two handicaps that are inherent inCoupling efficiency is the ratio of the power avail- the system. Firstly, idling drag tends to make theable at the turbine to the amount of power supplied vehicle creep forwards unless the parking brake isto the impeller. The difference between input and fully applied, and secondly there is always a smalloutput power, besides the power lost by fluid shock, amount of slip which is only slight under part loadfriction and heat, is due mainly to the relative slip (less than 2%) but becomes greater when transmit-between the two members (Fig. 4.4). A more useful ting anything near full torque.term is the percentage slip, which is defined as the These limitations have been overcome for largeratio of the difference in input and output speeds truck applications by combining a shoe and drumdivided by the input speed and multiplied by 100. centrifugally operated clutch to provide a positive NÀn lock-up at higher output speeds with a smalleri:e: 7 slip ˆ Â 100 coreless fluid coupling than would be necessary if N the drive was only to be through a fluid coupling. The percentage slip will be greatly influenced by The reduced size and volume of fluid circulation inthe engine speed and output turbine load conditions the coupling thereby eliminate residual idling drag(Fig. 4.5). A percentage of slip must always exist to (Fig. 4.6).create a sufficient rate of vortex circulation which is With this construction there is a shoe carrieressential to impart energy from the impeller to the between the impeller and flywheel attached to theturbine. The coupling efficiency is at best about 98% output shaft. Mounted on this carrier are four brakeunder light load high rotational speed conditions, shoes with friction material facings. They are eachbut this will be considerably reduced as turbine output pivoted (hinged) to the carrier member at one endload is increased or impeller speed is lowered. If the and a garter spring (coil springs shown on front viewoutput torque demand increases, more slip will occur to illustrate action) holds the shoes in their retrac-and this will increase the vortex circulation velocity tion position when the output shaft is at rest.which will correspondingly impart more kinetic When the engine is accelerated the fluid couplingenergy to the output turbine member, thus raising automatically takes up the drive with maximumthe torque capacity of the coupling. An additional smoothness. Towards maximum engine torquefeature of such couplings is that if the engine should speed the friction clutch shoes are thrown outwardstend to stall due to overloading when the vehicle by the centrifugal effect until they come into con-is accelerated from rest, the vortex circulation will tact with the flywheel drum. The frictional grip willimmediately slow down, preventing further torque now lock the input and output drives together.transfer until the engines speed has recovered. Subsequently the fluid vortex circulation stops Fluid coupling torque transmitting capacity for a and the fluid coupling ceases to function (Fig. 4.7).given slip varies as the fifth power of the impeller Relative slip between input and output member ininternal diameter and as the square of its speed. low gear is considerably reduced, due to the auto-i:e: T G D5 N 2 matic friction clutch engagement, and engine brak- ing is effectively retained down to idling speeds.where D ˆ impeller diameter N ˆ impeller speed (rev/min) 4.4 Hydrokinetic three element torque converter (Figs 4.8 and 4.9) Thus it can be seen that only a very small A three element torque converter coupling is com-increase in impeller diameter, or a slight increase prised of an input impeller casing enclosing the 102
  • 112. Fig. 4.6 Fluid friction coupling output turbine wheel. There are about 26 and 23 blades for the impeller and turbine elements respect- ively. Both of these elements and their blades are fabricated from low carbon steel pressings. The third element of the converter called the stator is usually an aluminium alloy casting which may have some- thing in the order of 15 blades (Figs 4.8 and 4.9). The working fluid within a converter when the engine is operating has two motions: 1 Fluid trapped in the impeller and turbine vane cells revolves bodily with these members about their axis of rotation. 2 Fluid trapped between the impeller and turbine vane cells and their central torus core rotates in a circular path in the section plane, this being known as its vortex motion.Fig. 4.7 Relationship of torque carrying capacity, effici- When the impeller is rotated by the engine, it actsency and output speed for a fluid coupling as a centrifugal pump drawing in fluid near the 103
  • 113. centre of rotation, forcing it radially outwards It therefore follows that the engine torquethrough the cell passages formed by the vanes to the delivered to the impeller and the reaction torqueimpeller peripheral exit. Here it is ejected due to its transferred by the fluid to the impeller are bothmomentum towards the turbine cell passages and in transmitted to the output turbine through the mediathe process acts at an angle against the vanes, thus of the fluid.imparting torque to the turbine member (Fig. 4.8). The fluid in the turbine cell passages moves i.e. Engine Reaction Output turbine ‡ ˆinwards to the turbine exit. It is then compelled to torque torque torqueflow between the fixed stator blades (Fig. 4.9). Thereaction of the fluids momentum as it glides over 4.4.1 Hydrokinetic three element torquethe curved surfaces of the blades is absorbed by the converter principle of operation (Fig. 4.8)casing to which the stator is held and in the process When the engine is running, the impeller acts asit is redirected towards the impeller entrance. It a centrifugal pump and forces fluid to flow radiallyenters the passages shaped by the impeller vanes. around the vortex passage made by the vanes andAs it acts on the drive side of the vanes, it imparts core of the three element converter. The rotationa torque equal to the stator reaction in the direction of the impeller by the engine converts the engineof rotation (Fig. 4.8). power into hydrokinetic energy which is utilized inFig. 4.8 Three element torque converter action 104
  • 114. Fig. 4.9 Three element torque converterproviding a smooth engine to transmission take-up to the impeller, the pressure generated in the impel-and in producing torque multiplication if a third ler will be far greater than in the turbine. Subse-fixed stator member is included. quently the fluid particle in the turbine curved An appreciation of the principle of the converter passage will be forced inwards to the exit point Dcan be obtained by following the movement and and in doing so will give up more of its kineticevents of a fluid particle as it circulates the vortex energy to the turbine wheel.passage (Fig. 4.8). The fluid particle, still possessing kinetic energy Consider a fluid particle initially at the small at the turbine exit, now moves to the stator bladesdiameter entrance point A in the impeller. As the entrance side to point E. Here it is guided by theimpeller is rotated by the engine, centrifugal force curvature of the blades to the exit point F.will push the fluid particle outwards to the impellers From the fixed stator (reactor) blades the fluidlargest exit diameter, point B. Since the particles path is again directed to the impeller entrance pointcircumferential distance moved every revolution A where it imparts its hydrokinetic energy to thewill be increased, its linear velocity will be greater impeller, this being quite separate to the kineticand hence it will have gained kinetic energy. energy produced by the engine rotating the impeller. Pressure caused by successive particles arriving Note that with the fluid coupling, the transfer ofat the impeller outermost cell exit will compel the fluid from the turbine exit to the impeller entranceparticle to be flung across the impeller±turbine is direct. Thus the kinetic energy gained by thejunction where it acts against the side of cell vane input impeller is that lost by the output turbineit has entered at point C and thereby transfers some and there is no additional gain in output turningof its kinetic energy to the turbine wheel. Because effort, as is the case when a fixed intermediatethe turbine wheel rotates at a lower speed relative stator is incorporated. 105
  • 115. 4.4.2 Hydrokinetic three element torque When the turbine is in the stalled condition andconverter velocity diagrams (Figs 4.9 and 4.10) the impeller is being driven by the engine, the directionThe direction of fluid leaving the turbine to enter of the fluid leaving the impeller will be determinedthe stator blades is influenced by the tangential exit entirely by the curvature and shape of the turbinevelocity which is itself determined by the vortex vanes. Under these conditions, the fluids directioncirculating speed and the linear velocity due to the of motion, Â1 , will make it move deep into the con-rotating turbine member (Fig. 4.10). cave side of the stator blades where it reacts and isFig. 4.10 Principle of the single stage torque converter 106
  • 116. made to flow towards the entrance of the impeller in zero. Above this speed the stator is freewheeled.a direction which provides the maximum thrust. This offers less resistance to the circulating fluid Once the turbine begins to rotate, the fluid will and therefore produces an improvement in coup-acquire a linear velocity so that the resultant ling efficiency (Figs 4.11 and 4.12).effective fluid velocity direction will now be Â2 . This description of the operating conditions of theA reduced backward reaction to the stator will be converter coupling shows that if the transmissionproduced so that the direction of the fluids is suddenly loaded the output turbine speed willmomentum will not be so effective. automatically drop, causing an increase in fluid As the turbine speed of rotation rises, the fluids circulation and correspondingly a rise in torquelinear forward velocity will also increase and, multiplication, but conversely a lowering of efficiencyassuming that the turbines tangential exit velocity due to the increased slip between input and outputdoes not alter, the resultant direction of the fluid members. When the output conditions have changedwill have changed to Â3 where it now acts on the and a reduction in load or an increase in turbine speedconvex (back) side of the stator blades. follows the reverse happens; the efficiency increases Above the critical speed, when the fluids thrust and the output to input torque ratio is reduced.changes from the concave to the convex side of theblades, the stator reaction torque will now act in theopposite sense and redirect the fluid. Thus its result- 4.5 Torque converter performance terminologyant direction towards the impeller entry passages will (Figs 4.11 and 4.12)hinder instead of assist the impeller motion. The result To understand the performance characteristics ofof this would be in effect to cancel out some of the a fluid drive (both coupling and converter), it isengines input torque with further speed increases. essential to identify and relate the following terms The inherent speed limitation of a hydrokinetic used in describing various relationshipsand conditions.converter is overcome by building into the statorhub a one way clutch (freewheel) device (Fig. 4.9). 4.5.1 Fluid drive efficiency (Figs 4.11 and 4.12)Therefore, when the direction of fluid flow changes A very convenient method of expressing the energysufficiently to impinge onto the back of the blades, losses, due mainly to fluid circulation within a fluidthe stator hub is released, allowing it to spin freely drive at some given output speed or speed ratio, isbetween the input and output members. The free-wheeling of the stator causes very little fluid inter-ference, thus the three element converter nowbecomes a two element coupling. This conditionprevents the decrease in torque for high outputspeeds and produces a sharp rise in efficiency atoutput speeds above the coupling point.4.4.3 Hydrokinetic torque convertercharacteristics (Figs. 4.11 and 4.12)Maximum torque multiplication occurs when thereis the largest speed difference between the impellerand turbine. A torque output to input ratio ofabout 2:1 normally occurs with a three elementconverter when the turbine is stationary. Undersuch conditions, the vortex rate of fluid circulationwill be at a peak. Subsequently the maximumhydrokinetic energy transfer from the impeller toturbine then stator to impeller again takes place(Figs 4.11 and 4.12). As the turbine output speedincreases relative to the impeller speed, the effi-ciency rises and the vortex velocity decreases andso does the output to input torque ratio until even-tually the circulation rate of fluid is so low that itcan only support a 1:1 output to input torque Fig. 4.11 Characteristic performance curves for a threeratio. At this point the reaction torque will be element converted coupling 107
  • 117. impeller to turbine speed variation, with the result that the vortex fluid circulation and correspond- ingly torque conversion are at a maximum, conver- sely converter efficiency is zero. Whilst these stall conditions prevail, torque conversion loading drags the engine speed down to something like 60±70% of the engines maximum torque speed, i.e. 1500±2500 rev/min. A converter should only be held in the stall condition for the minimum of time to prevent the fluid being overworked. 4.5.5 Design point (Figs 4.11 and 4.12) Torque converters are so designed that their inter- nal passages formed by the vanes are shaped so as to make the fluid circulate with the minimum of resistance as it passes from one member to another member at definite impeller to turbine speed ratio, known as the design point. A typical value might be 0.8:1. Above or below this optimum speed ratio, theFig. 4.12 Characteristic performance curves for a con-verter coupling plotted to a base of output (turbine speed) resultant angle and direction of fluid leaving oneto input (impeller speed) member to enter another will alter so that the flow from the exit of one member to the entry of another will no longer be parallel to the surfaces of the vanes,to measure its efficiency, that is, the percentage in fact it will strike the sides of the passage vanesratio of output to input work done. entered. When the exit and entry angles of the vanes do not match the effective direction of fluid motion, Output work donei:e: Efficiency ˆ Â 100 some of its momentum will be used up in entrance Input work done losses and consequently the efficiency declines as the speed ratio moves further away on either side of the4.5.2 Speed ratio (Fig. 4.12) design point. Other causes of momentum losses areIt is frequently necessary to compare the output internal fabrication finish, surface roughness andand input speed differences at which certain events inter-vane or blade thickness interference. If theoccur. This is normally defined in terms of a speed design point is shifted to a lower speed ratio, say 0.6,ratio of output (turbine) speed N2 to the input the torque multiplication will be improved at(impeller) speed N1. stall and lower speed conditions at the expense of an earlier fall-off in efficiency at the high speed ratio N2i:e: Speed ratio ˆ such as 0.8. There will be a reduction in the torque N1 ratio but high efficiency will be maintained in the upper speed ratio region.4.5.3 Torque ratio (Fig. 4.12)The torque multiplication within a fluid drive is 4.5.6 Coupling point (Figs 4.11 and 4.12)more conveniently expressed in terms of a torque As the turbine speed approaches or exceeds that ofratio of output (turbine) torque T2 to the input the impeller, the effective direction of fluid entering(impeller) torque T1. the passages between the stator blades changes T2 from pushing against the concave face to beingi:e: Torque ratio ˆ redirected towards the convex (back) side of the T1 blades. At this point, torque conversion due to fluid transfer from the fixed stator to the rotating4.5.4 Stall speed (Figs 4.11 and 4.12) impeller, ceases. The turbine speed when the direc-This is the maximum speed which the engine tion of the stator reaction is reversed is known as thereaches when the accelerator pedal is fully down, the coupling point and is normally between 80 and 90%transmission in drive and the foot brake is fully of the impeller speed. At this point the stator isapplied. Under such conditions there is the greatest released by the freewheel device and is then driven 108
  • 118. in the same direction as the impeller and turbine. torque (drive speed) is greater than that of theAt and above this speed the stator blades will spin output member. If the conditions are reversed andwith the impeller and turbine which then simply act the output members applied torque (or speed)as a fluid coupling, with the benefit of increasing becomes greater than that of the input, the outputefficiency as the turbine output speed approaches member will overrun the input member (rotatebut never reaches the input impeller speed. faster). Thus the lock between the two members will be automatically released. Immediately the4.5.7 Racing or run-away point (Fig. 4.12) drive will be discontinued which permits the inputIf the converter does not include a stator freewheel and output members to revolve independently todevice or if the mechanism is jammed, then the one another.direction of fluid leaving the stator would progres- Overrun clutches can be used for a number ofsively change from transferring fluid energy to applications, such as starter motor pre-engagementassist the impeller rotation to one of opposition drives, overdrives, torque converter stator release,as the turbine speed catches up with that of the automatic transmission drives and final differentialimpeller. Simultaneously, the vortex fluid circu- drives.lation will be declining so that the resultant torque Most overrun clutch devices take the form ofcapacity of the converter rapidly approaches zero. either the roller and wedge or sprag lock to engageUnder these conditions, with the accelerator pedal and disengage drive.fully down there is very little load to hold backthe engines speed so that it will tend to race orrun-away. Theoretically racing or run-away should 4.6.1 Overrun clutch with single diameter rollersoccur when both the impeller and turbine rotate at (Fig. 4.13)the same speed and the vortex circulation ceases, A roller clutch is comprised of an inner and outerbut due to the momentum losses caused by internal ring member and a series of cylindrical rollersfluid resistance, racing will tend to begin slightly spaced between them (see Fig. 4.13). Incorporatedbefore a 1:1 speed ratio (a typical value might be between the inner and outer members is a cage0.95:1). which positions the rollers and guides so that they roll up and down their ramps simultaneously. One4.5.8 Engine braking transmitted through of the members has a cylindrical surface concentricconverter or coupling on overrun with its axis, this is usually made the outer member.Torque converters are designed to maximize their The other member (inner one) has a separate wedgetorque multiplication from the impeller to the tur- ramp formed for each roller to react against. Thebine in the forward direction by adopting back- shape of these wedge ramps may be flat or curvedward swept rotating member circulating passage depending upon design. In operation each rollervanes. Unfortunately, in the reverse direction provides a line contact with both the outer internalwhen the turbine is made to drive the impeller on cylindrical track and the external wedge ramp tracktransmission overrun, the exit and entry vane guide of the inner member.angles of the members are unsuitable for hydro- When the input wedge member is rotatedkinetic energy transference, so that only a limited clockwise and the output cylindrical member isamount of engine braking torque can be absorbed prevented from rotating or rotates anticlockwiseby the converter except at high output overrun in the opposite direction, the rollers revolve andvehicle speeds. Conversely, a fluid coupling with climb up the wedge ramps, and thereby squeezeits flat radial vanes is able to transmit torque in themselves between the inner and outer membereither drive or overrun direction with equal effect. tracks. Eventually the elastic compressive and frictional forces created by the rollers against these tracks prevents further roller rotation.4.6 Overrun clutches Torque can now transfer from the input innerVarious names have been used for these mechan- member to the outer ring member by way of theseisms such as freewheel, one way clutch and overrun jammed (locked) rollers.clutch, each one signifying the nature of the device If the output outer member tries to rotate in theand is therefore equally appropriate. same direction but faster than the inner member, A freewheel device is a means whereby torque the rollers will tend to rotate and roll down theiris transmitted from one stationary or rotating ramps, thereby releasing (unlocking) the outermember to another member, provided that input member from that of the input drive. 109
  • 119. Fig. 4.13 Overrun freewheel single diameter roller type clutch4.6.2 Overrun clutch with triple diameter rollers If conditions change and the outer member(Fig. 4.14) overruns the inner member, the rollers will beThis is a modification of the single roller clutch in compelled to revolve in the opposite direction towhich the output outer member forms an internal when the drive was established towards the diver-cylindrical ring, whereas the input inner member ging end of the tracks. It thus releases the outerhas three identical external inclined plane profiles member and creates the freewheel phase.(see Fig. 4.14). Situated between the inner andouter tracks are groups of three different sized 4.6.3 Sprag overrun clutch (Fig. 4.15)rollers. An anchor block and energizing shoe is A very reliable, compact and large torque-carryingarranged, between each group of rollers; the blocks capacity overrun clutch is the sprag type clutch.are screwed to the inner member while the shoes This dispenses with the wedge ramps or inclined(with the assistance of the springs) push the rollers plane formed on the inner member which istogether and against their converging contact essential with roller type clutches (see Fig. 4.15).tracks. The inclined plane profile required to The sprag clutch consists of a pair of inner andmatch the different diameter rollers provides a outer ring members which have cylindrical externalvariable wedge angle for each size of roller. It is and internal track surfaces respectively. Interlinkingclaimed that the take-up load of each roller will the input and output members are circular rows ofbe progressive and spread more evenly than short struts known as sprags. Both ends of thewould be the case if all the rollers were of the sprags are semicircular with their radius of curva-same diameter. ture being offset to each other so that the sprags When the input inner ring takes up the drive, the appear lopsided. In addition a tapered waste isrollers revolve until they are wedged between the formed in their mid-region. Double cages are incor-inclined plane on the inner ring and the cylindrical porated between inner and outer members. Theseinternal track of the outer member. Consequently cages have rectangular slots formed to equallythe compressive load and the frictional force thus space and locate the sprags around the inner andcreated between the rollers and tracks locks solid outer tracks. During clutch engagement there willthe inner and outer members enabling them to be a slight shift between relative positions of thetransmit torque. two cages as the springs tilt, but the spacing will be 110
  • 120. Fig. 4.14 Overrun freewheel triple diameter roller type clutchaccurately kept. This ensures that each sprag turbine and stator members within the converter,equally contributes its share of wedge action so that there are more stages of conversionunder all operating conditions. In between the (Fig. 4.16).cages is a ribbon type spring which twists the sprags Consider the three stage torque converter. As showninto light contact with their respective track when in Fig. 4.17, it is comprised of one impeller, threethe clutch is in the overrun position. interlinked output turbines and two fixed stator When the inner ring member is rotated clockwise members.and the outer ring member is held stationary or is Tracing the conversion vortex circuit startingrotated anticlockwise, the spring tension lightly from the input rotating member (Fig. 4.18), fluidpresses the sprags against their track. This makes is pumped from the impeller P by centrifugal forcethe inner and outer members move in opposite to the two velocity components Vt and Vr, makingdirections. The sprags are thus forced to tilt anti- up the resultant velocity Vp which enters betweenclockwise, consequently wedging their inclined the first turbine blades T1 and so imparts some ofplanes hard against the tracks and thereby locking its hydrokinetic energy to the output. Fluid thenthe two drive and driven members together. passes with a velocity VT1 to the first fixed stator, As conditions change from drive to overrun and S1, where it is guided and redirected with a result-the outer member rotates faster than the inner one, ant velocity VS1 , made up from the radial andthe sprags will rotate clockwise and so release the tangential velocities Vr and Vt to the second set ofouter member: a freewheel condition is therefore turbine blades T2, so that momentum is given toestablished. this member. Fluid is now transferred from the exit of the second turbine T2 to the entrance of the second stator S2. Here the reaction of the curved4.7 Three stage hydrokinetic torque converter blades deflects the fluid towards the third turbine(Figs 4.16, 4.17 and 4.18) blades T3 which also absorb the fluids thrust.A disadvantage with the popular three element Finally the fluid completes its circulation cycle bytorque converter is that its stall torque ratio is again entering the impeller passages.only in the region of 2:1, which is insufficient The limitation of a multistage converter is thatfor some applications, but this torque multiplica- there are an increased number of entry and exittion can be doubled by increasing the number of junctions between various members which raise 111
  • 121. Fig. 4.15 (a and b) Overrun freewheel sprag type clutch the fluid flow resistance around the torus passages. Subsequently, efficiency drops off fairly rapidly with higher speed ratios compared to the three element converter (Fig. 4.16). 4.8 Polyphase hydrokinetic torque converter (Figs 4.19 and 4.20) The object of the polyphase converter is to extend the high efficiency speed range (Fig. 4.20) of the simple three element converter by altering the vane or blade shapes of one element. Normally the stator is chosen as the fluid entrance direction changes with increased turbine speed. To achieve this, the stator is divided into a number of separate parts, in this case three, each one being mounted on its own freewheel device built into its hub (Fig. 4.19). The turbine exit and linear velocities VE and VL produce an effective resultant velocity VR whichFig. 4.16 Characteristic performance curves of a three changes its direction of entry between stator bladesstage converter as the impeller and turbine relative speeds 112
  • 122. Fig. 4.17 Multistage (six element) torque converterapproach unity. It is this direction of fluid entering spin in the same direction as the input and outputbetween the stator blades which in phases releases elements. The two remaining fixed stators nowthe various stator members. form the optimum blade curvatures for high efficiency.Initial phaseUnder stall speed conditions, the fluid flow from Third phasethe turbine to the stator is such as to be directed With higher vehicle and turbine speeds, the fluidsonto the concave (rear) side of all three sections of resultant direction of entry to the two remainingthe divided stator blades, thus producing optimum held stators changes sufficiently to push from thestator reaction for maximum torque multiplication rear of the second set of stator blades S2. This sectionconditions. will now be released automatically to enable the third set of stator blades to operate with optimumSecond phase efficiency.As the turbine begins to rotate and the vehicle ispropelled forwards, the fluid changes its resultant Coupling phasedirection of entry to the stator blades so that it Towards unity speed ratio when the turbine speedimpinges against the rear convex side of the first has almost caught up with the impeller, the fluidstator blades S1. The reaction on this member is entering the third stator blades S3 will have alterednow reversed so that it is released and is able to its direction to such an extent that it releases this 113
  • 123. Fig. 4.18 Principle of the three stage torque converterlast fixed set of blades. Since there is no more The disadvantage of the early fall in efficiencyreaction torque, conversion ceases and the input with rising speed may be overcome by incorporatingand output elements act solely as a fluid coupling. a friction disc type clutch between the flywheel and converter which is hydraulically actuated by means of a servo piston (Fig. 4.21). This lock-up clutch is4.9 Torque converter with lock-up and gear designed to couple the flywheel and impellerchange friction clutches (Figs 4.21 and 4.22) assembly directly to the output turbine shaft eitherThe two major inherent limitations with the torque manually, at some output speed decided by theconverter drive are as follows: driver which would depend upon the vehicle load Firstly, the rapid efficiency decline once the and the road conditions or automatically, at a defi-relative impeller to turbine speed goes beyond nite input to output speed ratio normally in thethe design point, which implies higher input speeds region of the design point here where efficiency isfor a given output speed and increased fuel con- highest (Fig. 4.22).sumption. Secondly, the degree of fluid drag at idle To overcome the problem of fluid drag betweenspeed which would prevent gear changing with the input and output members of the torque con-constant mesh and synchromesh gearboxes. verter when working in conjunction with either 114
  • 124. Fig. 4.19 Principle of a polystage torque converter 115
  • 125. Fig. 4.22 Characteristic performance curves of a three element converter with lock-up clutchFig. 4.20 Relationship of speed ratio, torque ratio andefficiency for a polyphase stator torque converterFig. 4.21 Torque converter with lock-up and gear change function clutchesconstant mesh or synchromesh gearboxes, a its disengaged position, the gearbox input primaryconventional foot operated friction clutch can be shaft and the output main shaft may be unified,utilized between the converter and the gearbox. thereby enabling the gear ratio selected to beWhen the pedal is depressed and the clutch is in engaged both smoothly and silently. 116
  • 126. 5 Semi- and fully automatic transmission5.1 Automatic transmission considerations siderable and the large gear ratio steps of the con-Because it is difficult to achieve silent and smooth ventional transmission are reduced and smoothedgear ratio changes with a conventional constant out by the converters response between automaticmesh gear train, automatic transmissions com- gear shifts. Studying Fig. 5.1, whilst in first gear, themonly adopt some sort of epicyclic gear arrange- torque converter provides a maximum torque multi-ment, in which different gear ratios are selected by plication at stall pull away conditions which pro-the application of multiplate clutches and band gressively reduces with vehicle speed until thebrakes which either hold or couple various mem- converter coupling point is reached. At this point,bers of the gear train to produce the necessary the reaction member freewheels. With further speedspeed variations. The problem of a gradual torque increase, the converter changes to a simple fluidtake-up when moving away from a standstill has coupling so that torque multiplication ceases. Inalso been overcome with the introduction of a second gear the converter starts to operate nearertorque converter between the engine and transmis- the coupling point causing it to contribute far lesssion gearing so that engine to transmission slip is torque multiplication and in third and fourth gearautomatically reduced or increased according to the converter functions entirely beyond the couplingchanges in engine speed and road conditions. point as a fluid coupling. Consequently, there is noTorque converter performance characteristics have further torque multiplication.been discussed in Chapter 3. The actual speed at which gear ratio changes 5.2 Four speed and reverse longitudinallyoccur is provided by hydraulic pressure signals mounted automatic transmission mechanicalsupplied by the governor valve and a throttle power flow (Fig. 5.2)valve. The former senses vehicle speed whereas (Similar gear trains are adopted by some ZF,the latter senses engine load. Mercedes-Benz and Nissan transmissions) These pressure signals are directed to a hydraulic The epicyclic gear train is comprised of three pla-control block consisting of valves and pistons which netary gear sets, an overdrive gear set, a forwardcompute this information in terms of pressure gear set and a reverse gear set. Each gear set con-variations. The fluid pressure supplied by a pressure sists of an internally toothed outer annular ringpump then automatically directs fluid to the gear, a central externally toothed sun gear and avarious operating pistons causing their respective planet carrier which supports three intermediateclutch, clutches or band brakes to be applied. planet gears. The planet gears are spaced evenlyConsequently, gear upshifts and downshifts are between and around the outer annular gear andperformed independently of the driver and are so the central sun gear.made that they take into account the condition of The input to the planetary gear train is throughthe road, the available output of the engine and the a torque converter which has a lock-up clutch.requirements of the driver. Different parts of the gear train can be engaged or released by the application of three multiplate5.1.1 The torque converter (Fig. 5.1) clutches, two band brakes and one first gear oneThe torque converter provides a smooth automatic way roller clutch.drive take-up from a standstill and a torque multi- Table 5.1 simplifies the clutch and brake engage-plication in addition to that provided by the normal ment sequence for each gear ratio.mechanical gear transmission. The performance A list of key components and abbreviations usedcharacteristics of a hydrokinetic torque converter are as follows:incorporated between the engine and the gear train 1 Manual valve MVis shown in Fig. 5.1 for light throttle and full throttle 2 Vacuum throttle valve VTVmaximum output conditions over a vehicle speed 3 Governor valve GVrange. As can be seen, the initial torque mul- 4 Pressure regulating valve PRVtiplication when driving away from rest is con- 5 Torque converter TC 117
  • 127. Fig. 5.1 Torque multiplication and transmitted power performance relative to vehicle speed for a typical four speedautomatic transmission 6 1±2 shift valve (1±2)SV put shaft and pinion carrier. Torque is then split 7 2±3 shift valve (2±3)SV between the overdrive annular gear and the sun 8 3±4 shift valve (3±4)SV gear, both paths merging due to the engaged direct 9 Converter check valve CCV clutch. Therefore the overdrive pinion gears are10 Drive clutch DC prevented from rotating on their axes, causing the11 High and reverse multiplate clutch (H ‡ R)C overdrive gear set to revolve as a whole without any12 Forward clutch FC gear ratio reduction at this stage. Torque is then13 Overdrive band brake ODB conveyed from the overdrive annular gear to the14 Second gear band brake 2GB intermediate shaft where it passes through the15 Low and reverse multiplate brake (L ‡ R)B applied forward clutch plates to the annular gear16 First gear one way roller clutch OWC of the forward gear set. The clockwise rotation of17 Torque converter one way clutch OWCR the forward annular gear causes the forward planet18 Parking lock PL gears to rotate clockwise, driving the double sun gear counter clockwise. The forward planetary car-5.2.1 D drive range Ð first gear rier is attached to the output shaft so that the planet(Figs 5.3(a) and 5.4(a)) gears drive the sun gear instead of walking aroundWith the selector lever in D range, engine torque is the sun gear. This anticlockwise rotation of the suntransmitted to the overdrive pinion gears via the out- gear causes the reverse planet gears to rotate 118
  • 128. Fig. 5.2 Longitudinally mounted four speed automatic transmission layoutTable 5.1 Clutch and brake engagement sequence Drive High and Second gear Forward Overdrive Low and One way clutch reverse clutch band brake clutch brake reverse clutchRange DC (H ‡ R) C 2GB FC ODB brake (L ‡ R)B OWC RatioP and N ± ± ± ± ± ± ± ±First D Applied ± ± Applied ± ± Applied 2.4:1Second D Applied ± Applied Applied ± Applied ± 1.37:1Third D Applied Applied ± Applied ± ± ± 1:1Fourth D ± Applied ± Applied Applied ± ± 0.7:1Reverse R Applied Applied ± ± ± Applied ± 2.83:1 119
  • 129. Fig. 5.3 (a±e) Four speed and reverse automatic transmission for longitudinally mounted units 120
  • 130. Fig. 5.3 contdclockwise. With the one way roller clutch holding multiplate clutch and the forward multiplatethe reverse planet carrier, the reverse planetary gears clutch, both of which are applied. Subsequently,turn the reverse annular gear and output shaft clock- the high and reverse clutch will rotate the doublewise in a low speed ratio of approximately 2.46:1. sun gear clockwise and similarly the forward clutch will rotate the forward annular gear clockwise. This causes both external and internal gears on5.2.2 D drive range Ð second gear the forward gear set to revolve in the same direc-(Figs 5.3(b) and 5.4(b)) tion at similar speeds so that the bridging planetIn D range in second gear, both direct and forward gears become locked and the whole gear set there-clutches are engaged. At the same time the second fore revolves together as one. The output shaftgear band brake holds the double sun gear and drive via the reverse carrier therefore turns clock-reverse pinion carrier stationary. wise with no relative speed reduction to the input Engine torque is transmitted through the locked shaft, that is as a direct drive ratio 1:1.overdrive gear set similarly to first gear. It is thenconveyed through the applied forward clutch viaintermediate shaft to the forward annular gear. 5.2.4 D drive range Ð fourth or overdrive gearWith the double sun gear held by the applied second (Figs 5.3(d) and 5.4(d))gear band brake, the clockwise rotation of the In D range in fourth gear, the overdrive band brake,forward annular gear compels the pinion gears to the high and reverse clutch and the forward clutchrotate on their own axes and roll `walk around the are engaged. Under these conditions, torque is con-stationary sun gear in a clockwise direction. veyed from the input shaft to the overdrive carrier,Because the forward pinion gear pins are mounted causing the planet gears to rotate clockwise aroundon the pinion carrier, which is itself attached to the the held overdrive sun gear. As a result, the over-output shaft, the output shaft will be driven clock- drive annular gear will be forced to rotate clock-wise at a reduced speed ratio of approximately wise but at a higher speed than the input overdrive1.46. carrier. Torque is then transmitted via the inter- mediate shaft to the forward planetary gear set which are then locked together by the engagement5.2.3 D drive range Ð third or top gear of the high and reverse clutch and the forward(Figs 5.3(c) and 5.4(c)) clutch. Subsequently, the gear set is compelled toWith the selector lever in D range, hydraulic line rotate bodily as a rigid straight through drive. Thepressure will apply the direct clutch, high and torque then passes from the forward planet carrierreverse clutch and forward clutch. to the output shaft. Hence there is a gear ratio step As for first and second gear operating condi- up by the overdrive planetary gear set of roughlytions, the engine torque is transmitted through the 30%, that is, the output to input shaft gear ratio islocked overdrive gear set to the high and reverse about 0.7:1. 121
  • 131. Fig. 5.4 (a±e) Four speed and reverse epicycle gear set directional motion 122
  • 132. 5.2.5 R range Ð reverse gear the various valves and to energize the clutch and(Figs 5.3(e) and 5.4(e)) band servo pistons will vary under different work-With the selector lever in reverse position all three ing conditions. Therefore the fluid pressure gener-clutches and the low and reverse multiplate brake ated by the pump is unlikely to suit the manyare engaged. Subsequently, engine torque will be operating requirements. To overcome these diffi-transmitted from the input shaft through the locked culties, a pressure regulating valve is used whichoverdrive gear set through the locked forward gear automatically adjusts the pumps output pressureset via the intermediate shaft to the reverse sun gear to match the working requirements at any onein a clockwise direction. time. One of the functions of the pressure regulat- Because the reverse planet carrier is held by the ing valve is to raise the line pressure reaching thelow and reverse multiplate brake, the planet gears clutch and brake when the vehicle is driven hardare forced to rotate counterclockwise on their axes, with large throttle opening to prevent the frictionand in doing so compel the reverse annular gear to surfaces slipping. Conversely under light loads andalso rotate counterclockwise. As a result, the out- with a small throttle opening, a much lower lineput shaft, which is attached to the reverse annular pressure is adequate to clamp the friction plates orgear, rotates counterclockwise, that is, in the bands. By reducing the line pressure, fierce clutchreverse direction, to the input shaft at a reduction and brake engagements are eliminated which pro-ratio of approximately 2.18:1. motes smooth and gentle gear changes. Power con- sumption, which is needed to drive the hydraulic pump, is also reduced as actuating pressures are lowered. The pressure regulating valve is normally5.3 The fundamentals of a hydraulic control a spring-loaded spool type valve, that is, a plungersystem with one or more reduced diameter sectionsThe effective operation of an automatic transmis- along its length, positioned in a cylinder whichsion relies upon a hydraulic control circuit to has a number of passages intersecting the cylinderactuate the gear changes relative to the vehicles walls.road speed and acceleration pedal demands with When the engine speed, and correspondinglythe engine delivering power. Only a very small pump pressure, is low, fluid flows via the inletproportion of a transmissions operating time is port around the wasted section of the plunger andspent in performing gear changes. In fact, the out unrestricted along a passage leading to thehydraulic system is operational for less than 1% manual valve where it is distributed to the variousof the driving time. The transition time from one control valves and operating pistons. As the pumpgear ratio to the next takes roughly one second or pressure builds up with rising engine speed, lineless and therefore the hydraulic control valves must pressure is conveyed to the rear face of the plungerbe designed to direct fluid pressure to the appro- and will progressively move the plunger forwardpriate operating pistons which convert the fluid against the control spring, causing the middle landpressure into mechanical force and movement to to uncover an exhaust port which feeds back to theenergize the respective clutches and band brakes pumps intake. Hence as the pump output pressureinstantly and precisely. tends to rise, more fluid is passed back to the suc- An understanding of a basic hydraulic control tion intake of the pump. It therefore regulates thesystem can best be considered under the four output fluid pressure, known as line pressure,headings: according to the control spring stiffness. To enable 1 Pressure supply and regulating valves the line pressure to be varied to suit the operating 2 Speed and load sensing valves conditions, a throttle pressure is introduced to the 3 Gear shift valves spring end of the plunger which opposes the line 4 Clutch and brake coupling and hold devices pressure. Increasing the throttle pressure raises line pressure and vice versa.5.3.1 Pressure supply and regulating valve In addition to the main pressure regulating valve(Fig. 5.5) there is a secondary regulating valve which limitsThe essential input to the hydraulic control system the fluid flowing through to the torque converter.is fluid pressure generated by a pump and driven by Raising the torque converters fluid pressurethe engine. The pumps output pressure will increases its torque transmitting capacity which isincrease roughly in proportion to the engines desirable when driving in low gear or when thespeed. However, the pressure necessary to actuate engine is delivering its maximum torque. 123
  • 133. 5.3.2 Speed and load sensing valves opposing end forces acting on the spool valve end(Figs 5.5 and 5.6) faces.For gear changes to take place effectively at the Spring Throttle Governoroptimum engine and road speed, taking into ‡ ˆ load pressure load pressure loadaccount the drivers demands expressed in throttleopening, some means of sensing engine load and FS ‡ PT AT ˆ PG AGvehicle road speed must be provided. Engine out- But PA ˆ Fput torque is simply monitored by a throttle valve hence FS ‡ FT ˆ FGwhich is linked to the accelerator pedal, eitherdirectly or indirectly, via a vacuum diaphragm where FS ˆ Spring loadoperated linkage which senses the change in induc- FT ˆ Throttle pressure loadtion depression, which is a measure of the engine PT ˆ Throttle pressureload. The amount the accelerator pedal or mani- FG ˆ Governor pressure loadfold vacuum alters is relayed to the throttle valve AT ˆ CSA of plunger at throttle pressurewhich accordingly raises or lowers the output pres- end PG ˆ Governor pressuresure. This is then referred to as throttle pressure. AG ˆ CSA of plunger at governor pres- Road speed changes are measured by a centri- sure endfugal force-sensitive regulating valve which sensestransmission output shaft speed and transmits this Thus increasing or decreasing the spring stiffnessinformation in the form of a fluid pressure, referred or enlarging or reducing the diameter of the spoolto as governor pressure, which increases or decreases valve at one end considerably alters the conditionaccording to a corresponding variation in road speed. when the shift valve moves from one end toBoth throttle pressure and governor pressure are the other to redirect line pressure to and from thesignalled to each gear shift valve so that these may various clutch and brakes and so produce therespond to the external operating conditions necessary gear change.(i.e. engine torque developed and vehicle speed) Each shift valve control spring will have a parti-by permitting fluid pressure to be either applied or cular stiffness so that different governor pressures,released from the various clutch and brake actuating that is, road speeds, are required to cause eitherpiston chambers. a gear upshift or downshift for a given opposing throttle pressure. Conversely, different engine5.3.3 Gear shift valves (Fig. 5.5) power outputs will produce different throttle pres-Shift valves are of the spool plunger type, taking sures and will alter the governor pressure accord-the form of a cylindrical plunger reduced in dia- ingly when a particular gear shift occurs. Largemeter in one or more sections so as to divide its engine loads (high throttle pressure) will delaylength into a number of lands. When operating, gear upshifts whereas light engine load demandsthese valves shift from side to side and cover or (low throttle pressure) and high vehicle speedsuncover passages leading into the valve body so (high governor pressure) will produce earlythat different hydraulic circuits are switched on upshifts and prevent early downshift.and off under various operating conditions. To improve the quality of the time sequence The function of a shift valve is to direct the fluid of up or down gear shift, additional valves andpressure to the various clutch and brake servo pis- components are included to produce a smoothtons to effect gear changes when the appropriate transition from one gear to the next. Some ofload and speed conditions prevail. Shift valves are these extra devices are described in Section 5.6.controlled by line or throttle pressure, which isintroduced into the valve at the spring end, and 5.3.4 Clutch and brake coupling and hold devicesgovernor pressure, which is introduced directly to (Figs 5.5 and 2.16)the valve at the opposite end. Generally, the gover- Silent gear change synchronization is made possi-nor valve end is of a larger diameter than the spring ble by engaging or locking out various members ofend so that there will be a proportionally greater the epicyclic gear train gear sets with the enginesmovement response due to governor pressure vari- power being transmitted continuously. It thereforeation. Sometimes the shift valve plunger at the gov- requires a rapid and accurate gear change which isernor pressure end is referred to as the governor plug. achieved by utilizing multiplate clutches and band The position of the shift valve at any instant brakes. A gear up- or downshift therefore occursdepends upon the state of balance between the with the almost simultaneous energizing of one 124
  • 134. Fig. 5.5 (a and b) Basic multiplate clutch and band brake transmission hydraulic control systemclutch or brake and a corresponding de-energizing one shaft or member to another quickly andof another clutch or brake. smoothly. The rotating and fixed friction plates can be energized by an annular shaped, hydraulic-Multiplate clutch (Figs 2.16 and 5.5) Wet multi- ally operated piston either directly or indirectly byplate type clutches are very compact for their tor- a dished washer which acts also as a lever to multi-que transmitting and heat dissipating capacity. ply the operating clamping load. Return springsThey are used to lock any two members of a plan- are used to separate the pairs of rubbing facesetary gear set together or to transfer drive from when the fluid pressure is released. Wear and 125
  • 135. adjustment of the friction plate pack is automatic- With rising output shaft speed, the centrifugalally compensated by the piston being free to move force acting through the primary valve is sufficientfurther forward (see Chapter 2, Fig. 2.16). to overcome the hydraulic line pressure, which is acting against the shouldered groove face area andBand brake (Fig. 5.5) This form of brake consists will therefore progressively move outwards as theof a friction band encompassing an external drum rotational speed increases until the valve bordersso that when the brake is applied the band con- on an end stop. The opening of the governor valvetracts, thereby wrapping itself tightly around the outlet passage now allows fluid to flow out fromdrum until the drum holds. The application of the the governor, where it is then directed to the largeband is achieved through a double acting stepped diameter end of the shift valve. This output pres-servo cylinder and piston. Fluid line pressure is sure is known as governor pressure. With evenintroduced to the small diameter end of the piston higher rotational output shaft speed (vehicleto energize the band brake. To release the band, speed), greater centrifugal force will be imposedsimilar line pressure is directed to the spring cham- on the secondary valve until it is able to overcomeber side of the cylinder. Band release is obtained due the much larger hydraulic inward load imposed onto the larger piston area side producing a greater the large shoulder of this valve. The secondaryforce to free the band. This method of applying and valve will start to move out from the centre ofreleasing the band enables a more prolonged and rotation, uncovering the secondary valve outletcontrollable energizing and de-energizing action to passage so that increased governor pressure passesbe achieved. This class of brake is capable of absorb- to the shift valve.ing large torque reactions without occupying very This two stage governor valve action enables themuch space, which makes the band brake particu- governor to be more sensitive at the very lowlarly suitable for low gear high torque output gear speeds but not oversensitive at the higher speedssets. Band wear slackness can be taken up by exter- (Fig. 5.5(c)). Sensitivity refers to the amount ofnally adjusting the anchor screw. fluid pressure increase or decrease for a unit change in rotational speed. If there is a large increase or decrease in governor pressure per unit charge in5.4 Basic principle of a hydraulically controlled speed, then the governor is sensitive. If there isgearshift (Fig. 5.5) very little variation in governor pressure with a changeSelecting the drive D range positions the manual in rotational speed (i.e. vehicle speed), then thevalve spool so that line pressure from the pressure governor is insensitive and therefore not suitableregulator valve passes through to the shift valve, for signalling speed changes to the hydraulicthrottle valve and governor valve (Fig. 5.5(a)). control systems. Throttle pressure will be introduced to the spring The reason a single stage governor would not per-end of the shift valve via the throttle valve. Depress- form satisfactorily over the entire output shift speeding the accelerator pedal allows the spool valve to range is due to the centrifugal force square law: atmove outwards. This increases the valve opening so low speeds the build-up in centrifugal force forthat a high throttle pressure will be delivered to the a small increase in rotational speed is very small,shift valve. Conversely, depressing the accelerator whereas at higher speeds only a small rise in speedpedal partially restricts the flow of fluid and there- produces a considerable increase in centrifugalfore reduces the throttle pressure reaching the shift force. If the governor has the correct sensitivity atvalve (Fig. 5.5(b)). high speed it would be insensitive at low speed or if it At the same time, line pressure enters the gover- has the desired sensitivity at low speed it would benor valve, flows between the wasted region of both far too responsive to governor pressure changes inprimary and secondary spool valves and reacts the higher speed range.against the difference in the annular adjacent face Once the governor pressure end load (PG Â AG )areas of each spool valve. Both valves are forced equals the spring and throttle pressure loadinwards, covering up the two exits from the gover- (FS ‡ PT Â AT ) with rising vehicle speed, anynor valve housing. As the vehicle moves forwards, further speed increase will push the shift valvethe rotation of the governor causes a centrifugal plunger towards the spring end to the positionforce to act through the mass of each governor shown in Fig. 5.5(a). The fluid on the applied sidevalve so that it tends to draw the valve spools out- of the band brake servo piston will now exhaustwards in opposition to the hydraulic pressure (drain) through the shift valve to the inlet side ofwhich is pushing each valve inwards (Fig. 5.5(a)). the oil pump. Simultaneously, line pressure from 126
  • 136. the manual valve is directed via the shift valve to shift valve spring end is subjected to line pressureboth the release side of the band servo piston and from the manual valve.to the multiplate clutch piston which then energizes Whilst the transmission is in drive first gear thethe friction plates. one way clutch will engage, so preventing the Supply fluid to the spring side of the servo piston reverse planetary carrier from rotating (not shown(known as the release side), provides a more pro- in hydraulic system).gressive and controllable transition from one gearchange to another which is not possible when 5.5.2 Second gear (Fig. 5.7)relying only on the return spring. With the manual valve still in D, drive position, When the vehicles speed is reduced or the throt- hydraulic conditions will be similar to first gear,tle pressure is raised sufficiently, the shift valve that is, the overdrive and forward clutches areplunger will move to the governor pressure end of engaged, except that rising vehicle speed increasesthe valve (Fig. 5.5(a)). The line pressure trans- the governor pressure sufficiently to push the 1±2mitted to the shift valve is immediately blocked shift valve against both spring and line pressure endand both the multiplate clutch and the band loads. As a result, the 1±2 shift valve middle landbrake hydraulic feed passages are released of fluid uncovers the line pressure supply passage feedingpressure by the middle plunger land uncovering the from the manual valve. Line pressure is now direc-exhaust part. Simultaneously, as the same middle ted to the second gear band servo on the appliedland covers the right hand exhaust port and side, energizing the second gear brake and causinguncovers the line pressure passage feeding from both the forward and reverse sun gears to hold.the manual valve, fluid will flow to the applied If there is a reduction in vehicle speed or if theside of the band servo piston, causing the band to engine load is increased sufficiently, the resultingcontract and so energize the brake. imbalance between the spring and throttle pressure load as opposed to governor pressure acting on the 1±2 shift valve at opposite ends causes the shift5.5 Basic four speed hydraulic control system valve to move against the governor pressure. Con-A simplified hydraulic control system for a four sequently the hydraulic circuitry will switch backspeed automatic transmission will now be examined to first gear conditions, causing the transmissionfor the reader to obtain an appreciation of the overall to shift down from second to first gear again.function of the hydraulic computer (control) system.5.5.1 First gear (Fig. 5.6) 5.5.3 Third gear (Fig. 5.8)With the manual valve in D, drive position, fluid is At even higher road speeds in D, drive position, thedelivered from the oil pump to the pressure regu- governor pressure will have risen to a point where itlating valve. It then divides, some being delivered is able to overcome the spring and throttle pressureto the torque converter, the remainder passing out load of the 2±3 shift valve. This causes the spoolto the manual valve as regulating pressure (more valve to shift over so that the line pressure passagecommonly known as line pressure). Line pressure feed from the manual valve is uncovered. Linefrom the manual valve is then channelled to the pressure will now flow through the 2±3 shift valveforward clutch, which is energized, and to the where it is directed to the high and reverse clutch tooverdrive band servo on the applied side. At the energize the respective fixed and rotating frictionsame time, line pressure from the pressure regulat- plates. At the same time, line pressure passes toing valve passes through the 3±4 shift valve where it the second gear band servo on the release side tois directed to energize the drive clutch and to the disengage the band. Consequently both overdrivereleased side of the overdrive band servo, thus pre- and forward planetary gear sets lock-up, permit-venting the engagement of the band. Line pressure ting the input drive from the torque converter to beis also directed to both the governor valve and to transmitted directly through to the transmissionsthe vacuum throttle valve. The reduced pressure output shaft.output from the governor valve which is known The actual vehicle speed at which the 2±3 shiftas governor pressure is directed to the end faces valve switches over will be influenced by the throt-of each of the three shaft valves, whereas the output tle opening (throttle pressure). A low throttle pres-pressure from the throttle valve, known as throttle sure will cause an early gear upshift whereas a largepressure, is conveyed to the spring end of the 2±3 engine load (high throttle pressure) will raise theand 3±4 shift valves. On the other hand, the 1±2 upshift speed. 127
  • 137. Fig. 5.6 Hydraulic control system (D) range first gear5.5.4 Fourth gear (Fig. 5.9) the line pressure exhausts from the release side ofWith still higher road speeds in D, drive position, the overdrive band servo which then permits thethe increased governor pressure will actuate the 3±4 band to be energized. At the same time the driveshift valve, forcing it to shift across so that it covers clutch will be de-energized because of the collapseup the line pressure supply passage and at the same of line pressure as it is released through the 3±4time uncovers the exhaust or drain port. As a result, shift valve exhaust port. 128
  • 138. Fig. 5.7 Hydraulic control system (D) range second gear Under these operating conditions the overdrive 5.5.5 Reverse gear (Fig. 5.10)shaft planetary gear set reduces the intermediate With the manual valve in R, reverse position, lineshift speed and, since the forward clutch is in a pressure from the manual valve is directed via thestate of lock-up only, this speed step up is trans- 2±3 shift valve to the release side of the second gearmitted through to the output shaft. band servo, causing the band to disengage. At the 129
  • 139. Fig. 5.8 Hydraulic control system (D) range third gearsame time line pressure from the same supply pas- is directed via the 3±4 shift valve to the release sidesage engages the high and reverse clutch. The man- of the overdrive brake servo to disengage the bandual valve also supplies line pressure to the low and and to the drive clutch piston to engage the frictionreverse band brake via the 1±2 shift valve to hold plates. Note that both band brake servos on thethe reverse planetary carrier. In addition, line pres- applied sides have been exhausted of line pressuresure from the pressure regulating valve output side and so has the forward clutch piston chamber. 130
  • 140. Fig. 5.9 Hydraulic control system (D) range fourth gear5.5.6 Lock-up torque converter (Fig. 5.11) between the input pump impeller and the turbine output shaft. The benefits of this lock-up can onlyIntroduction To overcome the inherent relative be realised if the torque converter is allowed toslip which always occurs between the torque con- operate when light torque demands are made onverters pump impeller and the turbine runner, even the engine and only when the converter is operatingdriving at moderate speeds under light load condi- above its torque multiplication range that istions, a lock-up friction clutch may be incorporated beyond the coupling point. Consequently, converter 131
  • 141. Fig. 5.10 Hydraulic control system (D) reverse gearlock-up is only permitted to be implemented when the vehicle. Due to this net gain in power output,the transmission is in either third or fourth gear. fuel wastage will be reduced.The advantages of bypassing the power transferthrough the circulating fluid and instead trans- Lock-up clutch description The lock-up clutchmitting the engines output directly to the trans- consists of a sliding drive plate which performsmission input shaft eliminates drive slippage, two functions; firstly to provide the friction coup-thereby increasing the power actually propelling ling device and secondly to act as a hydraulic con- 132