小型SUV手動(dòng)五檔變速器設(shè)計(jì)
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Shifts in a dual-clutch transmission (DCT) are realized by torque transfer from one clutch to another without tractionThere has been a clear trend in the automotive industry in recent years towards increased ride comfort andautomatedlay-shaftgearing transmissions. One uses a single clutch and is basically amanualtransmission withanadded-oncontrolunitthatautomatestheclutchandshiftoperations.Inthisdesign,thereisaninterruptionof*Corresponding author. Tel.: +1 313 593 5539.E-mail address: anding (Y. Zhang).Mechanism and Machine Theory 42 (2007) 168182/locate/mechmtMechanismandMachine Theory0094-114X/$ - see front matter C211 2006 Elsevier Ltd. All rights reserved.fuel eciency. 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 oerdierent performance priorities when fit into a vehicle 1. Manual transmissions have an overall eciency of96.2%, which is the highest eciency value for any type of transmission. Current production automatics havebeen improved to provide an eciency of not more than 86.3%. Belt type CVTs have an overall eciency of84.6%,however,themajoradvantageofCVTisthatitallowstheenginetooperateinthemostfuel-ecientman-ner 2. Automated manual transmissions have the same eciency of manual transmissions and oer operationconvenience similar to conventional automatic transmissions. There exist two technically feasible designs forinterruption 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. 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 output torque in response to dierent 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 performance. Vehicle launch and shift process are bothsimulated to assess transmission shift quality and validate the eectiveness of the shift control.C211 2006 Elsevier Ltd. All rights reserved.Keywords: Dual-clutch transmission; Automatic transmissions1. IntroductionShift 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 2006Abstractdoi:10.1016/j.mechmachtheory.2006.03.002ntsM. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182 169torqueduringagearchangesincetheengineiscut-obytheclutchduringshift.Thistorqueinterruptionleadstounanticipatedpassengerfeltjerksduetovehicleaccelerationdiscontinuityandishighlyuncharacteristicofcon-ventionalautomatictransmissions.Theotherdesignusesadual-clutchsystembetweentheengineandthetrans-mission and overcomes the shortcomings of the single clutch version 3. The two clutches are engagedalternatively in dierent 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 ogoing clutch.Thisresultsinshiftcharacteristicsthataretypicalofclutch-to-clutchshiftscommonlyseeninconventionalauto-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 eorts have been focused on the modelingand control of vehicle transmission, such as conventional 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 equations 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 conventional 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 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 simulation 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 dierent clutch pressure profiles. Optimized clutch pressure profiles areobtained to achieve the best possible 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 eectiveness 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 positionedinside 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 between 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 change a gear the ogoing 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 mass 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, dierential and a vehicle model that 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 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 development of the model:nts170 M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182CL2 CL1SYN5RSYN66R5Final Drive Pinion 2Intermediate Shaft 2 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 hydraulic actuation system are not considered. Clutches are modeled as Coulomb friction elements. Temperature eects of the drivetrain are neglected.SYN13SYN2442 3 1Input ShaftOutputFinal Drive Pinion 1Intermediate Shaft 1Fig. 1. DCT stick diagram.4Output ShaftRCL1CL2131265Input ShaftIi/pIeEngineKmCmImepi /K1C1IhIsI1sI2imimhK2C2I3aiowFig. 2. DCT dynamic model.ntsThevehicl3.2. ClutchThedisconnectn is thelar velocitythe vehicleclutchM. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182 171clutch. The apply pressure is the control signal and the coecient of friction depends on the clutch slip, i.e., therelative velocity 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 freewheeling 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;Dx13 4where Tsynis the synchronizing torque, F is the normal force acting on the friction cone, l is the coecient offriction of the synchronizer ring and Dx13is the angular velocity dierence across a synchronizer, between thefirst andis running in particular speed with the clutch fully closed. As shown in the equation above, thetorque during a shift depends only on the clutch apply pressure and the coecient of friction for a givenoperation states and the torque transmitted in each state is described by the following equation:TCLClDxPappslippingT closed0 open8:3where 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 whennumber of friction discs. For simplicity, the clutch torque is modeled as a function of the relative angu-and the hydraulic pressure in the clutch piston in terms of a look-up table. The clutch has threeclutch torque is calculated as follows:TCL lFnn23C18C19R3oC0 R3iR2oC0 R2iC18C192where l is coecient 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,clutches in a DCT function as primary gear changing elements in addition to their normal purpose ofing the engine from the driving unit. Based on the clutch geometry and friction characteristics, the3.1. EngineThe mechanics of the engine assembly is modeled 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 output 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;xe 1where Teis the engine torque, xeis the engine angular velocity and TAis throttle angle.DCT system consists of complicated component level subsystems such as the engine, clutches, ande road load. These component models are described in the following.3. Component modelsthird gears as indicated by the subscripts.nts3.4. VehicleThelossesmodeled as coupled lumped masses. There are two sets of dynamic equations for the model. The model followsone setmechaniis nothe genericfora1 open172 M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182TCL1;TCL2Timengaged0 openC2613Ieqdxim TimitC0To14Ii=pdxi=pdt Ti=pC0TCL1 TCL2124.1. Operation in a particular gear4.1.1. Engine and input shaftImdxmdt ToC0 Tm7Tm Kmhm Cmxm8Iedxedt TeC0 Ti=p9Te TeTA;xe 10Ti=p K1heC0 hi=pC1xeC0 xi=p11where Tmis the torque at the engine mount, Ti/pis the input shaft torque, Kmand K1are the stiness 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 coecient of the input shaft and 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-transmission assembly, whosemagnitude is the same as the output torque.4.1.2. Transmission and output shaftof equations of motion when the vehicle is running in a particular gear and the transmission provides acallinkbetweentheengineandthewheels.Duringagearshiftitfollowsanothersetofequationsastheredirect link between the engine and the wheels as the system is in a dynamic shift state. Presented below aredynamic equations for model operation in any given gear and also during an upshift. The equationsspecificgearandaspecificgearchangecanbewrittenbychoosingthecorrectgearsandconsideringClutchin second, fourth and sixth gears and Clutch 2 open in first, third and fifth gear modes.TRL RARd6where Cdis the air drag coecient that depends on the body style and dimension, W is the vehicle weight, lfisthe rolling coecient, h is the grade angle and V is the vehicle velocity. 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 areroad loadvehicle road load model accounts for the rolling resistance, aerodynamic resistance and transmissionas formulated in the following:RA CdV2 lfW W sinh 5dt iantswhereangularexistingfollowswhereT f P ;Dx 25M. Kulkarni et al. / Mechanism and Machine Theory 42 (2007) 168182 173CL11 app1 CL1TCL2 f2Papp2;DxCL2 26DxCL1 xeC0xh27DxCL2 xeC0xs28ishiftxe29eTe fxe 23where 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=pC0TCL1 TCL224Ti=p K1heC0 hi=p C1xeC0xi=p21xeZTeC0 Ti=pIdt x0e22inertia of the two intermediate shafts assemblies. Ihand Isare the mass moments of inertia of the hollow andsolid shaft assemblies.4.1.3. 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 intermediate 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 TeC0 Ti=pC0C120C0C1C0C1transmission ratio. For example, Ieqfor the third speed and the sixth speed is calculated respectively as:Ieq I1 Ihi23for the third speed 18Ieq I2 Isi26for the sixth speed 19i3and i6are the gear ratios of the third and sixth speed, respectively. I1and I2are the mass moments ofinput shaft. TCL1and TCL2are the torques carried by the dual clutches. Timis the 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, respectively. 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 theTo K2hoC0 hwC2xoC0xw15ToC0 TRLRdWgdvdt16dvdt Rddxwdt17K2and C2are the stiness and damping coecients of the output shaft, respectively. hoand hware thedisplacements of the output shaft and wheel, respectively. Ii/pis the mass moment of inertia of thexwiantsIeqdximdtTCL1iodd TCL2ievenC0Toia30To K2hoC0hwC2xoC0 xw31ToC0 TRLRdWgdvdt32where xhand xsare the angular velocities of the hollow and solid shafts, respectively. 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) 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 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 combination 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
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