ADAPTIVE DIRECT PRESSURE SHIFT CONTROL FOR A MOTOR VEHICLE TRANSMISSION

Abstract: In an adaptive shift control system for regulating the operation of a fluid-operated torque-establishing device (28, 30, 32, 34) associated with a specified speed ratio in a motor vehicle automatic transmission (10), shifting from a currently engaged speed ratio to the specified speed ratio includes a completion phase during which fluid is supplied to the torque-establishing device (28, 30, 32, 34) in accordance with a predetermined pressure schedule to initiate and progressively increase the transmission of torque therethrough. The time required to complete a specified portion of the speed ratio progression is identified and compared with a reference time representative of the time required to complete the progression when the shift quality is not degraded. The predetermined pressure schedule is adjusted based on such comparison so as to adaptively compensate for sources of error which affect the increase of torque transmission through the torque-establishing device (28, 30, 32, 34) during the completion phase and cause the shift quality to be degraded.

Description: This invention relates to a method of adaptively correcting the operation of a shift control system for a motor vehicle multiple speed-ratio automatic transmission, as specified in the preamble of claim 1, for example as disclosed in US-A-4 283 970. Motor vehicle transmissions generally include selectively engageable gear elements for providing two or more forward-speed ratios by means of which engine output torque is applied to the vehicle drive wheels. In automatic transmissions, the gear elements which provide the various speed ratios are selectively activated by means of fluid-operated torque-establishing devices, such as clutches and brakes. The brake can be of the band or disc type; engineering personnel in the automotive art refer to disc-type brakes in transmissions as clutches, clutching devices or reaction clutches. Thus, shifting from one speed ratio to another generally involves releasing (disengaging) the clutching device associated with the current speed ratio, and applying (engaging) the clutching device associated with the desired speed ratio. The clutching device to be released is referred to as the off-going clutch, whereas the clutching device to be applied is referred to as the on-coming clutching device. There is generally a slight overlap between release and apply, and high-quality shifts are achieved only when the release and apply are properly timed and executed. Conventionally, the control of shifting in an automatic transmission is performed with hydraulic logic and servo elements responsive to various system parameters such as vehicle speed and throttle position. Fluid pressure signals representative of the various system parameters are processed to determine when a shift is in order, and spring elements and fluid orifices within the servo elements determine the timing calibration for the release and apply of the respective clutching devices. To overcome certain disadvantages of hydraulic control, it has been proposed to perform at least some of the transmission control functions electronically. For example, one suggestion has been to electronically determine the desired speed ratio based on measured system parameters, and directly control the supply of fluid to the respective clutching elements to effect shifting from one speed ratio to another. The advantages of electronic control include reduced hardware complexity, increased reliability and greater control flexibility. An example of an electronic control system for an automatic transmission is the system disclosed in US-A-3 688 607, the said system utilising a closed-loop control in which the speed rate of change of a specified transmission element is made to conform to a reference rate. The present invention, on the other hand, is concerned with an open-loop control. In an open-loop control, the fluid valves are controlled in accordance with a predetermined schedule to effect apply and release of the various transmission clutching devices, and the control is not changed in the course of a shift in accordance with a measure of a controlled parameter. Such pure open-loop control of the clutching devices is acceptable so long as there are no significant variations in the engine and transmission operating characteristics, and no significant assembly tolerances. However, engine and transmission operating characteristics do change with time, and the production assembly tolerances may result in significant vehicle-to-vehicle variability. As a result, control schedules that produce acceptable ratio shifting in one vehicle may produce unacceptable ratio shifting in another vehicle. The present invention is primarily concerned with the provision of an improved open-loop direct-pressure shift control method wherein the control is adaptively compensated in the course of its operation for variations in the operating characteristics of the engine and transmission. To this end a method of adaptively correcting the operation of a shift control system for a motor vehicle multiple speed-ratio automatic transmission in accordance with the present invention is characterised by the features specified in the characterising portion of claim 1. Such a method in accordance with the present invention permits adaptive correction of performance deficiencies in the control due to vehicle-to-vehicle variability and wear. In its basic form, the control system directs the supply of fluid pressure to the transmission clutching devices in accordance with empirically derived schedules as a function of operator demand and various parameters. In the course of shifting from one speed ratio to another, certain operating parameters of the transmission are monitored as an indication of the shift quality. If the monitored parameters indicate that a particular shift did not progress in an optimum manner, the controller develops corrections for the empirically derived schedules involved in the shift, so that when the shift is repeated at a later point it will be performed in a more nearly optimum manner. More particularly, the present invention addresses the completion phase of ratio shifting. The completion phase follows the fill or preparation phase in which the on-coming clutching device is readied to transmit torque. In the completion phase, the transmission of torque through the on-coming clutching device is initiated and progressively increased by the supply of fluid thereto in accordance with a predetermined pressure schedule. The completion phase comprises torque and inertia phases as in conventional usage, the torque phase being defined as the portion of the completion phase during which there is an exchange of torque between the off-going and on-coming clutches but no speed change, and the inertia phase being defined as the portion of the completion phase during which the speed change is effected. The predetermined pressure schedules are empirically derived to provide optimum shift quality, and they may be different for each clutching device. However, various sources of error affect the pressure of the supplied fluid and the required fluid volume. In this event, the predetermined pressure schedules are incorrect, and the shift quality may be degraded. Broadly, the adaptive control in conformity with the present invention adjusts the predetermined pressure schedules by detecting the time required to effect the speed change (or a specified portion thereof), comparing the detected time with a reference time, and if necessary developing a correction for the scheduled pressure. The corrections are developed and applied in a novel manner designed to maximise the speed of convergence while minimising the converged error. The present invention thereby makes available a direct pressure shift control in which adaptive schedule adjustment individually adjusts the pressure applied to each clutching device in the course of a shift, with such adjustment having no effect on the line pressure, such that fill time is correspondingly unaffected. Co-pending patent application no (Serial no ) of even date, based on USSN 802,676 filed 29 November 1985, is directed to an adaptive control of the fill or preparation phase of the speed ratio shift. Such control is also described in the present specification. In the drawings: Figures 1a and 1b schematically depict a computer-based electronic transmission control system for carrying out a method in accordance with the present invention; Figure 2 graphically depicts various parameters of the engine and transmission in the course of a typical upshift using the system shown in Figure 1; Figures 3 to 6 graphically depict certain of the parameters shown in Figure 2 for upshifts in which stored fill time and pressure schedules are in error; Figures 7 to 10 graphically depict adaptive compensation of an empirically determined fill time in the system shown in Figure 1, with Figure 7 depicting the stored fill time (T fill ) vs. working pressure ΔP relationship, Figure 8 depicting typical fill time error distributions for converged and non-converged systems, Figure 9 depicting the scheduling of fill time corrections, and Figure 10 depicting the application of the corrections to the stored t fill vs. ΔP relationship of Figure 7; Figures 11 and 12 graphically depict adaptive compensation of the empirically determined pressure schedules, with Figure 11 depicting the stored pressure ΔP vs. torque variable (T v ) vs. time (t) relationship, and Figure 12 depicting the measurement of a predefined inertia phase interval; and Figures 13 to 17 depict flow diagrams representative of suitable program instructions executed by the computer-based controller of Figure 1 for carrying out the control functions in conformity with the present invention, with Figure 13 depicting a main loop program, Figures 14 and 15 depicting pressure control algorithms for a typical power-on upshift, Figures 16a to 16c depicting the algorithm for adaptive fill time correction, and Figure 17 depicting the algorithm for adaptive pressure correction. With reference now to the drawings, and more particularly to Figures 1a and 1b, reference numeral 10 generally designates a motor vehicle drive-train including an engine 12 and a parallel-shaft transmission 14 having a reverse speed ratio and four forward speed ratios. The engine 12 includes a throttle mechanism 16 mechanically connected to an operator-manipulated device such as an accelerator pedal (not shown) for regulating engine output torque, such torque being applied to the transmission 14 by way of the engine output shaft 18. The transmission 14 transmits engine output torque to a pair of drive axles 20 and 22 by way of a torque converter 24 and one or more of a plurality of fluid-operated clutching devices 26 to 34, such clutching devices being applied or released according to a predetermined schedule for establishing the desired transmission speed ratio. With reference now more particularly to the transmission 14, an impeller (input member) 36 of the torque converter 24 is connected to be rotatably driven by the output shaft 18 of the engine 12 by way of an input shell 38. A turbine (output member) 40 of the torque converter 24 is rotatably driven by the impeller 36 by fluid transfer therebetween and is connected to rotatably drive a shaft 42. A stator member 44 redirects the fluid which couples the impeller 36 to the turbine 40, the stator being connected by way of a one-way device 46 to the housing of the transmission 14. The torque converter 24 also includes a clutching device 26 comprising a clutch plate 50 secured to the shaft 42. The clutch plate 50 has a friction surface 52 formed thereon that is engageable with the inner surface of the input shell 38 to form a direct mechanical drive between the engine output shaft 18 and the transmission shaft 42. The clutch plate 50 divides the space between the input shell 38 and the turbine 40 into two fluid chambers, namely an apply chamber 54 and a release chamber 56. When the fluid pressure in the apply chamber 54 exceeds that in the release chamber 56, the friction surface 52 of the clutch plate 50 is moved into engagement with the input shell 38 as shown in Figure 1, thereby engaging the clutching device 26 to provide a mechanical drive connection in parallel with the torque converter 24. In such a case, there is no slippage between the impeller 36 and the turbine 40. When the fluid pressure in the release chamber 56 exceeds that in the apply chamber 54, the friction surface 52 of the clutch plate 50 is moved out of engagement with the input shell 38, thereby uncoupling such mechanical drive connection and permitting slippage between the impeller 36 and the turbine 40. The circled numeral 5 represents a fluid connection to the apply chamber 54, and the circled numeral 6 represents a fluid connection to the release chamber 56. A positive-displacement hydraulic pump 60 is mechanically driven by the engine output shaft 18 by way of the input shell 38 and impeller 36 as indicated by the broken line 62. The pump 60 receives hydraulic fluid at low pressure from a fluid reservoir 64 and supplies pressurised fluid to the transmission control elements via an output line 66. A pressure regulator valve (PRV) 68 is connected to the pump output line 66 and serves to regulate the fluid pressure (hereinafter referred to as line pressure) in the line 66 by returning a controlled portion of the fluid therein to the reservoir 64 via the line 70. In addition, the pressure regulator valve 68 supplies fluid pressure for the torque converter 24 via a line 74. The pump and pressure regulator valve designs are not critical to the present invention: a representative pump is disclosed in US-A-4 342 545, and a representative pressure regulator valve is disclosed in US-A-4 283 970. The transmission shaft 42 and a further transmission shaft 90 each have a plurality of gear elements rotatably supported thereon. The gear elements 80 to 88 are supported on the shaft 42, and the gear elements 92 to 102 are supported on the shaft 90. The gear element 88 is rigidly connected to the shaft 42, and the gear elements 98 and 102 are rigidly connected to the shaft 90. The gear element 92 is connected to the shaft 90 by way of a freewheel (one-way device) 93. The gear elements 80, 84, 86 and 88 are maintained in meshing engagement with the gear elements 92, 96, 98 and 100 respectively, and the gear element 82 is coupled to the gear element 94 by way of a reverse idler gear 103. The shaft 90, in turn, is coupled to the drive axles 20 and 22 by way of gear elements 102 and 104 and a conventional differential gear set (DG) 106. A dog clutch 108 is splined on the shaft 90 so as to be axially slidable thereon, and serves to rigidly connect the shaft 90 either to the gear element 96 (as shown) or to the gear element 94. A forward speed relation between the gear element 84 and the shaft 90 is established when the dog clutch 108 connects the shaft 90 to the gear element 96, and a reverse speed relation between the gear element 82 and the shaft 90 is established when the dog clutch 108 connects the shaft 90 to the gear element 94. The clutching devices 28 to 34 each comprise an input member rigidly connected to a transmission shaft 42 or 90, and an output member rigidly connected to one or more gear elements such that engagement of a clutching device couples the respective gear element and shaft together to effect a driving connection between the shafts 42 and 90. The clutching device 28 couples the shaft 42 to the gear element 80; the clutching device 30 couples the shaft 42 to the gear elements 82 and 84; the clutching device 32 couples the shaft 90 to the gear element 100; and the clutching device 34 couples the shaft 42 to the gear element 86. Each of the clutching devices 28 to 34 is biased towards a disengaged state by means of a return spring (not shown). Engagement of the clutching device is effected by supplying fluid pressure to an apply chamber thereof. The resulting torque capacity of the clutching device is a function of the applied pressure less the return spring pressure, hereinafter referred to as the working pressure ΔP. The circled numeral 1 represents a fluid passage for supplying pressurised fluid to the apply chamber of the clutching device 28; the circled numeral 2 and letter R represent a fluid passage for supplying pressurised fluid to the apply chamber of the clutching device 30; the circled numeral 3 represents a fluid passage for supplying pressurised fluid to the apply chamber of the clutching device 32; and the circled numeral 4 represents a fluid passage for directing pressurised fluid to the apply chamber of the clutching device 34. The various gear elements 80 to 88 and 92 to 100 are relatively sized such that engagement of first, second, third and fourth forward speed ratios is effected by engaging the clutching devices 28, 30, 32 and 34 respectively, with the dog clutch 108 being in the position depicted in Figure 1 to obtain a forward speed ratio. A neutral speed ratio or an effective disconnection of the drive axles 20 and 22 from the engine output shaft 18 is effected by maintaining all of the clutching devices 28 to 34 in a released condition. The speed ratios defined by the various gear element pairs are generally characterised by the ratio of the turbine speed N t to the output speed N o . Representative N t /N o ratios for the transmission 14 are as follows: First - 2.368 Second - 1.273 Third - 0.808 Fourth - 0.585 Reverse - 1.880 As indicated above, shifting from a current forward speed ratio to a desired forward speed ratio requires that the clutching device associated with the current speed ratio (off-going) be disengaged and the clutching device associated with the desired speed ratio (on-coming) be engaged. For example, a shift from the first forward speed ratio to the second forward speed ratio involves disengagement of the clutching device 28 and engagement of the clutching device 30. As is explained below, the timing of such disengagement and engagement is critical to the attainment of high-quality shifting, and this invention is directed primarily to a method of utilising a control system for supplying fluid pressure to the various clutching devices 28 to 34 to achieve consistent high-quality shifting. The fluid control elements of the transmission 14 include a manual valve 140, a directional servo 160 and a plurality of electrically operated fluid valves 180 to 190. The manual valve 140 operates in response to operator demand, and serves, in conjunction with the directional servo 160, to direct regulated line pressure to the appropriate fluid valves 182 to 188. The fluid valves 182 to 188, in turn, are individually controlled to direct fluid pressure to the clutching devices 28 to 34. The fluid valve 180 is controlled to direct fluid pressure from the pump output line 66 to the pressure regulator valve 68, and the fluid valve 190 is controlled to direct fluid pressure from the line 74 to the clutching device 26 of the torque converter 24. The directional servo 160 operates in response to the condition of the manual valve 140, and serves to properly position the dog clutch 108. The manual valve 140 includes a shaft 142 for receiving axial mechanical input from the operator of the motor vehicle in relation to the speed range the operator desires. The shaft 142 is also connected to an indicator mechanism 144 by way of a suitable mechanical linkage as indicated generally by the broken line 146. Fluid pressure from the pump output line 66 is applied as an input to the manual valve 140 by way of the line 148, and the valve outputs include a forward (F) output line 150 for supplying fluid pressure for engaging the forward speed ratios, and a reverse (R) output line 152 for supplying fluid pressure for engaging the reverse speed ratio. Thus when the shaft 142 of the manual valve 140 is moved to the D4, D3, or D2 positions shown in the indicator mechanism 144, line pressure from the line 148 is directed to the forward (F) output line 150. When the shaft 142 is in the R position shown in the indicator mechanism 144, line pressure from the line 148 is directed to the reverse (R) output line 152. When the shaft 142 of the manual valve 140 is in the N (neutral) or P (park) positions, the input line 148 is isolated, and the forward and reverse output lines 150 and 152 are connected to an exhaust line 154 which is adapted to return any fluid therein to the fluid reservoir 64. The directional servo 160 is a fluid-operated device and includes an output shaft 162 connected to a shift fork 164 for axially shifting the dog clutch 108 on the shaft 90 to selectively enable either forward or reverse speed ratios. The output shaft 162 is connected to a piston 166 axially movable within the servo housing 168. The axial position of the piston 166 within the housing 168 is determined according to the fluid pressures supplied to the chambers 170 and 172. The forward output line 150 of the manual valve 140 is connected via a line 174 to the chamber 170, and the reverse output line 152 of the manual valve 140 is connected via a line 176 to the chamber 172. When the shaft 142 of the manual valve 140 is in a forward-range position, the fluid pressure in the chamber 170 urges the piston 166 rightwardly as viewed in Figure 1 to engage the dog clutch 108 with the gear element 96 for enabling engagement of a forward speed ratio. When the shaft 142 of the manual valve 140 is moved to the R position, the fluid pressure in the chamber 172 urges the piston 166 leftwardly as viewed in Figure 1 to engage the dog clutch 108 with the gear element 94 for enabling engagement of the reverse speed ratio. However, the actual engagement of the second or reverse speed ratio is not effected until the clutching device 30 is engaged. The directional servo 160 also operates as a fluid valve for enabling the reverse speed ratio. To this end, the directional servo 160 includes an output line 178 connected to an electrically operated fluid valve 186. When the operator selects a forward speed ratio and the piston 166 of the directional servo 160 is in the position depicted in Figure 1, the passage between the lines 176 and 178 is cut off, and conversely when the operator selects the reverse gear ratio, the passage between the lines 176 and 178 is open. The electrically operated fluid valves 180 to 190 each receive fluid pressure at an input passage thereof from the pump 60, and are individually controlled to direct fluid pressure to the pressure regulator valve 68 or the respective clutching devices 26 to 34. The fluid valve 180 receives line pressure directly from the pump output line 66, and is controlled to direct a variable amount of such pressure to the pressure regulator valve 68, as indicated by the circled letter V. The fluid valves 182, 186 and 188 receive fluid pressure from the forward output line 150 of the manual valve 140, and are controlled to direct variable amounts of such pressure to the clutching devices 34, 32 and 28 as indicated by the circled numerals 4, 3 and 1 respectively. The fluid valve 186 receives fluid pressure from the forward output line 150 and the directional servo output line 178, and is controlled to direct a variable amount of such pressure to the clutching device 30, as indicated by the circled numeral 2 and the circled letter R. The fluid valve 190 receives fluid pressure from line 74 of the pressure regulator valve 68, and is controlled to direct a variable amount of such pressure to the release chamber 56 of the clutching device 26, as indicated by the circled numeral 6. The apply chamber 54 of the clutching device 26 is supplied with fluid pressure from the output line 74 via the orifice 192, as indicated by the circled numeral 5. Each of the fluid valves 180 to 190 includes a spool element 210 to 220 respectively, axially movable within the respective valve body for directing fluid flow between the input and output passages thereof. When a respective spool element 210 to 220 is in the rightmost position as viewed in Figure 1, the input and output passages are interconnected. Each of the fluid valves 180 to 190 includes an exhaust passage, as indicated by the circled letters EX, such passage serving to drain fluid from the respective clutching device when the spool element is shifted to the leftmost position as viewed in Figure 1. In Figure 1, the spool elements 210 and 212 of the fluid valves 180 and 182 are shown in the rightmost position interconnecting the respective input and output lines, whereas the spool elements 214, 216, 218 and 220 of the fluid valves 184, 186, 188 and 190 are shown in the leftmost position interconnecting the respective output and exhaust lines. Each of the fluid valves 180 to 190 includes a solenoid 22 to 232 for controlling the position of its spool element 210 to 220. Each such solenoid 22 to 232 comprises a plunger 234 to 244 connected to the respective spool element 210 to 220 and a solenoid coil 246 to 256 surrounding the respective plunger. One terminal of each such solenoid coil 246 to 256 is connected to ground potential as shown, and the other terminal is connected to an output line 258 to 268 of a control unit 270 which governs the solenoid coil energisation. As is set forth hereinafter, the control unit 270 pulse-width-modulates the solenoid coils 246 to 256 according to a predetermined control algorithm to regulate the fluid pressure supplied to the pressure regulator 68 and the clutching devices 26 to 34, the duty cycle of such modulation being determined in relation to the desired magnitude of the supplied pressures. Although the fluid valves 180 to 190 have been illustrated as spool valves, other types of valve could be substituted therefor. By way of example, valves of the ball and seat type could be used. In general terms, the fluid valves 180 to 190 may be mechanised with any three-port pulse-width-­ modulated valving arrangement. Input signals for the control unit 270 are provided on the input lines 272 to 284. A position sensor (S) 286 responsive to movement of the manual valve shaft 142 provides an input signal to the control unit 270 via a line 272. Speed transducers 288, 290 and 292 sense the rotational velocity of various rotary members within the transmission 14, and supply speed signals in accordance therewith to the control unit 270 via lines 274, 276, and 278, respectively. The speed transducer 288 senses the velocity of the transmission shaft 42 and therefore the turbine or transmission input speed N t ; the speed transducer 290 senses the velocity of the drive axle 22 and therefore the transmission output speed N o ; and the speed transducer 292 senses the velocity of the engine output shaft 18 and therefore the engine speed N e . The position transducer 294 is responsive to the position of the engine throttle 16, and provides an electrical signal in accordance therewith to the control unit 270 via the line 280. A pressure transducer 296 senses the manifold absolute pressure (MAP) of the engine 12, and provides an electrical signal to the control unit 270 in accordance therewith via the line 282. A temperature sensor 298 senses the temperature of the oil in the transmission fluid reservoir 64, and provides an electrical signal in accordance therewith to the control unit 270 via the line 284. The control unit 270 responds to the input signals on the input lines 272 to 284 according to a predetermined control algorithm as set forth herein, for controlling the energisation of the fluid valve solenoid coils 246 to 256 via the output lines 258 to 268. As such, the control unit 270 includes an input/output (I/O) device 300 for receiving the input signals and outputting the various pulse-width-­ modulation signals, and a microcomputer 302 which communicates with the I/O device 300 via an address-and-control bus 304 and a bidirectional data bus 306. Flow diagrams representing suitable program instructions for developing the pulse-width-­ modulation outputs in conformity with the present invention are depicted in Figures 13 to 17. As indicated above, every shift from one speed ratio to another involves disengagement of an off-going clutching device and engagement of an on-coming clutching device. Each shift includes a fill phase during which the apply chamber of the on-coming clutching device is filled with fluid, a torque phase during which the torque capacity of the off-going clutching device is reduced and the torque capacity of the on-coming clutching device is increased, and an inertia phase during which the turbine is accelerated to a new velocity determined according to the new speed ratio. Such phases are defined in terms of times t₀ to t₄ for a typical 2-3 upshift in graphs A to D of Figure 2, each of the graphs having a common time base. Graph A depicts the turbine speed N t ; Graph B depicts the pressure command for the on-coming clutching device fluid valve; Graph C depicts the engine torque T e and the torque carried by the clutching devices 30 and 32; and Graph D depicts the transmission output torque T o . Prior to the shift activity, the relation between the turbine and output speeds N t and N o is static and is determined according to the second speed ratio. In addition, the output torque T o is substantially constant. In the course of the shift, the speed and torque relationships become dynamic as the engine torque T e is shifted from the clutching device 30 to the clutching device 32. Subsequent to the shift activity, the output torque is once again substantially constant, and the relation between N t and N o is determined according to the third speed ratio. At time t₀ when it is determined that a 2 - 3 ratio shift is desired, the solenoid coil 250 of the fluid valve 184 is energised at a duty cycle of 100%, to commence filling the apply chamber of the clutching device 32. This marks the beginning of the fill phase of the shift, as indicated below Graph D. Although not shown in Figure 2, the solenoid coil 252 of the fluid valve 186 is energised at a relatively high duty cycle during the fill phase to maintain engagement of the second speed ratio. At time t₁, t fill seconds after time t₀, the fluid pressure in the apply chamber of the clutching device 32 is sufficiently great to compress the clutch return spring, marking the end of the fill phase and the beginning of the torque phase, as indicated below Graph D. Thereafter, the pressure command is reduced to a value corresponding to an empirically derived initial pressure P x and progressively increased to a value corresponding to an empirically derived final pressure P y . During such time, the torque T cd32 carried by the on-coming clutching device 32 progressively increases, and the torque T cd30 carried by the off-going clutching device 30 progressively decreases, as seen in Graph C. The output torque T o in this interval is determined according to the sum of T cd30 and T cd32 as reflected by the respective speed ratios of the transmission 14, and progressively decreases as seen in Graph D. At time t₂, the torque T cd32 equals the engine torque T e , the torque T cd30 is reduced to zero, and the output torque T o begins to rise with T cd32 , as seen in Graphs C to D. After time t₂, the torque T cd32 continues to rise, and the torque differential between it and the engine torque T e urges the turbine to decelerate towards the third-ratio speed, designated by the trace 308 in the Graph A. At time t₃, the turbine speed N t begins to decrease, marking the end of the torque phase and the beginning of the inertia phase, as indicated below Graph D. As the turbine speed N t decreases, the engine torque T e increases, as seen in Graph C. At time t₄, the turbine speed joins the third speed trace 308, marking the end of the inertia phase and the shift, as indicated below Graph D. Since the clutching device 32 is no longer slipping at this point, the torque T cd32 drops to the level of the engine torque T e , and the output torque T o drops to the post-shift level. The shaded area 309 between the T e and T cd32 traces in Graph C is referred to as the inertia torque, and represents the amount of torque the on-coming clutching device must exert to effect the speed change. The on-coming clutch fill time and the clutch pressure schedules are individually determined for each ratio shift. If both are correct, and the various control elements each function as expected, the ratio shift will progress in the desired manner as depicted in Figure 2, with neither excessive harshness nor excessive slippage of the friction devices. These are the essential ingredients of open-loop ratio shifting. As indicated above, however, a certain amount of variation in the engine and transmission operating characteristics can be expected over the life of the vehicle, due to wear. Moreover, there may be some vehicle-to-vehicle variability due to assembly and component tolerances. If the on-coming clutching device begins developing torque capacity either before or after the end of the calculated fill time, the exchange of torque capacity between the off-going and on-coming clutching devices will not proceed according to schedule. In this regard, the consequences of overfill and underfill errors are graphically illustrated in Figures 3 and 4. Similarly, the shift quality is degraded if the clutch pressure during the torque and inertia phases is too high or too low for a given operating condition. The consequences of improperly low and high pressure scheduling are graphically illustrated in Figures 5 and 6. Figures 3 to 6 each include Graphs A, B and C corresponding to the Graphs A, C and D of Figure 2. To facilitate comparison of the various traces with the corresponding traces of Figure 2, each of the graphs of Figures 3 to 6 includes the time scale designations t₀ to t₄ as defined in reference to the normal high-quality shift of Figure 2. In addition, the static torque and speed levels shown in Figure 2 have been adopted in Figures 3 to 6. When the stored fill time t fill is too high, as illustrated in Figure 3, the on-coming clutching device 32 is overfilled, and begins transmitting torque prior to time t₁, as seen by the T cd32 trace in Graph B. In such case, the capacity T cd32 of the on-coming clutching device reaches the engine torque T e before the capacity T cd30 of the off-going clutching device is reduced to zero as seen at time t x in Graph B. As a result, the on-coming clutching device 32 is opposed by the off-going clutching device 30, resulting in what is known as bind-up, which bind-up reduces the output torque T o as compared with the shift of Figure 2. The magnitude of the output torque reduction is graphically represented by the shaded area 310 of Graph C. The bind-up also results in a momentary unwinding of the various transmission and drive-line shafts, as evidenced by the momentary reduction 311 in turbine speed N t . When the stored fill time t fill is too low, as illustrated in Figure 4, the on-coming clutching device 32 is underfilled, and does not begin transmitting torque until after time t₁, as seen by the T cd32 trace in Graph B. In such case, the output torque is reduced as compared with the shift depicted in Figure 2, the amount of such reduction being graphically represented by the shaded area 312 in Graph C. Moreover, the torque capacity C cd32 of the on-coming clutching device is not yet sufficient to transmit all of the engine torque T e when the off-going clutching device is completely released at time t₂. This causes a turbine speed flare, as indicated by the reference numeral 313 in Graph A. When the scheduled pressure for the on-coming clutching device is too low, as illustrated in Figure 5, the torque capacity C cd32 is reduced as compared with Figure 2. As a result, the duration of the inertia phase becomes excessively long, so degrading the shift quality and inducing excessive wear and heating of the clutching devices. For the example depicted in Figure 5, the length of the inertia phase is designated by the interval 314. When the scheduled pressure for the on-coming clutching device is too high, as illustrated by Figure 6, the torque capacity C cd32 is increased as compared with Figure 2, and the turbine is rapidly decelerated to its new speed as seen in Graph A. As a result, the duration of the inertia phase is relatively short, as designated by the interval 316. In addition, the rapid turbine deceleration causes a transient increase in the output torque T o as indicated by the shaded area 318 in Graph C, and produces an undesirably harsh shift. In conformity with the present invention, the empirically derived fill times and pressure schedules for the various clutching devices are adaptively compensated in the course of vehicle operation so as to achieve consistent high-quality ratio shifting. In each case, specified operating parameters are monitored during each upshift, and then compared with reference parameters to determine whether the shift progressed in a desired manner. If the comparison of the monitored and reference parameters indicates that the shift did not progress in the desired manner, the respective fill time and/or pressure schedule is adjusted in a corrective direction so that subsequent shifting will be carried out in a more nearly optimum manner. The empirically derived fill times are adaptively corrected by monitoring the time interval between the start of fill and the turn-down (reduction) in turbine speed during each upshift. Since the turbine speed turn-down marks the beginning of the inertia phase of the shift, such interval is referred to herein as the inertia phase delay, IPDELAY. The measured IPDELAY is compared with a reference desired delay, DESDELAY, to determine whether the on-coming clutching device was properly filled at time t₁. If the stored fill time t fill is correct, the on-coming clutching device will be properly filled at time t₁, and IPDELAY will be substantially equal to DESDELAY. If the stored fill time t fill is too short, and the on-coming clutching device is underfilled at time t₁, the turbine speed turn-down will be late, as shown in Figure 3, and IPDELAY will be significantly greater than DESDELAY. In this event, the control unit 270 operates to increase the fill time t fill for the respective clutching device so that subsequent shifts involving that clutching device will be performed in a more nearly optimum manner. If the stored fill time t fill is too long, and the on-coming clutching device is overfilled (already developing torque capacity) at time t₁, the resulting bind-up and momentary turbine speed reduction described in relation to Figure 4 will be sensed as an early turbine speed turn-down, and IPDELAY will be significantly less than DESDELAY. In this event, the control unit 270 operates to decrease the fill time t fill for the respective clutching device so that subsequent shifts involving that clutching device will be carried out in a more nearly optimum manner. In practice, the turbine speed flare characteristic associated with underfill is more easily identified than the momentary reduction associated with overfill. This is especially true in low-torque high turbine-speed shifts, since the momentary turbine speed changes associated with overfill are but a small percentage of the steady-state turbine speed. This difficulty is overcome in conformity with the present invention by periodically decrementing the fill time while the vehicle is being operated under conditions for which overfills cannot be accurately identified. When the incremental changes in fill time result in a detectable underfill, the control unit 270 operates as described above to increase the fill time. In this way, the stored fill times for the various clutching devices are maintained relatively close to the correct values even during periods of vehicle operation for which overfill detection may not be reliable. In addition, large overfill error indications are treated as small overfill indications until several (for example three) such error indications are successively sensed. The description of the mechanisms for identifying turn-down of turbine speed and for adaptively compensating the clutching device fill times in response thereto is prefaced by a description of the mechanism for computing the fill time. As briefly set forth above, the fill time for a given clutching device is determined primarily as a function of the requested line pressure, the geometry of the clutching device, and the viscosity of the fluid. Algebraically, the fill time t fill is given as follows: t fill = V / [A * (2ΔP/r) 1/2 ] where V is the volume of the apply chamber, A is the area of the clutch piston, ΔP is the apply pressure less the return spring pressure, and r is the fluid viscosity. To improve the fill time calculation efficiency, the present invention allows utilisation of a fill time vs. pressure (ΔP) function look-up table, as graphically depicted by the trace 320 of Figure 7. The trace 320 takes into account the clutching device geometry, and is in the form of an inverse square root function due to the ΔP dependence as set forth in the algebraic expression above. To avoid the need for storing the entire function, just the two fill time points (designated L and H) corresponding to the lowest and highest available line pressures ΔP L and ΔP H are stored by the control unit 270. The fill time is linearly interpolated along the broken line 322 interconnecting the fill time points L and H, and then mathematically adjusted to reflect the inverse square root form (1/) of the trace 320. The adjusted fill time is then modified by a oil temperature-dependent factor to compensate for the variations in the fluid viscosity. The time required to effect a turn-down in turbine speed in the course of an upshift is determined by starting a timer at the end of fill and stopping the timer upon detection of the turn-down. The timed interval may thus be viewed as the delay between the end of fill and the beginning of the inertia phase. The turn-down is identified by predicting a future turbine speed (through an extrapolation process) and comparing the actual turbine speed with the predicted turbine speed. Turbine speed is detected in terms of T/TP, this being the time between pulses received from the turbine speed transducer 288 of Figure 1. By the nature of its definition, T/TP varies inversely with turbine speed. The measured values of T/TP are averaged by a first-order lag function to determine the average time between turbine pulses, AT/TP. In turn, the difference (AT/TP - T/TP) is computed and subjected to a first-order lag function to determine the average change in time between turbine pulses, AΔT/TP. Algebraically, the predicted time between turbine pulses for a point (k+2) seconds in the future, PT/TP(k+2), is given by the expression: PT/TP(k+2) = AT/TP(k) - [AT/TP(k-4) - AT/TP(k)]/2 - [AΔT/TP(k-4) + AΔT/TP(k)] The term k represents several loop times of the control unit 270, and the predicted time between turbine pulses one loop time (L) in the future, PR/TP(L), is determined by linear interpolation between the calculated values. The error time between turbine pulses ET/TP -- i.e. the difference between actual and predicted time between turbine pulses T/TP(L) - PT/TP(L) -- is computed to identify the turn-down of turbine speed. Due to the inverse relation between turbine speed and time T, the turn-down is identified as a significant error ET/TP of positive sign. Several steps are taken to minimise the likelihood of false turn-down detection. The main concern in this regard is that noise or fluctuation of the turbine speed signal (due to bumps in the road surface, for example) causes some difference between the predicted and actual time between turbine pulses. Primarily, the likelihood of false detection is minimised by the employment of novel signal processing techniques, including (1) enabling the detection algorithm only in a specified time window during the shift, (2) defining a two-stage error threshold for identifying the turn-down, and (3) adjusting the two-stage error threshold in accordance with a measure of the turbine speed signal noise. The time window during which the detection algorithm is enabled is defined so that the algorithm is operative to detect the occurrence of turbine speed turn-down in worst-case overfill and underfill situations. The two-stage error threshold comprises a first relatively low threshold and a second relatively high threshold. When the first threshold is exceeded, the fill timer is sampled. If the second threshold is subsequently exceeded, the occurrence of turn-down is confirmed, and turn-down indication is given. If the second threshold is not subsequently exceeded, the error is assumed to be noise-induced, and the fill timer is permitted to continue counting. In this way, noise-induced error is distinguished from turn-down-induced error. The measure of the turbine speed noise for adjusting the first and second thresholds is obtained by applying a first-order lag function to the error signal ET/TP, such filtered signal being identified herein as FET/TP. The thresholds are increased with increasing FET/TP, and decreased with decreasing FET/TP. In spite of the above signal processing techniques, it is possible that turbine speed noise could cause consecutive significant error signals, and trigger an early turn-down indication. The consequence of such a false indication is mitigated (as explained below) by limiting the adaptive correction to a relatively small value until several consecutive overfills have been indicated. In the illustrated embodiment, three consecutive overfill indications are required before a large fill time correction can be issued. The adaptive compensation mechanism interfaces with the above described fill-time determination technique by suitable modification of the fill time end points L and H. The amount by which the end points L and H are modified is determined in relation to the fill time error (E ft ) between the detected and expected times to turbine speed turn-down (downturn). The expected time to turn-down is empirically determined in relation to the type of shift, and either the working pressure ΔP or the input torque T i . As an example, let it be assumed that the torque phase of a given upshift is expected to take 30 ms. It the actual turbine speed turn-down is detected 30ms after the end of the calculated fill time (E ft = 0ms), it is assumed that the calculated fill time is correct, and no adaptive compensation is attempted. However, if the turn-down is detected 60ms after the end of the calculated fill time (E ft = +30ms), it is assumed that the calculated fill time is too short -- that is, the on-coming clutching device was underfilled and developed torque capacity too late. On the other hand, if the turn-down is detected 10ms after the end of the calculated fill time (E ft = -20ms), it is assumed that the calculated fill time is too long -- that is, the on-coming clutching device was overfilled and developed torque capacity too early. The sign of the error indicates whether the on-coming clutching device was underfilled or overfilled, and the magnitude indicates the amount of error. The fill time error E ft as defined above is used to determine a fill time correction (C ft ) for adaptively compensating the stored fill time end points L and H as defined in Figure 7. The fill time correction amount is determined in accordance with a gain-scheduling technique designed to achieve fast convergence of the calculated fill times with minimum converged misadjustment. Functionally, the intent is to provide large adjustment when large errors of a given sign are sensed, and little or no adjustment when a distribution of errors is sensed. The randomness of the system operation is graphically depicted in Figure 8, where traces 330 and 332 represent typical distributions of sensed fill time error for a clutching device. The distribution trace 330 is centred around zero error, and therefore represents a system which is accurately calibrated and thus cannot be improved by adaptive compensation. The shape of the trace is influenced by the control algorithms of the control unit 270 and the physical control elements of the transmission 14. Presumably such algorithms and elements are designed to provide a sufficiently high level of repeatability for acceptable shift quality to be achieved within most, if not all, of the distribution range. For the example depicted in Figure 8, a fill time error of plus or minus e₁ would not result in unacceptable shift quality. The distribution trace 332 is centred around an error value e₂, and therefore represents a system which is not accurately calibrated and which could be improved by adaptive compensation. Nearly all the fill time errors within the randomness of the distribution trace 332 are greater than e₁, and would be likely to result in unacceptable shift quality. The aim of adaptive compensation is to move the distribution trace 332 to the left as viewed in Figure 8 by the amount of e₂ so as to achieve the shift quality associated with the distribution trace 330. However, the control unit 270 cannot determine the centre of the error distribution based on a given error measurement. For example, a measured error of e₁ could occur with a system represented by the distribution trace 330, a system represented by the distribution trace 332, or a system represented by any distribution trace therebetween. If the distribution trace 332 is representative of the system, a relatively large adaptive modification of the fill time would be proper. If the distribution trace 330 is representative of the system, an adaptive modification of the fill time would be a misadjustment. The difficulty set forth above is overcome in conformity with the present invention by establishing a non-linear base gain schedule which is relatively low for achieving relatively low converged misadjustment and a non-linear direction-sensitive dynamic gain modifier for increasing the base schedule gain in relation to the time integral of the measured error. The authority of the dynamic gain modifier is limited by a maximum overall gain which is dependent on the error magnitude, and the modifier is reset to zero when a significant error of the opposite sign is detected. In Figure 9, the base gain schedule and the maximum overall gain are graphically depicted as a function fill time error E ft . The base gain schedule is depicted by the trace 334, and the maximum overall gain is depicted by the trace 336. As indicated above, fill time error that is positive in sign (underfill) produces a positive correction for increasing the fill time end points, whereas fill time error that is negative in sign (overfill) produces a negative correction for decreasing the fill time end points. The dynamic gain modifier can increase the base gain correction (in either a positive or a negative sense) up to the maximum overall gain in relation to the integral of error signals in one direction. Graphically, the shaded areas between the traces 334 and 336 represent the authority range of the dynamic gain modifier. In this way, the fill time correction is determined primarily in accordance with the base schedule gain when the error distribution is centred at or near zero error. When the error distribution becomes significantly skewed in either direction, the dynamic gain modifier becomes active, and adds to the base gain to achieve fast correction of the error. Essentially, the adaptive corrections become greater with increased detected error, and increased time required to correct the error. The fill time correction C ft is apportioned between the fill time end points L and H in accordance with the fluid pressure applied to the on-coming clutching device during the upshift. The mechanisation of such apportionment is graphically depicted in Figure 10, where the trace 340 represents a gain factor G L for the end point L, and the trace 342 represents a gain factor G H for the end point H. After any upshift where a fill time correction C ft is in order, the end point L is adjusted by the amount (C ft * G L ), and the end point H is adjusted by the amount (C ft * G H ). In future shifts involving the subject clutching device, the calculated fill time t fill will more correctly reflect the actual time required to fill its apply chamber and stroke the return spring to develop torque capacity. As a result, changing conditions which affect the fill time of the clutching device are fully compensated for over a number of upshifts involving the clutching device. The empirically derived pressure schedules for the on-coming clutching devices are adaptively corrected by monitoring the inertia phase interval t ip during each upshift and comparing such interval with a reference interval t rip . If the stored pressure schedule is correct, the shift will progress in a desired manner, and t ip will be substantially equal to t rip . If the stored pressure schedule is too high, the shift will be too harsh, and t ip will be significantly less than t rip . In such event the control unit 270 will operate to decrease the stored pressure schedule so that subsequent shifts involving that clutching device will be carried out in a more nearly optimum manner. If the stored pressure schedule is too low, the shift will take too long, and t ip will be significantly greater than t rip . In such event, the control unit 270 will operate to increase the stored pressure schedule. In operation, the pressure schedules are determined as a function of a torque variable T v . The torque variable T v , in turn, is determined as a function of the gear set input torque T i and the entry turbine speed N te , N te being defined as the turbine speed N t at the end of the fill phase. The entry turbine speed, in combination with predicted turbine speed for the new speed ratio, provides an indication of the inertia torque required to effect the shift. With this information, the clutch pressures are scheduled so that, for any value of input torque T i , the time required to effect the shift varies in direct relation to ΔN t . However, some input torque dependency may be introduced if it is desired to stretch-out or soften off-pattern shifts, such as high-speed/low-torque upshifts. The status of the torque converter clutching device 26 also affects the scheduled pressure. If the clutching device 26 is disengaged during the shift, the torque converter 24 effectively isolates the inertia of the engine 12, and the on-coming clutching device need overcome only the turbine inertia. If the clutching device 26 is engaged during the shift, the inertia torque must be significantly greater, since both the engine and turbine inertias must be overcome. In mechanising the determination of T v , the gear-set input torque T i is calculated as a function of the engine manifold absolute pressure (MAP), the engine pumping efficiency (K), a mechanical friction term (T f ), the accessory load torque (T L ), and the torque multiplication ratio (T c ) of the torque converter 24, according to the following expression: T i = [(MAP x K) - T f - T L ] x Tc The engine MAP is determined from the sensor 296, and the efficiency K is stored based on previously determined data. The mechanical friction term T f is determined as a function of engine speed, and the load torque term T L is determined by loading indicators. The torque multiplication ratio T c is determined as a function of the speed ratio N t /N e . The desired pressures for the on-coming and off-going clutching devices are stored as a function of the torque variable T v and time, as graphically depicted in Figure 11. For any given value of the torque variable T v , the ΔP vs. time schedule is defined by a pair of pressure end-points, one such end-point corresponding to an initial time t i , and the other corresponding to a final time t f . The time t i marks the beginning of the torque phase, and the time t f marks the end of the inertia phase. If the calculated torque variable T v is zero or near-zero, for example, the ΔP vs. time schedule is defined by the line 350 interconnecting the pressure end-points P a and P b . If the calculated torque variable T v is very high, as designated by T v (max), the ΔP vs. time schedule is defined by the line 352 interconnecting the pressure end-points P c and P d . In practice, only the four pressure end-points P a , P b , P c and P d need be stored by the control unit 270. For any calculated torque variable value T v1 between zero and T v (max), the initial pressure P x is linearly interpolated along the line 354 interconnecting the initial pressure end-points P a and P c , and the final pressure P y is linearly interpolated along the line 356 interconnecting the final pressure end-points P b and P d . In such case, the ΔP vs. time schedule for the shift would be defined by the line 358 interconnecting the initial and final pressures P x and P y . The time (t f - t i ) for a given shift is empirically derived, and is stored in the memory of the control unit 270. The pressure schedules may be defined by three or more pressure end-points if desired, using the techniques described herein. Although the pressure control algorithm set forth above provides good results in a well-calibrated system, adaptive compensation will be needed to correct for variations in system performance which affect the inertia phase torque. In conformity with the present invention, the pressure schedule of Figure 11 is adaptively compensated by developing a reference inertia phase interval t rip and comparing it with a measure of the actual inertia phase interval t ip . If the comparison indicates that t ip is too long, the pressure schedule is corrected upwardly, and conversely if the comparison indicates that t ip is too short, the pressure schedule is corrected downwardly. If the scheduled clutch pressures are developed to result in constant shift time for a given entry turbine speed N et as described above, the reference interval t rip is determined solely in relation to the entry turbine speed N te . If some input torque dependency is included to soften the off-pattern shifts as mentioned above, the reference interval t rip is determined as a function of both T i and N te . The actual inertia phase interval t ip is determined in the course of each upshift by monitoring the speed ratio N t /N o . The initial and final ratios are known, and the control unit 270 continuously computes the percentage of ratio completion, %RAT. Algebraically, %RAT is given by the expression: %RAT = |RAT meas - RAT old |/|Rat new - RAT old | where RAT meas is the actual ratio, RAT old is the ratio of the previously engaged speed ratio, and RAT new is the ratio of the desired speed ratio. The speed ratio for a typical 2-3 ratio shift is graphically represented by the trace 360 of Figure 12. In such example, the ratio changes from the second-speed ratio value of 1.273 RPM/RPM to the third-speed ratio value of 0.808 RPM/RPM. Technically, the inertia phase of the shift begins at time t₀ when the turbine speed (and hence, the ratio) begins to change, and ends at time t₃ when the ratio reaches the third-speed ratio value of 0.808 RPM/RPM. However, the initial and final non-linearity of the trace makes the measurement of the interval t₀ to t₃ somewhat difficult. To obtain a more repeatable indication of the inertia phase interval t ip and to permit reliable extrapolation of the available data, t ip is defined as the interval between 20% and 80% of ratio completion. In the example of Figure 12, the ratio change is 20% complete (1.180 RPM/RPM) at time t₁, and 80% complete (0.901 RPM/RPM) at time t₂. When a significant difference between the measured inertia phase interval t ip and the reference inertia phase interval t rip is detected, the control unit 270 develops a pressure correction amount C p as a function of such difference, and apportions the correction amount C p among the four stored pressure end-points P a , P b , P c and P d defined in Figure 11. The pressure correction amount C p is determined in a manner similar to that described above in relation to the fill time correction amount C ft . That is, a non-linear base gain schedule and direction-sensitive dynamic gain modifier similar to those described above in relation to Figure 9 are used. The pressure correction amount C p is apportioned among the stored pressure end-points P a , P b , P c and P d as a function of the torque variable T v used to schedule the shift. One portion of the correction amount C p is applied equally to the end-points P a and P b , and the remaining portion is applied equally to the end-points P c and P d . The apportionment is performed in a manner similar to that of the fill time correction amount C ft (described above in relation to Figure 10) by developing a gain term G L for the end-points P a and P b , and a gain term G H for the end-points P c and P d . The end-points P a and P b are adjusted by the amount C p * G L ), and the end-points P c and P d are adjusted by the amount (C p * G H ). When the torque variable T v is relatively low, most of the correction amount C p will be applied to the end-points P a and P b . When the torque variable T v is relatively high, most of the correction amount C p will be applied to the end-points P c and P d . Due to the adaptive correction, changing conditions which affect the inertia phase interval are compensated for after a number of such upshifts. The flow diagrams depicted in Figures 13 to 17 represent program instructions to be executed by the microcomputer 302 of the control unit 270 in mechanising ratio shifting and the adaptive control functions of this invention. The flow diagram of Figure 13 represents a main or executive program which calls various sub-routines for executing particular control functions as necessary. The flow diagrams of Figures 14 to 17 represent the functions performed by those sub-routines which are pertinent to the present invention. With reference now more particularly to Figure 13, the reference numeral 370 designates a set of program instructions executed at the initiation of each period of vehicle operation for initialising the various registers, timers etc. used in carrying out the control functions of this invention. Following such initialisation, the instruction blocks 372 to 384 are repeatedly executed in sequence as designated by the flow-diagram lines connecting such instruction blocks and the return line 386. The instruction block 372 reads and conditions the various input signals applied to the I/O device 300 via the lines 272 to 284, and updates (increments) the various control unit timers. The instruction block 374 calculates various terms used in the control algorithms, including the input torque T i , the torque variable T v , and the speed ratio N o /N i . The algebraic expressions used to calculate the terms T i and T v are given above in relation to Figure 11. The instruction block 376 determines the desired speed ratio, R des , is accordance with a number of inputs including throttle position, vehicle speed and manual valve position. In the transmission control art, this function is generally referred to as shift pattern generation. The instruction block 378 determines the clutching device pressure commands for effecting a ratio shift, if required. The pressure commands for the pressure regulator valve (PRV) and non-shifting clutching devices are also determined. An expanded description of the instruction block 378 is set forth below in relation to the flow diagrams of Figures 14 to 15. The instruction block 380 converts the clutching device and PRV pressure commands to a PWM duty cycle based on the operating characteristics of the various actuators (empirically determined), and energises the actuator coils accordingly. The instruction block 382 relates to the determination of adaptive corrections for the empirically derived clutch pressure schedules, and is discussed in more detail below in relation to Figure 17. The instruction block 384 relates to the determination of adaptive corrections for the empirically derived clutch fill times, and is discussed in more detail below in relation to Figures 16a to 16c. As indicated above, the flow diagrams of Figures 14 and 15 set forth the clutch and PRV pressure determination algorithm generally referred to at the main loop instruction block 378 of Figure 13. On entering such algorithm, the blocks designated generally by the reference numeral 388 are executed to set up initial conditions if a shift is in order. If a shift is in order, the blocks designated generally by the reference numeral 390 are executed to develop pressure commands for the clutching devices involved in the shift. Thereafter, the instruction blocks 392 and 394 are executed to develop pressure commands for the non-shifting clutches and the pressure regulator valve PRV, so completing the routine. As indicated at instruction block 394, the pressure command for the regulator valve PRV is set equal to the highest of the pressure commands for the various clutching devices. The blocks designated by the reference numeral 388 include the decision block 396 for determining whether a shift is in progress, as indicated by the "SHIFT IN PROGRESS" flag; the decision block 398 for determining whether the actual speed ratio R act (that is, N o /N t ) is equal to the desired speed ratio R des determined at instruction block 376 of Figure 13; and the instruction block 400 for setting up the initial conditions for a ratio shift. The instruction block 400 is executed only when the decision blocks 396 and 398 are both answered in the negative. In such case, the instruction block 400 serves to set the old ratio variable, R old , equal to R act , to set the "SHIFT IN PROGRESS" flag, to clear the shift timers, and to calculate the fill time t fill for the on-coming clutching device. If a shift is in progress, the execution of the blocks 398 and 400 is skipped, as indicated by the flow diagram line 402. If no shift is in progress, and the decision block 398 is answered in the affirmative, the execution of the instruction block 400 and the blocks designated by the reference numeral 390 is skipped, as indicated by the flow diagram line 404. The blocks designated by the reference numeral 390 include the decision block 406 for determining whether the shift is an upshift or a downshift; the instruction block 408 for developing pressure commands for the active (shifting) clutching devices if the shift is an upshift; and the instruction block 410 for developing the pressure commands for the active clutching devices if the shift is a downshift. To illustrate how such pressure commands are developed, the steps involved in the development of a typical power-on upshift (i.e. instruction block 408) are set forth in the flow diagram of Figure 15. On entry into the flow diagram of Figure 15, the decision block 412 is first executed to determine whether the fill phase of the shift has been completed, as indicated by the "FILL COMP" flag. If not, the flow diagram branch generally designated by the reference numeral 414 is executed; or if so, the flow diagram branch generally designated by the reference numeral 416 is executed. The flow diagram branch 414 includes a fill initialising routine comprising the blocks 418 and 420, and a fill completion routine comprising the blocks 422 and 424. At the beginning of each shift, the "FILL COMP" flag is not set, and the decision block 418 of the fill initialising routine is executed to determine whether the fill phase has started, as indicated by the "FILL START" flag. Initially, the "FILL START" flag is not set, and the instruction block 420 is executed to set the energisation duty cycle of the on-coming clutching device, DC(ONC), equal to 100%, to set the "FILL START" flag, and to start the FILL TIMER and the adaptive fill timer, AFILL TIMER. Thereafter the decision block 418 is answered in the affirmative, and execution of the instruction block 420 is skipped, as indicated by the flow diagram line 426. The decision block 422 of the fill completion routine determines whether the count in FILL TIMER is greater than or equal to the fill time t fill determined at the instruction block 400 of Figure 14. If so, the instruction block 424 is executed to set DC(ONC) equal to 0%, to save the entry turbine speed N te , and to set the "FILL COMP" flag. If the decision block 422 is answered in the negative, the fill phase is incomplete, and execution of the instruction block 422 is skipped, as indicated by the flow diagram line 428. The flow diagram branch 416 includes a shift initialising routine comprising the blocks 430 to 436, and a shift completion routine comprising the blocks 438 to 444. The decision block 430 of the initialising routine determines whether the "FILL COMP" flag has just been set, as indicated by the status of the "FIRST FILL" flag. If so, the instruction blocks 432 and 434 are executed to set up the torque and inertia phases of the shift. The instruction block 432 determines the pressure parameters P i , P f , and t f for the on-coming (ONC) and off-going (OFG) clutching devices. The instruction block 434 calculates the reference inertia phase interval t rip as a function of N te , R old , and R des , starts the timer, IP TIMER, and resets the "FIRST FILL" flag. Thereafter, the decision block 430 is answered in the negative, and the instruction block 436 is executed to calculate the value of the term %RATCOMP for use in the adaptive pressure correction algorithm. In the inertia phase completion routine, the decision blocks 438 and 440 are executed to determine whether the count in IP TIMER is at a maximum value, MAX, or whether the term %RATCOMP is substantially equal to 100%. If either of the decision blocks 438 or 440 is answered in the affirmative, the shift is complete, and the instruction block 442 is executed to reset the "SHIFT IN PROGRESS" flag, to set the on-coming duty cycle, DC(ONC), equal to 100%, and to set the off-going duty cycle, DC(OFG), equal to 0%. If both decision blocks 438 and 440 are answered in the negative, the instruction block 444 is executed to determine the on-coming and off-going pressure commands, P(ONC) and P(OFG), as a function of the P i , P f , t f and IP TIMER values. The flow diagram of Figures 16a to 16c represents an algorithm for adaptively correcting the determination of fill time t fill in conformity with the present invention. As is set forth above in relation to Figures 7 to 10, such algorithm involves the detection of turbine speed turn-down in the course of an upshift, the determination of the error E ft between measured and desired inertia phase delays, IPDELAY and DESDELAY, and the application of an error-dependent correction amount C ft to the end-points L and H of the stored t fill vs. ΔP relationship. Generally, the portions of the flow diagram depicted in Figures 16a to 16b relate to turn-down detection, and the determination of error E ft , and the portion depicted in Figure 16c relates to the application of the correction amount C ft to the end-points L and H. The flow diagram portions are joined where indicated by the circled numerals 1, 2 and 3. With reference now more particularly to Figure 16a, the decision blocks 450 to 452 refer to initial conditions which must be satisfied before enabling the turn-down detection algorithm. The detection algorithm is enabled only if a single ratio upshift is in progress (block 450), and the turn-down has not yet been detected (as determined by a "TURNDOWN" flag at block 452). If either condition is not met, execution of the algorithm is skipped, as indicated by the flow diagram return line 456. The turbine speed turn-down detection algorithm includes an initialising routine comprising the blocks 458 to 462, and an end of fill (EOF) identification routine comprising the blocks 464 to 486. As is explained above, the turn-down detection involves determining the time between pulses, T/TP, of the turbine speed signal on the line 274, and novel signal processing of the measured times. The measurement of T/TP is made with a timer, PULSE TIMER, which is reset (enabled to start counting) each time a turbine speed pulse is identified. The initialising routine is executed only when the algorithm is first enabled in the course of a shift, as indicated by the "FIRST ENABLED" flag. Once the first turbine pulse has been identified by the decision block 460, the instruction block 462 is executed to start the PULSE TIMER and to reset the "FIRST ENABLED" flag. Thereafter, the decision block 458 is answered in the negative, and the EOF identification routine is entered. As with the initialising routine, the EOF identification routine includes a decision block 464 for identifying turbine speed pulses, and an instruction block 466 that is executed each time a pulse is identified, for resetting the PULSE TIMER. Prior to resetting of the PULSE TIMER, however, the time per turbine pulse, T/TP, counted by the PULSE TIMER is read and stored. Thereafter, the instruction block 468 is executed to compute the average time between turbine pulses, AT/TP; the average change in time between turbine pulses, AΔT/TP; the predicted time between turbine pulses, PT/TP(k+2) and between loop, PT/TP(L); the loop error time between turbine pulses, ET/TP(L); and the filtered loop error time between turbine pulses, FET/TP(L). The instruction block 470 is then executed to determine the first and second error thresholds E th (1) and E th (2) as a function of the calculated FET/TP(L). Thereafter, the decision block 472 is executed to determine whether the shift has progressed to within 200 ms of the expected end of fill. If not, the instruction block 474 is executed to read the AFILL TIMER, and the remainder of the routine is skipped, as indicated by the flow diagram line 488. If the decision block 472 is answered in the affirmative, the decision blocks 476 and/or 478 are executed, to compare the error time ET/TP(L) with the thresholds E th (1) and E th (2) for determining whether a turn-down has occurred. If the error time ET/TP(L) does not exceed the first threshold E th (1), the instruction block 474 is executed to read the AFILL TIMER, and the remainder of the routine is skipped, as indicated by the flow diagram return line 488. If the time error does exceed the first threshold, the decision block 478 is executed to compare the error time with the second threshold E th (2). If the second threshold is exceeded, the turn-down detection is assumed valid, and the instruction block 480 is executed to set the "TURNDOWN" flag. If the second threshold is not exceeded, the remainder of the routine is skipped, as indicated by the flow diagram line 488. Once the turbine speed turn-down has been detected, the decision block 484 is executed to determine whether the shift is suitable for the formulation of an adaptive correction. Examples of the indicia used to make such determination include stable throttle position, positive calculated input torque T i , and suitable transmission fluid temperature. If the various parameters are not indicative of a normal-pattern shift, the remainder of the algorithm is skipped, as indicated by the flow diagram return line 488. If the parameters are indicative of a normal pattern shift, the instruction block 486 is executed, to determine the measured inertia phase delay, IPDELAY, the desired inertia phase delay, DESDELAY, and the fill time error, E ft . As is indicated at instruction block 486, IPDELAY is computed according to the difference between the count in AFILL TIMER and the scheduled t fill ; DESDELAY is determined as a function of the line pressure command PL and the shift type; and E ft is computed according to the difference (IPDELAY - DESDELAY). As is described above, the sign of the fill time error E ft indicates whether the on-coming clutching device was overfilled (negative) or underfilled (positive), and the magnitude thereof indicates the amount of error. To lessen the likelihood of an erroneous fill time correction being made due to spurious error, and to prevent unnecessary correction of the fill time due to a pressure scheduling error, the adaptive fill algorithm includes a limiting routine comprising the blocks 490 to 512. The blocks 490 to 498 operate to limit fill time corrections in response to unusually high overfill error by comparing the error term E ft with a negative reference, -REF, which corresponds to a severe overfill indication. In the event of a severe overfill indication, the error E ft is limited to a relatively small value, -E sm , until three or more of such error indications have been determined in succession. A large overfill counter, LG OVF COUNTER, is used to keep track of the number of successive overfill indications. When a severe overfill is indicated (as sensed by the decision block 490), the instruction block 492 is executed to increment the LG OVF COUNTER; when a smaller overfill is indicated, the instruction block 494 is executed to decrement the LG OVF COUNTER. Until the LG OVF COUNTER has been incremented to three or greater (as determined by decision block 496), the instruction block 498 is executed to limit the error E ft to a relatively small overfill indication, -E sm . When the LG OVF COUNTER has been incremented to three or greater, the limit is no longer effective. The blocks 500 to 502 operate in response to a large positive inertia phase error E ip (explained below with reference to Figure 17) to limit positive (underfill) fill time error E ft to a reference small value, +E sm . As is illustrated in Figure 6, improperly low pressure scheduling in the torque and inertia phases reduces the torque available for decelerating the turbine and delays the turbine speed turn-down. In such case, the late turn-down detection may be misinterpreted as an underfill error by the fill time adaptive algorithm, even if the scheduled fill time is correct. To prevent significant correction of the fill time in response to such a misinterpretation, the decision block 500 compares the inertia phase error, E ip , with a positive reference, +REF, indicative of undesirably low pressure scheduling. If the error E ip exceeds the reference, +REF, the instruction block 502 is executed, to limit the fill time error E ft to a relatively small positive value, +E sm . If the reference +REF is not exceeded, the sensed fill time error E ft is not limited. The blocks 504 to 512 operate to limit fill time corrections when the vehicle speed is so high that overfill (bind-up) is difficult to determine reliably. Under such conditions, the normal fill time corrections are permitted only if turbine speed flare is sensed, or if the fill time error E ft indicates a relatively high underfill. In all other cases, a relatively small overfill error (-E sm ) is assumed. If the assumed overfill error is actually incorrect, underfill errors will be detected in successive shifting, and the correction will be reversed. The block 504 compares the vehicle speed N v with a reference high speed indication, REF HI . If N v exceeds REF HI , the decision block 506 is executed to determine whether turbine speed flare has been detected. If so, the underfill indication is assumed to be reliable, and the error E ft is not limited. If turbine speed flare is not detected, the decision block 508 is executed to determine whether the fill time error E ft is positive and greater than a relatively high reference value, +REF. If so, the instruction block 510 is executed to set the fill time error E ft equal to a moderate positive amount, +E mod . If E ft is less than +REF, the instruction block 512 is executed to set the fill time error E ft equal to the relatively small overfill indication, -E sm . If N v is not in excess of REF HI , the execution of the blocks 506 to 512 is skipped, as indicated by the flow diagram line 514. Subsequent to the limiting routine, the instruction blocks 516 to 518 are executed to correct the fill time end-points L and H in relation to the error E ft and the time of integral of E ft . The instruction block 516 updates the time integral of E ft and calculates a number of terms including the fill time correction C ft , the end-point gain factors G H and G L , and the end-point correction amounts C LEP and C HEP . The instruction block 518 applies the end-point correction amounts C LEP and C HEP to the end-points L and H respectively. As is described above in relation to Figure 9, the correction amount C ft is determined as a function of E ft and the time integral of E ft . As is described above in relation to Figure 10, the gain factors G L and G H are determined as a function of the line pressure command PL, the respective gain factors being multiplied by the correction amount C ft to determine the end-point correction amounts C LEP and C HEP . The adaptive pressure correction algorithm is depicted by the flow diagram of Figure 17. As is described above, the algorithm comprises the steps of obtaining a measure t ip of the inertia phase interval, comparing t ip with a reference interval t rip to obtain an inertia phase error term E ip , and correcting the stored pressure end-points in relation to E ip and the time integral of E ip . The measured interval begins when the ratio shift is 20% complete, and ends when the ratio shift is 80% complete, as judged by the term %RATCOMP. The algorithm includes an initialising routine, an interval measurement routine, and a correction routine. The initialising routine comprises the blocks 520 to 526; the interval measurement routine comprises the blocks 528 to 542; and the correction routine comprises the blocks 544 to 546. In the initialising routine, the decision blocks 520 and 522 are executed to determine whether a single ratio upshift is in progress, and whether the ratio shift is at least 20% complete, as judged by the term %RATCOMP. If either of the decision blocks 520 and 522 is answered in the negative, the remainder of the flow diagram is skipped, as indicated by the flow diagram return line 550. When both decision blocks are answered in the affirmative, the decision block 524 is executed to determine whether the IP flag is set. This flag marks the beginning of the measured inertia phase interval, and is set by the instruction block 526 the first time that the decision block 524 is executed. The instruction block 526 also serves to start the IP TIMER. Thereafter, the instruction block 524 is answered in the negative, and the measurement routine is entered. In the measurement routine, the decision block 528 is executed to compare the count in the IP TIMER with the reference interval, t rip . As long as the count in IP TIMER is less than t rip , the blocks 530 to 534 are executed to stop IP TIMER at 80% completion and to calculate the inertia phase error E ip according to the difference (IP TIMER - t rip ). However, when the count in IP TIMER exceeds t rip , the blocks 536 to 542 are executed to either (1) set the error E ip at a predetermined large value, E LG , if the shift is less than 50% complete, or (2) compute the error E ip in relation to the difference between t rip and a linear extrapolation of the inertia phase time, t ip . In the latter case, the time t ip is extrapolated from the current values of IP TIMER and %RATCOMP, as indicated at the instruction block 540 by the expression: t ip = (IPTIMER*.60)/(%RATCOMP-.20) Once the inertia phase error E ip has been determined, the decision block 542 is executed to determine whether the various parameters monitored in the course of the shift are indicative of a normal-pattern shift. As is described above in relation to the adaptive fill time correction, such parameters include stable throttle position, positive torque, and satisfactory oil temperature throughout the shift. If the decision block 542 is answered in the affirmative, an adaptive pressure correction may be reliably made, and the correction routine is entered. In the correction routine, the instruction blocks 544 and 546 are executed to correct the pressure end-points P a , P b , P c and P d in relation to the error E ip and the time integral of E ip . The instruction block 544 updates the time integral of E ip , and calculates a number of terms including the inertia phase pressure correction C ip , the end-point gain factors G H and G L , and the end-point correction amounts C LEP and C HEP . This instruction block 546 applies the end-point correction amounts C LEP and C HEP to the pressure end-points. As is described above in relation to Figure 9, the correction amount C ip is determined as a function of E ip and the time integral of E ip . As is described above in relation to Figure 10, the gain factors G L and G H are determined as a function of the torque variable T v , the respective gain factors being multiplied by the correction amount C ip to determine the end-point correction amounts C LEP and C HEP . The end-point correction amount C LEP is applied to the pressure end-points P a and P b , whereas the end-point correction amount C HEP is applied to the pressure end-points P c and P d . In future shifts, the pressure supplied to the subject clutching device will result in an inertia phase interval more nearly equal to the reference interval t rip , and a more nearly optimum quality shift.

Claims: 1. A method of adaptively correcting the operation of a shift control system for a motor vehicle multiple speed-ratio automatic transmission, in which the transmission (10) has a fluid-operated torque-establishing device (28,30,32,34) associated with a specified speed ratio, a source (60) of fluid pressure, and control means (270) for adaptive correction of the operation of a control system for effecting selective engagement of the torque-establishing device (28,30,32,34), characterised in that, for correcting the operation of a control system for the transmission (10) wherein shifting from a currently engaged speed ratio to the specified speed ratio includes a shift completion phase during which fluid is supplied to the torque-establishing device (28,30,32,34) in accordance with a predetermined pressure schedule to initiate and progressively increase the transmission of torque therethrough, during the shift completion phase fluid is supplied to the torque-establishing device (28,30,32,34) by means of an open-loop control technique, and the predetermined pressure schedule is adaptively corrected for sources of error which affect the increase of torque transmission through the torque-establishing device (28,30,32,34) during the completion phase, and thereby degrade the shift quality, by the steps of: repeatedly computing a measure of the speed ratio progression during the completion phase of the shift as a function of the transmission input and output speeds; measuring the time interval required for the computed measure of speed ratio progression to change from a first predefined value to a second predefined value, and comparing such measured time interval with a reference time interval representative of the time that would be required to effect such change when the shift quality is not degraded; and adjusting the predetermined pressure schedule in relation to the deviation between the measured and reference time intervals such that the scheduled pressure in subsequent shifts to the specified speed ratio is increased if the measured time interval is significantly longer than the reference time interval, and decreased if the measured time interval is significantly shorter than the reference time interval, thereby to improve the shift quality in such subsequent shifts. 2. A method according to claim 1, characterised in that the method includes the additional step of: determining the amount of adjustment of the predetermined pressure schedule in accordance with a first term dependent on the difference between the measured and reference time intervals and a second term dependent on the integral of such difference with respect to time. 3. A method according to claim 2, characterised in that the method includes the additional step of: resetting the second term to zero when the difference between the measured and reference time intervals changes sign and has a magnitude in excess of a reference amount. 4. A method according to any one of claims 1 to 3, characterised in that the method includes the additional step of: limiting the amount of adjustment of the predetermined pressure schedule to a maximum reference determined in relation to the difference between the measured and reference time intervals. 5. A method according to any one of claims 1 to 4, characterised in that the first and second predefined values represent a portion of the speed ratio progression which is substantially linear with respect to time. 6. A method according to claim 1, characterised in that the step of repeatedly computing a measure of the speed ratio progression during the completion phase of the shift is carried out by repeatedly computing a measure of the speed ratio completion as a function of the transmission input and output speeds, and that the step of measuring the time interval and comparing the same with a reference time interval is carried out by the steps of: establishing first and second reference values which define a selected portion of the speed ratio shift progression; initiating an elapsed time measurement when the measure of the speed ratio completion reaches the first reference value, and terminating the elapsed time measurement when the measure of the speed ratio completion reaches the second reference value; and developing a time error term in relation to the difference between the elapsed time measurement and a reference time interval representative of the time that would be required to effect such change when the shift quality is not degraded. 7. A method according to claim 6, characterised by the further step of: identifying the point at which the elapsed time measurement reaches the reference time interval; and thereafter setting the time error term to a predetermined high value if the measure of the speed ratio completion is less than a third reference value intermediate the first and second reference values. 8. A method according to claim 6, characterised by the further step of: identifying the point at which the elapsed time measurement reaches the reference time interval; and thereafter predicting a future value of the elapsed time measurement as a function of the current value of the elapsed time measurement and the current measure of speed ratio shift progression if the measure of speed ratio shift progression is greater than a third reference value intermediate the first and second reference values.

Inventors: DOWNS, ROBERT CHARLES, NITZ, LARRY THEODORE, WANAMAKER, JOSEPH LEE
Applicants: GENERAL MOTORS CORPORATION
Application year: 1986
Main classification: B60K41
IPC codes: B60W10, F16H61