Abstract:
The present invention provides a method and apparatus for mathematically calculating an initial value of an adaptive parameter and thereafter adaptively controlling a power-on downshift in an automatic transmission wherein a transmission aberration during a shift is diagnosed and corrected during subsequent power-on downshifts. The invention is carried out by monitoring transmission characteristics including input speed, output speed and shift duration during a power-on downshift, and identifying departures from acceptable patterns. Each type of departure calls for a particular remedy, and a suitable adjustment is calculated based on the times and/or the commanded pressures at certain times, the adjustment being implemented by changing one or more initial conditions for the next shift of the same type. The adjustments may have to be large to make a full or significant partial correction at the next shift. Conversely small increments may be necessary to avoid over-correction.

Description:
TECHNICAL FIELD  
       [0001]     The present invention relates to a method and apparatus for improving power-on downshifts of an automatic transmission.  
       BACKGROUND OF THE INVENTION  
       [0002]     Generally, a motor vehicle automatic transmission includes a number of gear elements coupling its input and output shafts, and a related number of torque establishing devices such as clutches and brakes that are selectively engageable to activate certain gear elements for establishing a desired speed ratio between the input and output shafts. As used herein, the terms “clutches” and “torque transmitting devices” will be used to refer to brakes as well as clutches.  
         [0003]     The transmission input shaft is connected to the vehicle engine through a fluid coupling such as a torque converter, and the output shaft is connected directly to the vehicle wheels. Shifting from one forward speed ratio to another is performed in response to engine throttle and vehicle speed, and generally involves releasing or disengaging the clutch (off-going) associated with the current speed ratio and applying or engaging the clutch (on-coming) associated with the desired speed ratio.  
         [0004]     The speed ratio is defined as the transmission input speed or turbine speed divided by the output speed. Thus, a low gear range has a high speed ratio and a higher gear range has a lower speed ratio. To perform a downshift, a shift is made from a low speed ratio to a high speed ratio. In the type of transmission involved in this invention, the downshift is accomplished by disengaging a clutch associated with the lower speed ratio and engaging a clutch associated with the higher speed ratio, to thereby reconfigure the gear set to operate at the higher speed ratio. Shifts performed in the above manner are termed clutch-to-clutch shifts and require precise timing in order to achieve high quality shifting.  
         [0005]     The quality of shift depends on the cooperative operation of several functions, such as pressure changes within on-coming and off-going clutch apply chambers and the timing of control events. Moreover, manufacturing tolerances in each transmission, changes due to wear, variations in oil quality and temperature, etc., lead to shift quality degradation.  
       SUMMARY OF THE INVENTION  
       [0006]     The invention provides a method and apparatus for calculating optimal values for off-going clutch torque and transmission input torque, and thereafter adaptively controlling a power-on downshift in an automatic transmission wherein a transmission aberration during a shift is diagnosed and corrected during subsequent power-on downshifts.  
         [0007]     The method of the present invention is capable of making both large and small corrections.  
         [0008]     The method of the invention is carried out by mathematically calculating optimal values for off-going clutch torque and transmission input torque through the shift event. The method of the invention also monitors transmission characteristics including input speed, output speed, and shift duration during a power-on downshift, and identifies departures from acceptable patterns. Each type of departure calls for a particular remedy, and a suitable adjustment is calculated and applied by changing certain parameters in the shift control to alter one or more conditions for the next shift of the same type. The adjustments may have to be large to make a full or significant partial correction at the next shift. Conversely, small increments may be necessary to avoid over-correction.  
         [0009]     The above objects, features and advantages, and other objects, features and advantages of the present invention are readily apparent from the following detailed description of the best mode for carrying out the invention when taken in connection with the accompanying drawings. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0010]      FIG. 1  is a schematic illustration of an automatic transmission;  
         [0011]      FIG. 1   a  is a schematic illustration of a valve of  FIG. 1 ;  
         [0012]      FIG. 2   a  is a block diagram illustrating a method of calculating off-going clutch pressure during the inertia phase of a ratio change;  
         [0013]      FIG. 2   b  is a block diagram illustrating a method of calculating input torque during the torque phase of a ratio change;  
         [0014]      FIG. 3   a  is a graphical depiction of turbine acceleration vs. time during an optimal power-on downshift;  
         [0015]      FIG. 3   b  is a graphical depiction of turbine speed vs. time during an optimal power-on downshift;  
         [0016]      FIG. 4  is a schematic illustration of an automatic transmission;  
         [0017]      FIG. 5   a  is a graphical depiction of turbine speed vs. time during an optimal power-on downshift;  
         [0018]      FIG. 5   b  is a graphical depiction of the off-going clutch pressure vs. time during the optimal power-on downshift of  FIG. 5   a;    
         [0019]      FIG. 5   c  is a graphical depiction of the on-coming clutch pressure vs. time during the optimal power-on downshift of  FIG. 5   a;    
         [0020]      FIG. 6   a  is a graphical depiction of turbine speed during the shift aberrations “slip early” and “slip late”;  
         [0021]      FIG. 6   b  is a graphical depiction of turbine speed during the shift aberration “flare”;  
         [0022]      FIG. 6   c  is a graphical depiction of turbine speed during the shift aberrations “short shift,” “long shift,” “closed loop increase,” and “closed loop decrease”;  
         [0023]      FIG. 6   d  is a graphical depiction of turbine speed during the shift aberration “underlap turbine float”;  
         [0024]      FIG. 7  is a block diagram illustrating a method of adjusting the off-going pressure adaptive parameter of the present invention; and  
         [0025]      FIG. 8  is a block diagram illustrating a method of adjusting the on-coming volume adaptive parameter of the present invention. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0026]     The control of this invention is described in the context of a multi-ratio power transmission having a planetary gear set 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, both of which are hereby incorporated by reference in their entireties. Accordingly, the gear set 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.  
         [0027]     Referring to  FIG. 1 , the reference numeral  10  generally designates a vehicle power train including engine  12 , transmission  14 , and a torque converter  16  providing a fluid coupling between engine  12  and transmission input shaft  18 . It should be appreciated that while the invention will be described as being used with a conventional engine  12 , alternate power sources such as an electric motor or hybrid electric/gas motor may be implemented as well.  
         [0028]     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.  
         [0029]     The transmission  14  has three inter-connected planetary gear sets, designated generally by the reference numerals  23 ,  24  and  25 . The planetary gear set  23  includes a sun gear member  28 , a ring gear member  29 , and a planet carrier assembly  30 . The planet carrier assembly  30  includes a plurality of pinion gears rotatably mounted on a carrier member and disposed in meshing relationship with both the sun gear member  28  and the ring gear member  29 . The planetary gear set  24  includes a sun gear member  31 , a ring gear member  32 , and a planet carrier assembly  33 . The planet carrier assembly  33  includes a plurality of pinion gears rotatably mounted on a carrier member and disposed in meshing relationship with both the sun gear member  31  and the ring gear member  32 . The planetary gear set  25  includes a sun gear member  34 , a ring gear member  35 , and a planet carrier assembly  36 . The planet carrier assembly  36  includes a plurality of pinion gears rotatably mounted on a carrier member and disposed in meshing relationship with both the sun gear member  34  and the ring gear member  35 .  
         [0030]     The input shaft  18  continuously drives the sun gear  28  of gear set  23 , selectively drives the sun gears  31 ,  34  of gear sets  24 ,  25  via clutch C 1 , and selectively drives the carrier  33  of gear set  24  via clutch C 2 . The ring gears  29 ,  32 ,  35  of gear sets  23 ,  24 ,  25  are selectively connected to ground  42  via clutches (i.e., brakes) C 3 , C 4  and C 5 , respectively.  
         [0031]     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 . Downshifting from one 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 downshifted from second to first by disengaging clutch C 4  while engaging clutch C 5 .  
         [0032]     The torque converter clutch  19  and the transmission clutches C 1 -C 5  are controlled by an electro-hydraulic control system, generally designated by 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 ,  60 ,  62  and  64 .  
         [0033]     The electronic portion of the electro-hydraulic control system  44  is primarily embodied in the transmission control unit  66 , or controller, which is microprocessor-based and conventional in architecture. The transmission control unit  66  controls the solenoid-operated fluid control valves  58 - 64  based on a number of inputs  68  to achieve a desired transmission speed ratio. Such inputs include, for example, signals representing the transmission input speed TIS, a driver torque command TQ, the transmission output speed TOS, and the hydraulic fluid temperature Tsump. Sensors for developing such signals may be conventional in nature, and have been omitted for simplicity.  
         [0034]     The control lever  82  of manual valve  56  is coupled to a sensor and display module  84  that produces a 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, 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.  
         [0035]     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, fluid control valves  60  are a set of three on/off relay valves, shown in  FIG. 1  as a consolidated block, and are utilized in concert with manual valve  56  to enable controlled engagement and disengagement of each of the clutches C 1 -C 5 . Valves  62 ,  64  are of the modulated type. 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 one of the modulated valves  62 ,  64  to the off-going clutch.  
         [0036]     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 motors (not shown). Fluid control 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 to shift from one speed ratio to another, develops corresponding force motor current commands, and then supplies current to the respective force motors in accordance with the current commands. Thus, the clutches C 1 -C 5  are responsive to the pressure commands via the valves  58 - 64  and their respective actuating elements (e.g., solenoids, current-controlled force motors).  
         [0037]     As indicated above, each shift from one speed ratio to another includes a fill or preparation phase during which an apply chamber  91  of the on-coming clutch is filled in preparation for torque transmission. Fluid supplied to the apply chamber compresses an internal return spring (not shown), thereby stroking a piston (not shown). Once the apply chamber is filled, the piston applies a force to the clutch plates, developing torque capacity beyond the initial return spring pressure. Thereafter, the clutch transmits torque in relation to the clutch pressure, and the shift can be completed using various control strategies. The usual control strategy involves commanding a maximum on-coming clutch pressure for an empirically determined fill time, and then proceeding with the subsequent phases of the shift. The volume of fluid required to fill an apply chamber and thereby cause the clutch to gain torque capacity is referred to as the “clutch volume.” 
         [0038]     The controller  66  determines the timing of the pressure commands based on an estimated on-coming clutch volume, i.e., an estimated volume of fluid required to fill the on-coming clutch apply chamber and thereby cause the on-coming clutch to gain torque capacity. An estimated on-coming clutch volume must be used because the actual on-coming clutch volume may vary over time as a result of wear, and may vary from transmission to transmission because of build variations and tolerances.  
         [0039]     The controller  66  calculates an estimated volume of fluid supplied to the on-coming clutch apply chamber as the chamber is being filled based on a mathematical model of the transmission hydraulic system, and compares the estimated volume of fluid supplied to the estimated clutch volume. When the estimated volume of fluid supplied to the apply chamber equals the estimated clutch volume, then the on-coming clutch should gain capacity. A hydraulic flow model for use in estimating the volume of fluid supplied to an apply chamber is described in U.S. Pat. No. 6,285,942, issued Sep. 4, 2001 to Steinmetz et al, which is hereby incorporated by reference in its entirety. The model inputs include the fill pressure, the shift type ST (for example, a 2-1 downshift), the speed of pump  46 , and the temperature Tsump of the hydraulic fluid. The output of the model is the on-coming clutch flow rate. The flow rate is integrated by an integrator to form the estimated cumulative volume of fluid supplied to the apply chamber. In a preferred embodiment, the controller  66  subtracts the estimated volume of fluid supplied from the estimated clutch volume to determine an estimated clutch volume remaining. If the controller is accurate, the estimated clutch volume remaining will be zero at the time the on-coming clutch gains torque capacity.  
         [0040]     Alternatively, instead of modulated valves  62 ,  64  and relay valves  60 , the transmission may include a plurality of individual control valves each operatively connected to a respective apply chamber  91 . Referring to  FIG. 1A , an exemplary fluid control valve  90  includes a regulator  92 , a solenoid  94  and a pressure sensor  96 . Each control valve  90  is configured to provide fluid to the apply chamber  91  of its respective clutch C 1 -C 5  at either a full feed state or a regulating state.  
         [0041]     As shown in  FIG. 2   a,  a method of the present invention calculates an optimal off-going clutch pressure during the inertia phase of a power-on downshift. The method shown in  FIG. 2   a  and described hereinafter is predicated on the assumption that the output acceleration to output torque ratio does not change during the ratio change. Additionally, for purposes of this disclosure the derivative of a reference character is represented by the reference character with a dot thereabove as is well known in the field of mathematics. For example, the reference character n t  represents turbine speed and the reference character {dot over (n)} t  represents the first derivative of turbine speed which is also known as turbine acceleration.  
         [0042]     At step  100 , the desired shift time is applied to establish a desired turbine acceleration profile as will be described in detail hereinafter. At step  102 , the desired turbine acceleration {dot over (n)} t  and the current transmission input torque Ti are used to calculate the corresponding desired output torque To_Blend, and the output torque is modified by a scalar to the value of the desired output torque. The scalar is a calibration allowing for different combinations of clutch torque and input torque during the inertia phase such that shift time is maintained. In other words the scalar may be calibrated to provide either a firm shift or a more gentle shift during the same shift time. After output torque has been modified, a corresponding clutch torque Tcl_Blend is calculated. At step  104 , clutch torque is limited and this limited torque value is used to recalculate input torque Ti_Clamp and output torque To_Clamp such that shift time is maintained. Also at step  104 , the recalculated input torque Ti_Clamp is adjusted by a multiplication factor representative of the torque converter and sent to the engine control module  107 . At step  106 , available turbine acceleration is calculated and limited to a final turbine acceleration value A final  described hereinafter. At step  108 , clutch torque and output torque are calculated with the limited turbine acceleration value established in step  106 . At step  110  clutch torque is converted to a pressure value.  
         [0043]     The turbine acceleration profile established at step  100  is shown in  FIG. 3 . More precisely,  FIG. 3   a  depicts a desired input acceleration trajectory for the inertia phase of a power-on downshift from an attained gear speed Ag to a commanded gear speed Cg, assuming a constant output acceleration during the shift, and  FIG. 3   b  depicts a corresponding input speed profile. As seen in  FIG. 3   b,  the input speed prior to the inertia phase is determined by the product (output speed)×Ag, whereas the input speed at the conclusion of the inertia phase is determined by the product (output speed)×Cg.  
         [0044]     The parameters of the acceleration trajectory of Graph A include the initial acceleration A init , the maximum acceleration A max , the final acceleration A final , and the times t init , t final , and t shift . The terms A final , t init , t final , and t shift  are determined by calibration as a function of one or more other parameters. For example, t shift  may be determined as a function of driver torque demand, whereas t init  and t final  may be predetermined percentages of t shift . The value of A final  is a calibrated value selected to achieve smooth shift completion. A init  is the turbine speed measured prior to a shift event. A max  is computed based on the acceleration trajectory parameters and speed difference across the on-coming clutch, referred to herein as the slip speed.  
         [0045]     The calculations performed in step  102  of  FIG. 2   a  start with the following two basic equations: 
 
 {dot over (n)}   t   =a   t   T   i   +b   t   T   cl   +c   t   T   o  
 
 {dot over (n)}   o   =a   o   T   i   +b   o   T   cl   +c   o   T   o  
 
         [0046]     The calibration constants a t , b t , c t , a o , b o , and c o  are derived by summing the forces about the components of a particular transmission. As an example,  FIG. 4  shows a free body diagram of an arbitrary transmission for which the calibration constants will be derived.  
         [0047]      FIG. 4  schematically illustrates a six-speed planetary transmission  150 . The transmission  150  includes an input shaft  152  connected directly with an engine (not shown), a multi-speed planetary gear arrangement  154 , and an output shaft  156  connected directly with final drive mechanism (not shown). Planetary gear arrangement  154  includes a compound planetary gearset  158 , two simple planetary gearsets  160  and  162 , three selectively engageable rotating torque transmitting mechanisms  164 ,  166  and  168  and a selectively engageable stationary torque transmitting mechanism  170 . In a preferred embodiment, the planetary gear arrangement  154  includes a 1-2 overrunning clutch “OWC”  172  installed between stationary housing  174  and common carrier assembly  176 , and a modified low/reverse starting clutch  178 .  
         [0048]     The first planetary gearset  158  is shown to include a sun gear  180 , a ring gear  182 , and a planet carrier assembly  176 . Meshed pairs of pinion gears  184  and  186  are rotatably supported on pinion shafts  188  and  190 , respectively, that extend between laterally-spaced carrier segments of carrier assembly  176 . Pinion gears  184  mesh with sun gear  180  while pinion gears  186  mesh with ring gear  182 .  
         [0049]     The second planetary gearset  160  includes a sun gear  192 , a ring gear  194 , and a plurality of pinion gears  196  that are meshed with both sun gear  192  and ring gear  194 . As seen, pinion gears  196  are rotatably supported on pinion shafts  188  that also extend between the laterally-spaced carrier segments of carrier assembly  176 . Thus, carrier assembly  176  is common to both first planetary gearset  158  and second planetary gearset  160 . A ring gear assembly  198  is defined by ring gear  182  of first gearset  158  and ring gear  194  of second planetary gearset  160  being connected together to rotate as a unitary component. Third planetary gearset  168  is shown to include a sun gear  200 , a ring gear  202 , and pinion gears  204  in meshed engagement with both sun gear  200  and ring gear  202 . Pinion gears  204  are rotatably supported on shafts  206  extending between components of a carrier assembly  208 . In addition, sun gear  200  is shown to be held stationary due to its direct connection to a stationary housing portion  174  of transmission  150 .  
         [0050]     The calibration constants a t , b t , c t , a o , b o , and c o  can be solved for the transmission of  FIG. 4  using Newton&#39;s second law for rotational dynamics and summing the forces at the input and output of each component. The equations derived in this manner from the transmission of  FIG. 4  are as follows:  
           I   202     *       ω   .     202       =       -     T   202       +     (       T   i     +     T   168       )           
           I   200     *       ω   .     200       =       T   200     -     T   ground           
           I   208     *       ω   .     208       =       T   202     -     T   200     -     T   164     -     T   166           
           I   204     *       ω   .     204       =           T   202       N   202       ⁢     (         N   202     -     N   200       2     )       +         T   200       N   200       ⁢     (         N   202     -     N   200       2     )             
           N   200     *       ω   .     200       =           ω   .     208     ⁡     (         N   202     +     N   200       2     )       -         ω   .     204     ⁡     (         N   202     -     N   200       2     )             
           N   202     *       ω   .     202       =           ω   .     208     ⁡     (         N   202     +     S   200       2     )       +         ω   .     204     ⁡     (         N   202     -     N   200       2     )             
           I   180     *       ω   .     180       =       -     T   180       +     T   164           
           I   176     *       ω   .     176       =       T   180     -     T   182     -     T   192     -     T   178     -     T   172     +     T   168           
           I   180     *       ω   .     180       =       -     T   180       +     T   164           
           I   192     *       ω   .     192       =       T   192     +     T   166     -     T   170           
           I   182     *       ω   .     182       =       T   182     -     T   156           
           I   184     *       ω   .     184       =         N   184     ⁢     F   184       -       T   180     ⁡     (       N   184       N   180       )             
                 I   196     *       ω   .     196       =       ⁢         -       T   182       N   182         ⁢     (         N   182     -     N   192       2     )       +         T   192       N   192       ⁢     (         N   182     -     N   192       2     )       +                     ⁢       F   184     ⁡     (         N   182     -     N   190       2     )                 
           N   192     *       ω   .     192       =           ω   .     176     ⁡     (       N   182     +     N   192       )       -         ω   .     182     ⁢     N   182             
           N   182     *       ω   .     182       =         N   180     *       ω   .     180       +         ω   .     a1     ⁡     (       N   182     -     N   180       )             
           N   182     *       ω   .     182       =           ω   .     196     ⁡     (         N   182     -     N   192       2     )       +       N   182     *       ω   .     176             
           N   184     *       ω   .     184       =         ω   .     186     ⁡     (         N   182     -     N   192       2     )           
           ω   .     208     =       ω   .     180         
 
 where T is a torque value, I is inertia, F is force, ω is rotational velocity, {dot over (ω)} is rotational acceleration and N is the number of teeth on a particular gear element. {dot over (n)} and {dot over (ω)} are both rotational acceleration values but are differentiated in that {dot over (n)} is measured in rpm/second 2  whereas {dot over (ω)} is measured in radians/second 2 . 
 
         [0051]     Having solved for the calibration constants associated with a particular transmission, corresponding values for {dot over (n)} t  and {dot over (n)} o  are calculated from the two basic equations provided hereinabove. At step  102  of  FIG. 2   a,  the values of {dot over (n)} t  and {dot over (n)} o  are then input into the following equation to solve for T o     —   Blend.  
           T   o     ⁢   _Blend     =           n   .     t     -       (       a   t     -       b   t     ⁢       a   o     /     b   o           )     ⁢     T   i           (           b   t     /       b   o     ⁡     (         n   .     o     /     T   o       )         ⁢   i     -       b   t     ⁢       c   o     /     b   o         +     c   t       )           
 
 It should be appreciated that assumption made for the torque phase of the ratio change described hereinabove, specifically that the output acceleration to output torque ratio does not change during the ratio change, is embodied by the term ({dot over (n)} o /T o ) i. Therefore this term becomes a constant measured only at the beginning of the ratio change. 
 
         [0052]     As the value of {dot over (n)} derived from the free body diagram of the transmission was based on the desired shift time, the corresponding value of To_Blend is similarly scaled to meet the desired shift time.  
         [0053]     At step  102  of  FIG. 2   a,  the value of To_Blend scaled to meet desired shift time is then input into the following equation to solve for Tcl_Blend, which is thereby also scaled to follow both the desired shift time and the scaled output torque.  
           T   cl     ⁢   _Blend     =       {         [         (         n   .     o     /     T   o       )     i     +       a   o     ⁢       c   t     /     a   t         -     c   o       ]     ⁢     T   o       -         a   o     /     a   t       *       n   .     t         }       (       b   o     -       a   o     ⁢       b   t     /     a   t           )           
 
         [0054]     At step  104  of  FIG. 2   a,  a limited value of output torque To_Clamp is recalculated with the limited value of clutch torque Tcl according to the equation:  
           T   o     ⁢   _Clamp     =       [         T   cl     ⁡     (       b   o     -       a   o     ⁢       b   t     /     a   t           )       +         a   o     /     a   t       *       n   .     t         ]       [         (         n   .     o     /     T   o       )     i     +       a   o     ⁢       c   t     /     a   t         -     c   o       ]           
 
         [0055]     The recalculated value of output torque T o     —   Clamp and the limited value of clutch torque T cl     —   Blend are input into the following equation to derive a base input torque T i     —   Clamp required to achieve the desired shift time.  
           T   i     ⁢   _Clamp     =           n   .     t     -       b   t     ⁢     T   cl       +       c   t     ⁢     T   o           a   t           
 
         [0056]     This value of input torque is limited to levels that the engine can produce, which thereby may necessitate modification of the desired shift time.  
         [0057]     At step  106  of  FIG. 2   a,  the limited input torque T i     —   Clamp and, if necessary, the modified desired shift time are input into the following equation to generate an attainable turbine acceleration {dot over (n)} t . 
 
 {dot over (n)}   t   ={a   t   +c   t   a   o /[( {dot over (n)}   o   /T   o ) i   −c   o   ]}*T   i   +{b   t   +c   t   b   o /[( {dot over (n)}   o   /T   o ) i   −c   o   ]}*T   cl  
 
         [0058]     At step  108  of  FIG. 2   a,  values for clutch torque and output torque required to meet constraints identified hereinabove are respectively calculated according to the following equations:  
         T   cl     =           [         (         n   .     o     /     T   o       )     i     -     c   o       ]     ⁢       n   .     t       -       a   t     ⁢       T   i     ⁡     [         (         n   .     o     /     T   o       )     i     -     c   o       ]         -       c   t     ⁢     a   o     ⁢     T   i           {         [         (         n   .     o     /     T   o       )     i     -     c   o       ]     ⁢     b   t       +       c   t     ⁢     b   o         }           
         T   o     =         [         T   cl     ⁡     (       b   o     -       a   o     ⁢       b   t     /     a   t           )       +         a   o     /     a   t       *       n   .     t         ]       [         (         n   .     o     /     T   o       )     i     +       a   o     ⁢       c   t     /     a   t         -     c   o       ]       .         
 
         [0059]     At step  110 , the torque value for the off-going clutch T cl  is converted to a pressure value P off .  
         [0060]      FIG. 2   b  illustrates a method for calculating an optimal value for transmission input torque during the torque phase of the ratio change. Engine output may then be altered by an amount necessary to change the actual value of the transmission input torque to the calculated optimal value of transmission input torque.  FIG. 2   b  is distinguishable from  FIG. 2   a  in part because  FIG. 2   a  is implemented during the inertia phase and  FIG. 2   b  is implemented during the torque phase of a shift event.  
         [0061]     At step  112  of  FIG. 2   b,  the off-going clutch torque T cl  calculated according to the method of  FIG. 2   a  is ramped to zero over the duration of the torque phase time to produce a ramped off-going clutch torque T off . At step  114 , which is performed generally simultaneously with step  112 , on-coming clutch torque T oncl  is ramped from a calibration threshold to a value representing the holding torque for the next gear ratio over the duration of the torque phase time. The ramped on-coming clutch torque derived at step  114  is identified by reference character T on . At step  116 , the torque phase input torque T i  is calculated. Also at step  116 , the recalculated input torque Ti (Desired) is adjusted by a multiplication factor representative of the torque converter and sent to the engine control module  107 . At step  118 , the torque values for the on-coming and off-going clutch T on  and T off  are converted to corresponding pressure values P on  and P off . In an alternate embodiment, at step  118  the torque value off-going clutch T off  is converted to corresponding pressure value P off , and the pressure value used for P on  is that which is achieved by filling the on-coming clutch at the maximum fill rate. At step  120 , the torque phase output torque T o  is calculated.  
         [0062]     At step  116  of  FIG. 2   b,  the following two equations are used to calculate the torque phase input torque T i : 
 
 T   on   =k   61   T   o   +k   62   T   i   +k   63   {dot over (n)}   t  
 
 T   off   =k   64   T   o   +k   65   T   i   +k   66   {dot over (n)}   t  
 
         [0063]     The values k 61 , k 62 , k 63 , k 64 , k 65  and k 66  are calibration constants which are solved for a particular transmission in a manner similar to that described hereinabove for the calibration constants a t , b t , c t , a o , b o , and c o . Input torque is then solved for using the equation:  
         T   i     =       -     (           k   61     /     k   64       *     T   off       -     T   on     -       k   66     ⁢       k   61     /     k   64       *       n   .     t       +       k   63     *       n   .     t         )         (         -     k   61       ⁢       k   65     /     k   64         +     k   62       )           
 
         [0064]     The value of input torque derived from this equation represents the amount of engine torque necessary at synchronization. In a preferred embodiment, a reduction of engine torque is accomplished by spark arrest and an increase of engine torque is accomplished by opening the throttle. It should be appreciated, however, that there are numerous methods for increasing and/or decreasing engine torque.  
         [0065]     The method of the present invention establishes two adaptive parameters for each power-on downshift. The adaptive parameters include an off-going clutch pressure adaptive parameter, and an on-coming clutch volume adaptive parameter. The adaptive parameters are so named because they are monitored and may be adapted to improve subsequent downshifts.  
         [0066]      FIGS. 5   a - 5   c  show a predefined optimal power-on downshift. More precisely,  FIG. 5   a  shows an optimal torque converter turbine speed n t1  transitioning from the attained gear speed Ag to the commanded gear speed Cg. Those skilled in the art will recognize that the turbine and input shaft are interconnected, and, accordingly, the turbine speed is the same as the input shaft speed. Those skilled in the art will also recognize that the attained gear speed Ag is the transmission output speed multiplied by the currently selected gear ratio, whereas the commanded gear speed Cg is the transmission output speed multiplied by the commanded gear ratio. Accordingly, during a power-on 4-3 downshift, Ag is transmission output speed multiplied by the fourth gear ratio and Cg is the transmission output speed multiplied by the third gear ratio.  
         [0067]      FIG. 5   c  shows on-coming clutch pressure during the power-on downshift, including the fill phase in which the on-coming clutch apply chamber is filled and wherein on-coming torque is zero. Similarly,  FIG. 5   b  shows off-going clutch pressure during the power-on downshift. During an optimal power-on downshift, there is zero on-coming clutch torque until turbine speed Ts reaches the point of synchronization identified in  FIG. 5   a.  It should also be appreciated that the point of synchronization also represents the beginning of the torque phase.  
         [0068]     Torque applied by the off-going clutch is preferably converted from off-going clutch pressure according to a table. The table provides a torque versus pressure curve defined by multiple points or cells. In a preferred embodiment, the table is a three-place table defined by three cells. This provides flexibility by allowing adaptive correction of the torque to pressure relationship at a specific point on the curve without altering the remainder of the curve. In other words, only the cells attributable to a particular aberration are updated.  
         [0069]     As seen in  FIG. 5   c,  at the point of synchronization on-coming clutch pressure is equal to the pressure applied by the on-coming clutch return spring (not shown) and zero torque is therefore being applied by the on-coming clutch. Immediately after the point of synchronization, the on-coming clutch is generating some torque but not enough to prevent a past-synchronization condition, hereinafter called engine flare, wherein the turbine speed n t  exceeds the commanded gear speed Cg. The method of the present invention therefore implements engine torque management at the point of synchronization to prevent engine flare.  
         [0070]     The shift aberrations, i.e., deviations, from the predefined optimal shift of  FIG. 5   a  that are correctable by adjusting the off-going pressure adaptive parameter are graphically represented in  FIGS. 6   a - c.  In  FIG. 6   a,  turbine speed n t2  represents the shift aberrations “slip early” and turbine speed n t3  represents the shift aberration “slip late.” Slip early and slip late are both potentially attributable to inadequate off-going clutch pressure.  
         [0071]     Deviation of turbine speed T s  from attained gear speed A g  is monitored by the control unit to determine the occurrence of slip early or slip late. If turbine speed n t  prematurely rises more than a predetermined amount, e.g., 50 rpm, above attained gear speed A g , slip early is indicated. Conversely, if turbine speed n t  is delayed in rising more than a predetermined amount, e.g., 50 rpm, above attained gear speed A g , slip late is indicated.  
         [0072]     As shown in  FIG. 6   b,  flare is a shift aberration wherein the turbine speed n t4  rises more than a predetermined amount, e.g., 50 rpm, above commanded gear speed C g .  
         [0073]     The turbine speed during a short shift and a long shift are graphically depicted by line n t5  and line n t6  of  FIG. 6   c,  respectively, and are contrasted by the solid line representation of turbine speed n t1  during the predefined optimal power-on downshift. A short shift or long shift is identified by comparing the duration of the inertia phase with a predetermined optimal shift time. The duration of the inertia phase is the period of time beginning when the turbine speed is a predetermined amount, e.g., 50 rpm, greater than the attained gear speed Ag and ending when the turbine speed is a predetermined amount, e.g., 50 rpm, less than the commanded gear speed Cg. Insufficient inertia phase duration, i.e., in comparison to the predetermined optimal shift time, is indicative of a short shift, and excessive inertia phase duration is indicative of a long shift.  
         [0074]     The controller is configured for closed-loop control of commanded pressure. Accordingly, the controller is configured to recognize deviation between intended pressure and actual pressure based on deviation between actual turbine speed and intended turbine speed. Previously addressed shift aberrations are detected by the controller comparing the actual characteristics of a shift to a predefined optimal shift. The controller is further configured to analyze information obtained from the closed loop control to adjust the off-going pressure adaptive parameter accordingly.  
         [0075]     The turbine speed during a closed loop increase and a closed loop decrease is graphically similar to a short shift and long shift, respectively. Therefore, referring to  FIG. 6   c,  the turbine speed during a closed loop increase is graphically depicted by line n t5 , and the turbine speed during a closed loop decrease is graphically depicted by line n t6 . As error between actual turbine speed profile and intended turbine speed profile increases, the closed loop control causes the commanded pressure to proportionally increase to correct the error. A “closed loop increase” or a “closed loop decrease” occurs when the commanded pressure increases or decreases by more than a predetermined maximum threshold.  
         [0076]     A method for addressing the shift aberrations identified hereinabove by adjusting the off-going pressure adaptive parameter is shown in  FIG. 7 . At step  121 , if slip early is detected the off-going pressure adaptive parameter is increased. At step  122 , if either a short shift or a closed loop increase is detected and flare is detected, the off-going pressure adaptive parameter is increased. According to a preferred embodiment of the present invention, the off-going adaptive parameter is the multi-place table described hereinabove and the cells are increased at steps  121  and  122  in proportion to their degree of responsibility for a particular aberration. At step  124 , if slip late is detected the off-going pressure adaptive parameter is decreased. At step  126 , if either a long shift or a closed loop decrease is detected the off-going pressure adaptive parameter is decreased. At step  128 , if either a short shift or a closed loop increase is detected and the criteria for steps  120 - 126  are not met, the off-going pressure adaptive parameter is decreased. At step  130 , if flare is detected and there is not a short shift or closed loop increase, the off-going pressure adaptive parameter is increased. At step  132 , if the criteria for steps  120 - 130  are not met the off-going pressure adaptive parameter is incrementally decreased to produce flare as will be described in detail hereinafter.  
         [0077]     To correct the adaptive parameter at step  132 , the off-going pressure adaptive parameter is revised after a predetermined number of shifts during which the criteria for steps  120 - 130  are not met. More precisely, if a predetermined number of shifts occur without meeting the criteria for steps  120 - 130 , the low torque cell point of the multi-place off-going pressure adaptive parameter is incrementally reduced during subsequent shifts until any increase aberration is detected or minimum clamp is achieved, and thereafter the low torque cell point is incrementally increased until the aberration no longer exists. In the preferred embodiment of the present invention, the off-going pressure adaptive parameter is composed of the three-place table described hereinabove and the correction at step  132  is applied only to the low torque cell, however it should be appreciated that in alternate embodiments such correction may be applied to additional cells.  
         [0078]     The off-going pressure adaptive parameter is preferably increased or decreased according to the method of  FIG. 7  by a corrective value obtained by the following equation: (full correction)(scalar)(gain)(gain 2 ). Full correction is either a calibration or measured signal, such as from turbine speed, that gives a term to correct the adaptive problem. The scalar is a function of the shift aberration type, since some shift aberrations require more aggressive corrective action than others. The gain is related to an adaptive error counter that tracks the direction the off-going pressure adaptive parameter is moving. Gain 2  is a variable adapted to assign a weighted correction to the specific cells in the off-going clutch multi-place adaptive that are attributable to a given aberration. Accordingly, Gain 2  corrects the cells of the off-going clutch multi-place adaptive in proportion to their degree of responsibility for the aberration such that the correction is not necessarily evenly applied. If the off-going pressure adaptive parameter increases during consecutive downshifts, the adaptive error counter is increased by one each shift to a predetermined maximum value, e.g., seven. Similarly, if the off-going pressure adaptive parameter decreases during consecutive downshifts, the adaptive error counter is decreased by one each shift to a predetermined minimum value, e.g., negative seven. The gain is established based on the adaptive error counter value such that the magnitude of the gain is proportional to the absolute value of the adaptive error counter. In other words, each consecutive increase or decrease in the adaptive error counter gives rise to a larger gain. In this manner the degree of adaptive correction can be increased if the off-going pressure adaptive parameter has been commanded to change in one direction, i.e., increased or decreased, during consecutive downshifts. Thus, the corrective value varies in response to the quantity of consecutive monitored downshifts in which a shift aberration occurs. If the off-going pressure adaptive parameter is increased and then subsequently decreased, or vice versa, the adaptive error counter is reset to zero and the gain becomes its minimal value. Additionally, it should be appreciated that the volume adaptive parameters are increased and decreased in a similar manner.  
         [0079]     Having described the off-going pressure corrections in detail hereinabove, the following will discuss the on-coming volume adaptive parameter. Referring to  FIG. 6   d,  a shift aberration, i.e., deviation, from the predefined optimal shift of  FIG. 5   a  that is correctable by adjusting the on-coming volume is shown. More precisely,  FIG. 6   d  shows the aberration underlap turbine float which is a shift aberration wherein the turbine speed n t7  floats at a value below the commanded gear speed Cg and thereby fails to reach the commanded gear speed Cg in the desired time.  
         [0080]     A method for adjusting the on-coming volume adaptive parameter is shown in  FIG. 8 . At step  134 , the on-coming volume adaptive parameter is increased when flare is detected, neither short shift nor closed loop increase are detected, the commanded gear turbine acceleration is below a predefined minimum value, and slip early is not detected. At step  136 , the on-coming volume adaptive parameter is increased when flare control is invoked. Flare control is invoked when T s  exceeds the commanded gear speed Cg by an amount deemed excessive. In a preferred embodiment of the present invention, flare control is invoked when T s  exceeds the commanded gear speed Cg by more than, for example, 300 rpm. At step  138 , the on-coming volume adaptive parameter is decreased when there is an extreme short shift detected and neither flare nor long shift are detected, or when underlap turbine float is detected and slip late is not detected. At step  140 , the on-coming volume adaptive parameter is incrementally decreased. The incremental decrease at step  132  is performed in the same manner as that described hereinabove for the off-going pressure adaptive parameter at step  122 . Additionally, it should be appreciated that the incremental decrease of steps  132  and  122  are preferably configured to alternate during subsequent shifts such that only one or the other is performed during a single shift.  
         [0081]     The aberrations flare and short shift may be attributable to either inadequate pressure or inadequate calculated volume. Therefore at step  134  there is an upper limit for the on-coming volume adaptive parameter intended to prevent on-coming volume correction of an aberration caused by an incorrect off-going pressure value. More precisely, an aberration that suggests an increase of the learned volume above the maximum limit is likely attributable to off-going pressure rather than volume, and the problem is addressed by the off-going pressure adaptive described hereinabove. The limit applied at step  134  is preferably implemented with a pressure switch (not shown) adapted to inhibit an increase in on-coming volume above a predefined maximum value. In this manner the off-going pressure and on-coming volume adaptives work together to identify which is responsible for the aberration and thereafter address the aberration in the appropriate manner.  
         [0082]     While the best mode for carrying out the invention has been described in detail, those familiar with the art to which this invention relates will recognize various alternative designs and embodiments for practicing the invention within the scope of the appended claims.