Patent Grant
10166989
The invention relates to a multiple ratio transmission mechanism in a powertrain for an automotive vehicle and to a strategy for achieving smooth engagement and release of friction torque establishing elements during a transmission ratio upshift
A step-ratio automatic transmission uses multiple friction elements for automatic gear ratio shifting. A ratio change from a low gear ratio to a high gear ratio occurs in a synchronous clutch-to-clutch upshift as one friction element is engaged and a second friction element is disengaged. One friction element may be referred to as an off-going clutch (OGC). It is released while a second friction element, which may be referred to as an oncoming clutch (OCC), engages to create the upshift. The upshift event is divided into a preparatory phase, a torque phase and an inertia phase. During the preparatory phase, the OCC actuator is stroked to prepare for its engagement, while the OGC torque-holding capacity is reduced to prepare for its release. During the torque phase, the OCC torque is raised in a controlled manner while the OGC is still engaged or allowed to slip at a controlled slip rate.
Simultaneous engagement of the OCC and release of OGC in a conventional transmission upshift may result in a momentary activation of two torque flow paths through the gearing. During the torque phase, the lower gear speed ratio from input to output is maintained. However, as the OCC gains torque capacity and the OGC loses it, more of the input torque is routed through higher gear path until, when the OGC no longer has any capacity, all of the torque is routed through the higher gear path, which has a lower torque ratio. Thus, in the small timespan of the torque transfer, the input torque goes from being multiplied by a higher amount to a lower amount before the inertia of the subsequent speed change raises the output torque again. This momentary dropping and subsequent rise of output torque is known as the “torque hole.” This is perceived by a vehicle occupant as an unpleasant shift shock. The inertia phase begins when the OGC is released or has no significant torque capacity.
An objective of the present invention is to eliminate or reduce the torque hole effect while reducing transient torque disturbances during an upshift event. A transmission controller can provide estimated friction element torque targets using friction element actuator pressures in the case of a transmission control system with pressure operated actuators. The controller executes control algorithms in a software control strategy without knowing actual torque profiles for the oncoming and off-going friction elements.
Powertrain sensors, in a control system embodying the present invention, provide direct reading of operating variables such as output torque. They are used, together with physical properties and functions of the transmission and driveline components, algorithms governing those functions and appropriate transfer functions, to provide accurate torque values for the oncoming and off-going clutches. The sensors provide torque feedback signals for correcting estimates of friction element torque in a closed loop fashion during calculations of actuator pressures.
The invention includes a control strategy for coordinating the actuators to achieve minimal torque disturbance at the output shaft. The algorithms, if used without the actual torque feedback provided by the sensors, do not have the ability to accurately estimate in real time the clutch torques during an upshift. They determine what the desired clutch torque would be and assume, based on a clutch model, that this torque will be delivered using a calibrated transfer function between clutch pressure command and clutch torque. The clutch actuators, however, are non-linear and their response to control pressures is affected by variables such as transmission oil temperature and other environmental factors. This can result in OCC torque transients or disturbances.
The invention uses a feedback control that uses one or more sensors (e.g., torque sensors) to develop an actual, real time sensor feedback (e.g., torque feedback) to calculate an oncoming friction element torque to ensure that the oncoming friction element torque tracks a target torque and to obtain a desired off-going friction element torque to obtain a controlled slip of the off-going friction element. A torque sensor signal that is used to calculate current corrected oncoming and off-going friction element torque values is a direct torque measurement. For example, a torque sensor can be located at a transmission torque input shaft or at a transmission torque output shaft, or at both locations. Torque at other locations can be calculated using the direct readings for the sensors.
When the transmission input and output torques are known, the friction element torques can be calculated during the shift using a technique that is disclosed in U.S. application Ser. No. 12/861,387, filed Aug. 23, 2010, now U.S. Pat. No. 8,510,003, issued Aug. 13, 2013, which is assigned to the assignee of the present invention. Reference also may be made to U.S. Patent Publication 2010/0262344, filed Apr. 9, 2009, now U.S. Pat. No. 8,255,130, issued Oct. 14, 2010 which also is assigned to the assignee of the present invention. Those references explain, for example, how to estimate the input shaft torque if only the output shaft torque is measured, and vice versa.
By knowing the friction element torques, performance and predictability of the algorithms can be improved because it is possible to determine if a friction element torque is actually achieved and to provide accurate modulation of the OCC actuator pressures so that torque transients at the OCC are minimized as the OGC has a controlled slip. The target level of the OCC torque capacity is determined using governing equations to achieve a seamless output shaft torque transition from the torque phase of an upshift to an inertia phase.
A companion co-pending continuation-in-part patent application, which is assigned to the assignee of the present invention, discloses a control strategy for achieving a smooth upshift in a multiple ratio transmission without sensor feedback. The co-pending patent application is application Ser. No. 12/858,468, filed Aug. 18, 2010, now U.S. Pat. No. 8,328,688, issued Dec. 11, 2012. The present application has some features that are common to that co-pending application.
As required, detailed embodiments of the present invention are disclosed herein; however, it is to be understood that the disclosed embodiments are merely exemplary of the invention that may be embodied in various and alternative forms. The figures are not necessarily to scale; some features may be exaggerated or minimized to show details of particular components. Therefore, specific structural and functional details disclosed herein are not to be interpreted as limiting, but merely as a representative basis for teaching one skilled in the art to variously employ the present invention.
Numeral 10 represents a power input shaft drivably connected to torque source 12. Input shaft 10 drives a clutch housing 14, which carries torque input driving discs 16 situated in interdigital relationship with respect to driven discs 18 and 20. A fluid pressure actuator or electro-mechanical actuator of any known design is used to selectively engage driven discs 18 and 20 with respect to driving discs 16. Discs 20 are connected to a central torque input shaft 22 and discs 18 are connected to torque input sleeve shaft 24. Although only one disc 18 and only one disc 20 are shown in the schematic view of
Drive gear elements 26 and 28 are connected drivably to the sleeve shaft 24. Gear element 26 has a smaller pitch diameter than gear element 28.
Central power input shaft 22 is drivably connected to drive gear element 30, gear element 32 and gear element 34, which have decreasing pitch diameters.
When driving clutch discs 20 are engaged, driving torque is distributed through engaged clutch discs 20 to the gear elements 30, 32 and 34. Clutch discs 20 and 18 are part of a clutch structure that may be referred to as a tandem or dual clutch 36.
When clutch discs 18 are engaged by the tandem clutch 36, torque from the torque source is distributed directly to torque input gears 26 and 28.
The layshaft transmission of
Countershaft 40 rotatably supports countershaft gear elements 48, 50 and 52, which have progressively decreasing pitch diameters. Countershaft gear element 48 is a first ratio gear element, countershaft gear element 50 is a fifth ratio gear element and countershaft gear element 52 is a sixth ratio gear element.
Countershaft gear elements 54 and 56 also are rotatably supported by countershaft 40. Gear element 54 drivably engages gear element 26 during second ratio operation. Countershaft gear element 56 drivably engages a reverse drive pinion (not shown), which in turn drivably engages reverse gear element 44 during reverse drive operation. Gear element 46 connected to countershaft 38 is drivably connected to gear element 58, which is drivably connected to countershaft 40, for example, through torque transfer gearing (not shown in
Gear 58 is connected drivably to torque output gear 60, which is drivably connected to vehicle traction wheels.
During first gear ratio operation, gear 48 is connected drivably through synchronizer clutch 62 to countershaft 40, and clutch 36 engages discs 20 as discs 18 are disengaged. At that time, second ratio synchronizer clutch 64 drivably engages gear element 54 to precondition gear element 54 for second ratio operation. Power then is delivered from the torque source through clutch discs 20 to central shaft 22 so that torque is delivered from gear 34, to countershaft 40 and engaged gears 58 and 60.
An upshift is made from the first gear ratio to the second gear ratio by disengaging clutch discs 20 and engaging clutch discs 18 for the tandem clutch. To make a smooth transition from the first gear ratio to the second gear ratio, discs 18 are engaged as discs 20 are slowly disengaged to allow for clutch slip. At this time, third ratio synchronizer clutch 66 is engaged thereby connecting countershaft gear element 40 to countershaft 38. This preselects third ratio while the transmission operates in the second ratio. An upshift to the third ratio is achieved by tandem clutch 36 as clutch discs 20 are engaged and clutch discs 18 are disengaged. At this time, the fourth ratio synchronizer clutch 68 is engaged to preselect the fourth ratio. An upshift from the third gear ratio to the fourth gear ratio then is achieved by disengaging clutch discs 20 and engaging clutch discs 18. At this time, fifth gear ratio is preselected by engaging synchronizer clutch 70. An upshift to the fifth ratio then is achieved by engaging friction discs 20 and disengaging friction discs 18. At this time, the sixth ratio is preselected by engaging synchronizer clutch 72.
An upshift to the sixth ratio is achieved by again trading engagement of the discs for the tandem clutch 36. Clutch discs 20 are disengaged as clutch discs 18 are engaged.
Reverse drive is obtained by disengaging the forward drive synchronizer clutch and engaging reverse drive synchronizer clutch 74. Reverse driving torque then is delivered through sleeve shaft 24, gear 26, gear element 54 and gear element 56, reverse drive pinion gearing, countershaft 38 and torque transfer gear elements 46 and 58.
If the torque source is an internal combustion engine, the upshift controls would include a microprocessor 75, which may be of conventional design, an electronic engine control 77, including an engine fuel and spark retard controller, and a transmission control module 83.
The microprocessor 75, when the torque source is an engine, receives input signals such as driver desired input torque (Te_des) input speed (Ne), driver-selected ratio range (PRNDL), transmission input speed (Ninput), engine throttle position (Tp) if the torque source is a throttle-controlled engine, and transmission output speed (Noutput). The input signals are received by random access memory (RAM) from data input ports. A central processor unit (CPU) receives the input signals that are stored in RAM and uses the information fetched from RAM to execute algorithms that define control strategies stored in ROM. Output signals are delivered from signal output ports to the controllers 77 and 83. Actuating pressure for the clutches is supplied by pump 85 driven by engine 12 or by an electro-magnetic force actuator.
a show a schematic representation of a planetary type transmission that may embody the present invention. A torque source may be an engine 76 that drives a ring gear 80 of a simple planetary gear unit 82, which has a sun gear 84 and a planetary carrier 86. A hydrokinetic torque converter may be included in the transmission if a design objective requires it. It is shown at 78 in
During low gear ratio operation, friction brake 100 is disengaged. Brake 100 may be referred to as clutch #1. This corresponds to tandem clutch 36 of
When the gearing of
For purposes of this description, it will be assumed that if the powertrain has no hydrokinetic torque converter, torque input to the transmission will be referred to as engine torque (Te). If the powertrain has a torque converter, the engine torque would be replaced by converter turbine torque.
In the example of a planetary transmission shown in
During intermediate ratio operation, the sun gear for gear unit 25 is anchored to the housing 35 by intermediate coupling 39.
During direct drive, the transmission input shaft 13′ is clutched by direct coupling 41 to the sun gear for gear unit 25, thus establishing a one-to-one driving ratio through the planetary gearing. Overdrive coupling 43, when engaged, directly connects the carrier for gear unit 25 and the ring gear for gear unit 23 to the input shaft 13′.
As previously mentioned, torque sensors in the disclosed embodiments of the invention are used to obtain direct-reading oncoming and off-going clutch torques. In the case of the layshaft transmission of
In the case of the planetary transmission of
The shift event is divided into a preparatory phase, a torque phase, and an inertia phase. During the preparatory phase, torque capacity of clutch 20, which is the off-going clutch, is reduced, as shown at 86, to prepare for its release. However, enough clutch torque capacity is maintained at 88 to only allow a small incipient slip near the end of the preparatory phase, as shown by the small separation between the dotted input torque line 106 and OGC line 86. Transmission controller 82 adjusts an actuator for clutch 18 (clutch #2), which is referred to as the oncoming clutch, to prepare for its engagement. At that point, the oncoming clutch 18, in a synchronous upshift event, is yet to carry significant torque.
During the torque phase of the control shown in
During the torque phase of the shift characteristic shown in
The inertia phase begins when the off-going clutch capacity is reduced to a non-significant level, as shown at 98. Oncoming clutch (clutch #2) carries enough torque capacity, as shown at 100, to pull down engine speed, as shown at 102, closer to that of the speed of shaft #2, as indicated at 104.
The shift event is completed, as shown in
In contrast to the upshift characteristics shown in
During the torque phase, the controller 83 increases oncoming clutch target torque, as shown at 112 in
Input torque is increased, as shown at 114, while allowing clutch discs 20 to slip at a controlled level. Slipping the off-going clutch discs 20 causes input speed to be slightly greater than the shaft speed shown at 116, as shown at 124. This is true for a transmission having a slipping off-going clutch, but it is not true for a transmission with a locked off-going clutch.
When the off-going clutch 20 slips, its frictional torque is transmitted to shaft 22. Thus, the transmission controller can actively manage torque level that drives the gears coupled to the gearing connected to shaft 22 by adjusting the off-going clutch torque capacity 118. Similarly, when the oncoming clutch slips during the torque phase, its torque capacity, shown at 112, is transmitted to shaft 24, which drives the gearing (gearset #2) connected to shaft 24. Thus, when both the off-going clutch (OGC) and the oncoming clutch (OCC) slip during the torque phase, output shaft torque τos can be mathematically described as:
τos=Gonτon+Goffτoff, Eq. (1)
where τon is OCC torque capacity, τoff is OGC torque capacity, Goff is gear ratio for low gear operation and Gon is gear ratio for high gear operation. Equation (1) can be rearranged as:
Rewriting τos as τos,des, Eq. (2) can be expressed as:
where τos,des is a desired output shaft torque. The governing equation (3) of the present invention provides a systematic means to self-calibrate a level of OCC torque capacity Ton for achieving a desired output torque profile τos,des while OGC slips during the torque phase. More specifically, torque profile τos,des can be specified to smoothly transition output shaft torque 120 before and after the torque phase, from point 71 to point 73 and after point 73, thereby eliminating or reducing the torque hole. OGC torque capacity τoff can be estimated and actively adjusted based on OGC actuator position or clamping force. Thus, for a given τoff, Eq. (2) specifies a level of OCC torque capacity τon (112) required for achieving a desired output shaft torque 120.
During the torque phase, powertrain controller 75 and engine controller 77 control engine torque 114 or input shaft torque in order to maintain OGC slip at a desired level. This can be achieved, for example, by adjusting engine torque 114 using a closed-loop throttle control, valve timing control or fuel control or engine spark timing control based on OGC slip measurements independently from OCC and OGC torque control in a separate control loop or background loop, for the controller.
The transmission controller 83 (
Output shaft torque is described as:
τos=Goffτin+(Gon−Goff)τon, Eq. (4)
where input shaft torque τin can be equated to input torque τe (when the transmission has no torque converter). Replacing τos with a desired torque profile τos,des, Eq. (4) can be rearranged as:
Torque variables τos and τe can be represented as:
τos,des=τos
where τes0 and Te0 are the output shaft torque and engine torque at the beginning of the torque phase, respectively. Δτos and Δτe represent the change in output shaft torque and engine torque, respectively, at the elapsed time Δt after the torque phase begins. Substituting Eq. (6) into Eq. (5) yields:
OCC torque τon can be written as:
τon=τon
where τon0 is i the OCC torque capacity at the beginning of the torque phase and Δτon is the change in OCC torque at Δt. Substituting Eq. (8) into Eq. (7) results in:
where Δτoff=τe−Δτon. (Note that Eq. (9) takes the same form as Eq. (3), which is the governing equation for slipping OGC.)
The governing equations (5), (7) and (9) provide a systematic means to self-calibrate a level of OCC torque capacity (τon) for achieving a desired output torque profile (τos,des) during torque phase if OGC remains locked. More specifically, a torque profile τos,des can be specified to smoothly transition the output shaft torque 120 from a time before the torque phase to a time after the torque phase, thereby eliminating or reducing a torque hole. For a given τin or τe, Eq. (5) specifies a level of OCC torque capacity τon required for achieving the target output torque profile τos,des.
Alternatively, for a given oncoming clutch torque, Eq. (5) may be used to systematically determine a target engine torque τe or τin required for achieving desired output shaft torque τos,des. Once the target level is determined, τe or τin can be controlled through engine throttle control, spark timing control, intake and exhaust valve timing control, or through an auxiliary torque source such as an electric motor. (Note that engine torque control is coupled to OCC torque control in Eq. (5)).
The inertia phase begins at 73 in
Thus, the output shaft torque τos (120) in the inertia phase is primarily affected by OCC torque capacity τon (122). According to the present invention, Equation (10) is used to provide a target OCC torque capacity τon, during the inertia phase, that is required to achieve a seamless output shaft torque profile τos,des (120) from the torque phase to the inertia phase. τon is a feed-forward term. In addition, there is a feedback as well as an effect of a change in engine torque.
Engine torque can be actively and independently managed at 140 through a closed loop control to achieve a desired OGC slip speed. OGC torque capacity is adjusted through either closed loop control or open-loop control of its actuator position or actuator force. During a torque phase, a controller first chooses a desired level of output shaft torque (138). It also chooses desired OGC torque at 143.
Having chosen the desired OGC torque at 143, the engine is controlled at 140, as previously described, to achieve the desired slip. Simultaneously, the OGC actuator is adjusted at 144 to achieve the desired OGC torque.
A feedback torque correction, (τon,fb(k)), is calculated at 145 based on a measurement of oncoming clutch torque. Alternatively, (τon,fb)(k)) can be determined from calculated OCC torque based on torque measurements at other locations such as an output shaft. This correction is needed because of the inherent variability in the development of clutch torque. As previously indicated, the variability is due to unforeseen or uncharacteristic variation, or irregularities in the clutch actuator transfer function. Further, irregularities can be due, for example, to temperature changes, viscosity changes, wear of mechanical elements in the actuator structure, debris, rate of cooling of actuator fluid, etc. The increasing oncoming clutch torque shown at 93 in the plot for a synchronous clutch-to-clutch upshift is based upon a theoretical model. In actual practice, the response of the clutch actuator to a pressure command is affected by environmental factors, as mentioned.
The plot, as shown at 112′ in
Correcting for the difference between the commanded torque in a previous processor control loop (k-1) and the current measurement in the current processor control loop (k) is carried out at step 148 in
The oncoming clutch feedback torque can be calculated also using other sensors, such as an input shaft speed and an output shaft speed. Thus, the oncoming clutch feedback torque can be expressed as a function of the input shaft torque sensor reading, the output shaft sensor reading, the input shaft speed sensor reading and the output shaft speed sensor reading. The equations for accomplishing this are set out in the co-pending patent applications previously described; i.e., application Ser. No. 12/861,387 and Patent Publication 2010/0262344, which are assigned to the assignee of the present application.
After the controller uses Equation (3) to self-calibrate the required level of OCC torque capacity at 146, it adjusts OCC actuator position at 148 or its torque capacity to realize the desired output shaft torque. The controller evaluates whether the end of the torque phase is reached at 150 based upon a calibrated threshold OGC torque. If it is not, it repeats the control loop, as shown at 153. It re-estimates the desired output shaft torque at 138 and chooses OGC torque at 143 for the next controller loop time step k+1.
The end of the torque phase is reached when OGC torque becomes sufficiently small or less than a pre-specified threshold, τthresoff, at 150. The controller then releases the OGC clutch at 152 and moves to the inertia phase control at 154. Equation (10) is used to determine a target OCC torque at 154 for a seamless output shaft torque transition from the torque phase to the inertia phase.
In
After the desired slip is determined at block 214, a target input torque is determined at block 215. This input torque (τi,tgt) is a function of desired output shaft torque. The target input torque is that torque that exists for each control loop of the controller until the shift sequence reaches the end of the torque phase. If the sum of the target input torque and the desired slip torque is less than a precalibrated maximum value, as shown at block 216, the routine will continue to block 218 where a change in input torque (Δτi) at any instant during the torque phase is equal to the target input torque (Ti,tgt) minus the change in input torque (Δτi) at the beginning of the torque phase. If the sum of the target input torque and the slipping clutch torque at 216 is greater than τi maximum, the routine is recalculated at 217 until the inquiry at 216 is true.
A desired off-going clutch torque τoff(k) is chosen at 219, and delta off-going torque Δtoff also is calculated at 219. The off-going clutch actuator is adjusted accordingly at 227. Oncoming clutch target torque, τon,τgt, is calculated at 220 using the equation τon,τgt=Δτoff+Δτi, which is ramped to τon,ff(k). A feedback correction τon,fb(k) based on measured OCC torque (torque sensor output) is determined at 228 and the OCC actuator is adjusted at 230 to achieve τon(k), which is equal to τon,ff(k)+τon,fb(k). The input torque τi(k) then is ramped at 223 toward target input torque τi,tgt. The subscript ff designates a feed-forward term, the subscript fb is a feed-back term and k is a control loop indicator. The engine controller is adjusted to achieve engine torque τe(k).
If τoff is less than a calibrated threshold at block 224, the routine will return to the beginning and then repeat in the next control loop k+1. Otherwise, the OGC will be released at 225, where desired OCC torque is determined by the equation τon=τos,des/Gon. “G” is gear ratio of gearing in the OCC torque flow path.
The oncoming clutch target torque (τon,tgt) is computed by determining the sum of the delta off-going clutch torque at 219 (change of torque) and the delta input torque calculated at 218 at the beginning of the torque phase. The OGC actuator is adjusted at 227 to achieve the OGC torque chosen at step 219. The input torque then is ramped upwardly to the feed-forward target. This is the value for oncoming clutch torque at the end of the torque phase.
The step of ramping the input torque is shown at 223 in
The routine 311 of
If the target input torque is greater than the maximum calibrated input torque, as shown at 316, the target input torque and the oncoming clutch torque target torque are recalibrated at 317 before the routine will continue.
If the inquiry at block 316 is true, the routine will advance to block 318 where a desired off-going clutch torque is chosen. This is the value at the end of the torque phase. Having established the desired off-going clutch torque, the oncoming clutch torque is ramped toward the target oncoming clutch torque at 319. The clutch actuator for the oncoming clutch torque is adjusted at 321 to achieve the target oncoming clutch torque. The routine then will continue to block 320 in
A test then is made at 323, as in the case of the routine of
The control routine steps carried out at action blocks 328, 320, and 322 in
The clutch actuators may be fluid pressure actuators with a servo piston wherein piston movement during clutch engagement can be measured. During pre-calibration, a transfer function between actuator position and clutch slip torque is obtained by bench testing. The transfer function is stored in the memory of microprocessor 75 for vehicle control, including torque hole filling control. The transfer function is shown in
It is difficult to determine a position “X0” where a clutch actually starts assuming non-zero torque T0. Point “X0” is affected by unit-to-unit hardware variability, assembly process and clutch plate wear. An error in “X0” results in an inaccurate clutch torque estimate that affects torque hole fill control.
When torque measurements are available at the input shaft between the engine and the clutches, “X0” can be accurately determined because when the oncoming clutch starts assuming non-zero torque at T0, the measured input shaft torque momentarily increases because the clutch exerts additional load on the shaft. When “X0” is accurately known for both clutches, slipping oncoming clutch torque and slipping off-going clutch torque can be readily calculated using their transfer functions. Then their torque values are adjusted to be consistent with overall input shaft torque measurements.
For example, the torque sensor 10′ in
It is to be understood that this invention is not limited to the exact shift control steps illustrated and described. Various modifications and equivalents thereof, including revisions to the governing equations (3), (5), (7) and (9), may be made by persons skilled in the art without departing from the spirit and the scope of the invention to make this invention applicable to all types of automatic transmissions, including both a lay-shaft type and a planetary type.
This application is a continuation of U.S. application Ser. No. 14/295,411 filed Jun. 4, 2014, now U.S. Pat. No. 9,260,102, which is a continuation of U.S. application Ser. No. 13/155,867, filed Jun. 8, 2011, now U.S. Pat. No. 8,775,044, the disclosures of which are hereby incorporated in their entirety by reference herein.
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| Child | 14987466 | US | |
| Parent | 13155867 | Jun 2011 | US |
| Child | 14295411 | US |
