《外文资料--Shift dynamics and control of dual-clutchtransmissions》由会员分享，可在线阅读，更多相关《外文资料--Shift dynamics and control of dual-clutchtransmissions（15页珍藏版）》请在金锄头文库上搜索。
1、Shift dynamics and control of dual-clutch transmissionsManish Kulkarni, Taehyun Shim, Yi Zhang*Department of Mechanical Engineering, University of Michigan-Dearborn, Dearborn MI 48128, United StatesReceived 4 October 2005; accepted 1 March 2006Available online 18 May 2006AbstractShifts in a dual-clu
2、tch transmission (DCT) are realized by torque transfer from one clutch to another without tractioninterruption due to the controlled slippage of the clutches. The timing of engagement and disengagement of the twoclutches is critical for achieving a smooth shift without engine flare and clutch tie-up
3、. This paper presents an analyticalmodel for the simulation, analysis and control of shift dynamics for DCT vehicles. A dynamic model and the control logicfor the integrated vehicle have been developed using Matlab/Simulink as the simulation platform. The model has been usedto study the variation in
4、 output torque in response to different clutch pressure profiles during shifts. Optimized clutch pres-sure profiles have been created for the best possible shift quality based on model simulation. As a numerical example, themodel is used for a DCT vehicle to simulate the wide-open throttle performan
5、ce. Vehicle launch and shift process are bothsimulated to assess transmission shift quality and validate the effectiveness of the shift control.? 2006 Elsevier Ltd. All rights reserved.Keywords: Dual-clutch transmission; Automatic transmissions1. IntroductionThere has been a clear trend in the autom
6、otive industry in recent years towards increased ride comfort andfuel efficiency. As the power transmission unit, transmissions play an important role in vehicle performanceand fuel economy. There are currently several types of transmissions and the associated technologies that offerdifferent perfor
7、mance priorities when fit into a vehicle 1. Manual transmissions have an overall efficiency of96.2%, which is the highest efficiency value for any type of transmission. Current production automatics havebeen improved to provide an efficiency of not more than 86.3%. Belt type CVTs have an overall eff
8、iciency of84.6%,however,themajoradvantageofCVTisthatitallowstheenginetooperateinthemostfuel-efficientman-ner 2. Automated manual transmissions have the same efficiency of manual transmissions and offer operationconvenience similar to conventional automatic transmissions. There exist two technically
9、feasible designs forautomated lay-shaft gearing transmissions. One uses a single clutch and is basically a manual transmission withanadded-oncontrolunitthatautomatestheclutchandshiftoperations.Inthisdesign,thereisaninterruptionof0094-114X/$ - see front matter ? 2006 Elsevier Ltd. All rights reserved
10、.doi:10.1016/j.mechmachtheory.2006.03.002*Corresponding author. Tel.: +1 313 593 5539.E-mail address: andingumich.edu (Y. Zhang).Mechanism and Machine Theory 42 (2007) Theorytorqueduringagearchangesincetheengineiscut-offbytheclutchduringshift.Thistorqueinterruptionleadstounanticipated passenger fel
11、t jerks due to vehicle acceleration discontinuity and is highly uncharacteristic of con-ventionalautomatictransmissions. Theotherdesignusesadual-clutch systembetweentheengineandthetrans-mission and overcomes the shortcomings of the single clutch version 3. The two clutches are engagedalternatively i
12、n different speeds and power transmission continues during a shift through the control of clutchslippage. A shift process involves the engagement of the oncoming clutch and the release of the offgoing clutch.Thisresultsinshiftcharacteristicsthataretypicalofclutch-to-clutchshiftscommonlyseeninconvent
13、ionalauto-matic transmissions.It is a common practice in the automotive industry to use analytical models for the prediction and assess-ment of new types of powertrain systems. A great deal of research efforts have been focused on the modelingand control of vehicle transmission, such as conventional
14、 automatic transmissions 46, continuously variabletransmissions 7,8, and hybrid systems 912. Various formulation methods and programming techniqueshave been used in these researches to model the dynamics of vehicle powertrain and simulate the performanceof transmission control. Typically, the equati
15、ons of motion are first derived separately at the component leveland then integrated into the overall vehicle system. The integrated system models are either implemented ingenerically developed codes or in object oriented programming environment. As compared with the maturityof technologies for conv
16、entional automatic transmissions, the modeling and control of dual-clutch transmis-sions is still a new area and technologies associated with the DCT design and control are still at the early stageof development.This paper presents an analytical model for the simulation, analysis and control of the
17、launch and shift pro-cesses of DCT vehicles. The research work is concentrated on modeling the vehicle dynamics during shiftingand establishing a simulation tool for the analysis and optimization of shift control using clutch pressure pro-files as the control signals. Matlab/Simulink is used as a si
18、mulation platform to develop the dynamic modeland control logic for the integrated vehicle powertrain model. The simulation model is used to study the var-iation in output torque in response to different clutch pressure profiles. Optimized clutch pressure profiles areobtained to achieve the best pos
19、sible shift quality based on model simulation. As a numerical example, themodel is used for a vehicle equipped with a dual-clutch transmission to simulate the wide-open throttle per-formance. Vehicle launch and shift process are also simulated to assess transmission shift quality and validatethe eff
20、ectiveness of the transmission control strategy.2. DCT structure and model descriptionThe dual-clutch transmission is shown schematically in Fig. 1. The transmission has six forward speeds anda reverse speed. The transmission input shafts are designed as quill-shafts, with one solid shaft positioned
21、inside another hollow shaft. The solid shaft carries on it the second, fourth, sixth and the reverse gear, whilethe hollow shaft carries the first, third, fifth gear. Clutch 1 (CL1) connects all the odd gears and Clutch 2 (CL2)connects all the even gears to the input. Synchronizers are positioned be
22、tween two gears similar to conven-tional manual transmissions. When in a particular gear, the respective clutch and synchronizer are engagedand power flows from the engine through the clutch and synchronizer to the output shaft. The other clutchremains open and the remaining gears freewheel. To chan
23、ge a gear the offgoing clutch is slowly releasedand the oncoming clutch simultaneously engaged. It is this feature of DCT that allows for uninterruptedtorque transfer even during a gearshift.The powertrain is modeled as an integrated multi-degree-of-freedom system in which each element is alumped ma
24、ss model, as shown in Fig. 2. The engine is modeled as a mass inertia and accepts throttle angleas the input to produce a mean torque at the crankshaft. The drivetrain consists of component models ofthe dual clutches, layshaft transmission gearsets, driveshaft, differential and a vehicle model that
25、incorporatesthe road load and aerodynamic drag. All gears and synchronizers are modeled as non-compliant elements andare represented as mass inertias. The input and output shafts are modeled as compliances and are representedby torsion spring-damper assemblies. Clutches and synchronizers are modeled
26、 as friction elements withhydraulic pressure as control signals. The output of the simulation model is the torque at the axle and vehiclespeed. The relative angular velocity across the two clutches is monitored for decision making in the controllogic. The following assumptions are made in the develo
27、pment of the model:M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182169 The engine output torque is interpolated in terms of the throttle angle and RPM from the engine map. Gears have no backlash. All the mechanical losses are modeled as a part of the vehicle drag. Delays due to hyd
28、raulic actuation system are not considered. Clutches are modeled as Coulomb friction elements. Temperature effects of the drivetrain are neglected.CL2CL1SYN5RSYN6SYN13SYN244231Input ShaftOutput6R5Final Drive Pinion 2Final Drive Pinion 1Intermediate Shaft 1Intermediate Shaft 2SYN5RSYN6SYN13SYN24Input
29、 ShaftOutput6R5SYN5RSYN6SYN13SYN24Input ShaftOutput6R5Final Drive Pinion 2Final Drive Pinion 1Intermediate Shaft 1Intermediate Shaft 2Fig. 1. DCT stick diagram.4Output Shaft4RCL1CL2131265Input ShaftIi/pIeCL1312EngineKmCmImepi /K1C1IhIsI1sI2imimhK2C2I3aiowFig. 2. DCT dynamic model.170M. Kulkarni et a
30、l. / Mechanism and Machine Theory 42 (2007) 1681823. Component modelsThe DCT system consists of complicated component level subsystems such as the engine, clutches, andvehicle road load. These component models are described in the following.3.1. EngineThe mechanics of the engine assembly is modeled
31、with a two degree-of-freedom system: one is the rota-tional inertia of the moving parts and the other is the inertia of the engine and transmission supported on theirmounts. As an assumption, the engine is modeled as a mean value torque generator that does not include theengine transients. Engine ou
32、tput torque is interpolated corresponding to engine speed and throttle positionfrom an engine map modeled as a look-up table. For each throttle opening (TA), the engine torque from idleto redline is a function of engine angular velocity (xe) such that,Te fTA;xe1where Teis the engine torque, xeis the
33、 engine angular velocity and TAis throttle angle.3.2. ClutchThe clutches in a DCT function as primary gear changing elements in addition to their normal purpose ofdisconnecting the engine from the driving unit. Based on the clutch geometry and friction characteristics, theclutch torque is calculated
34、 as follows:TCL lFnn23? ?R3o? R3iR2o? R2i?2where l is coefficient of friction that can be formulated as a function of clutch slip, Fnis normal force on clutchface that depends on the apply pressure, Rois outside radius of friction disc, Riis inside radius of friction disc,n is the number of friction
35、 discs. For simplicity, the clutch torque is modeled as a function of the relative angu-lar velocity and the hydraulic pressure in the clutch piston in terms of a look-up table. The clutch has threeoperation states and the torque transmitted in each state is described by the following equation:TCLCl
36、DxPappslippingTclosed0open8:3where C is the constant reflecting the clutch dimension, Dx is the relative angular velocity between the clutchinput and output ends and Pappis the pressure in the clutch piston. T is the torque applied on the clutch whenthe vehicle is running in particular speed with th
37、e clutch fully closed. As shown in the equation above, theclutch torque during a shift depends only on the clutch apply pressure and the coefficient of friction for a givenclutch. The apply pressure is the control signal and the coefficient of friction depends on the clutch slip, i.e., therelative v
38、elocity of the two clutch sides.3.3. SynchronizerIn the dual clutch design, gears of the next speed are pre-engaged while the vehicle is running in a currentspeed. Gear engagement and synchronization do not occur simultaneously with the shifting process. The twogears to be synchronized are freewheel
39、ing on the shafts and hence, the synchronizer is modeled as a simplefriction element. The synchronizer torque is proportional to the friction force as shown below:Tsyn flF;Dx134where Tsynis the synchronizing torque, F is the normal force acting on the friction cone, l is the coefficient offriction o
40、f the synchronizer ring and Dx13is the angular velocity difference across a synchronizer, between thefirst and third gears as indicated by the subscripts.M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 1681821713.4. Vehicle road loadThe vehicle road load model accounts for the rolling re
41、sistance, aerodynamic resistance and transmissionlosses as formulated in the following:RA CdV2 lfW W sinh5TRL RARd6where Cdis the air drag coefficient that depends on the body style and dimension, W is the vehicle weight, lfisthe rolling coefficient, h is the grade angle and V is the vehicle velocit
42、y. TRLis the road load torque on thewheel and Rdis the wheel radius.4. System modelThe dynamic model for a dual-clutch transmission is illustrated in Fig. 2. The engine mounts, input and out-put shafts are modeled as spring-dampers to account for the compliance of these components. Gear shafts aremo
43、deled as coupled lumped masses. There are two sets of dynamic equations for the model. The model followsone set of equations of motion when the vehicle is running in a particular gear and the transmission provides amechanicallinkbetweentheengineandthewheels.Duringagearshiftitfollowsanothersetofequat
44、ionsasthereis no direct link between the engine and the wheels as the system is in a dynamic shift state. Presented below arethe generic dynamic equations for model operation in any given gear and also during an upshift. The equationsfor aspecific gear anda specific gear change can be written by cho
45、osing the correct gears and considering Clutch1 open in second, fourth and sixth gears and Clutch 2 open in first, third and fifth gear modes.4.1. Operation in a particular gear4.1.1. Engine and input shaftImdxmdt To? Tm7Tm Kmhm Cmxm8Iedxedt Te? Ti=p9Te TeTA;xe10Ti=p K1he? hi=p C1xe? xi=p11where Tmi
46、s the torque at the engine mount, Ti/pis the input shaft torque, Kmand K1are the stiffness of theengine mounts and input shaft respectively, Ieand Imare the mass moments of inertia of the engine outputshaft and the engine-transmission block, C1and Cmare the damping coefficient of the input shaft and
47、 theengine mount, hm, heand hi/pare the angular displacements of the engine mount, engine crankshaft and inputshaft respectively, xm, xeand xi/pare the angular velocities of the engine mounts, engine crankshaft and inputshaft, respectively. Tois the reaction torque applied on the mount of the engine
48、-transmission assembly, whosemagnitude is the same as the output torque.4.1.2. Transmission and output shaftIi=pdxi=pdt Ti=p? TCL1 TCL212TCL1;TCL2Timengaged0open?13Ieqdximdt Timit?Toia14172M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182To K2ho? hw C2xo? xw15To? TRLRdWgdvdt16dvdt R
49、ddxwdt17where K2and C2are the stiffness and damping coefficients of the output shaft, respectively. hoand hware theangular displacements of the output shaft and wheel, respectively. Ii/pis the mass moment of inertia of theinput shaft. TCL1and TCL2are the torques carried by the dual clutches. Timis t
50、he torque on intermediate shaft1 for the first, second, third and fourth speeds and on intermediate shaft 2 for the fifth, sixth, and reversespeeds as shown in Fig. 1. ximis the angular velocity of the intermediate shaft. xoand xware the angularvelocities of the output shaft and the wheel, respectiv
51、ely. Tois the axle output torque. Ieqis the equivalent massmoment of inertia on the intermediate shaft that includes all rotational masses in the power flow path. itis theexisting transmission ratio. For example, Ieqfor the third speed and the sixth speed is calculated respectively asfollows:Ieq I1
52、Ihi23for the third speed18Ieq I2 Isi26for the sixth speed19where i3and i6are the gear ratios of the third and sixth speed, respectively. I1and I2are the mass moments ofinertia of the two intermediate shafts assemblies. Ihand Isare the mass moments of inertia of the hollow andsolid shaft assemblies.4
53、.1.3. Equations of motion during a shiftDuring a shift, the clutch is no longer closed. For an upshift, Clutch 1 is being released and at the same timeClutch 2 is being applied such that both of them are slipping during the shift period. Hence, the engine torqueis not directly transmitted to the int
54、ermediate shaft by a closed clutch. It is the friction torque in the clutch thatis now been transmitted. Since there is no mechanical link between the engine and wheel, the powertraindynamics are governed by another set of the equations represented as follows.4.1.4. EngineIedxedt Te? Ti=p?20Ti=p K1h
55、e? hi=p? C1xe? xi=p?21xeZTe? Ti=pIedt x0e22Te fxe23where x0eis the engine angular velocity at the beginning of the shift. The engine is a function of its angularvelocity alone since the throttle angle is controlled at a fixed value during shift.4.1.5. TransmissionIi=pdxi=pdt Ti=p? TCL1 TCL224TCL1 f1
56、Papp1;DxCL125TCL2 f2Papp2;DxCL226DxCL1 xe? xh27DxCL2 xe? xs28ishiftxexwia29M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182173Ieqdximdt TCL1iodd TCL2ieven ?Toia30To K2ho? hw C2xo? xw31To? TRLRdWgdvdt32where xhand xsare the angular velocities of the hollow and solid shafts, respecti
57、vely. ioddand ievenare thegear ratios of the current and next speeds involved in the shift. ishiftis the gear ratio during a shift and is afunction of shift time. Papp1and Papp2are the clutch pressure signals that serve as the shift control input.The equivalent mass moment of inertia Ieqin Eq. (30)
58、depends on the related shafts and gear ratios involvedin the speed change. For instance, Ieqis calculated for a 12 shift as follows:Ieq I2 Ihi21 Isi22335. Shift control logicThe dynamic equations derived above are incorporated in the simulation model in Simulink. Each equationis treated as a matrix
59、of elements and has one quantity that is derived from known values. Sometimes externalinputs are used to calculate the unknown and in some cases an algebraic loop is formed, where the output of afuture event acts as an input to its driver. Gear-shifting strategy can be defined as the logical combina
60、tion of aseries of interdependent events, which dictate the accurate instance of a gear upshift or downshift as well as theoptimum duration of a gear change. Fig. 3 shows the decision block used to build the control logic for theDCT modeling. A shift controller receives input from various sensors li
61、ke the throttle position, current vehiclespeed, current gear position, clutch schedule and engine speed. Based on these inputs the controller makes aGEAR POSITIONDECISION BLOCK FOR CLUTCH ACTIVATIONNEUTRAL? INGEAR? CHANGE GEAR? 1, 3, 5? ODDTOEVEN?EVENTOODD? CL1 ONCOMING (RISING CURVE) CL2 OFFGOING (
62、FALLING CURVE) CL1 OFFGOING (FALLING CURVE) CL2 ONCOMING (RISING CURVE) CL1ACTIVECL2ACTIVE2, 4, 6? YES YES NOCL1 = 0 CL2= 0YES Fig. 3. Decision logic block.174M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182decision of an upshift, downshift or gear position holding. A gear shift sc
63、hedule used in this paper is illustratedin Fig. 4. At every simulation time step, the threshold shift speeds on the six shift curves, 12, 21, 23, 32, upto 56 and 65 are calculated according to the current throttle angle. To implement shift logic for the dualclutch operation, the following three crit
64、eria need to be addressed:1. To detect the time the shift is initiated.2. To create a specific rate of engagement and disengagement of the clutches.3. To determine the completion of shift.The continuous dynamics of the gearshift event is modeled in Stateflow/Simulink. A pulse generator isplaced in t
65、he Stateflow subsystem such that it generates a spike every time the gear state changes to a newstate. To determine the end of shift, the difference in angular velocity across the oncoming clutch is monitored.Since this clutch is open before the start of shift, there is a velocity differential acros
66、s the clutch. When theclutch locks-up, both sides rotate at the same speed such that the relative angular velocity across the clutchis zero. In this paper each clutch is modeled as a look-up table consisting of a rising pressure curve and a fall-ing pressure curve for the clutch to function as oncom
67、ing clutch and offgoing clutch, respectively. A counter isdesigned such that it generates an input ramp curve for the clutch pressure look-up tables once the gear changepulse is detected. The counter always compares the angular velocity across the oncoming clutch and creates alock-up signal until th
68、e relative velocity becomes zero. The rate of engagement of the clutch thus becomes afunction of how quickly the angular velocity drops to zero. For a quick shift, engagement curve will be steeperthan that for a slow shift. A downshift is carried out in exactly the same manner, except the clutch pre
69、ssurecurves are reversed and so are the oncoming and offgoing clutches.The ratio change during a shift can be qualitatively explained using a 12 upshift as an example. In the firstgear operation Clutch 1 is in the completely engaged condition and Clutch 2 is disengaged. The second gear isalready eng
70、aged by the 24 synchronizer. In the first gear the output torque of the transmission is a function ofthe engine torque, engine speed, gear ratio and the driveline inertias. If the damping effects of the input shaftare not considered for simplicity, the transmission output torque in the first gear is
71、ToTei1? I3 I2i22 I1i21 _ xeia3401020304050607080901000102030405060708090100110Vehicle Speed (mph)Throttle Position (%)2-11-23-22-34-33-45-44-56-55-6Fig. 4. Gear shift schedule.M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182175The output angular velocity is related to the engine an
72、gular velocity:xoxei?1ia35For a fixed throttle position, the engine torque in a particular gear is calculated based on the engine speed anda look-up table which has the engine map incorporated in it. This is possible since the clutch is connected andthere exists a mechanical link between the engine
73、and the wheels. However, during a gear shift, both theclutches are slipping and are partially engaged. The engine speed becomes a function of the clutch engagementschedule and hence the engine torque also varies as the speed. The engine speed is calculated as follows:xeRTe? TCL1 TCL2?dtIe36Two impor
74、tant observations can be made from this equation:(1) Engine angular velocity is not constant and hence the gear ratio follows a waterfall like curve based onthe engine speed.ii?2xex?oia37(2) The torque developed by the off-going and on-coming clutch is deducted from the engine torque. Hence,the torq
75、ue reduction during a gear change depends on the timing and engagement characteristics of thetwo clutches. The most efficient condition would be when the two clutches can smoothly take up the tor-que differential between the initial gear and the target gear output torque.6. Simulation results and an
76、alysisThe developed vehicle model is used for the shift control of a midsize passenger car equipped with a dual-clutch transmission. The vehicle data involved in the model are shown in Table 1.Table 1Vehicle dataParameterValueEngine size3.0L V6Vehicle mass3000 kgTire radius0.312 mEffective tire roll
77、ing radius0.308 mDifferential gears moment of inertia0.0047 kg m2Final drive gear ratio3.07Final drive gears moment of inertia0.0002 kg m2Transmission gear ratiosFirst 3.14, second 1.98, third 1.37, fourth 1.00,fifth 0.76, sixth 0.6, Rev 4.04Odd gears moment of inertia0.0023 kg m2Even gears moment o
78、f inertia0.0009 kg m2Engine mount damping constant135 N m s/radEngine mount spring constant16269 N m/radEngine moment of inertia2.7 kg m2Input shaft moment of inertia0.004 kg m2Input shaft spring constant4920 N m/radInput shaft damping constant13.6 N m s/radOutput shaft spring constant16300 N m/radO
79、utput shaft spring constant31.1 N m s/radSolid shaft moment of inertia0.002 kg m2Hollow shaft moment of inertia0.001 kg m2Intermediate shaft moment of inertia0.008 kg m2176M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182The output torque profile of the dual-clutch transmission for
80、a typical upshift is shown in Fig. 5. As seen inthe nature of the output torque, clutch-to-clutch shifts consist of two phases, a torque phase and an inertiaphase. When a gear shift signal is initiated, the engine torque is transferred through two clutches simulta-neously. During this time neither o
81、f the two clutches is completely engaged. The transmission output torquedrops to the lowest value during the torque phase due to shift transients. The speed ratio is not changed yetduring the torque phase. In the inertia phase that follows the torque phase, the engine has to be decelerated inorder t
82、o achieve speed synchronization with the target gear. During this phase, the kinetic energy of the pow-ertrain components is transferred to the vehicle while engaging the next gear, causing a brief rise in the outputtorque. The inertia phase is responsible for most of the torque recovery and this is
83、 where the torque overshootsoccur. The two key aspects considered in the clutch pressure control are:(a) As the oncoming clutch reaches its critical capacity, the offgoing clutch pressure is brought to zero. Thisensures that the clutches will not tie-up and prevent causing a sharp torque overshoot,
84、which is felt as avehicle jerk by the driver.(b) Before the oncoming clutch reached its critical capacity, the offgoing clutch is held above its criticalcapacity to maintain continuity in torque transmission and prevent the engine from flaring that is feltas a power loss by the driver.6.1. LaunchTo
85、launch a vehicle from neutral, Clutch 1 is controlled to slip with the first gear engaged. The launch con-trol is through a look-up table for the clutch pressure and the normalized pressure in the clutch cylinder ismodulated. Various output torque profiles at launch, shown in Fig. 6, can be obtained
86、 based on differentlaunch pressure profiles of the clutch cylinder as shown in Fig. 7. For an aggressive launch, the slip torqueReduce engine torque Start ofShiftEnd ofShiftOffgoing clutchdisengagingOscillationsOncoming clutchengagingPeakovershootTorque PhaseInertia PhaseOffgoing clutchdisengaedTime
87、Output TorqueTorqueRe-instatementFig. 5. Output torque profile for typical DCT shift.M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182177profile has a steep slope and clutch locking torque is achieved in a very short time. Likewise, for a slow startthe same torque profile has a gent
88、le slope and the time to lock the clutch is greatly increased. An optimizedFig. 6. Output torque during launch.Fig. 7. Clutch pressure profiles for launch.178M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182launch is achieved between the above two extremes. Although steady state tor
89、que is seen to be achieved in thesame launch time, the clutch is hooked up quite early in case of option 2. This results in a pronounced vehiclejerk, however such a situation might be desirable when the vehicle is being driven in Sport mode.6.2. 12 UpshiftThe simulation was run for an upshift from f
90、irst gear to second gear at wide open throttle. The vehicle speedis approximately 20 mph when the shift is initiated. The output torque during shift is plotted in Fig. 8. At thebeginning of the shift, the pressure on the offgoing clutch is sharply reduced to the threshold for disengage-ment. This ca
91、uses the output torque to drop, which depicts the torque phase. The oncoming clutch pressureis ramped up to the desired pressure level and the output torque begins to regain its value in the inertia phase.Fig. 9 shows the relative angular velocity across the oncoming and the offgoing clutches.The ou
92、tput torque variations mainly depend on the clutch piston fill pressures. The simulation model wastested for three different clutch pressures and the effect on output torque profiles was studied. Fig. 10 showsthe three clutch pressure profiles for the shift control. Each set of twin clutch pressure
93、curves has a differentpoint for offgoing clutch disengagement and also a different rate of engagement for the oncoming clutch. Thepressure profile in red is optimized in terms of the shift smoothness. For a higher than optimum pressureapplied at the clutch, the torque overshoot is clearly observed.
94、On the other hand, a large torque dip calledtorque hole is observed when the clutch fill pressures are lower than optimum. The three torque profilesare plotted for comparison in Fig. 11.6.3. 32 DownshiftA 32 downshift is simulated and the results are shown in Fig. 12. A downshift is initiated either
95、 when thethrottle is suddenly depressed for a passing maneuver or when a brake is applied. In this model, the signalFig. 8. Output torque in a 12 upshift.M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182179builder incorporates a brake model. In the Stateflow gear selector, an additi
96、onal condition is included to checkif brake signal is prevailing. When the brake signal is detected the transmission is ready for a downshift.Fig. 9. Relative angular velocities of the dual clutches.Fig. 10. Clutch pressure profiles for 12 shift.180M. Kulkarni et al. / Mechanism and Machine Theory 4
97、2 (2007) 1681827. ConclusionThis paper is aimed at studying the dynamic characteristics of the shift events of dual clutch automatedlayshaft transmissions. A simulation model integrating the engine, the transmission and the vehicle environ-Fig. 11. Effects of pressure profiles on output torque.Fig.
98、12. Output torque during a 32 downshift.M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182181ment has been built using Matlab/Simulink software tool. The main objective was to create an integratedengine-driveline system model and simulate the behavior of the DCT in events like launch
99、, upshift or down-shift. The importance of clutch apply timing and clutch pressure control signals are quantitatively analyzedand explained using model simulation in this paper. To understand the effects of clutch apply timing, threedifferent clutch pressure profiles are modeled and the resulting to
100、rque outputs are comparatively studied asa case study. By modifying the clutch pressure profiles interactively with the simulation results, optimizedclutch control can be achieved for smooth vehicle launch and gear shifts. It can be concluded that an effectiveand detailed analysis of the DCT shift c
101、haracteristics can be performed using the model presented in the paper.References1 G. Lechner, H. Naunheimer, Automotive Transmissions: Fundamentals, Selection, Design and Application, Springer, Berlin, NewYork, 1999.2 M.A. Kluger, Denis M. Long, An Overview of Current Automatic, Manual and Continuo
102、usly Variable Transmission Efficiencies andTheir Projected Future Improvements, SAE paper 1999-01-1259, 1999.3 W. Grobpietsch, T. Sudau, Dual Clutch for Power-Shift Transmissions A Traditional Engaging Element with New Future, VDI-Berichte Nr. 1565, 2000, pp. 259273.4 D.H. Kim, K.J. Yang, K.S. Hong,
103、 J.O. Hahn, K.I. Lee, Smooth shift control of automatic transmissions using a Robust adaptivescheme with intelligent supervision, International Journal of Vehicle Design 32 (3/4) (2003) 250272.5 Y. Zhang, Z. Zou, X. Chen, X. Zhang, W. Tobler, Simulation and analysis of transmission shift dynamics, I
104、nternational Journal ofVehicle Design 32 (3/4) (2003) 273289.6 T.W. Megli, M. Haghgooie, D.S. Colbin, Shift Characteristics of A 4-Speed Automatic Transmission, SAE paper 1999-01-1060, 1999.7 Z. Zou, Y. Zhang, X. Zhang, W. Tobler, Modeling and simulation of traction drive dynamics and control, Journ
105、al of MechanicalDesign, ASME Transaction 123 (4) (2001) 556561.8 M. Yasuoka, M. Uchida, S. Katakuta, T. Yoshino, An Integrated control Algorithm for an SI Engine and CVT, SAE paper 1999-01-0752, 1999.9 M. Goetz, M.C. Levesley, D.A. Corolla, Integrated Powertrain Control of Gearshifts on Twin Clutch
106、Transmissions, SAE paper2004-01-1637, 2004.10 L.W. Tsai, G. Schultz, N. Higuchi, A novel parallel hybrid transmission, Journal of Mechanical Design, ASME Transaction 123 (2)(2001) 161168.11 K.L. Butler, M. Ehsani, P. Kamath, A Matlab-based modeling and simulation package for electric and hybrid elec
107、tric vehicle design,IEEE Transactions on Vehicular Technology 48 (6) (1999) 17701778.12 B.K. Powell, K.E. Bailey, S.R. Cikanek, Dynamic modeling and control of hybrid electric vehicle powertrain systems, IEEE ControlSystems Magazines 18 (5) (1998) 1733.182M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182
