Abstract:
An improved control for an automatic transmission power-on downshift, wherein a dynamic model of the transmission is used to schedule the on-coming and off-going clutch pressures based on the transmission input torque and a desired trajectory of the input shaft during the shift. The shift is initiated with the off-going clutch by using the dynamic model achieve consistent initiation of clutch slippage, and to conform the input speed to the desired trajectory. When the input speed nears a synchronization speed for the target speed ratio, the dynamic model is used to engage the on-coming clutch at a rate based on the input torque while maintaining the input speed in synchronism with the target speed ratio. As a result, the control responds appropriately to dynamic changes in input torque, the input speed more accurately tracks the desired trajectory, and the shifts are completed at or near synchronism. Additionally, scheduling the clutch pressures based on the dynamic model achieves more consistent shift feel and improved adaptability to different powertrain and vehicle-type configurations while reducing the number of calibrated parameters requiring adaptive correction.

Description:
TECHNICAL FIELD 
     This invention relates to a shift control for an automatic transmission, and more particularly to a model-based clutch pressure control for carrying out a power-on downshift. 
     BACKGROUND OF THE INVENTION 
     In general, a motor vehicle automatic transmission includes a number of gear elements and selectively engageable friction elements (referred to herein as clutches) that are controlled to establish one of several forward speed ratios between the transmission input and output shafts. The input shaft is coupled to the vehicle engine through a fluid coupling such as a torque converter, and the output shaft is coupled to the vehicle drive wheels through a differential gearset. Shifting from a currently established speed ratio to new speed ratio involves, in most cases, disengaging a clutch (off-going clutch) associated with the current speed ratio and engaging a clutch (on-coming clutch) associated with the new speed ratio. 
     Various techniques have been used for electronically controlling the on-coming and off-going clutches during a power-on downshift. For example, the U.S. Pat. Nos. 5,029,494 and 5,070,747 to Lentz et al. disclose a downshift control in which the on-coming clutch is filled in preparation for engagement, while the off-going clutch is controlled in a series of steps including (1) progressively releasing the off-going pressure until the off-going clutch begins to slip, (2) controlling the off-going pressure to achieve a desired input speed profile, and (3) controlling the off-going pressure to hold the input speed substantially at the post-shift speed; and thereafter, engaging the on-coming clutch and dis-engaging the off-going clutch. Similar control techniques are also described in the U.S. Pat. No. 4,653,351 to Downs et al., the U.S. Pat. Nos. 4,796,490 and 5,079,970 to Butts et al. 
     The above-described controls tend to involve numerous calibrated parameters requiring adaptive adjustment to compensate for variations, and have a relatively limited ability to react to changes in input torque during the shift. In practice, a margin of safety is frequently provided by intentionally overlapping the on-coming and off-going clutches to some degree, at the expense of clutch heating and shift quality. 
     SUMMARY OF THE INVENTION 
     The present invention is directed to an improved control for an automatic transmission power-on downshift, wherein a dynamic model of the transmission is used to schedule the on-coming and off-going clutch pressures based on the transmission input torque and a desired trajectory of the input shaft during the shift. The shift is initiated with the off-going clutch by using the dynamic model to achieve consistent initiation of clutch slippage, and to conform the input speed to the desired trajectory. When the input speed nears a synchronization speed for the target speed ratio, the dynamic model is used to engage the on-coming clutch at a rate based on the input torque while maintaining the input speed in synchronism with the target speed ratio. As a result, the control responds appropriately to dynamic changes in input torque, the input speed more accurately tracks the desired trajectory, and the shifts are completed at or near synchronism. Additionally, scheduling the clutch pressures based on the dynamic model achieves more consistent shift feel and improved adaptability to different powertrain and vehicle-type configurations while reducing the number of calibrated parameters requiring adaptive correction. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a diagram of an automatic transmission and microprocessor-based control unit for carrying out the control of this invention. 
     FIG. 2 is a table indicating a relationship between transmission clutch activation and corresponding speed ratio. 
     FIGS. 3 and 4 graphically depict power-on downshifts carried out according to this invention. In each figure, Graph A depicts the transmission input speed, Graph B depicts the on-coming clutch pressure, and Graph C depicts the off-going clutch pressure. 
     FIG. 5 is a state diagram illustrating an off-going clutch pressure control according to this invention. 
     FIG. 6 is a state diagram illustrating an on-coming clutch pressure control according to this invention. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     The control of this invention is described in the context of a multi-ratio power transmission having a planetary gearset of the type described in the U.S. Pat. No. 4,070,927 to Polak, and having an electro-hydraulic control of the type described in U.S. Pat. No. 5,601,506 to Long et al. Accordingly, the gearset and control elements shown in FIG. 1 hereof have been greatly simplified, it being understood that further information regarding the fluid pressure routings and so on may be found in the aforementioned patents. 
     Referring to FIG. 1, the reference numeral  10  generally designates a vehicle powertrain including engine  12 , transmission  14 , and a torque converter  16  providing a fluid coupling between engine  12  and transmission input shaft  18 . A torque converter clutch  19  is selectively engaged under certain conditions to provide a mechanical coupling between engine  12  and transmission input shaft  18 . The transmission output shaft  20  is coupled to the driving wheels of the vehicle in one of several conventional ways. The illustrated embodiment depicts a four-wheel-drive (FWD) application in which the output shaft  20  is connected to a transfer case  21  that is also coupled to a rear drive shaft R and a front drive shaft F. Typically, the transfer case  21  is manually shiftable to selectively establish one of several drive conditions, including various combinations of two-wheel-drive and four-wheel drive, and high or low speed range, with a neutral condition occurring intermediate the two and four wheel drive conditions. 
     The transmission  14  has three inter-connected planetary gearsets, designated generally by the reference numerals  23 ,  24  and  25 . The input shaft  18  continuously drives a sun gear  28  of gearset  23 , selectively drives the sun gears  30 ,  32  of gearsets  24 ,  25  via clutch C 1 , and selectively drives the carrier  34  of gearset  24  via clutch C 2 . The ring gears  36 ,  38 ,  40  of gearsets  23 ,  24 ,  25  are selectively connected to ground  42  via clutches C 3 , C 4  and C 5 , respectively. 
     As diagrammed in FIG. 2, the state of the clutches C 1 -C 5  (i.e., engaged or disengaged) can be controlled to provide six forward speed ratios (1, 2, 3, 4, 5, 6), a reverse speed ratio (R) or a neutral condition (N). For example, the first forward speed ratio is achieved by engaging clutches C 1  and C 5 . Shifting from one speed forward speed ratio to another is generally achieved by disengaging one clutch (referred to as the off-going clutch) while engaging another clutch (referred to as the on-coming clutch). For example the transmission  14  is shifted from first to second by disengaging clutch C 5  while engaging clutch C 4 . 
     The torque converter clutch  19  and the transmission clutches C 1 -C 5  are controlled by an electro-hydraulic control system, generally designated by the reference numeral  44 . The hydraulic portions of the control system  44  include a pump  46  which draws hydraulic fluid from a reservoir  48 , a pressure regulator  50  which returns a portion of the pump output to reservoir  48  to develop a regulated pressure in line  52 , a secondary pressure regulator valve  54 , a manual valve  56  manipulated by the driver of the vehicle and a number of solenoid operated fluid control valves  58 - 64 . 
     The electronic portion of the control is primarily embodied in the engine control unit  65  and the transmission control unit  66 , illustrated in FIG. 1 as two separate modules. Both control units  65 ,  66  are microprocessor-based, and may be conventional in architecture. The engine control unit  65  controls the operation of engine functions such as fuel, spark timing, and so on depending on the control variables afforded by engine  12 , and the transmission control unit  66  controls the solenoid operated fluid control valves  58 - 64  based on a number of inputs to achieve a desired transmission speed ratio. The transmission control unit inputs include signals representing the transmission input speed TIS, a driver torque command TQ, and the transmission output speed TOS. Sensors for developing such signals may be conventional in nature, and have been omitted for simplicity. Additionally, the engine control unit  65  supplies an engine output torque signal EOT to transmission control unit  66 . 
     The control lever  82  of manual valve  56  is coupled to a sensor and display module  84  that produces an diagnostic signal on line  86  based on the control lever position; such signal is conventionally referred to as a PRNDL signal, since it indicates which of the transmission ranges (P, R, N, D or L) has been selected by the vehicle driver. Finally, the fluid control valves  60  are provided with pressure switches  74 ,  76 ,  78  for supplying diagnostic signals to control unit  66  on lines  80  based on the respective relay valve positions. The control unit  66 , in turn, monitors the various diagnostic signals for the purpose of electrically verifying proper operation of the controlled elements. 
     The solenoid operated fluid control valves  58 - 64  are generally characterized as being either of the on/off or modulated type. To reduce cost, the electro-hydraulic control system  44  is configured to minimize the number of modulated fluid control valves, as modulated valves are generally more expensive to implement. To this end, a set of three on/off relay valves, shown in FIG. 1 as a consolidated block  60 , are utilized in concert with manual valve  56  to enable controlled engagement and disengagement of each of the clutches C 1 -C 5  with only two modulated valves  62 ,  64 . For any selected ratio, the control unit  66  activates a particular combination of relay valves  60  for coupling one of the modulated valves  62 ,  64  to the on-coming clutch, and the other modulated valve  62 ,  64  to the off-going clutch. 
     The modulated valves  62 ,  64  each comprise a conventional pressure regulator valve biased by a variable pilot pressure that is developed by current controlled force motor. The fluid controlled valve  58  is also a modulated valve, and controls the fluid supply path to converter clutch  19  in lines  70 ,  72  for selectively engaging and disengaging the converter clutch  19 . The transmission control unit  66  determines pressure commands for smoothly engaging the on-coming clutch while smoothly disengaging the off-going clutch, develops corresponding force motor current commands, and then supplies current to the respective force motors in accordance with the current commands. 
     In a power-on downshift, the transmission speed ratio (TIS/TOS) is increased, which requires that the engine  12  accelerate the transmission input shaft  18  from a pre-shift speed defined by the product (TOS*SRold) to a synchronization speed defined by the product (TOS*SRnew), where SRold is the old or current speed ratio, and SRnew is the new or desired speed ratio. In general, this can be achieved by controllably releasing the off-going clutch pressure while preparing the on-coming clutch for engagement, and then releasing the off-going clutch as the on-coming clutch engages. Indeed, this is the general control premise of the aforementioned U.S. Pat. Nos. 5,029,494, 5,070,747, 4,653,351, 4,796,490 and 5,079,970. As indicated above however, such controls involve numerous calibrated parameters that must be adaptively adjusted, and have a relatively limited ability to react to changes in input torque during the shift. The present invention, on the other hand utilizes a dynamic model of the transmission that takes into account dynamic variations in input torque, and that provides improved and more consistent shift quality with less reliance on extensive adaptive correction. 
     FIGS. 3-4 graphically illustrate a power-on downshift carried out according to this invention. In each figure, Graphs A-C respectively depict the transmission input speed TIS, the on-coming clutch pressure command Ponc, and the off-going clutch pressure command Pofg. As explained below, FIG. 3 represents a shift with ideal timing, and FIG. 4 represents a shift with input speed flare. 
     In general, the shift is initiated by progressively reducing Pofg, resulting in off-going clutch slippage at time t 2 . At such point, Pofg is adjusted to a new level and then controlled to allow the engine to raise the input speed TIS at a desired rate to a synchronization speed (SYNC) defined by the product of the new speed ratio (SRnew) and the output speed TOS. Meanwhile, the Ponc is controlled so that the on-coming clutch will be ready for engagement when TIS nears SYNC at time t 4 . The time t 4  may be defined in terms of a predetermined on-coming clutch slip (that is, SYNC−TIS), but is preferably defined in terms of a predicted time until TIS reaches SYNC, based on measured input acceleration and on-coming clutch slip. In the ensuing interval t 4 -t 7 , Ponc and Pofg are controlled to maintain TIS at or near SYNC, and the shift is completed at time t 7  by fully engaging the on-coming clutch and fully disengaging the off-going clutch. 
     On a more detailed level, the off-going pressure control is designed to initiate off-going clutch slip (SLIPofg) a given time Tofg_slip(des) after shift initiation. This is achieved by ramping Pofg downward at a ramp rate RR 1  so that its value at the conclusion of the interval Tofg_slip(des)—that is, at time t 2 —produces a torque capacity corresponding to the minimum reaction torque (TQofg_reaction) required to hold the off-going clutch without slipping. This reaction torque may be computed as a pressure Pofg_reaction according to the expression: 
     
       
           Pofg _reaction= Kcl*Klv*TQin   (1) 
       
     
     where Kcl is pressure-to-torque gain of the off-going clutch, Klv is the leverage gain corresponding to the old or current speed ratio, and TQin is the transmission input torque. The input torque TQin is subject to variation during the shift, and may be computed based on the engine output torque EOT and the torque ratio TR of torque converter  19 . The ramp rate RR 1  is set by calibration, and Pofg during the interval t 0 -t 2  is subject to adaptive adjustment based on a detected deviation between the time Tofg_slip(des) and the time when off-going clutch slip actually occurs. Thus, the off-going pressure in the interval t 0 -t 2  is given by the equation: 
     
       
           Pofg=Pofg _reaction+( RR   1 * t )+ Pofg _adaptive  (2) 
       
     
     where Pofg_adaptive is the adaptive correction, and t is the accumulated time ramping at the rate RR 1 . 
     Once the off-going clutch begins to slip, the off-going pressure is controlled so that the input speed will rise toward the SYNC speed at a desired rate, referred to herein as TIS_DOT(des). This control involves the combination of a feed-forward component based on the dynamic model of the transmission, and a feed-back component based on a detected input speed error. The dynamic model of the transmission during this phase of the control comprehends the inertial effects of the engine and transmission, and is given according to the equation: 
     
       
           Pofg   —   accel=Kcl[ ( Klv*TQin )+( Kin*TIS   —   DOT ( des ))+( Keng*ES   —   DOT ( des ))]  (3) 
       
     
     where Pofg_accel is the off-going pressure required to accelerate the input shaft at the desired acceleration TIS_DOT(des), ES_DOT(des) is the engine acceleration corresponding to TIS_DOT(des), Kin is an inertia coefficient for the input shaft  18 , and Keng is an inertia coefficient for the engine  12 . The coefficients Kin and Keng are negative in sign, reflecting the fact that an increase in TIS_DOT(des) necessitates a decrease in Pofg_accel. 
     Although the off-going clutch will theoretically begin slipping at time t 2  when Pofg falls to the computed value of Pofg_reaction, slipping may actually begin slightly before or after time t 2  due to modeling inaccuracies of equation (1). Regardless of such inaccuracies, it is certain that the offgoing clutch reaction torque TQofg_reaction is in equilibrium with the input torque TQin at the instant of off-going clutch slippage. Accordingly, the off-going clutch pressure for input acceleration control is adjusted relative to the modeled off-going pressure when slip is detected. Such pressure is designated as Pofg_at_slip, and is determined as follows: 
     
       
           Pofg   —   at _slip= Pofg   —   accel (at slip)+ Pofg   —   hyd _delay  (4) 
       
     
     where Pofg_accel(at slip) is the value of Pofg_accel (equation 3) when the off-going clutch begins to slip, and Pofg_hyd_delay is a pressure offset due to the hydraulic response delay of the off-going clutch. The hydraulic response delay HD is calibrated for any given clutch, and the term Pofg_hyd_delay is given according to the product (HD*RR 1 ), where RR 1  is the off-going pressure ramp rate prior to off-going slip detection. Thus, when off-going clutch slippage is detected, the off-going pressure Pofg is changed by a value ΔP determined according to the equation: 
     
       
         Δ P=Pofg _reaction− Pofg   —   accel (at slip)−( HD*RR   1 )  (5) 
       
     
     The pressure change may be made in a single step as illustrated in FIG. 3, or may be made in a series of smaller steps if desired. Thereafter, the off-going pressure Pofg is repeatedly computed in the interval t 2 -t 5  according to the equation: 
     
       
           Pofg (new)= Pofg (old)+ Pofg   —   accel (new)− Pofg   —   accel (at slip)+( K*SPD   —   ERR )  (6) 
       
     
     where Pofg(old) is the previous off-going pressure command, Pofg_accel(new) is the evaluation of equation (3) based on the current value of input torque TQin, Pofg_accel(at slip) is the evaluation of equation (3) at the initiation of off-going clutch slippage, K is a closed-loop proportional gain constant, and SPD_ERR is the closed-loop speed error between the TIS and a desired input speed corresponding to TIS_DOT(des). 
     While the off-going clutch pressure is being controlled to initiate and then control off-going clutch slip, the on-coming clutch is prepared for engagement by setting Ponc to a fill pressure Pfill for a predetermined fill interval (t 1 -t 3 ), and then lowering Ponc to a trim value sufficient to maintain the on-coming clutch in readiness for engagement. In the preferred embodiment, the combined duration of the fill and trim periods (that is, the interval t 1 -t 4 ) is designed to be substantially constant for a given shift, regardless of the engine speed or torque. Consequently, the on-coming pressure control is initiated after a variable fill delay Tdelay (defined by the interval t 0 -t 1  in FIG. 3) computed as follows: 
     
       
           T delay= T shift−( T fill+ T trim)  (7) 
       
     
     where Tfill is the fill interval t 1 -t 3 , Ttrim is the low pressure trim interval t 3 -t 4 , and Tshift is the estimated time required to accelerate the input speed to the synchronization speed SYNC, given the desired acceleration TIS_DOT(des). Thus, Tshift may be given by the equation: 
     
       
           T shift= Tofg _slip( des )+( SYNC−TIS   —   init )/TIS —   DOT ( des )  (8) 
       
     
     where TISinit is the input speed TIS at time t 2 . 
     The fill pressure Pfill is typically scheduled as a function of fluid temperature Tsump, and the fill time Tfill for any given clutch is determined according to the product of a calibrated fill time Tcal and a factor F representing the percent of fluid exhausted from the clutch since the last shift involving that clutch. Thus, the factor F accounts for any fluid remaining in the clutch, and is given according to the ratio of the time that the clutch has been exhausted to a calibrated time required to fully exhaust the clutch fluid, not to exceed a value of one. Preferably, Pfill is adaptively adjusted by iterative reduction to ensure that TIS will not significantly exceed SYNC; an adaptive adjustment of this type is disclosed in the aforementioned U.S. Pat. No. 5,070,747, which is incorporated herein by reference. The time Ttrim is calibrated, and the pressure Ptrim may be calibrated or determined through adaptive learning. 
     When the input speed TIS nears SYNC, the on-coming and off-going clutch pressures are controlled to maintain the input speed at SYNC. The on-coming pressure control is initiated when the estimated time to reach SYNC falls to a predetermined time, represented by the interval t 4 -t 5  in Graph B. As indicated above, the time-to-SYNC may be estimated based on the measured on-coming clutch slip speed (SYNC−TIS) divided by the input shaft acceleration (measured or desired). In the interval t 4 -t 7 , the dynamic model of the transmission is used to schedule on-coming pressure based on the input torque TQin and an inertia torque component designed to decelerate TIS in case TIS exceeds SYNC. The model equation for the on-coming synchronization control pressure Ponc_sync is given by: 
     
       
           Ponc   —   sync=Kcl[ ( Klv*TQin )+( Kin*TIS   —   DOT ( sync ))+( Keng*ES   —   DOT ( sync ))]  (9) 
       
     
     where Kcl[(Klv*TQin)] is the input torque dependent component, and Kcl[(Kin*TIS_DOT(sync))+(Keng*ES_DOT(sync))] is the inertia torque component that is used if TIS exceeds SYNC. The terms TIS DOT(sync) and ES_DOT(sync) represent desired input and engine pull-down rates for the case where TIS exceeds SYNC. 
     The off-going clutch pressure when TIS reaches SYNC at time t 5  is a combination of a calibrated open-loop ramp and a closed-loop term that increases the pressure as required to hold TIS at SYNC. Specifically, the pressure is repeatedly computed using the equation: 
     
       
           Pofg (new)= Pofg ( sync )−( RR   2 * Tsync )+( Kp*SLIPonc )+( Ki*Σ ( SLIPonc ))  (10) 
       
     
     where Pofg(sync) is the pressure command value at the beginning of the SYNC control period at time t 5 , RR 2  is the open-loop ramp rate, Tsync is the accumulated time in the SYNC control logic (defined as the current time t minus the SYNC initiation time t 5 ), Kp and Ki are proportional and integral closed-loop gain terms, and SLIPonc is the on-coming clutch slip. 
     In the shift of FIG. 3, the input speed TIS does not exceed SYNC; accordingly, the on-coming pressure in the interval t 5 -t 7  is based solely on input torque, and the off-going pressure is based solely on the calibrated ramp rate RR 2 . In the shift of FIG. 4, however, TIS exceeds SYNC; in this case, the inertia torque component of equation (9) raises the on-coming pressure to drive TIS back to SYNC at the calibrated rate TIS_DOT(sync), and the closed-loop components (proportional and integral) of equation (10) raise the off-going pressure in relation to SLIPonc, driving TIS back to SYNC at time t 6 . In either case, the shift is completed at time t 7  when TIS is substantially equal to SYNC for a calibrated period of time. 
     FIGS. 5 and 6 respectively depict the above-described off-going and on-coming pressure controls as a succession of states or phases. Referring to FIG. 5, the off-going control includes first, second, third, fourth and fifth states. The first state (off-going ramp) is entered at shift initiation, and is characterized by equation (2); the second state (off-going step) is entered on detection of off-going clutch slippage, and is characterized by the ΔPvalue of equation (5); the third state (off-going slipping) is entered on expiration of a hold period after the ΔP is activated, and is characterized by equation (6); the fourth state (off-going synchronization) is entered when TIS reaches SYNC, and is characterized by equation (10); and the fifth state (shift completion) is entered when TIS is substantially equal to SYNC for a calibrated period, and involves exhausting the off-going clutch. Referring to FIG. 6, the on-coming control also includes first, second, third, fourth and fifth states. The first state (oncoming fill delay) is entered at shift initiation, and involves a delay characterized by equation (7); the second state (on-coming fill pressure) is entered following the fill delay, and involves filling the on-coming clutch at Pfill for a predetermined period Tfill; the third state (on-coming trim) is entered following the fill period Tfill, and involves holding the pressure at a reduced value for a trim period Ttrim; the fourth state (on-coming synchronization) is entered when the estimated time to synchronization reaches a reference time, and is characterized by equation (9); and the fifth state (shift completion) is entered when TIS is substantially equal to SYNC for a calibrated period, and involves applying full pressure to the on-coming clutch. 
     In summary, the control of this invention uses a dynamic model of the transmission to schedule the on-coming and off-going clutch pressures in a power-on downshift based on the input torque and a desired input speed trajectory. As a result, the control responds appropriately to dynamic changes in input torque, the input speed more accurately tracks the desired trajectory, and the number of calibrated parameter requiring adaptive corrections are significantly reduced. 
     While described in reference to the illustrated embodiment, it will be understood that various modifications in addition to those mentioned above will occur to those skilled in the art. For example, the desired input acceleration may be scheduled as a function of time during the third state of the off-going clutch control, or alternately, the time Tshift may be specified instead of a constant desired acceleration. Additionally, various parameters, such as engine torque, vehicle loading, and the torque converter characterization may be determined by alternative methods than disclosed herein. Thus, it will be understood that controls incorporating these and other modifications may fall within the scope of this invention, which is defined by the appended claims.