CLUTCH-TO-CLUTCH POWER-ON DOWNSHIFTING IN A MOTOR VEHICLE AUTOMATIC TRANSMISSION

Abstract: A clutch-to-clutch power-on downshift in a motor vehicle automatic transmission (l4) is effected by reducing the torque capacity of the off-going clutching device (28,30,32) to permit the transmission input speed to increase towards a target speed which it will finally attain subsequent to the shift; increasing the torque capacity of the off-going clutching device and preparing the on-coming clutching device when the input speed approaches its target speed, so that the input speed is held near the target speed by the time the on-coming clutching device is prepared to transmit torque; and concurrently applying and releasing the on-coming and off-going clutching devices to complete the shift. Since the input speed is substantially at its target value when the interchange of the clutching devices occurs, a high-quality shift with minimal driveline torque disruption can be obtained.

Description: This invention relates to a method of effecting clutch-to clutch speed ratio control in a motor vehicle automatic transmission. The invention is more particularly concerned with a method of effecting ratio downshifts when the vehicle is being operated in a positive-torque, power-on mode. A motor vehicle automatic transmission generally includes a number of gear elements coupling its input and output shafts together, and a related number of torque-establishing devices such as clutches and brakes which are selectively engageable to activate selected gear elements for establishing a desired speed ratio between the input and output shafts. The brake can be of the band type or of the disc type: engineering personnel in the automotive art tend to refer to disc-type brakes in transmissions as "clutches" or "reaction clutches" The input shaft is connected to the vehicle engine by way of a fluid coupling such as a hydrodynamic 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 setting and vehicle speed, and generally involves releasing or disengaging the clutch or brake (off-going) associated with the current speed ratio, and applying or engaging the clutch or brake (on-coming) associated with the desired speed ratio. Shifts performed in the above manner, using clutches and/or reaction clutches, are termed clutch-­ to-clutch shifts. Shifts of this type present particular control difficulty in situations in which a downshift to a lower speed ratio is required for acceleration of the vehicle, or to maintain vehicle speed up an incline. Such downshifts typically involve relatively high transmission input torque, and are referred to herein as power-on downshifts. In the context of a motor vehicle electronic transmission control system, the present invention is concerned with achieving high-quality clutch-to-clutch power-on downshifts without the need for free-wheel (one-way) torque-transmitting elements. To this end a method of operating torque-establishing devices of a motor vehicle automatic transmission in accordance with the present invention is characterised by the combination of features specified in claim l. A motor vehicle electronic transmission control system is advantageous in practice due to its potential to continuously monitor various engine and transmission operating parameters and precisely control the pressure supplied to the clutching devices involved in a shift. In a preferred embodiment of a method of operating torque-establishing devices of a motor vehicle automatic transmission in accordance with the present invention, a power-on downshift is initiated by reducing the torque capacity of the off-going clutching device so as to permit the transmission input speed to increase towards a target speed which it will finally attain subsequent to the shift. Thereupon, as the input speed is increasing, the value of the target speed is continuously determined as a function of the current vehicle speed and the downshifted speed ratio. When the input speed approaches its target speed, hydraulic fluid is directed to the on-coming clutching device to prepare it for engagement, and the torque capacity of the off-going clutching device is increased to a value sufficient to maintain the the input speed substantially at its target value. When the on-coming clutching device is prepared for engagement, it is applied as the off-going clutching device is released, so completing the shift. Since the input speed is substantially at its target value when the exchange of clutching devices occurs, a high-quality shift with minimal driveline torque disruption is obtainable. The above shifting technique relies on the ability of the control system to accurately control the torque capacities of the various clutching devices during the shift. To this end, various calibrated pressure and timing parameters used in the shift are adaptively updated on the basis of the performance of the system in each such shift. As a result, inaccuracies in such parameters due to mis-calibration, transmission wear etc are compensated for over a number of shifts, thereby making it possible to achieve consistent high-quality shifting. In the drawings: Figure l schematically depicts a computer-based electronic vehicular transmission control system to which a method in accordance with the present invention is to be applied; Figure 2 graphically depicts various parameters of an engine and the transmission of the vehicle in the course of a power-on downshift in conformity with the present invention; Figures 3 to 6 graphically depict terms used in the generation of a pressure command P(OFG) for an off-going clutching device in a method in accordance with the present invention; and Figures 7 to ll depict flow diagrams representative of suitable program instructions executed by the computer-based controller of Figure l for carrying out the control functions in a method in accordance with the present invention, with Figure 7 depicting a main loop program and Figures 8 to ll depicting pressure control functions for a power-on downshift. With reference now to the drawings, and more particularly to Figures la and lb thereof, reference numeral l0 generally designates a motor vehicle drive train including an engine l2 and a parallel-shaft transmission l4 having a reverse speed ratio and four forward speed ratios. The engine l2 includes a throttle mechanism l6 mechanically connected to an operator-actuatable device such as an accelerator pedal (not shown) for regulating engine output torque, such torque being applied to the transmission l4 by way of the engine output shaft l8. The transmission l4 transmits engine output torque to a pair of drive axles 20 and 22 by way of a hydrodynamic torque converter 24 and one or more of a series of fluid-operated clutching devices 26 to 34, such clutching devices being applied or released in accordance with a predetermined schedule, for establishing the desired transmission speed ratio. With reference now more particularly to the transmission l4, the impeller (input member) 36 of the torque converter 24 is connected to be rotatably driven by the output shaft l8 of the engine l2 by way of an input shell 38. The 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 (reactor) 44 redirects the fluid which couples the impeller 36 to the turbine 40, the stator member being connected by way of a one-way device 46 to the housing of the transmission l4. 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 l8 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 seen in Figure l, 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 l8 by way of the input shell 38 and impeller 36, as indicated by a 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 output line 66 by returning a controlled portion of the fluid therein to the fluid reservoir 64 by way of the line 70. In addition, the pressure regulator valve 68 supplies fluid pressure for the torque converter 24 by way of a line 74. The precise designs of the pump and pressure regulator valve are not critical to the present invention. A representative form of pump is disclosed in US-A-4 342 545, and a representative form of 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. Specifically, gear elements 80 to 88 are supported on the shaft 42, and gear elements 92 to l02 are supported on the shaft 90. The gear element 88 is rigidly connected to the shaft 42, and the gear elements 98 and l02 are rigidly connected to the shaft 90. The gear element 92 is connected to the shaft 90 by way of a free-wheel (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 l00 respectively, and the gear element 82 is coupled to the gear element 94 by way of a reverse idler gear l03. The shaft 90, in turn, is coupled to the drive axles 20 and 22 by way of gear elements l02 and l04 and a conventional differential gear set (DG) l06. A dog clutch l08 is splined on the shaft 90 so as to be axially slidable thereon, and serves to rigidly connect the shaft 90 to either the gear element 96 (as shown) or the gear element 94. A forward-speed relation is established between the gear element 84 and the shaft 90 when the dog clutch l08 connects the shaft 90 to the gear element 96, and a reverse-speed relation between the gear element 82 and shaft 90 is established when the dog clutch l08 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 l00; and the clutching device 34 couples the shaft 42 to the gear element 86. The clutching devices 28 to 34 are each 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 delta P. The circled numeral l 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 l00 are relatively sized such that the first, second, third and fourth forward-speed ratios are engaged by engaging the clutching devices 28, 30, 32 and 34 respectively, it being necessary for the dog clutch l08 to be in the position depicted in Figure l for a forward-speed ratio to be obtained. A neutral-speed ratio, involving effective disconnection of the drive axles 20 and 22 from the engine output shaft l8, is effected by maintaining all the clutching devices 28 to 34 in a released condition. The speed ratios defined by the various pairs of gear elements are generally characterised by the ratio of the turbine speed N t to output speed N o . Representative N t /N o ratios for the transmission l4 are as follows: First - 2.368 Second - l.273 Third - 0.808 Fourth - 0.585 Reverse - l.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 that 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 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 control method for supplying fluid pressure to the various clutching devices 28 to 34 to achieve consistent high-quality power-on downshifts. The fluid control elements for the tranmission l4 include a manual valve l40, a directional servo l60 and a plurality of electrically operated fluid valves l80 to l90. The manual valve l40 operates in response to operator demand, and, in conjunction with the directional servo l60, serves to direct regulated line pressure to the appropriate fluid valves l82 to l88. The fluid valves l82 to l88 are individually controlled to direct fluid pressure to the clutching devices 28 to 34. The fluid valve l80 is controlled to direct fluid pressure from the pump output line 66 to the pressure regulator valve 68, and the fluid valve l90 is controlled to direct fluid pressure from the line 74 to the clutching device 26 of the torque converter 24. The directional servo l60 operates in response to the condition of the manual valve l40, and serves to properly position the dog clutch l08. The manual valve l40 includes a shaft l42 for receiving axial mechanical input from the operator of the motor vehicle indicative of the speed range the operator desires. The shaft l42 is also connected to an indicator mechanism l44 by way of a suitable mechanical linkage, as indicated generally by the broken line l46. Fluid pressure from the pump output line 66 is applied as an input to the manual valve l40 by way of a line l48, and the valve outputs include a forward (F) output line l50 for supplying fluid pressure for the engagement of forward speed ratios and a reverse (R) output line l52 for supplying fluid pressure for engaging the reverse speed ratio. Thus when the shaft l42 of the manual valve l40 is moved to the D4, D3 or D2 position shown on the indicator mechanism l44, line pressure from the line l48 is directed to the forward (F) output line l50. When the shaft l42 is in the R position shown on the indicator mechanism l44, line pressure from the line l48 is directed to the reverse (R) output line l52. When the shaft l42 of manual valve l40 is in the N (neutral) or P (park) position, the input line l48 is isolated, and the forward and reverse output lines l50 and l52 are connected to an exhaust line l54 which serves to return any fluid therein to the fluid reservoir 64. The directional servo l60 is a fluid-operated device, and includes an output shaft l62 connected to a shift fork l64 for axially shifting the dog clutch l08 on the shaft 90 to selectively enable either a forward or a reverse speed ratio. The output shaft l62 is connected to a piston l66 axially movable within the servo housing l68. The axial position of the piston l66 within the housing l68 is determined according to the fluid pressures supplied to the chambers l70 and l72. The forward output line l50 of the manual valve l40 is connected by way of a line l74 to the chamber l70, and the reverse output line l52 of the manual valve l40 is connected by way of a line l76 to the chamber l72. When the shaft l42 of the manual valve l40 is in a forward-range position, the fluid pressure in the chamber l70 urges the piston l66 rightwardly as viewed in Figure l, to engage the dog clutch l08 with the gear element 96 for enabling engagement of a forward-speed ratio. When the shaft l42 of the manual valve l40 is moved to the R position, the fluid pressure in the chamber l72 urges the piston l66 leftwardly as viewed in Figure l, to engage the dog clutch l08 with the gear element 94 for enabling engagement of the reverse-speed ratio. In each case, the actual engagement of the second or the reverse-speed ratio is not effected until the clutching device 30 is engaged. The directional servo l60 also operates as a fluid valve for enabling the reverse-speed ratio. To this end, the directional servo l60 includes an output line l78 connected to the electrically operated fluid valve l86. When the operator selects a forward-speed ratio and the piston l66 of the directional servo l60 is in the position depicted in Figure l, the passage between the lines l76 and l78 is cut off; alternatively, when the operator selects the reverse gear ratio, the passage between the lines l76 and l78 is open. The electrically operated fluid valves l80 to l90 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 to respective clutching devices 26 to 34. The fluid valve l80 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 l82, l86 and l88 receive fluid pressure from the forward output line l50 of the manual valve l40, 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 l respectively. The fluid valve l86 receives fluid pressure from the forward output line l50 and the directional servo output line l78, 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 l90 receives fluid pressure from the 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 by way of the orifice l92, as indicated by the circled numeral 5. Each of the fluid valves l80 to l90 includes a spool element 2l0 to 220 respectively, axially movable within the respective valve body for directing fluid flow between input and output passages thereof. When a respective one of the spool elements 2l0 to 220 is in the rightmost position as viewed in Figure l, the input and output passages thereof are interconnected. Each of the fluid valves l80 to l90 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 lb. In Figure lb, the spool elements 2l0 and 2l2 of the fluid valves l80 and l82 are shown in the rightmost position, interconnecting the respective input and output lines, whereas the spool elements 2l4, 2l6, 2l8 and 220 of the fluid valves l84, l86, l88 and l90 are shown in the leftmost position, interconnecting the respective output and exhaust lines. Each of the fluid valves l80 to l90 includes a solenoid 222 to 232 for controlling the position of its spool element 2l0 to 220. Each such solenoid 222 to 232 comprises a plunger 234 to 244 connected to the respective spool element 2l0 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 energisation of the solenoid coil. As will be described, the control unit 270 pulse-width-modulates the solenoid coils 246 to 256 in accordance with 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. The fluid valves l80 to l90 have been illustrated as spool valves, but other types of valve could be substituted therefor. By way of example, valves of the ball and seat type could be used. Stated generally, the fluid valves l80 to l90 may be implemented using 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 l42 provides an input signal to the control unit 270 by way of the line 272. Speed transducers 288, 290 and 292 sense the rotational velocity of various rotary members within the transmission l4, and supply speed signals in accordance therewith to the control unit 270 by way of 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 l8 and therefore the engine speed N e . The position transducer 294 is responsive to the position of the engine throttle l6, and provides an electrical signal in accordance therewith to the control unit 270 by way of the line 280. A pressure transducer 296 senses the manifold absolute pressure (MAP) of the engine l2, and provides an electrical signal to the control unit 270 in accordance therewith by way of 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 by way of the line 284. The control unit 270 responds to the input signals on the input lines 272 to 284 in accordance with a predetermined control algorithm as stated herein, for controlling the energisation of the fluid valve solenoid coils 246 to 256 by way of 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 by way of an address-and-control bus 304 and a bi-directional data bus 306. Flow diagrams representing suitable program instructions for developing the pulse-width-modulation outputs in a method in accordance with the present invention are depicted in Figures 7 to ll. Transmission operating parameters pertinent to the power-on downshift method of this invention are graphically depicted on a common time base in Graphs a to d of Figure 2. Graph 2a depicts the pressure command P(ONC) for the on-coming clutching device; Graph 2b depicts the pressure command P(OFG) for the off-going clutching device; Graph 2c depicts the turbine speed N t ; and Graph 2d depicts the axle (output) torque T o . As is indicated in the graphs, a power-on downshift in conformity with the present invention is functionally divided into four phases A, B, C and D. Additionally, the designations t₀ to t₇ appear on the time axes to facilitate description of the sequence of events in the shift. Prior to time t₀, a relatively high speed ratio, for example third ratio, is engaged, to generate an axle torque T₀. Accordingly, the pressure command P(OFG) for the active (off-going) clutching device 32 is maintained at a relatively high steady-state value P ss , as seen in Graph 2b. Such pressure command is obtained from a predetermined and torque-dependent schedule designed to prevent any slippage. Although the clutching device 26 for the torque converter 24 could be applied, it is assumed for the purpose of this example that it is released. At time t₀, the control unit 270 commands a 3-2 downshift, marking the beginning of the A phase as so designated in Figure 2. The shift may be commanded in response to either a decrease in vehicle speed or an increase in the setting of the engine throttle l6, and is thus considered to be a power-on downshift. In the A phase of the shift, the pressure command P(OFG) for the off-going clutching device 32 is progressively reduced in accordance with a predetermined torque and time-dependent schedule, as seen in Graph 2b. Such schedule is referred to herein as the off-going release schedule, and is graphically depicted in Figure 3, where the pressure command is designated P(OFG) and the torque variable is designated T v . The pressure schedule shown in Figure 3 is matched to the steady-state pressure schedule referred to in relation to Graph 2b, so that the initial pressure command P(OFG) obtained from the phase A schedule of Figure 3 (at time t₀) corresponds to the steady-state pressure command P ss in effect prior to the shift. The clutching device 32 begins to slip at time t₁, and the turbine speed N t begins to increase with engine speed N e as seen in Graph 2c. Due to the slipping, the axle torque T o begins to decrease, as seen in Graph 2d. When the slip, delta N ofg , between the friction elements of the off-going clutching device 32 exceeds a turbine speed-dependent reference, REF ofg , the A phase of the shift is ended, and the B phase has begun. The slip, delta N ofg , may be thought of as the difference between the turbine speed N t and the output speed n o as reflected by way of the differential gearset l06 to the transmission shaft 90. Accordingly, the slip reference REF ofg is graphically depicted with respect to an extension of the turbine speed trace in Graph 2c. In phase B of the shift, the reduction in the pressure command P(OFG) for the clutching device 32 ceases, so allowing the turbine speed N t to continue increasing, as seen in Graphs 2b and 2c. During this time, the axle torque T o remains relatively constant with the pressure command P(OFG), as seen in Graph 2d. With reference to the off-going release schedule of Figure 3, the phase B pressure command P(OFG) is obtained by freezing the time variable at the A-B transition. Thus the pressure command P(OFG) retains its torque dependency after time t₂, and can vary along the trace 3l0 in Figure 3. At time t₃, the turbine speed N t rises to within a reference amount REF onc of its target value N tar for the downshifted (second-speed) ratio, so ending the B phase of the shift. Such target value is represented by the broken trace 3l2 in Graph 2c, and is easily computed as a function of the output speed N o and the speed ratios provided by the differential gear set l06 and the gear elements 84 and 96 effective for the second ratio. The reference amount REF onc is determined as a function of the turbine speed, and physically refers to the slip between the friction elements of the on-coming clutching device 30: it is graphically depicted in Graph 2c with respect to the target speed trace 3l2. As indicated above, the references REF onc and REF ofg are turbine speed-dependent. More particularly, the references REF onc and REF ofg increase with increasing turbine speed, so permitting a faster rate of increase in turbine speed during relatively high-speed shifting. This tends to shorten the duration of the B phase of the shift with increasing turbine speed, thereby preventing the duration of high-speed shifts from becoming excessive. In phase C of the shift, the pressure command P(ONC) for the on-coming clutching device 30 is increased to a torque-dependent level P f , as seen in Graph 2a, to fill the apply chamber thereof in preparation for clutch engagement. At the same time, the pressure command P(OFG) for the off-going clutching device 32 is increased to a dynamic value P dyn , as seen in Graph 2b, for holding the turbine speed N t slightly higher than the second-ratio target speed depicted by the trace 3l2. As explained below, the dynamic pressure command P dyn is determined in accordance with an open-loop term P ol , a closed-loop term P cl and an adaptive term P ad . The open-loop term P ol is determined in accordance with a predetermined schedule as a function of torque and time as with the off-going release schedule of Figure 3; the closed-loop term P cl is determined as a function of the difference between the turbine speed N t and the target speed N tar for the downshifted speed ratio; and the adaptive term P ad is determined as a function of turbine speed N bc at the transition between the B phase and the C phase. The increasing pressure command causes an increase in the torque capacity of the off-going clutching device 32, and the axle torque T o returns approximately to its pre-shift value T₀. In practice, it has been found that the turbine speed N t should be held slightly above the target speed N tar for a minimum hold time t hold regardless of the time required to fill the on-coming clutching device. However, it is not desirable to hold the on-coming clutching device in a filled condition. Thus, the start of fill may be delayed until after time t₃ if the computed fill time t fill is less than the minimum hold time t hold . At time t₄, the on-coming clutching device 30 is filled, marking the end of phase C and the beginning of phase D. At such time, the pressure command P(ONC) for the on-coming clutching device 30 is reduced to a scheduled (torque-dependent) initial value P i , and then progressively increased as a function of the torque variable T v and time, as seen in Graph 2a. As the torque capacity of the on-coming clutching device 30 increases, the turbine speed N t decreases towards the target speed for the downshifted speed ratio, and the axle torque T o increases towards its post-shift level, T f . At time t₅, a calibrated time after time t₄, the pressure command P(OFG) is progressively reduced to zero, as seen in Graph 2b, to progressively release the off-going clutching device 32. At time t₆, the off-going clutching device 32 is fully released, and at time t₇, the on-coming clutching device 30 is fully applied, so ending phase D and completing the 3-2 downshift. The control method just described permits the achievement of consistently high-quality clutch-to-clutch power-on downshifts without the use of free-wheel (one-way) torque-transmitting elements. The ability to electronically schedule the apply forces of the various torque-establishing devices is used to bring the transmission input speed (turbine speed N t ) to a predetermined value in relation to its target speed in the downshifted speed ratio prior to the engagement of the on-coming clutching device, thereby minimising the driveline torque disruption associated with the shift. The adaptive pressure command term P ad ensures that the shift quality will not be degraded due to miscalibration, wear or other sources of error which affect the ability of the controls to accurately supply the commanded pressures. The schedule for the adaptive term P ad of the pressure command P(OFG) is graphically depicted in Figure 4. As has been indicated above, the adaptive term P ad is scheduled as a function of the turbine speed N bc at the transition between the B and C phases of the shift. As is set forth in Figure 4, the adaptive term P ad is defined by the two points L and H which correspond to the reference low speed N lo and the reference high speed N hi . The adaptive term P ad for any value of N bc falling between N lo and N hi is linearly interpolated along the line 3l6 connecting the points L and H. Each of the points L and H may be raised or lowered as indicated in Figure 4 to adjust the value of the adaptive term P ad if desired. The adjustment of the adaptive term P ad used in the development of the pressure command P(OFG) in the course of a power-on downshift is graphically depicted in Figures 5 and 6. At the transition from the C phase to the D phase, the difference, delta N t , between the turbine speed N t and the target speed N tar for the downshifted speed ratio is measured. As is noted above in relation to Graph 2c, the pressure command P(OFG) is designed to maintain the turbine speed N t slightly higher than the target speed N tar . At the end of each power-on downshift, the measured difference, delta N t , is compared with a reference difference REF dif , to arrive at a turbine speed error term E ts , which in turn is used to adjust the adaptive pressure term P ad . The graph of Figure 5 depicts an adaptive pressure correction term C ad as a function of the turbine speed error N ts , and the graph of Figure 6 illustrates how the correction term C ad is applied to the points L and H defined in Figure 4. With reference particularly to Figure 5, the pressure correction term C ad is determined as a function of turbine speed error E ts in accordance with a non-linear base gain schedule and a non-linear direction-sensitive dynamic gain modifier. The base gain schedule is depicted by the trace 320, and is relatively low for achieving relatively low converged misadjustment of the adaptive pressure term P ad ; the dynamic gain modifier serves to increase the base 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 depicted by the trace 322, which gain is dependent on the error magnitude. The modifier is reset to zero when a significant error of the opposite sign is detected. Turbine speed error that is positive in sign (N diff too large) produces a positive correction, for increasing the adaptive pressure term P ad . Turbine speed error that is negative in sign (N diff too small) produces a negative correction, for decreasing the adaptive pressure term P ad . 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 the error signal E ts in one direction. Graphically, the shaded areas between the traces 320 and 322 represent the authority range of the dynamic gain modifier. In this way, the adaptive pressure correction C ad is determined primarily in accordance with the base schedule gain when the distribution of sensed error is centred at or near zero error. When the distribution of sensed error becomes significantly skewed in either the positive or the negative 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 sensed error and increased time required to correct the error. With reference particularly to Figure 6, the adaptive pressure correction C ad is apportioned between the points L and H of Figure 4 in accordance with the turbine speed N tbc at the transition between the B and C phases of the shift. The implementation of such apportionment is graphically depicted in Figure 6 by the traces 324 and 326. The trace 324 represents a gain factor G L for the end point L, and the trace 326 represents a gain factor G H for the end point H. After each power-on downshift for which an adaptive pressure correction is in order, the end point L is adjusted by the amount (C ad * G L ), and the end point H is adjusted by the amount (C ad * G H ). Separate pressure terms P ad and corresponding points L and H are provided for each type of power-on downshift (eg 3-2 or 4-2), so that in future power-on downshifts the calculated pressure command P(OFG) will cause the speed difference N diff to more nearly correspond to the reference difference REF dif . As a result, changing conditions which affect the ability of the control system elements to accurately schedule the off-going pressure command P(OFG) will be fully compensated for over a number of power-on downshifts of a given type. The flow diagrams depicted in Figures 7 to ll represent program instructions to be executed by the microcomputer 302 of the control unit 270 in implementing the power-on downshift control method of this invention. The flow diagram of Figure 7 represents a main or executive program which calls various sub-routines for executing particular control functions as necessary. The flow diagrams of Figures 8 to ll represent the functions performed by those sub-routines which are pertinent to the present invention. With reference now more particularly to Figure 7, 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 380 are repeatedly executed in sequence, as designated by the flow diagram lines connecting such instruction blocks and the return line 382. The instruction block 372 reads and conditions the various input signals applied to the I/O device 300 by way of 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 torque variable T v . The algebraic expression used to calculate the term T v is given in EP-A-86308549.4 (Serial No ), which is directed to the development of upshift pressure commands. The instruction block 376 determines the desired speed ratio, R des , in accordance with a number of inputs including throttle position, vehicle speed and manual valve position. In transmission control, 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 given below in relation to the flow diagrams of Figures 8 to ll. 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 flow diagrams of Figures 8 to ll set forth the clutch and PRV pressure determination algorithm generally referred to at the main loop instruction block 378 of Figure 7. With reference particularly to Figure 8, 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 is indicated at the 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 the instruction block 376 of Figure 7; 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 a 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 4l0 for developing the pressure commands for the active clutching devices if the shift is a downshift. Since this invention is concerned with the development of pressure commands for power-on downshifts, the instruction block 4l0 is set forth in greater detail in the flow diagrams of Figures 9 to ll. On entry into the flow diagram of Figure 9, the decision block 4l2 is first executed, to determine whether the commanded downshift is at low torque and in response to a re-positioning of the manual valve l40. If so, the downshift can be performed in the traditional manner by merely releasing the active clutching device and applying the clutching device for the downshifted ratio. This type of shift is referred to as a power-off downshift, and the instruction block 4l4 is executed, to set the "POWER-OFF" flag. The decision block 4l6 determines whether the "POWER-OFF" flag has otherwise been set. If it is determined that a power-off downshift is in order, the remainder of the flow diagram of Figure 9 is executed, to fill the on-coming clutching device, and then to release the off-going clutching device and apply the on-coming clutching device. This procedure is carried out by first determining at decision block 4l8 whether the fill phase of the shift is completed, as indicated by the "FILL COMP" flag. If not, the flow diagram branch generally designated by the reference numeral 420 is executed; if so, the flow diagram branch generally designated by the reference numeral 422 is executed. The flow diagram branch 420 includes a fill initialising routine comprising the blocks 424 and 426, and a fill completion routine comprising the blocks 428 and 430. At the beginning of each power-off downshift, the "FILL COMP" flag is not set, and the decision block 424 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 426 is executed, to set the energisation duty cycle of the on-coming clutching device, DC(ONC), equal to l00%, to set the "FILL START" flag, and to start the FILL TIMER. Thereafter the decision block 424 is answered in the affirmative, and execution of the instruction block 426 is skipped, as indicated by the flow diagram line 432. The decision block 428 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 8. If so, the instruction block 430 is executed, to set DC(ONC) equal to 0%, and to set the "FILL COMP" flag. If the decision block 428 is answered in the negative, the fill phase is incomplete, and execution of the instruction block 430 is skipped, as indicated by the flow diagram line 434. The flow diagram branch 422 includes a shift initialising routine comprising the blocks 436 to 442, and a shift completion routine comprising the blocks 444 to 450. The decision block 436 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 438 and 440 are executed, to set up the torque and inertia phases of the shift. The instruction block 438 determines the pressure parameters P i , P f and t f for the on-coming (ONC) and off-going (OFG) clutching devices in accordance with a predetermined, torque-dependent pressure schedule of the general type depicted in Figure 3. The instruction block 440 starts the inertia phase timer IP TIMER, and resets the "FIRST FILL" flag. Thereafter, the decision block 436 is answered in the negative, and the instruction block 442 is executed, to calculate the percentage of ratio completion, %RATCOMP. In the inertia phase completion routine, the decision blocks 444 and 446 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 l00%. If either of the decision blocks 444 or 446 is answered in the affirmative, the shift is complete, and the instruction block 448 is executed, to reset the "SHIFT IN PROGRESS" flag, to set the on-coming duty cycle DC(ONC) equal to l00%, to set the off-going duty cycle DC(OFG) equal to 0%, and to reset the "POWER-OFF" flag. If both decision blocks 444 and 446 are answered in the negative, the instruction block 450 is executed, to determine the on-coming and off-going pressure commands, P(ONC) and P(OFG), as a function of the scheduled P i , P f , t f and IP TIMER values. If the "POWER-OFF" flag is not set, the commanded shift is a power-on downshift, and the flow diagrams of Figures l0 to ll are executed, to effect the shift as indicated by the circled numeral one ("l") in Figures 9 and l0, and the circled letters A to D in Figures l0 and ll. The flow diagram of Figure l0 serves to monitor the progress of the shift, and to establish the different phases of the shift; it is referred to herein as the shift phase-establishing routine. Four flags -- "A", "B", "C" and "D" -- are used to identify which phase of the shift is in effect, such flags corresponding to the letters A, B, C and D used in the graphs of Figure 2. The flow diagrams of Figures lla to lld develop the pressure commands P(ONC) and P(OFG) during the phases A to D respectively. The flow diagram of Figure l0 is initiated by determining at the decision block 452 whether the "A", "B", "C" or "D" flags are set. If not, the shift has just been commanded, and the instruction block 454 is executed, to set the "A" flag. Then the decision block 456 is executed, to determine whether the slip speed of the off-going clutching device delta N ofg is negative in sign and greater in magnitude than a reference, -REF. If so, the power-on downshift control is not appropriate, and the instruction block 458 is executed, to set the "POWER-OFF" flag and to clear the various timers, so that the shift can be re-started as a power-off downshift. As indicated by the circled letter A, the flow diagram of Figure lla is then executed, to schedule the pressure command P(OFG) for progressive release of the off-going clutching device as described above in relation to Graph 2b. With reference to that flow diagram, the instruction block 460 is executed, to increment an off-going timer, OFG TIMER. Then the decision block 462 is set, to determine whether the count in OFG TIMER is greater than a relatively long time reference T ref . If not, the shift is proceeding normally, and the instruction block 464 is executed, to determine the pressure command P(OFG) for the off-going clutching device from the predetermined schedule of Figure 3 as a function of the torque variable T v and the count (t) in the OFG TIMER. If the decision block 462 is answered in the affirmative, it is assumed that the shift should be performed with the power-off routine of Figure 9, and the instruction block 466 is executed, to set the "POWER-OFF" flag, and to clear the various timers so that the shift can be re-started as a power-off shift. As long as the "A" flag is set, the decision block 468 of the shift phase-establishing routine of Figure l0 is executed, to determine whether the "B", "C" or "D" flags are set. If the decision block 468 is answered in the negative, the shift is in the A phase, and the decision block 470 is executed, to compare the slip speed of the off-going clutching device delta N ofg with the reference slip REF ofg . If the measured slip is less than or equal to the reference slip, the phase A of the shift is still appropriate, and the flow diagram of Figure lla is re-executed. If the measured slip is greater than the reference slip, the instruction block 472 is executed, to make the transition from phase A to phase B by setting the "B" flag and storing the count from the OFG TIMER in the term t ab . As is indicated by the circled letter B, the flow diagram of Figure llb is then executed, to hold the pressure command P(OFG) whilst the turbine speed N t rises with the engine speed N e , as described above in relation to Graphs 2b and 2c. With reference to that flow diagram, the instruction block 474 is executed, to determine the pressure command P(OFG) from the predetermined schedule of Figure 3 as a function of the torque variable T v and the stored time t ab . Then the decision block 476 is executed, to compare the current turbine speed N t with a prior turbine speed value N t (old). If the current turbine speed value is not greater than the prior turbine speed value, the shift is not progressing as desired, and the instruction block 478 is executed, to effect a further reduction in the pressure command P(OFG) by setting P(OFG) = k * P(OFG), where k is a calibrated constant of less than unity (l.0). Further such reductions in the pressure command. P(OFG) are effected until the current turbine speed N t exceeds the prior turbine speed N t (old). At such time, the execution of the instruction block 478 is skipped, as indicated by the flow diagram line 480. As long as the "B" flag is set, the decision block 482 of the shift phase-establishing routine of Figure l0 is executed, to determine whether the "C" or "D" flags are set. If the decision block 482 is answered in the negative, the shift is in the B phase, and the decision block 484 is executed, to compare the slip speed of the on-coming clutching device delta N onc with the reference slip value REF onc . If the measured slip is greater than the reference slip, the B phase is still appropriate, and the flow diagram of Figure llb is re-executed, as indicated by the circled letter B. If the measured slip is less than or equal to the reference slip, the instruction block 486 is executed, to make the transition from the B phase to the C phase by setting the "C" flag and storing the turbine speed N t at the transition in the term N bc . As is indicated by the circled letter C, the flow diagram of Figure llc is then executed, to start the fill of the on-coming clutching device, and to increase the pressure command P(OFG) of the off-going clutching device for holding the turbine speed slightly above the target speed for the downshifted speed ratio, as described above in relation to Graphs 2a to 2c. With reference to that flow diagram, the instruction block 488 is executed, to determine the pressure command P(OFG) in accordance with an open-loop term P ol , a closed-loop term P cl and an adaptive term P ad . The open-loop term P ol is determined in accordance with a predetermined schedule as a function of the torque variable T v and time, as with the off-going release schedule of Figure 3; the closed-loop term P cl is determined as a function of the difference between the turbine speed N t and the target speed for the downshifted speed ratio; and the adaptive term P ad is determined as a function of turbine speed N bc stored by the instruction block 486 at the transition between the B phase and the C phase. The decision block 490 is then executed, to determine whether the "FILL START" flag is set. As described in relation to the power-off downshift routine of Figure 9, the "FILL START" flag is initially reset, and the decision block 490 is answered in the negative. In such a case, the decision block 492 is then executed, to determine whether the fill time t fill computed at the instruction block 400 of Figure 8 is less than the predetermined hold time, t hold . If not, the instruction block 494 is executed, to begin filling the on-coming clutching device by setting the "FILL START" flag, setting the duty cycle DC(ONC) of the solenoid valve associated with the on-coming clutching device to l00%, and starting the FILL TIMER. If the fill time t fill is less than the hold time t hold , the instruction block 496 is executed, to increment a HOLD TIMER, and the decision block 498 is executed, to determine whether the count in the HOLD TIMER is greater than or equal to the difference (t hold - t fill ). If the decision block 498 is answered in the affirmative, the instruction block 494 is executed, to start the fill as described above. If the decision block 498 is answered in the negative, the fill is delayed. From the above description, it will be evident that the fill of the on-coming clutching device does not necessarily begin at the start of the C phase of the shift. Rather, the start of fill is determined in relation to the predetermined hold time t hold , so that the duration of the C phase will be at least t hold . As indicated above in relation to Graph 2c, this is done to ensure that the turbine speed is stabilised at a value slightly greater than the target speed N tar by the time of the transition between the C and D phases. Once the instruction block 494 has been executed, to start the fill of the on-coming clutching device, the decision block 490 is answered in the affirmative, and the decision block 500 is executed, to determine whether the count in FILL TIMER is at least as great as the fill time t fill . If so, the fill is complete, and the instruction block 502 is executed, to set the duty cycle of the solenoid valve associated with the on-coming clutching device to 0%, and to set the "FILL COMP" flag. If the fill is not complete, the execution of the instruction block 502 is skipped, as indicated by the flow diagram line 504. As long as the "C" flag is set, the decision block 506 of the shift phase-establishing routine of Figure l0 is executed, to determine whether the "D" flag is set. If the decision block 506 is answered in the negative, the shift is in the C phase, and the decision block 508 is executed, to determine whether the "FILL COMP" flag is set. If not, the C phase is still appropriate, and the flow diagram of Figure llc is re-executed, as indicated by the circled letter C. If the "FILL COMP" flag is set, the instruction block 5l0 is executed, to make the transition from the C phase to the D phase, by setting the "D" flag and storing the difference between the turbine speed N t and the target speed N tar at the transition in the term delta N t . Such speed difference is used to adjust the adaptive term P ad of the pressure command P(OFG), as indicated above in relation to Figures 4 to 6. Then, as indicated by the circled letter D, the flow diagram of Figure lld is executed, to apply the on-coming clutching device and, a predetermined time thereafter, to release the off-going clutching device, as described above in relation to Graphs 2a and 2b. In the flow diagram of Figure lld, the decision block 5l2 determines whether an "ADAPTIVE" flag is set, such flag being used to identify the first execution of the D routine in a given shift. Initially, the "ADAPTIVE" flag is not set, and the instruction block 5l4 is executed, to set the "ADAPTIVE" flag, to compute the turbine speed error term E ts in accordance with the difference (delta N t - REF diff ), to update the adaptive points L and H, as described in relation to Figures 4 to 6 above, and to start a D PHASE TIMER. In subsequent execution of the D routine during the shift, the "ADAPTIVE" flag is set, and the execution of instruction block 5l4 is skipped, as indicated by the flow diagram line 5l6. Then the instruction block 5l8 is executed, to determine the pressure command P(ONC) for the on-coming clutching device as a function of the torque variable T v and the count t in the D PHASE TIMER, as described above in relation to Graph 2a. When the count in the D PHASE TIMER exceeds a reference delay t del , as determined at the decision block 520, the instruction block 522 is executed, to determine the pressure command P(OFG) for the off-going clutching device in accordance with the expression P(OFG)= P(OFG)*K, where K₁ is a constant and K=K-K₁. Until the count t in the D PHASE TIMER exceeds t del , the execution of the instruction block 522 is skipped, as indicated by the flow diagram line 524. As long as the "D" flag is set, the decision block 526 of the shift phase-establishing routine of Figure l0 is executed, to determine whether the on-coming clutching device is fully engaged. This may be determined by comparing the pressure command P(ONC) with the scheduled final pressure P f . Once the on-coming clutching device is fully engaged, the instruction block 528 is executed, to reset the "A", "B", "C" and "D" flags, so completing the power-on downshift routine.

Claims: 1. A method of operating torque-establishing devices so as to downshift a transmission from an upper speed ratio to a lower speed ratio when a downshift is needed to satisfy operator demand, in a motor vehicle including an engine (l2) connected to drive the vehicle by way of an automatic transmission (l4), the engine (l2) being effective to develop output torque in relation to operator demand and the transmission (l4) including a fluid-operated torque-establishing device (28,30,32) associated with each of an upper and a lower speed ratio mechanism, which devices (28,30,32) are selectively engageable to transmit engine torque by way of their respective speed ratio mechanisms, the method comprising the steps of: progressively reducing the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio, thereby allowing the engine (l2) to increase the transmission input speed; increasing the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio to a holding value when the transmission input speed increases substantially to a target speed corresponding to the transmission input speed that would be in effect if the torque-establishing device (28,30,32) associated with the lower speed ratio were engaged, the holding value being scheduled to hold the transmission input speed substantially at the said target speed; and progressively engaging and disengaging the torque-establishing devices (28,30,32) associated with the lower and upper speed ratio mechanisms respectively, thereby to complete the downshift. 2. A method of operating torque-establishing devices according to claim l, characterised in that the method comprises the steps of: initiating a first mode of operation (A) for effecting the said progressive reduction in the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio, thereby allowing the engine (l2) to effect a progressive increase in the transmission input speed; initiating a second mode of operation (B) for stopping the reduction of fluid pressure effected during the said first mode of operation (A) when such reduction results in a significant increase in the transmission input speed, thereby allowing the transmission input speed to continue its progressive rate of increase; initiating a third mode of operation (C) for effecting the said increase in the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio, the holding value being such that the transmission input speed is held substantially constant at a value determined in relation to the said target speed; and initiating a fourth mode of operation (D) for progressively increasing and decreasing the fluid pressure supplied to the torque-establishing devices (28,30,32) associated with the lower and upper speed ratio mechanisms respectively after the transmission input speed has been held at the target speed-related value for at least a predetermined time, thereby to complete the downshift by progressively effecting the said engagement of the lower speed ratio torque-­ establishing device (28,30,32) and disengagement of the upper speed ratio torque-establishing device (28,30,32). 3. A method according to claim 2, characterised by the further steps of: monitoring the change in transmission input speed during the first mode of operation (A), and comparing such change with a first reference value indicative of a significant amount of slippage across the torque-establishing device (28,30,32) associated with the upper speed ratio; and initiating the second mode of operation (B) when the monitored speed change exceeds the first reference value. 4. A method according to claim 3, characterised by the further step of: scheduling the first reference value as a function of the transmission speed, the first reference value increasing with increasing transmission input speed so as to effect an increase in the rate of change of transmission input speed during the second mode of operation (B), thereby to reduce the downshift time during relatively high-speed operation. 5. A method according to any one of claims 2 to 4, characterised by the further steps of: monitoring the transmission input speed relative to the target speed during the second mode of operation (B), and comparing the difference therebetween with a second reference value indicative of a predetermined amount of slippage across the torque-establishing device (28,30,32) associated with the lower speed ratio; and initiating the third mode of operation (C) when the comparison indicates that the slippage across the said torque-establishing device (28,30,32) is less than the predetermined amount. 6. A method according to claim 5, characterised by the further step of: scheduling the second reference value as a function of the transmission input speed, the second reference value increasing with increasing transmission input speed so as to cause the third mode of operation (C) to be initiated earlier than would otherwise occur, thereby to reduce the downshift time during relatively high-speed operation. 7. A method according to any one of claims 2 to 6, characterised by the further steps of: determining an engine torque-related variable during the course of the downshift; reducing the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio during the first mode of operation (A) in accordance with a pressure command previously determined and stored as a combined function of time and the said torque variable; and freezing the time dependency of the said pressure command at the initiation of the second mode of operation (B), whereafter the pressure command retains its torque-variable dependency, and results in a controlled increase in the transmission input speed during the second mode of operation. 8. A method according to any one of claims 2 to 7, characterised by the further step of: controlling the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio during the third mode of operation (C) in accordance with a pressure command determined as a combined function of l) an empirically derived and stored open-loop pressure term, and 2) a closed-loop pressure term determined in relation to the deviation between a reference difference and the actual difference between the transmission input speed and the target speed, thereby to maintain a predetermined relation between the target speed and the transmission input speed during the third mode of operation (C) irrespective of any speed disturbances occurring therein. 9. A method according to claim 8, characterised by the further steps of: determining the pressure command during the third mode of operation (C) as an additional function of an adaptive pressure term; and periodically updating the value of the adaptive pressure term following the initiation of the third mode of operation (C) of a downshift in accordance with the difference between the transmission input speed and the target speed at the initiation of the fourth mode of operation, thereby to compensate for changing conditions which could affect the accuracy of the pressure command. l0. A method according to any one of claims 2 to 9, characterised in that the engagement of the torque-establishing device (28,30,32) associated with the lower speed ratio includes a fill phase during which fluid is supplied to the torque-establishing device (28,30,32) for a predetermined fill time to prepare it for engagement, and a completion phase during which the fluid pressure supplied to the device (28,30,32) is progressively increased to complete the engagement thereof, and that the method includes the further steps of: establishing a comparison between the difference between the predetermined fill time of the torque-establishing device (28,30,32) associated with the lower speed ratio and the predetermined hold time for which the transmission input speed is held in relation to the target speed in the third mode of operation (C); and initiating the fill phase of the torque-establishing device (28,30,32) associated with the lower speed ratio in relation to the said comparison such that the end of the fill phase coincides with the end of the predetermined hold time, whereby the torque-establishing device (28,30,32) associated with the lower speed ratio is prepared for engagement at the initiation of the fourth mode of operation (D). 11. A method according to any one of claims 2 to l0, characterised by the further steps of: timing the operation of the first mode of operation (A); and overriding the second, third and fourth modes of operation (B), (C) and (D), and forcing completion of the downshift if the timed duration exceeds a reference duration. 12. A method according to any one of claims 2 to ll, characterised by the further steps of: sensing the rate of change of transmission inout speed during the second mode of operation (B) relative to a relatively low reference rate of change; and decreasing the fluid pressure supplied to the torque-establishing device (28,30,32) associated with the upper speed ratio as long as the sensed rate is less than the relatively low reference rate, to thereby force the progression of the downshift.

Inventors: DOWNS, ROBERT CHARLES, NITZ, LARRY THEODORE
Applicants: GENERAL MOTORS CORPORATION
Application year: 1987
Main classification: B60K41
IPC codes: F16H59, F16H61, B60W10, B60K20, F16D48