Patent Publication Number: US-8529405-B2

Title: Ratio shift control system and method for a multiple-ratio automatic transmission

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of application Ser. No. 12/858,468, filed Aug. 18, 2010, and issued as U.S. Pat. No. 8,328,688 on Dec. 10, 2012, which is a continuation-in-part of application Ser. No. 12/693,086, now abandoned, filed Jan. 25, 2010, which is assigned to the assignee of the present application and the disclosures of which are incorporated in their entirety by reference herein. 
    
    
     BACKGROUND 
     The invention relates to a multiple-ratio transmission mechanism in a powertrain for an automotive vehicle and to a control strategy for achieving smooth engagement and release of friction torque establishing elements during a transmission upshift event. 
     In a geared automatic transmission in an automotive vehicle powertrain having an engine or other torque source, a ratio change may be made from a so-called low ratio to a so-called higher ratio when a friction torque establishing element, such as a clutch or brake, is engaged in synchronism with disengagement of a companion friction torque establishing element. This is referred to as a ratio upshift. The friction torque establishing elements involved in the upshift may be referred to as an oncoming clutch or brake and an off-going clutch or brake. The upshift event is characterized by a preparatory phase, a torque phase and an inertia phase as the vehicle accelerates from a standing start. 
     In a conventional automatic transmission in a vehicle powertrain, the oncoming clutch torque capacity is controlled to increase from a low value during the torque phase. Simultaneous engagement of one clutch or brake and release of another results in a momentary activation of two torque flow paths through the gearing, causing a gear tie-up in which transmission output shaft torque decreases momentarily. This condition may be referred to as a “torque hole”. It occurs before the off-going clutch totally disengages. 
     Friction elements, such as disc clutches, band brakes and disc brakes, typically are actuated hydraulically under the control of a transmission control module, which disengages an off-going friction clutch or brake while simultaneously engaging an oncoming friction clutch or brake during an upshift in order to lower speed ratio. For purposes of the present description of the invention, the clutch and the brake will be referred to as friction elements. 
     During the preparatory phase, an automatic transmission control reduces off-going friction element torque capacity to prepare it for release as an actuator for the oncoming friction element is adjusted to prepare for its engagement. During the torque phase, the controller increases oncoming friction element torque capacity, which causes torque transmitted through the off-going friction element to drop quickly due to the transient gear tie up. 
     As torque is transmitted through the off-going friction element deceases, the automatic transmission output shaft torque drops, which causes the so-called torque hole. This is perceived by a vehicle occupant as an unpleasant shift shock. The inertia phase begins when the off-going clutch is released with no significant torque capacity. 
     SUMMARY 
     The invention comprises a transmission ratio control system and method that eliminates or reduces a so-called torque hole during upshifting of transmission gearing of a step ratio automatic transmission. The automatic transmission, for example, can be either a layshaft transmission with two torque input friction elements between a torque source and the transmission gearing, or a step ratio automatic transmission with planetary gearing, wherein a ratio change in the gearing during an upshifting event is effected by engaging one torque input friction element for the gearing and simultaneously disengaging another torque input friction element for the gearing. For purposes of describing the present invention, reference will be made to a lay-shaft type transmission. 
     The invention includes a strategy for execution of control algorithms that will achieve a desired output shaft torque profile that will avoid significant output shaft torque disturbances. 
     In the case of a powertrain with an internal combustion engine, torque input to the automatic transmission is increased during the torque phase of the shifting event. This is achieved by engine throttle control, spark timing adjustment for the engine (torque source), intake and exhaust valve timing control for the engine or by other means, such as by using auxiliary electric motor torque, based on an open loop control, a closed loop control, or a combination of both using engine speed, off-going and oncoming clutch slip speed measurements, and clutch actuator position measurements. 
     According to one aspect of the invention, a software-based controller is provided to self-calibrate a level of oncoming clutch torque capacity using algorithms in the form of algebraic equations whereby a desired output shaft torque profile is achieved while the off-going clutch slips during the torque phase in a controlled manner. 
     According to another aspect of the invention, the desired output shaft torque profile is achieved for a chosen off-going clutch torque capacity. 
     The invention, in executing the foregoing control features, may decouple control of engine torque or the input shaft torque from an oncoming clutch torque control during the torque phase, while the off-going clutch slips, and to achieve a desired off-going clutch slip based on a closed loop control of input shaft torque or engine torque. The end of the torque phase is determined based on torque level transmitted through the off-going clutch. 
     According to a further aspect of the invention, governing algebraic equations are used to determine a level of the oncoming clutch torque capacity to achieve a seamless transition from the torque phase to the inertia phase. This involves a self-calibration of a level of oncoming clutch torque capacity during the inertia phase to achieve a desired output shaft torque level. 
     In one embodiment of the invention, the off-going friction element is allowed to slip during the torque phase of a shift event as slip of the oncoming friction element is controlled. 
     According to another aspect of the invention, input torque may be increased during the torque phase, and the change in torque may be used in a determination of torque capacity of the off-going friction element during the torque phase. Torque of the torque source is reduced during the inertia phase and then restored, at least partially, after the inertia phase. 
     According to another aspect of the invention, control of the oncoming clutch torque control is decoupled from engine or input shaft control during the torque phase, and a desired off-going clutch slip is achieved based on a closed-loop control of input shaft torque (e.g., engine torque). 
     According to another aspect of the invention, the end of the torque phase is determined based on the torque level transmitted through the off-going clutch. 
     According to another aspect of the invention, a target level of oncoming clutch torque capacity is determined using governing equations to achieve a seamless output shaft torque transition from the torque phase to the inertia phase. 
     According to another aspect of the invention, a target level of oncoming clutch torque capacity is determined during the inertia phase using governing equations to achieve a desired output shaft torque level. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic illustration of a layshaft transmission in a first gear or low gear operating mode, which includes tandem torque input clutches that are selectively and alternately engaged and released. 
         FIG. 1   a  is a schematic illustration of the gearing arrangement of  FIG. 1  wherein the elements of the gearing are conditioned for high or second gear operating mode. 
         FIG. 2  is a schematic representation of a planetary type transmission that is capable of embodying the invention wherein the elements of the planetary gearing are conditioned for a low or first gear operating mode. 
         FIG. 2   a  is a schematic representation corresponding to  FIG. 2  wherein the elements are conditioned for a second or high gear operating mode. 
         FIG. 2   b  is a schematic representation of another planetary transmission that is capable of embodying the invention. 
         FIG. 3  is a time plot for a synchronous clutch-to-clutch upshift control characterized by a so-called torque hole at the output shaft. 
         FIG. 4  is a time plot corresponding to  FIG. 3  for the synchronous upshift control of the present invention wherein the tome hole is eliminated. 
         FIG. 5  is a flowchart showing the control strategy of the synchronous upshift control of the present invention when the off-going clutch is slipping. 
         FIG. 6  is a flowchart showing an alternate control strategy for a non-synchronous upshift when the off-going clutch is slipping. 
         FIG. 7  is a flowchart showing another alternate control strategy for a non-synchronous upshift when the off-going clutch is slipping. 
     
    
    
     PARTICULAR DESCRIPTION OF AN EMBODIMENT OF THE INVENTION 
     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. 
       FIG. 1  shows a schematic form of a lay-shaft transmission capable of embodying the invention together with a schematic representation of the transmission components involved in gear ratio changes. 
     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 inter-digital 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  FIGS. 1 and 1   a , several discs in a friction disc assembly may be used. 
     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 the 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  FIG. 1  has two countershafts, shown at  38  and  40 . Countershaft  38  supports rotatably a third ratio countershaft gear element  40 , a fourth ratio countershaft gear element  42  and a reverse countershaft gear element  44 . A torque transfer gear element  46  is directly connected to the countershaft  38 . 
     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  FIG. 1 ). The countershafts and the central shaft  22  actually are not in the same plane, so torque transfer gearing and the reverse drive pinions are not illustrated in the schematic illustration of  FIG. 1 . 
     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 disc  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  receives, when the torque source is an engine, 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. 
       FIG. 1   a  shows the gearing configuration during operation of the transmission in second gear ratio, which is the upshifted ratio. When the transmission operates in the second ratio, torque is delivered, as previously mentioned, to sleeve shaft  24  and through a second gear set, which comprises gear  26 , gear element  54  and transfer gears  58  and  60 . This gearing may be referred to as the second gear set. The gearing previously described with respect to  FIG. 1  for first gear operation hereafter may be referred to as the first gear set. 
       FIGS. 2 and 2   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  FIGS. 2 and 2   a  with phantom dotted lines since some designs capable of using the invention do no need a torque converter. If a torque converter is included, the converter turbine torque would be the input torque. The torque converter could be deleted if it is not needed. Carrier  86  supports planetary pinions that engage ring gear  80  and sun gear  84 . The output torque from the carrier drives sun gear  88  of a compound planetary gear set  90 . Compound planetary pinions  92  and  94  supported on a common carrier  96  engage respectively ring gear  90  and sun gear  88 . The ring gear is connected to the output shaft  98 . 
     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  FIGS. 1 and 1   a  when clutch discs  18  are released or disengaged. Brake  102  in  FIG. 2 , which is engaged in low speed ratio operation, corresponds to tandem clutch  36  shown in  FIGS. 1 and 1   a  when clutch discs  20  are engaged. Clutch # 2  in  FIG. 2  (brake  102 ) provides a reaction point for the carrier  96 . Sun gear, shown at  104 , which drivably engages with compound planetary pinion  92 , merely idles during low speed ratio operation. 
     When the gearing of  FIGS. 2 and 2   a  is operating in the second ratio, sun gear  104  is anchored by brake  100  so that the ring gear for compound planetary gear unit  92  is driven at an increased rate relative to the carrier speed of the simple planetary gear set  82 . 
     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. 
       FIG. 2   b  shows an example of another planetary step-ratio automatic transmission that may embody the invention. It comprises an engine driven torque input shaft  11  and a transmission input shaft  13 . A transmission output shaft  15  delivers torque to transmission torque output gearing  17 . A torque converter may be disposed between engine driven torque input shaft  11  and a transmission input shaft  13 , as shown at  19 . A torque converter impeller  11  is in fluid flow relationship with respect to turbine  13 . A stator  15  is disposed between the flow inlet section of impeller  11  and the flow exit section of turbine  13 . 
     In the example of a planetary transmission shown in  FIG. 2   b , there are three simply planetary gear units  21 ,  23  and  25 . Output torque is delivered from the carrier  27  to the torque output gearing. Carrier  27  is connected to the ring gear for gear unit  25  and to output shaft  15 . An overrunning coupling  29  anchors the carrier  31  of planetary gear unit  25  against rotation in one direction, but free wheeling motion is provided in the opposite direction. During reverse and during low ratio operation, carrier  31  is braked by coupling  33  against the transmission housing  35 . During forward drive operation, the sun gear for gear unit  21  is anchored to the housing through forward drive coupling  37 . 
     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 input shaft  13 , 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 .  FIG. 1   a  shows an engine  12  which acts as a source of torque for the transmission. If the transmission has a torque converter, engine speed will equal speed of converter impeller  22  and transmission input speed would equal converter turbine speed. 
       FIG. 3  shows a strategy for a typical known upshift event from a low gear configuration (i.e., high torque ratio) to a high gear configuration (i.e., low torque ratio) when the engine has a constant throttle setting, in accordance with a conventional upshift control method for a lay-shaft transmission of the type shown in  FIGS. 1 and 1   a . This strategy of the invention would apply also to a transmission such as the compound planetary transmission of  FIGS. 2 and 2   a  and the planetary transmission of  FIG. 2   b.    
     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 allows 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 piston 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 capacity. 
     During the torque phase of the control shown in  FIG. 3 , off-going clutch capacity is further reduced, as shown at  91 , while the controller  82  increases oncoming clutch torque capacity, as shown at  93 . Engine speed and input shaft speed are the same if the transmission has no torque converter between the engine and the clutch  36 . However, as will be explained subsequently in a discussion of  FIG. 4 , off-going clutch torque capacity may be controlled to induce a small target level slip at  91 , which allows engine speed  95  to be higher than the speed of shaft  22 . When the off-going clutch slips, off-going clutch torque  91 , or frictional torque generated by slipping, drives shaft  22 , seen in  FIGS. 1 and 1   a  and the downstream gear elements (gearset # 1 ), all the way to the output shaft. Increasing oncoming clutch torque  93  starts balancing torque distributed from the engine and reduces the off-going clutch torque capacity requirement at  91 . Thus, the off-going clutch and the oncoming clutch work together to maintain off-going clutch target level slip as the off-going clutch torque decreases as shown at  91 . 
     During the torque phase of the shift characteristic shown in  FIG. 3 , an increase in oncoming clutch torque capacity (clutch # 2  capacity) reduces net torque flow through the off-going clutch when the off-going clutch remains engaged. Thus, the output shaft torque drops significantly, as shown at  97 , creating a so-called torque hole. A large torque hole can be perceived by a vehicle occupant as a sluggish powertrain performance or an unpleasant shift shock. 
     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 . 
       FIG. 3  shows reduced input torque during the inertia phase, as shown at  106 . This is typically due to engine spark timing control, which is common practice in the conventional shift control method, to enable the oncoming clutch to engage within a target shift duration without excessive torque capacity. 
     The shift event is completed, as shown in  FIG. 3 , when clutch # 2  (the oncoming clutch) is engaged. The input shaft then is securely coupled to shaft  24 , seen in  FIG. 1 , thereby matching engine speed  102  to shaft speed  104 . The engine torque reduction at  106  is removed at  108  and the output shaft torque returns to the level that corresponds to an engine torque level during the high gear configuration. 
     In contrast to the upshift characteristics shown in  FIG. 3 ,  FIG. 4  shows the upshift characteristics of an embodiment of the upshift control method of the invention. During the preparatory phase, the controller  83  reduces the torque capacity of the off-going clutch (discs  20 ) to prepare for its release, as shown at  110 . The controller also adjusts the actuator piston for clutch  18  (the oncoming clutch) to prepare for its engagement. 
     During the torque phase, the controller  83  increases oncoming clutch torque capacity, as shown at  112 , to prepare for its engagement. 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, as shown at  124 , than the shaft speed, shown at  116 . 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 torque capacity or 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:
 
τ os   =G   on τ on   +G   off τ off ,  Eq. (1)
 
     where τ on  is OCC torque capacity, τ off  is OGC torque capacity, G off  is gear ratio for low gear operation and G on  is gear ratio for high gear operation. Equation (1) can be rearranged as: 
     
       
         
           
             
               
                 
                   
                     τ 
                     on 
                   
                   = 
                   
                     
                       
                         τ 
                         os 
                       
                       - 
                       
                         
                           G 
                           off 
                         
                         ⁢ 
                         
                           τ 
                           off 
                         
                       
                     
                     
                       G 
                       on 
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     2 
                     ) 
                   
                 
               
             
           
         
       
     
     Rewriting τ os  as τ os,des , Eq. (2) can be expressed as: 
     
       
         
           
             
               
                 
                   
                     
                       τ 
                       on 
                     
                     = 
                     
                       
                         
                           τ 
                           
                             os 
                             , 
                             des 
                           
                         
                         - 
                         
                           
                             G 
                             off 
                           
                           ⁢ 
                           
                             τ 
                             off 
                           
                         
                       
                       
                         G 
                         on 
                       
                     
                   
                   , 
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     3 
                     ) 
                   
                 
               
             
           
         
       
     
     where τ os,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 τ on  for achieving a desired output torque profile τ os,des  while OGC slips during the torque phase. More specifically, torque profile τ os,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 τ off  can be estimated and actively adjusted based on OGC actuator position or clamping force. Thus, for a given τ off , Eq. (2) specifies a level of OCC torque capacity τ on  ( 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  ( FIG. 1 ) could maintain enough OGC torque capacity during the torque phase without allowing OGC to slip. In this case, OGC still transmits a part of engine torque  114  to shaft # 1  ( 22 ). 
     Output shaft torque is described as:
 
τ os   =G   off τ in +( G   on   −G   off )τ on ,  Eq. (4)
 
     where input shaft torque τ in  can be equated to input torque τ e  (when the transmission has no torque converter). Replacing τ os  with a desired torque profile τ os,des , Eq. (4) can be rearranged as: 
     
       
         
           
             
               
                 
                   
                     τ 
                     on 
                   
                   = 
                   
                     
                       
                         
                           
                             τ 
                             
                               os 
                               , 
                               des 
                             
                           
                           - 
                           
                             
                               G 
                               off 
                             
                             ⁢ 
                             
                               τ 
                               e 
                             
                           
                         
                         
                           
                             G 
                             on 
                           
                           - 
                           
                             G 
                             off 
                           
                         
                       
                       ⁢ 
                       
                           
                       
                       ⁢ 
                       or 
                       ⁢ 
                       
                           
                       
                       ⁢ 
                       
                         τ 
                         e 
                       
                     
                     = 
                     
                       
                         
                           
                             τ 
                             
                               os 
                               , 
                               des 
                             
                           
                           - 
                           
                             
                               ( 
                               
                                 
                                   G 
                                   on 
                                 
                                 - 
                                 
                                   G 
                                   off 
                                 
                               
                               ) 
                             
                             ⁢ 
                             
                               τ 
                               on 
                             
                           
                         
                         
                           G 
                           off 
                         
                       
                       . 
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     5 
                     ) 
                   
                 
               
             
           
         
       
     
     Torque variables τ os  and τ e  can be represented as:
 
τ os,des =τ os     0   −Δτ os  and τ e =τ e     0   +Δτ e ,  Eq. (6)
 
     where τ os0  and τ e0  are the output shaft torque and engine torque at the beginning of the torque phase, respectively. Δτ os  and Δτ e  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: 
     
       
         
           
             
               
                 
                   
                     τ 
                     on 
                   
                   = 
                   
                     
                       
                         
                           Δτ 
                           
                             os 
                             , 
                             des 
                           
                         
                         + 
                         
                           
                             G 
                             off 
                           
                           ⁢ 
                           
                             Δτ 
                             e 
                           
                         
                       
                       
                         
                           G 
                           off 
                         
                         - 
                         
                           G 
                           on 
                         
                       
                     
                     . 
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     7 
                     ) 
                   
                 
               
             
           
         
       
     
     OCC torque τ on  can be written as:
 
τ on =τ on     0   +Δτ on ,  Eq. (8)
 
     where τ on0  is the OCC torque capacity at the beginning of the torque phase and Δτ on  is the change in OCC torque at Δt. Substituting Eq. (8) into Eq. (7) results in: 
     
       
         
           
             
               
                 
                   
                     
                       Δ 
                       ⁢ 
                       
                           
                       
                       ⁢ 
                       
                         τ 
                         on 
                       
                     
                     = 
                     
                       
                         
                           Δτ 
                           
                             os 
                             , 
                             des 
                           
                         
                         - 
                         
                           
                             G 
                             off 
                           
                           ⁢ 
                           
                             Δτ 
                             off 
                           
                         
                       
                       
                         G 
                         on 
                       
                     
                   
                   , 
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     9 
                     ) 
                   
                 
               
             
           
         
       
     
     where Δτ off ≡τ e −Δτ on . (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 when 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 at  73  to a time after the torque phase, thereby eliminating or reducing a torque hole. For a given τin or τe at  114 , Eq. (5) specifies a level of OCC torque capacity τon ( 112 ) required for achieving the target output torque profile τos,des ( 120 ). 
     Alternatively, for a given on ( 112 ), Eq. (5) may be used to systematically determine a target τe ( 114 ) or in required for achieving desired output shaft torque τos,des ( 120 ). 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  FIG. 4  when OGC is released. OGC transmits torque only at a non-significant level while OCC carries enough torque capacity, as shown at  122 , to slow down input speed  124  so that it is closer to shaft # 2  speed, as shown at  126 . Under this condition, both Eq. (3) and Eq. (5) can be reduced to: 
     
       
         
           
             
               
                 
                   
                     τ 
                     on 
                   
                   = 
                   
                     
                       
                         τ 
                         
                           os 
                           , 
                           des 
                         
                       
                       
                         G 
                         on 
                       
                     
                     . 
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                   ⁢ 
                   
                     ( 
                     10 
                     ) 
                   
                 
               
             
           
         
       
     
     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. Ton is a feed-forward term. In addition, there is a feed back as well as an effect of a change in engine torque. 
       FIG. 4  shows reduced input torque during the inertia phase. This is typically due to engine spark timing control according to a common practice in a conventional shift control method, enabling OCC to engage within a target shift duration without requiring excessive torque capacity. The shift event is completed when OCC is securely engaged, thereby coupling input shaft  10  and shaft # 2  ( 24 ). The engine torque reduction then is removed at  130  and the output shaft torque returns to a level  132 , which corresponds to an engine torque level in the high gear configuration. 
       FIG. 5  shows a control flow chart for the synchronous shift control of the present invention when the OGC is slipped during a torque phase. It describes a systematic approach to enable the shift control shown in  FIG. 4 . As previously stated, one of the advantages of this invention is the decoupling of OCC control, shown inside the dashed line  136 , from engine control  140  and OGC control  144 . 
     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 . Then, 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 OGC torque capacity level. If it is not, it repeats the control loop at  153 . It re-estimates the desired output shaft torque at  138  and chooses OGC torque capacity at  143  for the next controller 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  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. 
       FIG. 6  illustrates an alternate control strategy that will achieve the oncoming clutch torque characteristics, the off-going clutch torque characteristics and the engine torque characteristics that will avoid output shaft torque disturbances previously described. As previously indicated, in the strategy of  FIG. 5 , the output shaft torque that is chosen is used to calculate an oncoming clutch torque as shown at block  146  in  FIG. 5 . Regardless of whether the strategy of  FIG. 5  or the strategy of  FIG. 6  is used, the objective is to ensure that the engine torque will be higher throughout the duration of the torque phase than the oncoming clutch torque. The engine speed will remain above the off-going clutch speed during the torque exchange that occurs during the torque phase as seen in  FIG. 4 . This prevents a torque reversal. 
     In  FIG. 6 , prior to the start of the torque phase at block  212 , the off-going clutch torque will have decreased to a value that is slightly less than the input torque. This occurs during the preparatory stage as seen in  FIG. 4 . A desired output shaft torque then is chosen as shown at  213  rather than choosing a desired off-going clutch torque following the step at  213 . As in the case of the  FIG. 5  strategy routine, a desired slip is chosen at  214  as seen in  FIG. 6 . The value chosen is a value that will prevent torque source input speed flare during the torque phase. The slip torque depends upon the rate of change of engine speed (α) as well as engine inertia (I) if an engine is the torque source. 
     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 a 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) in 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. 
     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 end of the torque phase. The input torque then is ramped upwardly to the 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  FIG. 6 . If the result of the ramping at  223  is an off-going clutch torque that is less than the off-going clutch threshold value, which is precalibrated, the off-going clutch will be released at shown at  225 . As in the case of the routine of  FIG. 5 , the routine proceeds through the inertia phase where the desired oncoming clutch torque is determined by the equation shown at  226 . 
     The routine  311  of  FIG. 7  is somewhat similar to the routine  211  of  FIG. 6  except that, for example, a desired target oncoming clutch torque is chosen following the start of the torque phase at  312 . This is shown at block  313  in  FIG. 7 . In contrast, the desired output shaft torque is chosen in the case of  FIG. 6  starting at the beginning of the torque phase. After choosing a desired slip at  314 , the routine of  FIG. 7  will calculate an input torque at  315  so that the input torque will be sufficiently increased to compensate for the target oncoming clutch torque. This is evident by the rising slope of the input torque plot of  FIG. 4  during the torque phase. 
     If the target input torque is less 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  FIG. 7  where the input torque is ramped toward the target torque at the end of the torque phase, followed by a controller adjustment at  322  to achieve the target. 
     A test then is made at  323 , as in the case of the routine of  FIG. 6 , to determine whether the off-going clutch torque is less than a precalibrated off-going clutch torque threshold. The threshold torque is determined so that a residual torque will be maintained in the clutch actuator rather than having the off-going clutch torque fall to zero. The off-going clutch torque then is released and the routine continues to the inertia phase as shown at  324  and  325 . 
     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. 
     While exemplary embodiments are described above, it is not intended that these embodiments describe all possible forms of the invention. Rather, the words used in the specification are words of description rather than limitation, and it is understood that various changes may be made without departing from the spirit and scope of the invention. Additionally, the features of various implementing embodiments may be combined to form further embodiments of the invention.