汽車、奇瑞A3轎車兩軸式五擋手動(dòng)變速器設(shè)計(jì)含8張CAD圖

汽車、奇瑞A3轎車兩軸式五擋手動(dòng)變速器設(shè)計(jì)含8張CAD圖,汽車,A3,轎車,兩軸式五擋,手動(dòng),變速器,設(shè)計(jì),CAD J. Cent. South Univ. (2012) 19: 27882796 DOI: 10.1007/s11771-012-1343-4 Application of clutch to clutch gear shift technology for a new automatic transmission LU Xi(魯曦), WANG Shu-han(王書(shū)翰), LIU Yan-fang(劉艷芳), XU Xiang-yang(徐向陽(yáng)) School of Transportation Science and Engineering, Beihang University, Beijing 100191, China Central South University Press and Springer-Verlag Berlin Heidelberg 2012 Abstract: For the purpose of engineering development for a new 8-step speed automatic transmission, a simplified dynamic model for this gearbox was established and key parameters which affected the shift quality were analyzed. Aiming at four different shift types, the ideal characteristics of shift clutch and engine control were set up. By using torque estimation method, PI slip control algorithm and engine coordinated control principle, the control model and transmission controller were well developed for three shift phases which included rapid-fill phase, torque phase and inertia phase. The testing environment on the rig and prototype vehicle level was built and the testing results obtained in ultimate condition could verify the accuracy and feasibility of this shift control strategy. The peak jerk during shift process was controlled within 2 g/s where the smooth gearshift was obtained. The development proposal and algorithm have a high value for engineering application. Key words: automatic transmission; shift clutch; shift strategy; vehicle testing; peak jerk; gearshift; control algorithm 1 Introduction Clutch to clutch gear shift strategy is a key technology in design and control of automatic transmissions in modern automobile industry. Many efforts had been made aiming to improve gear shift quality especially upshift quality by establishing mathematical model and dynamic model. Among these studies, the most fundamental principle was the research of optimizing command pressure through model simulation and rig testing. Refs. 13 presented the structure of wet clutch and fundamentals of gear shift in automatic transmission by using a rigid body model which was the basic analysis for further researches. In Refs. 46, modeling and simulation technology had been mainly used for controlling command pressure at two active clutches. One disadvantage was obvious since the simulation procedure can not fully reflect hydraulic hysteresis characteristics. GOETZ et al 78 detailed the relatively more overall illustration for shift quality control but no reference verified control strategies in series vehicle. Some literatures 911 had introduced relevant dynamic models and shift characteristic optimization methods to obtain better shift quality, but without considering power-off shifts. MARANO et al 12 introduced the equivalent rotational inertia of the gearbox and driveline components as affection for shift quality. CHENG et al 1314 were interested in developing and optimizing the shift control algorithm through vehicle testing and calibration. These researches had made a great progress aiming at engineering application and it would be better if engine effect handling as shift progress could be considered to satisfy all shift types requirements. In this work, shift process analysis was introduced at first and then combined control strategy had been well developed for all of four main shift conditions including power-on upshift, power-on downshift, power-off upshift and power-off downshift. Torque estimated method and PI control were used to get an ideal torque handover profile and clutch slip control results. Calibrating and optimizing of command pressure in the prototype vehicle had also been concentrated on in this note. Finally, the testing results were obtained under the worst testing conditions which were power-on upshift and skip downshift between low gears at WOT wide open throttle so as to further validate the operation of control algorithm. 2 Simplified dynamic model The prototype transmission is an 8-step speed automatic transmission. Table 1 gives the transmission Foundation item: Project(51105017) supported by the National Natural Science Foundation of China; Project(2011BAG09B00) supported by the National Science and Technology Support Program of China; Project(2010DFB80020) supported by the Technology Major Project of the Ministry of Science and Technology of China Received date: 20110929; Accepted date: 20120502 Corresponding author: WANG Shu-han, PhD, Lecturer; Tel: +861082338121; E-mail: J. Cent. South Univ. (2012) 19: 27882796 2789 Table 1 Shift logic Engaged Clutches Gear B1 C1 C2 C3 C4 Reverse 1 2 3 4 5 6 7 8 shift logic showing totally five shift elements, one brake and four clutches, which are used to achieve eight forward gears and one reverse gear. As shown in Fig. 1, all 1-step and 2-step shifting and many far step shifting are called as “simple shift” because by disengaging offgoing clutch (current gear) and engaging oncoming clutch (target gear), a gear shift can be accomplished. Since the other three clutches will remain the previous states, keeping open or fully closed, the torque transmitted through these components can be considered as constant during each rapid shift process. Torque handover is only required from offgoing clutch to oncoming clutch. Therefore, a simplified dynamic model can be established for better understanding “simple shift” logic and its process (see Fig. 2 10). The simplified model consists of engine, torque converter, torque converter clutch (TCC), gearbox including two clutches K1 and K2, and vehicle body module. By using free body analysis, dynamic model can be mainly separated into several parts and their force Fig. 1 Simple shift logic moment equilibrium calculations are as follows. When vehicle starts to upshift, K1 gradually releases oil in torque phase and remains transmitting inertia torque before clutch slipping. After K2 has completed rapid-fill phase, it starts to transmit friction torque. tttK11if22i/JTTiTi (1) ooK1 1of2 2ooJT iT iT (2) 2i 1 t21f22i oo2iot()()i iTii Ti TiJJ (3) where Tt is turbine torque, TK1 is transferring inertia torque through K1, To is load torque at transmission output side, t is turbine angular speed and equals transmission input shaft speed. o is output shaft angular speed. Jt is equivalent inertia of turbine-active side of clutches part. Jo is equivalent inertia of passive side of clutches-output shaft part. i1i is drive ratio from input shaft to active side of K1 while i1o is the one from passive side of K1 to output shaft. One can make i1 by multiplying i1i and i1o. i2i is the drive ratio from input shaft to active side of K2 and i2o is the one from passive side of K2 to output shaft. One can make i2 by multiplying i2i and i2o. Friction torques at K1 and K2 are defined as Tf1 and Tf2, respectively, and can be calculated by Eq. (4). Fig. 2 Simplified dynamic model J. Cent. South Univ. (2012) 19: 27882796 2790 33oiff22oi2sgn()3RRTPA ZRR (4) where is friction coefficient, sgn() is direction of clutch friction torque which can be calculated by the relative rotational speed between two sides of the clutch, P is surface pressure of the clutch, Af is area of a friction element, Z represents the number of the friction surfaces. Ro and Ri represent outer and inner radius of the friction element, respectively. The direction of friction torque can be decided by relative angular speed between active and passive side of clutches. With continuous depressing of K1 pressure, two clutches K1 and K2 are both in slipping phase and transmit friction torque f1 1of2 2ooooT iT iTJ (5) In inertia phase, K1 is open and K2 transmits friction torque. f2 2ooooT iTJ (6) The degree of gear shift impact during shift process can be defined as Jerk, which is the rate of change of vehicle longitudinal acceleration as follows: 22or22d2 60ddd1 000ddravjttt (7) where a is vehicle longitudinal acceleration, v is the vehicle speed, rr is the radius of driving wheel. On the assumption that the change rate of To remains constant during shift process, Eqs. (36) can be combined into Eq. (7), respectively and then the jerk of vehicle can be only decided and controlled by torque change rate and pressure control during the shift process. eK1K2012ddd()()()dddTPPjfffttt (8) Therefore, the key factor of achieving a high-level shift quality is to precisely control engine torque change rate by torque request control and torque handover process at oncoming clutch and its opposite offgoing clutch by controlling the surface pressure PK1 and PK2 at both clutches. 3 Shift process analysis According to the different shift conditions of positive or negative engine torque, gear shift can be separated into four main shift types which are called as power-on upshift, power-on downshift, power-off upshift and power-off downshift, respectively. Except for rapid-fill phase, the control strategies for the above shift types are different. All shifts consist of two main phases during the shift process, and a torque phase where torque is transmitted from offgoing clutch to oncoming clutch and an inertia phase where engine speed is transmitted from the level of current gear to target gear will be completed. Torque phase appears first on power-on upshift and is followed by inertia phase. The sequence of two phases is reversed on the power-on downshift. In this work, only power-on shift is analyzed as power-off upshift is basically controlled like power-on downshift and power-off downshift is controlled like power-on upshift. 3.1 Upshift The control principle for power-on upshift is shown in Fig. 3. During this gear shift from current gear to target gear with a positive engine torque, torque handover is performed by disengaging offgoing clutch and engaging oncoming clutch. The torque handover needs to be precisely controlled since if oncoming clutch is engaged too late, then output torque will drop down and a power interruption will take place. On the other hand, if offgoing clutch is disengaged too late while oncoming clutch transfers most of engine torque, then offgoing clutch will generate a negative torque and a sharp rise of output torque will cause bump to the powertrain when it is released 8. Fig. 3 Control principle for power-on upshift: (a) Engine speed control principle for power-on upshift; (b) Engine and clutches torque control principles for power-on up shift J. Cent. South Univ. (2012) 19: 27882796 2791 In the following inertia phase, the engine speed will be decelerated for achieving synchronization to the level of target gear. In the control schemes, this is achieved by increasing the pressure at the oncoming clutch to follow a reducing speed profile. To avoid a bumpy shift caused by increasing pressure at oncoming clutch because of inertial torque which is transferred to the wheels due to the engine deceleration, an engine torque reduction is required during the inertial phase. 3.2 Downshift The control principle for power-on downshift is shown in Fig. 4. Downshift starts with an inertial phase first where the engine needs to be accelerated for synchronization to the level of projected engine speed at target gear. Engine acceleration in inertial phase can be achieved by decreasing hydraulic pressure at offgoing clutch which brings the clutch to the state of slip. Output torque will drop severely and affect the drivability of vehicle because of a partial disengagement of offgoing clutch. Therefore, the engine torque is required to increase and the oncoming clutch is required to fill for preparation of transferring the torque during inertia phase. After the synchronization has been achieved, the torque phase can be entered when the engine torque is transferred from offgoing clutch to oncoming clutch. Again, an increasing command pressure is applied at oncoming clutch while a decreasing command pressure at offgoing clutch for the sake of torque handover can accord with the ideal characteristics to avoid shift intervention or power transfer interruption or the occurrence of engine speed flare because of decreasing torque or engine speed tie up due to an uncontrolled increasing torque from load side. 4 Gear shift control strategy Based on the above analysis, the key issue of clutch-to-clutch shift technology is to control the command pressure at target offgoing and oncoming clutches by controlling control current on corresponding shift solenoids with considering of demand torque and target slip calculation. A control rule of gear shift is illustrated in Fig. 5. Fig. 4 Control principle for power-on downshift: (a) Engine speed control principle for power-on downshift; (b) Engine and clutches torque control principles for power-on downshift Fig. 5 Control rule of gear shift J. Cent. South Univ. (2012) 19: 27882796 2792 4.1 Clutch kiss point testing in spin rig Before transferring torque, oncoming clutch needs to be filled to reduce clearance of clutch piston and to achieve kiss point pressure, at this point oncoming clutch just transmits torque. Therefore, command pressure at oncoming clutch will be set as a kiss point pressure which can be calibrated as follows. The input speed of spin rig is set to be 1 000 r/min manually and kept constant throughout the testing. First gear, 1st gear, 5th gear, 2nd gear and 1st gear are selected for testing Clutch B1, C1, C2, C3 and C4 kiss point pressure, respectively. As shown in Fig. 6, when selecting 1st gear, clutches C1 and C4 are closed completely by applying 1 000 mA control current to the solenoid and control current at B1 valve is gradually increased by steps of 50 mA. To get a higher accuracy of kiss point pressure, control current can be gradually increased by steps of 2 mA. Increment control current of B1 valve causes pressure increment in B1 valve and at point when output shaft starts rotating pressure and current signal can be read to find out clutch kiss point. The same process was repeated to get the clutch kiss point pressure for each clutch (Table 2). Fig. 6 B1 clutch kiss point pressure testing in spin rig Table 2 Clutch kiss point data from spin rig Solenoid B1 at 1st gear C1 at 1st gear C2 at 5th gear C3 at gear2nd gear C4 at1st geari/mA 350 356 360 346 362 p/105Pa 1.48 1.45 1.43 1.51 1.54 4.2 Opposite clutch torque estimated and control In torque phase, no matter of upshift or downshift, the control algorithm is quite similar. Cooperative control of oncoming clutch and offgoing clutch is necessary. An increasing command pressure at oncoming clutch will be set with a certain calibrateable ramp rate of torque change. The controller only manipulates a decreasing command pressure at offgoing clutch based on torque estimation control. An estimated torque and rate of torque change at offgoing clutch are calculated to make sure demand torque at this clutch will be decreased as a similar rate as actual torque increases at oncoming clutch. Toc(n), the estimated torque at oncoming clutch in sample time n can be calculated by the following equation: ococbockocock( )( ), if( )0, ifPPnP nPPTnPP (9) where Poc(n) is the actual pressure at oncoming clutch in sample time n captured by pressure sensor, Pb(n) is a balanced compensation pressure in sample time n calculated by engine speed to avoid engine stall, Poc is a set of factors in look-up table of pressure versus torque, is a relative factor calculated by micro-slip state of clutch which is normally less than 0.8, and Pk is the kiss point pressure at oncoming clutch. Toc(n), which is defined as ramp rate of torque change at oncoming clutch in sample time n, can be obtained by ocococ( )( )(1)TnTnTn (10) where Toc(n1) is the estimated torque at opposite clutch in sample time n1. Therefore, Tog(n), which is the demand torque at offgoing clutch in sample time n, can be calculated as ogocogeocmax(1)( )( )min( )( )TnTnTnT n KTnT (11) where Tog(n1) is demand torque at offgoing clutch in sample time n1, Te(n) is engine torque without intervention in sample time n, K is torque ratio of torque converter, is an experience factor which can be calibrated. Tmax is the maximum allowed torque capacity of offgoing clutch. Then command pressure at offgoing clutch can be calculated based on the Tog by using a calibrateable look-up table of torque versus pressure. 4.3 Slip control and engine torque reduction In inertial phase, slip control strategy for power-on upshift and power-off downshift is different from the one for power-on downshift and power-off upshift. For power-on upshift and power-off downshift, slip controller manipulates command pressure at oncoming clutch based on target slip by PI control since pressure at offgoing clutch is zero during inertial phase and only needs to control the oncoming clutch to get a smooth shift during the process of oncoming clutch engagement and engine deceleration. The PI system can be defined by the following equation 15. ( )( )( )du tP e tI e tt (12) where P is the proportion coefficient, I is the integral coefficient, u(t) is the PI term torque as the control value, J. Cent. South Univ. (2012) 19: 27882796 2793 e(t) is the error which can be calculated by csliptslip( )( )( )e ttt (13) For these two shift types, the target slip tslip is always calculated by using the target turbine speed tt, output shaft speed o, and target gear ratio itgear when the target gear is determined in the very beginning of torque phase (Fig. 4). tslipttotgeari (14) The actual slip cslip can be obtained by using the current turbine speed ct, output shaft speed o, and current gear ratio icgear as follows: cslipctocgeari (15) When torque phase finishes, PI controller will adjust the actual slip approaching to target slip in inertial phase in order that oncoming clutch can be gradually engaged with a reducing speed profile. Meanwhile, due to the reason explained in previous section, a reduction of engine torque needs to be accomplished during inertial phase by sending a torque request Ter with air combined with spark control to the engine management system. ereeRollglslmax(,)TTTTT (16) where TeRoll can be calibrated during torque roll off and torque roll on phase and calculated gradient limit torque Tgl, limit torque to control the remain slip speed in inertial phase Tsl can be obtained as follows: glecslipslcsliptslipNmprpm()TJTR (17) where RNmprpm is the remained slip speed calibrated rate. For power-on downshift and power-off upshift, since the inertial phase comes first and oncoming clutch needs to be filled in the initial of inertial phase, the controller only manipulates the command pressure at offgoing clutch based on target micro slip of offgoing clutch to reduce its pressure smoothly for engine acceleration and keeping offgoing clutch slipping. 5 Hardware platform The transmission control unit is a well-developed prototype controller with necessary input and output interface for engineering application (Fig. 7). 6 Results analysis The worst case for shift quality control normally happens on power-on upshift or on power-on down shift between low gears both at WOT-wide open throttle since the gear step ratios between these gears are larger Fig. 7 I/O interface of transmission control unit than the one between high gears and engine generates maximum output torque to powertrain system when driver requires wide open throttle which is called “kickdown” driving. This aggressive sporty driving in low gears always requires high-level of gear shift control and calibration solution, otherwise, the driver or passengers in testing vehicle will feel obviously bumpy as gear shift progresses. 6.1 Upshift Figure 8 displays the testing results of a power-on upshift from 1st to 2nd gear at a vehicle speed around 35 km/h and wide open throttle. Fill phase starts at 0.34 s while inertial phase finishes at 1.2 s. Figure 8(a) shows command pressure and actual pressure at both clutches. Only very small deviation occurs between command pressure and hydraulic pressure. The whole system has a good response to control signals with