Patent Publication Number: US-7212898-B2

Title: Method and apparatus for adaptive control of power-on skip through neutral downshifts

Description:
TECHNICAL FIELD 
     The present invention relates to a method and apparatus for improving power-on downshifts that skip through neutral in an automatic transmission. 
     BACKGROUND OF THE INVENTION 
     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. 
     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 one or more clutches (off-going) associated with the current speed ratio and engaging one or more clutches (on-coming) associated with the desired speed ratio. Downshifts that skip through neutral include those wherein two off-going clutches associated with the current speed ratio are released and two on-coming clutches associated with the desired speed ratio are engaged during the ratio change such as, for example, a sixth gear to third gear ratio change or a fifth gear to second gear ratio change. 
     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 two clutches associated with the lower speed ratio and engaging two clutches 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 require precise timing in order to achieve high quality shifting. 
     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 
     The invention provides a method and apparatus for calculating optimal values for transmission input torque during the inertia phase and the torque phase of the ratio change, and thereafter adaptively controlling a power-on downshift that skips through neutral in an automatic transmission wherein a transmission aberration during a shift is diagnosed and corrected during subsequent downshifts. 
     The method of the invention is carried out by mathematically calculating optimal values for transmission input torque during the inertia phase and the torque phase of the ratio change. Additionally, the method of the invention calculates when to release the two off-going clutches, as well as when and how quickly to engage the two on-coming clutches to optimize the power-on skip through neutral downshift. 
     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. 
     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 
         FIG. 1  is a schematic illustration of an automatic transmission; 
         FIG. 1   a  is a schematic illustration of a valve of  FIG. 1 ; 
         FIG. 2   a  is a block diagram illustrating a method for calculating desired transmission input torque during the inertia phase of a ratio change; 
         FIG. 2   b  is a block diagram illustrating a method for calculating desired transmission input torque during the torque phase of a ratio change; 
         FIG. 3   a  is a graphical depiction of turbine acceleration vs. time during an optimal downshift; 
         FIG. 3   b  is a graphical depiction of turbine speed vs. time during an optimal downshift; 
         FIG. 4  is a schematic illustration of an automatic transmission; 
         FIG. 5  is a graphical depiction of clutch pressure vs. time during the optimal downshift of  FIG. 3   b ; 
         FIG. 6   a  is a graphical depiction of turbine speed during the shift aberrations “slip early” and “slip late”; 
         FIG. 6   b  is a graphical depiction of turbine speed during the shift aberration “flare”; 
         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”; 
         FIG. 6   d  is a graphical depiction of turbine speed during the shift aberration “underlap turbine float”; and 
         FIG. 7  is a block diagram illustrating a method of adjusting a primary on-coming volume adaptive parameter of the present invention. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     The method and apparatus for improving power-on downshifts that skip through neutral in an automatic transmission may be implemented with the pending patent application entitled “METHOD AND APPARATUS FOR ADAPTIVE CONTROL OF POWER-ON DOWNSHIFTS IN AN AUTOMATIC TRANSMISSION”, filed Oct. 22, 2004, to Matthew Whitton, Ser. No. 10/972,067, which is hereby incorporated by reference in its entirety. 
     Power-on downshifts include downshifts that take place while engine torque is positive. Downshifts that skip through neutral include those wherein two off-going clutches associated with the current speed ratio are released and two on-coming clutches associated with the desired speed ratio are engaged during the ratio change such as, for example, a sixth gear to third gear ratio change (6-3 ratio change) or a fifth gear to second gear ratio change (5-2 ratio change). 
     It should also be appreciated that the present invention may apply to any other ratio change by inducing a skip through neutral condition. A skip through neutral condition may be induced by releasing a common clutch element during the ratio change and thereafter re-engaging the common clutch. A common clutch is defined for purposes of the present invention as a clutch common to both the attained gear and the commanded gear. For example, during a 6-2 ratio change, the clutch C 4  is common to both sixth gear and second gear, and remains engaged during a downshift that does not skip through neutral. A skip through neutral condition may, however, be induced during the 6-2 downshift by releasing the C 4  clutch along with the other off-going clutch (C 2  clutch) and thereafter engaging the C 4  clutch and the other on-coming clutch (C 1  clutch) in a manner described in detail hereinafter. 
     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. 
     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. 
     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 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 . 
     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. 
     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 . Skip through neutral downshifting from one forward speed ratio to another is achieved by disengaging two clutches (referred to as the off-going clutches) associated with the current speed ratio while engaging two other clutches (referred to as the on-coming clutches) associated with the desired speed ratio. For example, the transmission  14  is downshifted from sixth gear to third gear by disengaging clutches C 2  and C 4 , and engaging clutches C 1  and C 3 . 
     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 . 
     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. 
     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. 
     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. 
     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 clutches while smoothly disengaging the off-going clutches 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). 
     As indicated above, each shift from one speed ratio to another includes a fill or preparation phase during which apply chambers (i.e. apply chamber  91 ) of the on-coming clutches are filled in preparation for torque transmission. Fluid supplied to the apply chambers compresses an internal return spring (not shown), thereby stroking a piston (not shown). Once the apply chambers are filled, the pistons apply a force to the clutch plates, developing torque capacity beyond the initial return spring pressure. Thereafter, the clutches transmit 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.” 
     The controller  66  determines the timing of the pressure commands based on an estimated volume for each on-coming clutch, i.e., an estimated volume of fluid required to fill each on-coming clutch apply chamber and thereby cause the on-coming clutches to gain torque capacity. Estimated on-coming clutch volumes must be used because the actual on-coming clutch volumes may vary over time as a result of wear, and may vary from transmission to transmission because of build variations and tolerances. 
     The controller  66  calculates an estimated volume of fluid supplied to each of the on-coming clutch apply chambers as the chambers are 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 one of the apply chambers equals the estimated clutch volume, then the respective 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. 
     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. 
     As shown in  FIG. 2   a , a method of the present invention calculates an optimal transmission input torque during the inertia phase of a power-on skip through neutral downshift. A power-on skip through neutral downshift is a subset of the more general power-on downshift. Accordingly, the method of the present invention initially derives a series of equations drawn to the more general power-on downshift disclosed in the incorporated patent application Ser. No. 10/972,067. The derived equations are thereafter solved using assumptions specifically applying to the inertia phase of power-on downshifts that skip through neutral. Such assumptions include that during the inertia phase of a power-on skip through neutral downshift the transmission output torque is zero and the torque applied by both off-going clutches is zero. 
     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 nt 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. 
     At step  100  of  FIG. 2   a , the desired shift time is applied to establish a desired turbine acceleration profile as will be described in detail hereinafter. The desired shift time is a predefined time period selected to provide optimal shift feel and, as best shown in  FIG. 3   b , is the time between the point at which the torque converter turbine speed n t1  first leaves the attained gear speed Ag and the point at which the torque converter turbine speed n t1  reaches the commanded gear speed Cg (also known as the point of synchronization). The following steps  102 – 104  will initially be described without applying the assumptions for power-on skip through neutral downshifts to explain the derivation of the equations, and thereafter the resultant equations will be solved using these assumptions. 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 . 
     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. 
     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. 
     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 
 
     The calibration constants a t , b t , c t , a o , b o , and c o  are derived by performing a torque analysis of the transmission to which the method of the present invention are being applied. More precisely, the torque analysis includes summing the forces at the input and output of each component of the transmission. As an example,  FIG. 4  shows a free body diagram of an arbitrary transmission for which the calibration constants will be derived. 
       FIG. 4  shows a six-speed planetary transmission  150 . Transmission  150  is shown to illustrate a method of the invention, and it should be appreciated that the method of the present invention applies to any transmission configuration. 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 . 
     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 . 
     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 . 
     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: 
     
       
         
           
             
               
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               * 
               
                 
                   ω 
                   . 
                 
                 182 
               
             
             = 
             
               
                 
                   N 
                   180 
                 
                 * 
                 
                   
                     ω 
                     . 
                   
                   180 
                 
               
               + 
               
                 
                   
                     ω 
                     . 
                   
                   
                     a 
                     ⁢ 
                     
                         
                     
                     ⁢ 
                     1 
                   
                 
                 ⁡ 
                 
                   ( 
                   
                     
                       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 . 
     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 To_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 
                   
                 
                 ) 
               
             
           
         
       
     
     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. 
     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 
                       
                     
                   
                 
                 ) 
               
             
           
         
       
     
     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 
                     
                   
                 
                 ] 
               
             
           
         
       
     
     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 
               
             
           
         
       
     
     This value of input torque is limited to levels that the engine can produce, which thereby may necessitate modification of the desired shift time. As indicated hereinabove, this equation is applicable to all power-on downshifts. The equation for Ti_Clamp as it applies to power-on skip through neutral downshifts may be obtained by incorporating the assumptions identified hereinabove. Therefore, assuming there is zero output torque and zero off-going clutch torque during the inertia phase of a power-on skip through neutral down shift, the equation for T i     —   Clamp becomes T i     —   Clamp={dot over (n)} t /a t . The value of T i     —   Clamp may be calculated in this manner and represents the optimal transmission input torque that meets the desired shift time. 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. 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 transmission input torque and that the present invention is not limited to applications incorporating engines but may be implemented in conjunction with any number of additional power sources as well. 
       FIG. 2   b  illustrates a method for calculating an optimal value for transmission input torque during the torque phase of the ratio change.  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. 
     The method of  FIG. 2   b  initially involves deriving a series of equations drawn to the more general power-on downshift disclosed in the incorporated patent application Ser. No. 10/972,067. The derived equations are thereafter solved by assuming that off-going clutch torque is zero during the torque phase of a power-on skip through neutral downshift. 
     As with the steps  102 – 104  of  FIG. 2   a , the following steps  112 – 120  will initially be described without applying the assumption for power-on skip through neutral downshifts to explain the derivation of the equations, and thereafter the resultant equations will be solved using this assumption. 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 torque phase 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  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 
 
     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 
                   
                 
                 ) 
               
             
           
         
       
     
     As indicated hereinabove, this equation is applicable to all power-on downshifts. The equation for T i  as it applies to the torque phase of a power-on skip through neutral downshift may be obtained by incorporating the assumptions identified hereinabove. Therefore, assuming there is zero off-going clutch torque during the torque phase of a power-on skip through neutral down shift, the equation for T i  becomes: 
     
       
         
           
             
               T 
               i 
             
             = 
             
               
                 - 
                 
                   ( 
                   
                     
                       - 
                       
                         T 
                         on 
                       
                     
                     - 
                     
                       
                         k 
                         66 
                       
                       ⁢ 
                       
                         
                           k 
                           61 
                         
                         / 
                         
                           k 
                           64 
                         
                       
                       * 
                       
                         
                           n 
                           . 
                         
                         t 
                       
                     
                     + 
                     
                       
                         k 
                         63 
                       
                       * 
                       
                         
                           n 
                           . 
                         
                         t 
                       
                     
                   
                   ) 
                 
               
               
                 ( 
                 
                   
                     
                       - 
                       
                         k 
                         61 
                       
                     
                     ⁢ 
                     
                       
                         k 
                         65 
                       
                       / 
                       
                         k 
                         64 
                       
                     
                   
                   + 
                   
                     k 
                     62 
                   
                 
                 ) 
               
             
           
         
       
     
     The value of T i  may be calculated in this manner and represents the optimal transmission input torque required to hold synchronization with minimal on-coming clutch torque. 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 using known physical characteristics of a torque converter such as, for example, converter torque ratio. 
     Another aspect of the present invention provides a method for optimally timing the two off-going and two on-coming clutches. To initiate a power-on skip through neutral downshift according to the method of the present invention both off-going clutches are simultaneously released. The release of the off-going clutches may be delayed by an amount calculated to meet the desired shift time as will be described in detail hereinafter. The engagement of a primary on-coming clutch is delayed by an amount calculated to maintain zero primary on-coming capacity until the point of synchronization and immediately thereafter generate positive torque. The engagement of a secondary on-coming clutch is similarly delayed by an amount calculated to reach full or nearly full capacity before the point of synchronization. Tie-up, a condition wherein the secondary on-coming clutch gains capacity before the off-going clutches are completely released, is prevented in the calculation of the desired shift time. More precisely, a desired shift time is selected that is long enough to prevent tie-up by providing adequate time to completely release the off-going clutches, and thereafter engage the secondary on-coming clutch such that it reaches full capacity prior to synchronization and within the desired shift time. 
     To calculate the off-going and on-coming clutch delay periods, the method of the present invention establishes an off-going delay adaptive parameter, a primary volume adaptive parameter and a secondary volume adaptive parameter (learned from a previous power-on downshift as described in detail hereinafter). The off-going delay adaptive parameter is an estimated value representing the expected amount of time between the commanded release of the two off-going clutches and slip. Slip, as will be described in detail hereinafter, is the point at which the turbine speed exceeds the attained gear speed to initiate the ratio change. The primary volume adaptive parameter and the secondary volume adaptive parameter are estimated values representing the volumes of the primary and secondary on-coming clutch apply chambers, respectively. The off-going delay and primary on-coming volume adaptive parameters are variable and may be adapted or corrected according to a method of the present invention to more accurately represent their respective estimated value when additional information becomes available and as described in detail hereinafter. The secondary on-coming volume adaptive parameter is learned from a previous power-on down shift that does not skip through neutral (has single off-going and on-coming clutches) and that includes the same clutch. For example, in a 6-3 power-on skip through neutral downshift the secondary on-coming clutch C 1  may be learned from a previous 5-4 power-on downshift. It should be appreciated that the off-going delay is proportional to input torque and may be adapted in a manner reflecting this relationship such as, for example, with a multi-place adaptive parameter. 
     The primary and secondary volume adaptive parameters may be converted to an estimate of the time required to fill the respective apply chambers based on a full-feed fill rate. An estimate of the time required to completely engage the secondary on-coming clutch can be calculated by adding the time required to fill the secondary on-coming apply chamber to a compliant element calibration. The compliant element calibration represents the minimum time required for the secondary on-coming clutch to reach full capacity after the apply chamber is full. Therefore, the minimum time required to perform a power-on skip through neutral downshift according to the method of the present invention may be defined by the off-going delay plus the greater of the two estimated values respectively representing the time required for the secondary on-coming clutch to reach full capacity, and the time required to completely fill the primary on-coming apply chamber. 
     If the desired shift time is long enough to allow the secondary on-coming clutch to reach full capacity prior to synchronization, and the desired shift time is also long enough to allow the primary on-coming apply chamber to be completely filled prior to synchronization, the timing of the clutch elements proceeds as follows. The power-on skip through neutral downshift is initiated by immediately releasing both off-going clutches. Actuation of the secondary on-coming clutch is delayed by an amount calculated to ensure full capacity prior to synchronization. Actuation of the primary on-coming clutch is delayed by an amount calculated to fill the primary on-coming apply chamber without generating any torque precisely at the point of synchronization such that immediately thereafter positive torque is generated. 
     If the desired shift time is long enough to allow the secondary on-coming clutch to reach full capacity prior to synchronization, but the desired shift time is not long enough to allow the primary on-coming apply chamber to be completely filled prior to synchronization, the timing of the clutch elements proceeds as follows. The control unit  66  (shown in  FIG. 1 ) begins filling the primary on-coming clutch apply chamber before the off-going clutches are released such that the primary on-coming apply chamber is completely filled precisely at the point of synchronization. Thereafter, the release of both off-going clutches is delayed by an amount adapted to meet the desired shift time. After the off-going clutches have been released, actuation of the secondary on-coming clutch is delayed by an amount calculated to ensure full capacity prior to synchronization. 
     If the desired shift time is long enough to allow the primary on-coming apply chamber to be completely filled prior to synchronization, but the desired shift time is not long enough to allow the secondary on-coming clutch to reach full capacity prior to synchronization, the timing of the clutch elements proceeds as follows. The control unit  66  (shown in  FIG. 1 ) begins filling the secondary on-coming clutch apply chamber before the off-going clutches are released such that full capacity of the secondary on-coming clutch is reached prior to synchronization. Thereafter, the release of both off-going clutches is delayed by an amount adapted to meet the desired shift time. Actuation of the primary on-coming clutch is delayed by an amount calculated to fill the primary on-coming apply chamber without generating any torque precisely at the point of synchronization such that immediately thereafter positive torque is generated. 
     If the desired shift time is not long enough to allow the secondary on-coming clutch to reach full capacity prior to synchronization, and the desired shift time is also not long enough to allow the primary on-coming apply chamber to be completely filled prior to synchronization, the timing of the clutch elements proceeds as follows. The control unit  66  (shown in  FIG. 1 ) begins filling the primary and secondary on-coming clutch apply chambers before the off-going clutches are released such that full capacity of the secondary on-coming clutch is reached prior to synchronization and the primary on-coming apply chamber is completely filled precisely at the point of synchronization. Thereafter, the release of both off-going clutches is delayed by an amount adapted to meet the desired shift time. 
     Referring again to  FIG. 3   b , a predefined optimal power-on skip through neutral downshift is shown. More precisely,  FIG. 3   b  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. For example, during a power-on skip through neutral 6-3 downshift, Ag is transmission output speed multiplied by the sixth gear ratio and Cg is the transmission output speed multiplied by the third gear ratio. 
       FIG. 5  shows off-going and on-coming clutch pressures superimposed on the optimal torque converter turbine speed of  FIG. 3   b . As shown in  FIG. 5   b , the ratio change is initiated when both off-going clutches are released such that the turbine speed n t1  slips from the attained gear speed Ag thereby starting the inertia phase of the shift. The secondary on-coming clutch gains capacity after the off-going clutches have been released, and is at full or nearly full capacity prior to synchronization. The primary on-coming clutch pressure reaches the offset pressure P offset  at the point of synchronization, and increases thereafter generating positive on-coming clutch torque. The offset pressure P offset  represents the pressure applied by the primary on-coming clutch return spring (not shown). Immediately after the point of synchronization, the primary on-coming clutch is generating some torque but generally 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 as described hereinabove with reference to  FIG. 2   b  to prevent engine flare. 
     In addition to the adaptive parameters identified hereinabove, including the off-going delay; primary on-coming volume; and secondary oncoming volume, the method of the present invention also establishes an engine torque adaptive parameter. The engine torque adaptive parameter represents the correction or adjustment to the calculated value of transmission input torque derived hereinabove. The adaptive parameters are variable and may be corrected in response to shift aberrations, or deviations from the predefined optimal shift of  FIG. 5   a , which indicate that one or more of the parameters are inaccurate. The adaptive parameters of the present invention will be described as single element adaptive parameters, however it should be appreciated that the adaptive parameters may include multiple elements as described in the incorporated patent application Ser. No. 10/972,067. 
     The shift aberrations that are correctable by adjusting one or more adaptive parameter are graphically represented in  FIGS. 6   a–d  . 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.” The slip early and slip late aberrations may be corrected by adjusting the off-going delay adaptive parameter. More precisely, the off-going delay adaptive parameter may be increased if slip is detected later than expected, and the off-going delay adaptive parameter may be decreased if slip is detected earlier than expected. 
     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 nt is delayed in rising more than a predetermined amount, e.g., 50 rpm, above attained gear speed A g , slip late is indicated. 
     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 . The flare aberration may be corrected by adjusting the primary on-coming volume adaptive parameter as will be described in detail hereinafter. 
     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 skip through neutral 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. The short shift and long shift aberrations may be corrected by adjusting the engine torque adaptive parameter. More precisely, the engine torque adaptive parameter may be increased if a short shift is detected, and the engine torque adaptive parameter may be decreased if a long shift is detected. 
     As shown in  FIG. 6   d , underlap turbine float is a shift aberration wherein the primary on-coming clutch does not have enough capacity to pull the turbine speed n t7  up to the commanded gear speed Cg, and the turbine speed n t7  therefore “floats” at a speed below the commanded gear speed Cg. The underlap turbine float aberration may be corrected by adjusting the primary on-coming volume adaptive parameter as will be described in detail hereinafter. 
     A method for addressing one or more of the shift aberrations identified hereinabove by adjusting the primary on-coming volume adaptive parameter is shown in  FIG. 7 . At step  122 , if flare or underlap turbine float is detected the primary on-coming volume adaptive parameter is increased. At step  124 , if flare control is invoked the primary on-coming volume adaptive parameter is increased. For purposes of the present invention, flare control is defined as the process of reducing engine output to address flare (shown in  FIG. 6   b ). According to a preferred embodiment, the primary on-coming volume adaptive parameter increase of steps  122  and  124  may not exceed a predefined volume limit representing the maximum primary on-coming apply chamber volume. At step  126 , if there is an extreme short shift, no flare, no underlap turbine float, and no long shift aberrations the primary on-coming volume adaptive parameter is decreased. For purposes of this disclosure, an extreme short shift is a shift that is short by more than a predefined amount. At step  128 , if the criteria for steps  122 – 126  are not met after a predefined number of shifts and the engine torque adaptive parameter is either decreased or remains constant, the primary on-coming volume adaptive parameter is incrementally decreased to produce flare. The optimal value for the primary on-coming volume adaptive parameter is that which is just enough to prevent flare. Therefore, the incremental decrease of step  128  periodically recalibrates the primary on-coming volume adaptive parameter to its optimal or nearly optimal value. 
     The primary on-coming volume 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). 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 may require more aggressive corrective action than others. The gain is related to an adaptive error counter that tracks the direction the primary on-coming volume adaptive parameter is moving. If the on-coming volume 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 delay 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 on-coming volume 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 on-coming volume 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 other adaptive parameters may be increased and decreased in a similar manner. 
     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.