Abstract:
A load sensor assembly for measuring an amount of torque transmitted through a torque establishing element includes a core mounted on a transmission housing and a load sensor mounted on the core. The load sensor is positioned against a portion of the torque establishing element whereby a portion of the amount of torque transmitted through the torque establishing element travels through the load sensor and is measured. A cable is connected to the load sensor for transmitting a signal representative of the amount of torque to a transmission controller.

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
       [0001]    This application is a divisional of U.S. patent application Ser. No. 12/421,379 filed Apr. 9, 2009. The entire contents of this application is incorporated herein by reference. 
     
    
     FIELD OF THE INVENTION 
       [0002]    The present invention pertains to the field of automatic transmissions for motor vehicles and, more particularly, to a friction element load sensor that directly measures torque transmitted by a friction element of an automatic transmission. 
       BACKGROUND OF THE INVENTION 
       [0003]    A step-ratio automatic transmission system in a vehicle utilizes multiple friction elements for automatic gear ratio shifting. Broadly speaking, these friction elements may be described as torque establishing elements although more commonly they are referred to as clutches or brakes. The friction elements function to establish power flow paths from an internal combustion engine to vehicle traction wheels. During acceleration of the vehicle, the overall speed ratio, which is the ratio of a transmission input shaft speed to a transmission output shaft speed, is reduced during a ratio upshift as vehicle speed increases for a given engine throttle setting. A downshift to achieve a higher speed ratio occurs as an engine throttle setting increases for any given vehicle speed, or when the vehicle speed decreases as the engine throttle setting is decreased. 
         [0004]    Various planetary gear configurations are found in modern automatic transmissions. However the basic principle of shift kinematics remains similar. Shifting a step-ratio automatic transmission having multiple planetary gearsets is accompanied by applying and/or releasing friction elements to change speed and torque relationships by altering the torque path through the planetary gearsets. Friction elements are usually actuated either hydraulically or mechanically. 
         [0005]    In the case of a synchronous friction element-to-friction element upshift, a first pressure actuated torque establishing element, referred to as an off-going friction element, is released while a second pressure actuated torque establishing element, referred to as an on-coming friction element, engages in order to lower a transmission gear ratio. A typical upshift event is divided into preparatory, torque and inertia phases. During the preparatory phase, an on-coming friction element piston is stroked to prepare for its engagement while an off-going friction element torque-holding capacity is reduced as a step toward its release. During the torque phase, which may be referred to as a torque transfer phase, on-coming friction element torque is raised while the off-going friction element is still engaged. The output shaft torque of the automatic transmission typically drops during the torque phase, creating a so-called torque hole. When the on-coming friction element develops enough torque, the off-going friction element is released, marking the end of the torque phase and the beginning of the inertia phase. During the inertia phase, the on-coming friction element torque is adjusted to reduce its slip speed toward zero. When the on-coming friction element slip speed reaches zero, the shift event is completed. 
         [0006]    In a synchronous shift, the timing of the off-going friction element release must be synchronized with the on-coming friction element torque level to deliver a consistent shift feel. A premature release leads to engine speed flare and a deeper torque hole, causing perceptible shift shock for a vehicle occupant. A delayed release causes a tie-up of gear elements, also resulting in a deep and wide torque hole for inconsistent shift feel. A conventional shift control relies on speed measurements of the powertrain components, such as an engine and a transmission input shaft, to control the off-going friction element release process during the torque phase. A conventional torque phase control method releases the off-going friction element from its locked state through an open-loop control based on a pre-calibrated timing, following a pre-determined off-going friction element actuator force profile. This conventional method does not ensure optimal off-going friction element release timing and therefore results in inconsistent shift feel. 
         [0007]    Alternatively, a controller may utilize speed signals to gauge off-going friction element release timing. That is, the off-going friction element is released if the controller detects a sign of gear tie-up, which may be manifested as a measurable drop in input shaft speed. When a release of the off-going friction element is initiated prematurely before the on-coming friction element develops enough torque, engine speed or automatic transmission input shaft speed may rises rapidly in an uncontrolled manner. If this so-called engine speed flair is detected, the controller may increase off-going friction element control force to quickly bring down automatic transmission input speed or off-going friction element slip speed. This speed-based or slip-based approach often results in a hunting behavior between gear tie-up and engine flair, leading to inconsistent shift feel. Furthermore, off-going friction element slip control is extremely difficult because of its high sensitivity to slip conditions and a discontinuity between static and dynamic frictional forces. A failure to achieve a seamless slip control during the torque phase leads to undesirable shift shock. 
         [0008]    In the case of a non-synchronous automatic transmission, the upshifting event involves engagement control of only an on-coming friction element, while a companion clutching component, typically a one-way coupling, automatically disengages to reduce the speed ratio. The non-synchronous upshift event can also be divided into three phases, which may also be referred to as a preparatory phase, a torque phase and an inertia phase. The preparatory phase for the non-synchronous upshift is a time period prior to the torque phase. The torque phase for the non-synchronous shift is a time period when the on-corning friction element torque is purposely raised for its engagement until the one-way coupling starts slipping or overrunning. This definition differs from that for the synchronous shift because the non-synchronous shift does not involve active control of a one-way coupling or the off-going friction element. The inertia phase for the non-synchronous upshift is a time period when the one-way coupling starts to slip, following the torque phase. According to a conventional upshift control, during the torque phase of the upshifting event for a non-synchronous automatic transmission, the torque transmitted through the oncoming friction element increases as it begins to engage. A kinematic structure of a non-synchronous upshift automatic transmission is designed in such a way that torque transmitted through the one-way coupling automatically decreases in response to increasing oncoming friction element torque. As a result of this interaction, the automatic transmission output shaft torque drops during the torque phase, which again creates a so-called “torque hole.” Before the one-way coupling disengages, as in the case previously described, a large torque hole can be perceived by a vehicle occupant as an unpleasant shift shock. An example of a prior art shift control arrangement can be found in U.S. Pat. No. 7,351,183 hereby incorporated by reference. 
         [0009]    A transmission schematically illustrated at  2  in  FIG. 1  is an example of a prior art multiple-ratio transmission with a controller  4  wherein ratio changes are controlled by friction elements acting on individual gear elements. Engine torque from vehicle engine  5  is distributed to torque input element  10  of hydrokinetic torque converter  12 . An impeller  14  of torque converter  12  develops turbine torque on a turbine  16  in a known fashion. Turbine torque is distributed to a turbine shaft, which is also transmission input shaft  18 . Transmission  2  of  FIG. 1  includes a simple planetary gearset  20  and a compound planetary gearset  21 . Gearset  20  has a permanently fixed sun gear S 1 , a ring gear R 1  and planetary pinions P 1  rotatably supported on a carrier  22 . Transmission input shaft  18  is drivably connected to ring gear R 1 . Compound planetary gearset  21 , sometimes referred to as a Ravagineaux gearset, has a small pitch diameter sun gear S 3 , a torque output ring gear R 3 , a large pitch diameter sun gear S 2  and compound planetary pinions. The compound planetary pinions include long pinions P 2 / 3 , which drivably engage short planetary pinions P 3  and torque output ring gear R 3 . Long planetary pinions P 2 / 3  also drivably engage short planetary pinions P 3 . Short planetary pinions P 3  further engage sun gear S 3 . Planetary pinions P 2 / 3 , P 3  of gearset  21  are rotatably supported on compound carrier  23 . Ring gear R 3  is drivably connected to a torque output shaft  24 , which is drivably connected to vehicle traction wheels through a differential and axle assembly (not shown). Gearset  20  is an underdrive ratio gearset arranged in series with respect to compound gearset  21 . Typically, transmission  2  preferably includes a lockup or torque converter bypass clutch, as shown at  25 , to directly connect transmission input shaft  18  to engine  5  after a torque converter torque multiplication mode is completed and a hydrokinetic coupling mode begins.  FIG. 2  is a chart showing a clutch and brake friction element engagement and release pattern for establishing each of six forward driving ratios and a single reverse ratio for transmission  2 . 
         [0010]    During operation in the first four forward driving ratios, carrier P 1  is drivably connected to sun gear S 3  through shaft  26  and forward friction element A. During operation in the third ratio, fifth ratio and reverse, direct friction element B drivably connects carrier  22  to shaft  27 , which is connected to large pitch diameter sun gear S 2 . During operation in the fourth, fifth and sixth forward driving ratios, overdrive friction element B connects turbine shaft  18  to compound carrier  23  through shaft  28 . Friction element C acts as a reaction brake for sun gear S 2  during operation in second and sixth forward driving ratios. During operation of the third forward driving ratio, direct friction element B is applied together with forward friction element A. The elements of gearset  21  then are locked together to effect a direct driving connection between shaft  28  and output shaft  26 . The torque output side of forward friction element A is connected through torque transfer element  29  to the torque input side of direct friction element B, during forward drive. The torque output side of direct friction element B, during forward drive, is connected to shaft  27  through torque transfer element  30 . Reverse drive is established by applying low-and-reverse brake D and friction element B. 
         [0011]    For the purpose of illustrating one example of a synchronous ratio upshift for the transmission of  FIG. 1 , it will be assumed that an upshift will occur between the first ratio and the second ratio. On such a 1-2 upshift, friction element C starts in the released position before the shift and is engaged during the shift while low/reverse friction element D starts in the engaged position before the shift and is released during the shift. Forward friction element A stays engaged while friction element B and overdrive friction element E stay disengaged throughout the shift. More details of this type of transmission arrangement are found in U.S. Pat. No. 7,216,025, which is hereby incorporated by reference. 
         [0012]      FIG. 3  depicts a general process of a synchronous friction element-to-friction element upshift event from a low gear configuration to a high gear configuration for the automatic transmission system of  FIG. 1 . For example, the process has been described in relation to a 1-2 synchronous ratio upshift above wherein friction element C is an oncoming friction element and low/reverse friction element D is an off-going friction element, but it is not intended to illustrate a specific control scheme. 
         [0013]    The shift event is divided into three phases: a preparatory phase  31 , a torque phase  32  and an inertia phase  33 . During preparatory phase  31 , an on-coming friction element piston is stroked (not shown) to prepare for its engagement. At the same time, off-going friction element control force is reduced as shown at  34  as a step toward its release. In this example, off-going friction element D still retains enough torque capacity shown at  35  to keep it from slipping, maintaining transmission  2  in the low gear configuration. However, increasing on-coming friction element control force shown at  36  reduces net torque flow within gearset  21 . Thus, the output shaft torque drops significantly during torque phase  32 , creating a so-called torque hole  37 . A large torque hole can be perceived by a vehicle occupant as an unpleasant shift shock. Toward the end of torque phase  32 , off-going friction element control force is dropped to zero as shown at  38  while on-coming friction element apply force continues to rise as shown at  39 . Torque phase  32  ends and inertia phase  33  begins when off-going friction element D starts slipping as shown at  40 . During inertia phase  33 , off-going friction element slip speed rises as shown at  41  while on-coming friction element slip speed decreases as shown at  42  toward zero at  43 . The engine speed and transmission input speed  44  drops as the planetary gear configuration changes. During inertia phase  33 , output shaft torque indicated by profile  45  is primarily affected by on-coming friction element C torque capacity indirectly indicated by force profile  46 . When on-coming friction element C completes engagement or when its slip speed becomes zero at  43 , inertia phase  33  ends, completing the shift event. 
         [0014]      FIG. 4  shows a general process of a synchronous friction element-to-friction element upshift event from the low gear configuration to the high gear configuration in which off-going friction element D is released prematurely as shown at  51  compared with the case shown in  FIG. 3 . When off-going friction element  1 ) is released, it breaks a path between automatic transmission input shaft  18  and automatic transmission output shaft  24 , depicted in  FIG. 1 , no longer transmitting torque to automatic transmission output shaft at the low gear ratio. Since on-coming friction element C is yet to carry enough engagement torque as indicated by a low apply force at  52 , automatic transmission output shaft torque drops largely, creating a deep torque hole  53  which can be felt as a shift shock. At the same time, engine speed or transmission input speed rapidly increases as shown at  54 , causing a condition commonly referred to as engine flare. A large level of engine flare can be audible to a vehicle occupant as unpleasant noise. Once on-coming friction element C develops sufficient engagement torque as indicated by a rising control force at  55 , automatic transmission input speed comes down and the output torque rapidly moves to a level at  56  that corresponds to on-coming friction element control force  55 . Under certain conditions, this may lead to a torque oscillation  57  that can be perceptible to a vehicle occupant as unpleasant shift shock.  FIG. 5  shows a general process of a friction element-to-friction element upshift event from the low gear configuration to the high gear configuration in which off-going friction element release is delayed as shown at  61  compared with the case shown in  FIG. 3 . Off-going friction element D remains engaged even after on-coming friction element C develops a large amount of torque as indicated by a large actual control force at  65 . Thus, transmission input torque continues to be primarily transmitted to output shaft  24  at the low gear ratio. However, large on-coming friction element control force  65  results in a drag torque, lowering automatic transmission output shaft torque, creating a deep and wide torque hole  63 . This condition is commonly referred to as a tie-up of gear elements. A severe tie-up can be felt as a shift shock or loss of power by a vehicle occupant. 
         [0015]    As illustrated in  FIGS. 3 ,  4 , and  5  a missed synchronization of off-going friction element release timing with respect to on-coming friction element torque capacity leads to engine flare or tie-up. Both conditions lead to varying torque levels and profiles at automatic transmission torque output shaft  24  during shifting. If these conditions are severe, they result in undesirable driving experience such as inconsistent shift feel or perceptible shift shock. The prior art methodology attempts to mitigate the level of missed-synchronization by use of an open loop off-going friction element release control based on speed signal measurements. It also attempts to achieve a consistent on-coming friction element engagement torque by an open-loop approach during a torque phase under dynamically-changing shift conditions. 
         [0016]      FIG. 6  illustrates a prior art methodology for controlling a friction element-to-friction element upshift from a low gear configuration to a high gear configuration for automatic transmission  2  in  FIG. 1 . The prior art on-coming control depicted in  FIG. 6  applies to a conventional torque phase control utilized for either a synchronous or non-synchronous shift. In this example off-going friction element D remains engaged until the end of torque phase  32 . Although the focus is placed on torque phase control,  FIG. 6  depicts the entire shift control process. As shown the shift event is divided into three phases: a preparatory phase  31 , a torque phase  32  and an inertia phase  33 . During preparatory phase  31 , an on-coming friction element piston is stroked (not shown) to prepare for its engagement. At the same time, off-going friction element control force is reduced as shown at  34  as a step toward its release. During torque phase  32  controller  4  commands an on-coming friction element actuator to follow a prescribed on-coming friction element control force profile  64  through an open-loop based approach. Actual on-coming friction element control force  65  may differ from prescribed profile  64  due to control system variability. Even if actual control force  65  closely follows prescribed profile  64 , on-coming friction element engagement torque may still vary largely from is shift to shift due to the sensitivity of the on-coming friction element engagement process to engagement conditions such as lubrication oil flow and friction surface temperature. Controller  4  commands enough off-going element control force  61  to keep off-going element D from slipping, maintaining the planetary gearset in the low gear configuration until the end of torque phase  32 . Increasing on-coming friction element control force  65  or engagement torque reduces net torque flow within the low-gear configuration. Thus, output shaft torque  66  drops significantly during torque phase  32 , creating so-called torque hole  63 . If the variability in on-coming friction element engagement torque significantly alters a shape and depth of torque hole  63 , a vehicle occupant may experience inconsistent shift feel. Controller  4  reduces off-going friction element actuator force at  38 , following a pre-calibrated profile, in order to release it at a pre-determined timing  67 . The release timing may be based on a commanded value of on-coming friction element control force  62 . Alternatively, off-going friction element D is released if controller  4  detects a sign of significant gear tie-up, which may be manifested as a detectable drop in input shaft speed  44 . Inertia phase  33  begins when off-going friction element D is released and starts slipping as shown at  67 . During inertia phase  33 , off-going friction element slip speed rises as shown at  68  while on-coming friction element slip speed decreases toward zero as shown at  69 . Transmission input speed  44  drops as the planetary gear configuration changes. During inertia phase  33 , output shaft torque  66  is primarily affected by on-coming friction element torque capacity or control force  65 . The shift event completes when the on-coming friction element comes into a locked or engaged position with no slip as shown at  70 . 
         [0017]      FIG. 7  illustrates another prior art methodology for controlling torque phase  32  of a synchronous upshift process from the low gear configuration to the high gear configuration. In this example, controller  4  allows off-going friction element D to slip during torque phase  32 . Although the focus is placed on torque phase control,  FIG. 7  depicts the entire shift event. During preparatory phase  31 , an on-coming friction element piston is stroked to prepare for its engagement. At the same time, off-going friction element control force  86  is reduced as a step toward its slip. During torque phase  32 , on-coming friction element control force is raised in a controlled manner. More specifically, controller  4  commands on-coming friction element actuator to follow a prescribed on-coming friction element control force profile  87  through an open-loop based approach. An actual on-coming friction element control force  88  may differ from the commanded profile  87  due to control system variability. Even if actual control force  88  closely follows commanded profile  87 , on-coming friction element engagement torque may still vary largely from shift to shift due to the sensitivity of on-coming friction element engagement process to engagement conditions such as lubrication oil flow and friction surface temperature. Increasing on-coming friction element control force  88  or on-coming friction element engagement torque reduces net torque flow within the low-gear configuration. This contributes to output shaft torque  99  being reduced during torque phase  32 , creating a so-called torque hole  85 . 
         [0018]    If the variability in on-corning friction element engagement torque significantly alters the shape and depth of torque hole  85 , the vehicle occupant may experience inconsistent shift feel. A deep torque hole may be perceived as an unpleasant shift shock. During torque phase  32 , off-going friction element control force is reduced as shown at  82  to induce an incipient slip  83 . Controller  4  attempts to maintain off-going friction element slip at a target level through a closed-loop control based on off-going friction element speed  96  which may be directly measured or indirectly derived from speed measurements at pre-determined locations. A variability in off-going friction element control force  82  of off-going element slip torque may alter the shape and depth of torque hole  85 , thus affecting shift feel. If controller  4  inadvertently allows a sudden increase in off-going friction element slip level, automatic transmission input speed or engine speed  90  may surge momentarily, causing the so-called engine speed flair or engine flair. The engine flair may be perceived by a vehicle occupant as an unpleasant sound. 
         [0019]    Controller  4  initiates off-going friction element release process at a predetermined timing shown at which may be based on a commanded value of on-corning friction element control force  93 . Controller  4  lowers off-going friction element control force, following a pre-calibrated profile  94 . If a release of off-going friction element D is initiated prematurely before on-coming friction element C develops enough torque, engine speed or input shaft speed may rise rapidly in an uncontrolled manner. If this engine speed flair  90  is detected, controller  4  increases off-going friction element control force to delay off-going friction element release process. Alternatively to the pre-determined off-going friction element release timing, controller  4  may utilize speed signals to determine a final off-going friction element release timing. When a sign of significant gear tie-up, which may be manifested as a measurable drop in input shaft speed, is detected, off-going friction element D is released following a pre-calibrated force profile. Inertia phase  33  begins when off-going friction element torque capacity or control force drops to non-significant level  95 . During inertia phase  33 , off-going friction element slip speed rises  96  while on-coming friction element slip speed decreases  97  toward zero. The transmission input shaft speed drops as shown at  98  as the planetary gear configuration changes. During inertia phase  33 , the output shaft torque  99  is primarily affected by on-coming friction element torque capacity, which is indicated by its control force  100 . When on-coming friction element C becomes securely engaged at  101 , the shift event completes. 
         [0020]    In summary, a prior art methodology, which is based on an open-loop on-coming friction element control during a torque phase, cannot account for control system variability and dynamically-changing shift conditions during the torque phase, resulting in inconsistent shift feel or unpleasant shift shock. A pre-determined off-going friction element release timing with a pre-calibrated control force profile cannot ensure an optimal timing under dynamically changing shift conditions, resulting in inconsistent shift feel or unpleasant shift shock. The alternative approach to gauge off-going friction element release timing based on speed signals often results in a hunting behavior between gear tie-up and engine flair, leading to inconsistent shift feel. Furthermore, off-going friction element slip control is extremely difficult because of its high sensitivity to slip conditions. In addition, a large discontinuity exists between static and dynamic friction coefficients, introducing a large torque disturbance during an incipient slip control. A failure to achieve a seamless off-going friction element slip control during the torque phase leads to undesirable shift shock. 
         [0021]    As can be seen from the above discussion the controllability of both off-going friction element and on-coming friction element is desirable in order to deliver a consistent and seamless shift quality. The prior art does not have a cost effective design solution to the problem of directly measuring torque passing through either a multiple disc clutch or a band brake and therefore is a need in the art for a transmission control system that minimizes shift shock during a gear ratio change that does not rely solely on traditional speed signal measurement or a predetermined open-loop control and instead relies on measuring friction element load level in either a multiple plate clutch or a band brake for consistently controlling its torque level through a closed loop approach. 
       SUMMARY OF THE INVENTION 
       [0022]    The present invention is directed to a load sensor assembly for measuring an amount of torque transmitted through a torque establishing element of an automatic transmission. The assembly comprises a core mounted on a transmission housing and a load sensor mounted on the core and positioned against a portion of the torque establishing element whereby a portion of the amount of torque transmitted through the torque establishing element travels through the load sensor and is measured by the load sensor assembly. Preferably, a cable is connected to the load sensor for transmitting a signal representative of the amount of torque to a transmission controller. A cover or sleeve extends over the core and the sensor. 
         [0023]    In a preferred embodiment, the torque establishing element is a multiple disk friction element including an end plate and a spline connection between the transmission case and the end plate. The connection has teeth that extend from the transmission case and cooperate with teeth extending from the end plate. The load sensor assembly is mounted on the transmission housing between two spline teeth extending from the end plate and in a location where a spline tooth would normally be located. Preferably, the core is made of metal and the sleeve is made from one of rubber, plastic and metal. The sensor may have several different configurations. In one configuration, a pin is fixed to the end plate and the load sensor is placed against the pin. In another configuration, the force sensor is a load-resistive elastomer deposited on a thin film and the core is a tooth of a friction element plate. An example of such a thin film force sensor can be found in U.S. Pat. No. 6,272,936, which is incorporated herein by reference. In yet another configuration, the core is a metal beam securely anchored to the transmission case and the load sensor is a strain sensor that measures an amount of strain on the beam caused by the torque. 
         [0024]    In another embodiment the torque establishing element is a band brake including an anchor bracket and a band brake strap. The core may engage the strap in many ways. In one configuration, the band brake strap has a block extending therefrom and the core passes through the transmission housing and engages the block. The load sensor is located between the core and the block. In another configuration, the band brake strap has a hook extending therefrom formed by punching a hole in the strap. The core passes through the transmission housing and engages the hook and the load sensor is located between the core and the hook. In yet another configuration, the anchor bracket has a pin extending therefrom. The core passes through the transmission housing and engages the pin. The load sensor is located between the core and the pin. Preferably, a cushion is located between the load sensor and the cover. 
         [0025]    In yet another embodiment, the torque establishing element is a band brake including an anchor bracket and a hand brake strap while the core is an anchor pin, which does not necessarily have a cover, mounted in the transmission case. The anchor pin extends out of the transmission case and engages the anchor bracket. The load sensor is mounted between the anchor pin and the transmission case whereby torque is transferred to the band strap, pushes on the anchor pin and is sensed by the load sensor. Preferably, a cushion is located between the load sensor and the anchor pin. The core includes an anchor pin mounted in the transmission case. The core extends out of the transmission case and is connected to an anchor strut which, in turn, engages the anchor bracket. The load sensor is mounted between the anchor pin and the transmission case. Torque is transferred to the band strap where it pushes on both the anchor strut and pin, with the torque being sensed by the load sensor. Preferably, the transmission housing includes a hole for supporting the anchor pin. A nut is mounted in one end of the hole and secures the anchor pin to the housing. A plug and a support are located between the nut and the anchor pin. With this arrangement, torque passing through the friction elements of a transmission may be directly measured and shift shock and engine flair may be reduced. 
         [0026]    Additional objects, features and advantages of the present invention will become more readily apparent from the following detailed description of preferred embodiments when taken in conjunction with the drawings, wherein like reference numerals refer to corresponding parts in the several views. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0027]      FIG. 1  is a schematic diagram of a gearing arrangement for an automatic transmission system according to the prior art; 
           [0028]      FIG. 2  is a chart showing a clutch and brake friction element engagement and release pattern for establishing each of six forward driving ratios and a single reverse ratio for the transmission schematically illustrated in  FIG. 1 ; 
           [0029]      FIG. 3  is a plot of a general process of a synchronous friction element-to-friction element upshift event from a low gear configuration to a high gear configuration for the prior art automatic transmission system of  FIG. 1 ; 
           [0030]      FIG. 4  is a plot of the general process of a synchronous friction element-to-friction element upshift event from the low gear configuration to the high gear configuration in which the off-going friction element is released prematurely compared with the case shown in  FIG. 3 ; 
           [0031]      FIG. 5  is a plot of the general process of a synchronous friction element-to-friction element upshift event from the low gear configuration to the high gear configuration in which off-going friction element release is delayed compared with the case shown in  FIG. 3 ; 
           [0032]      FIG. 6  is plot of a prior art synchronous friction element-to-friction element upshift control from a low gear configuration to a high gear configuration based on speed measurements of powertrain components for the automatic transmission system in  FIG. 1  wherein an off-going friction element remains locked during the torque phase; 
           [0033]      FIG. 7  is plot of a prior art synchronous friction element-to-friction element upshift control from a low gear configuration to a high gear configuration based on speed measurements of powertrain components for the automatic transmission system in  FIG. 1 , wherein an off-going friction element is slipped during the torque phase; 
           [0034]      FIG. 8  is a schematic diagram of a gearing arrangement for an automatic transmission system including load sensor locations in accordance with a first preferred embodiment of the invention; 
           [0035]      FIG. 9  is a plot of a synchronous friction element to friction element upshift control from a low gear configuration to a high gear configuration for the automatic control system in  FIG. 8  based on direct measurements or estimates of torsional load exerted onto an off-going friction element in accordance with a preferred embodiment of the invention; 
           [0036]      FIG. 10  is a flow chart showing an on-coming friction element control method in accordance with a preferred embodiment of the invention; 
           [0037]      FIG. 11  is a flow chart showing an off-going element release control method in accordance with a preferred embodiment of the invention; 
           [0038]      FIG. 12  is a plot of the process used to determine an ideal release timing of the off-going friction element in accordance with first preferred embodiment of the invention; 
           [0039]      FIG. 13  is a flow chart showing a shift control method in accordance with a preferred embodiment of the invention; 
           [0040]      FIG. 14  is a plot of a synchronous friction element-to-friction element upshift from a low gear configuration to a high gear configuration for the automatic transmission control system in  FIG. 8  based on the direct measurements or estimates of torsional load exerted onto an off-going friction element and an on-coming element in accordance with another preferred embodiment of the invention; 
           [0041]      FIG. 15  is a flow chart showing an on-coming friction element shift control method in accordance with another preferred embodiment of the invention; 
           [0042]      FIG. 16A  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed between two teeth of a endplate of a friction element for measuring a relative load level on the friction element; 
           [0043]      FIG. 16B  depicts the load sensor assembly of  FIG. 16A  installed in a transmission case; 
           [0044]      FIG. 17A  depicts a load sensor assembly in accordance with another preferred embodiment of the invention placed against a pin extending from an endplate of a friction element for measuring a relative load level on the off-going friction element; 
           [0045]      FIG. 17B  depicts the load sensor assembly of  FIG. 17A  installed in a transmission case; 
           [0046]      FIG. 18  depicts a load sensor in accordance with another preferred embodiment of the invention formed of a thin film-type load sensor and attached to a tooth for measuring a relative load level on the off-going friction element; 
           [0047]      FIG. 19  depicts a load sensor assembly in accordance with another preferred embodiment of the invention formed of a metal beam for measuring a relative load level on the off-going friction element; 
           [0048]      FIG. 20  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0049]      FIGS. 21A-21C  depict a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0050]      FIGS. 22A and 22B  depict a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0051]      FIG. 23  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0052]      FIG. 24  depicts a chart in accordance with another preferred embodiment of the invention; 
           [0053]      FIG. 25  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0054]      FIG. 26  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0055]      FIG. 27  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; 
           [0056]      FIG. 28  depicts a load sensor assembly in accordance with another preferred embodiment of the invention installed on a band brake type friction element for measuring a relative load level on the friction element; and 
           [0057]      FIG. 29  depicts a chart in accordance with another preferred embodiment of the invention. 
       
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
       [0058]    With initial reference to  FIG. 8 , there is shown an automotive transmission employing the invention. As this automatic transmission arrangement is similar to the one schematically illustrated in  FIG. 1  all the same parts have been indicated with corresponding reference numbers and therefore a duplicate discussion of these parts will not be made here. Instead, of particular importance is the addition of a torque sensor  120  located in friction element C, a load sensor  130  located in friction element D, and a torque sensor  131  located in transmission output shaft  24 , all connected to controller  4  for controlling various functions of transmission  2  as will be more fully discussed below. 
         [0059]      FIG. 9  shows a torque phase control method according to a preferred embodiment of the invention for a synchronous friction element-to-friction element upshift from a low gear configuration to a high gear configuration for the automatic transmission system in  FIG. 8 . The on-coming friction element control method illustrated here is also applicable to non-synchronous shift control. The shift event is divided into 3 phases: preparatory phase  31 , torque phase  32  and inertia phase  33 . During preparatory phase  31 , an on-coming friction element piston is stroked to prepare for its engagement. At the same time, off-going friction element control force or its torque capacity is reduced as shown at  404  as a step toward its release. During torque phase  32 , on-coming friction element control force is raised in a controlled manner as shown at  405 . More specifically, controller  4  commands on-coming friction element actuator to follow a target on-coming friction element engagement torque profile  406  through a closed-loop control directly based on the measurements of on-coming friction element engagement torque  407  during torque phase  32 . On-coming friction element torque  407  may be directly measured using a load sensor according to this invention as more fully described below. On-coming friction element engagement torque directly affects transmission output torque that is transmitted to the vehicle wheels. This torque-based close-loop control eliminates or significantly reduces the undesirable effects of on-coming friction element engagement torque sensitivity to hardware variability and shift conditions, achieving a consistent shift feel, regardless of shift conditions. 
         [0060]    Alternatively to the direct measurements, on-coming friction element torque can be determined from the measurements of transmission output shaft torque using torque sensor  131  depicted in  FIG. 8 . Mathematically, on-coming friction element torque T OCE  can be described as a function of measured output shaft torque T OS  as: 
         [0000]        T   OCE ( t )= G   OCE   T   OS ( t )  Eq. (1)
 
         [0000]    Where G OCE  can be readily obtained based on a given gear set geometry. 
         [0061]    Yet alternatively, on-coming friction element torque T OCE  can be estimated through the following Eq. (2), based on a slight change in transmission component speeds ω i  at pre-determined locations (i=1, 2, . . . , n), 
         [0000]        T   OCE ( t )= F   trans (ω i   ,t )  Eq. (2)
 
         [0000]    where t indicates time and F trans  represents a mathematical description of a transmission system. More specifically, as on-coming friction element engagement torque rises  407 , torque levels transmitted through various transmission components change. This creates small, but detectable changes in ω i . A transmission model, F trans , can be readily derived to estimate on-coming friction element engagement torque when off-going friction element remains locked during torque phase  32 . 
         [0062]    Controller  4  commands enough off-going friction element control force  408  to keep it from slipping, maintaining the planetary gearset in the low gear configuration during torque phase  32 . As on-coming friction element engagement torque  407  increases, a reaction torque goes against a component that is grounded to a transmission case. More specifically, in this case, torque transmitted through off-going friction element or torsional load  409  exerted onto off-going friction element D decreases proportionally. Off-going friction element load level  409  can be directly monitored using a torque sensor such as is more fully discussed below. Alternatively, off-going friction element load level T OGE    409  can be calculated from measured or estimated on-coming friction element engagement torque T OCE    407  when off-going friction element remains locked during torque phase  32  according to: 
         [0000]        T   OGE ( t )= F   OCE/OGE ( T   OCE ( t ))  Eq. (3)
 
         [0000]    where F OCE/OGE  represents a torque ratio between on-corning friction element C and off-going friction element D at the low gear configuration and can be obtained based on gear set geometry. According to this invention, off-going friction element D is released at an ideal timing when torque load exerted onto off-going friction element D becomes zero or a near-zero level. Transmission controller  4  initiates a release process of off-going friction element D as shown at  410  as off-going friction element load  409  approaches zero at  411 . Off-going friction element torque is dropped quickly as shown at  412  with no slip control. Since no off-going friction element slip control is involved, the method is insensitive to off-going friction element break-away friction coefficient variability. In addition, the quick release of off-going friction element D shown at  412  induces little disruption in output shaft torque at  413  because off-going friction element load level is near zero as shown at  411  at the moment of release. Off-going friction element D starts slipping  411  once its control force reaches a non-significant level. During inertia phase  33 , a conventional control approach may be utilized based on on-coming friction element slip measurements. Off-going friction element slip speed increases as shown at  415  while on-coming friction element slip speed decreases as shown at  416 . The transmission input speed drops as shown at  417  as the planetary gear configuration changes. During inertia phase  33 , output shaft torque  418  is primarily affected by on-coming friction element torque level  419 . Alternatively to the conventional control, a closed loop control that is based on measured or estimated on-corning friction element torque may continue to be employed. When on-coming friction element C completes engagement or when its slip speed becomes zero as shown at  420 , the shift event completes. 
         [0063]      FIG. 10  shows a flow chart of closed-loop on-coming friction element engagement torque control during the torque phase depicted in  FIG. 9 . Step  430  is the beginning of torque phase  32 . Controller  4  chooses a desired on-coming element torque at step  431  and measures or estimates an actual torque at step  432 . At step  433 , the on-coming friction element actuator is then adjusted by controller  4  based on the difference between the measured/estimated torque level and the actual torque level. At step  434 , controller  4  determines if torque phase has ended and if so controller  4  starts inertia phase  33  at  436 . 
         [0064]      FIG. 11  shows a flow chart of an off-going friction element torque control process during torque phase  32  depicted in  FIG. 9 . The process starts at step  440  at the beginning of torque phase  32 . A load transmitted through locked off-going friction element D is directly measured or estimated at step  441 . At step  442 , when its load level drops below a predetermined level, off-going friction element D is promptly released at step  444 . The control process ends at step  445  at the end of torque phase  32 . 
         [0065]    Alternatively to the measurements or estimates of absolute load levels,  FIG. 12  illustrates the process to determine the ideal release timing of off-going friction element D based on relative load measurements or estimates according to this invention.  FIG. 12  depicts an actual load profile  451  exerted on off-going friction element D and a relative load profile L(t)  452  measured by torque sensor  130  during the upshift event in  FIG. 9 . The preferred embodiment requires only relative load profile L(t)  452 . Relative load profile L(t)  452  is preferably constructed from uncalibrated sensor output that reflects actual load profile  451 , but not its absolute levels. This feature eliminates the need of a full sensor calibration across the entire load range. It also makes the preferred embodiment insensitive to sensor output drift over time. However, the preferred embodiment relies on knowledge of sensor measurement L 0    453  which corresponds to zero off-going friction element load level  454 . Sensor measurement L 0    453  can be readily identified, as often as required, by sampling sensor output while vehicle transmission  2  is in a neutral or a similar condition where no load is exerted onto off-going friction element D. Transmission controller  4  collects relative load data  455  during torque phase  32  to dynamically construct relative load profile L(t)  452 . Then, controller  4  extrapolates L(t) to predict t 0    457  where L(t 0 )=L 0 . Once t 0    457  is obtained in advance, controller  4  predicts when to initiate an off-going friction element release process. Specifically controller  4  starts the release process at a time equal to t 0 −Δt shown at  458 , where Δt is the time required to quickly drop off-going friction element control force to zero. In this way, off-going friction element D starts slipping at or near ideal timing t 0    457  when the actual off-going friction element load level is at or close to zero as shown by reference numeral  454 . 
         [0066]      FIG. 13  presents a flow chart of the new upshift control method according to this invention. During preparatory phase  31  at step  461  of a synchronous upshift event, off-going friction element torque capacity or apply force is reduced to a holding level without allowing any slip at step  462  while on-coming friction element piston is stroked at step  463 . During torque phase  32 , transmission controller  4  measures at step  465  a relative load level exerted onto off-going friction element D at a pre-specified sampling frequency using torque sensor  130  described further below. Controller  4  repeats this measurement step  465  until enough data points are collected at step  466  for dynamically constructing a relative load profile at step  467  that shows load as a function of time L(t). Once relative load profile L(t) is obtained, controller  4  predicts the ideal off-going friction element release timing t 0  at step  468  so that L(t 0 )=L 0  where L 0  corresponds to a substantially zero load level on off-going friction element D. Controller  4  initiates an off-going friction element release process at t 0 −Δt as shown as step  469  where Δt is a pre-specified time required to quickly drop off-going friction element apply force to zero. Alternatively, controller  4  may initiate the off-going friction element release process at t thres  such that L(t thres )=L thres  where L thres  is a predetermined threshold. No slip control is required for off-going friction element D during torque phase  32 . Inertia phase  33  starts when off-going friction element D is released. The control methodology illustrated in  FIG. 10  is preferably applied to on-coming friction element C during torque phase  32 . A conventional on-coming friction element control may be applied during inertia phase  33  based on speed signals. When on-coming friction element C becomes securely engaged at step  473 , the shift event completes at step  474 . 
         [0067]      FIG. 14  illustrates another preferred embodiment of the invention relating to a transmission system with an on-coming friction element actuator which may not have a sufficient control bandwidth compared with a sampling time of load measurements. At the beginning of torque phase  32 , a transmission controller raises on-coming friction element actuator force based on a pre-calibrated slope  480  over a time interval Δt between t 0  and t 1  as shown at interval  481 . During interval  481 , on-coming friction element load is either measured or estimated with a sampling time finer than Δt to construct an engagement torque profile  482 . If the measured or estimated torque profile  482  indicates a slow rise compared with a target torque profile  483 , controller  4  increases a slope of commanded on-coming friction element control force for a next interval  485  between t 1  and t 2 . On the other hand, if the actual torque is rising faster than a target profile, controller  4  reduces a slope of commanded on-coming friction element control force. For example, during interval  485  between t 1  and t 2 , on-coming friction element load is either measured or estimated with a sampling time finer than Δt to construct an engagement torque profile  486 . The measured or estimated slope  486  of the engagement torque is compared against a target profile  487  to determine a slope  488  of commanded force profile for the following control interval. This process is repeated until the end of torque phase  32 . The off-going friction element release control remains the same as that shown in  FIG. 9 . 
         [0068]      FIG. 15  shows a flow chart of alternative closed-loop on-coming friction element engagement torque control during torque phase depicted in  FIG. 14 . The start of torque phase  32  is shown at step  520 . Following path  521 , the off-coming friction element torque is measured or estimated at step  522  and torque profile  482  is created therefrom at step  523 . The method may have to go through several iterations as shown by decision block  524  and return loop  525 . Torque slope profile  486  or an average derivative of torque profile  482  is calculated at  526  and while a desired target slope profile  487  is calculated at  527  and compared with torque slope profile  486  at  528 . The actuator force slope is increased  529  or decreased  530  and the process continues  531 ,  532  until the end of torque phase  32 . The process then proceeds to inertia phase  33  at  533 . 
         [0069]    While the shift control has been discussed above, attention is now directed to the structure of the various load sensor assemblies.  FIG. 16A ,  16 B,  17 A,  17 B,  18  and  19  depict several preferred embodiments of load sensor assemblies for measuring a relative load level exerted on off-going friction element D or on-coming element C according to preferred embodiments of the invention.  FIG. 16A  shows a cross-sectional view of a load sensor assembly  601  design according to a preferred embodiment. In  FIG. 16A , sensor assembly  601  is installed between two teeth  602 ,  603  of an end plate  604  of off-going friction element D. Assembly  601  includes a core  605 , a load sensor  606  and a sleeve  607 . Core  605  is preferably made from a metal, such as steel or aluminum, and is securely grounded to a transmission case  608  through anchor bolts  609 . Load sensor  606  is preferably a film-type sensor constructed with a pressure-resistive material. Sensor  606  generates an electrical signal that corresponds to a relative level of loading force  610 . Sleeve  607 , which protects sensor  606 , is preferably made from rubber, plastic or metal. While cover  607  is referred to as either a sleeve or a cover, it is to be understood that the terms are interchangeable.  FIG. 16B  illustrates an installation of sensor assembly  601  in transmission case  608 . Sensor assembly  601  is securely positioned in a location where a spline tooth is normally located otherwise. When off-going friction element plates are installed, end plate  604  fits snugly around sensor assembly  601 , providing a preload to sensor  606 . That is, sensor  606  preferably indicates non-zero output L 0  even when no load is exerted on off-going friction element D or its end plate  604 . When the torque load is exerted as shown by arrow  610  during a shift event, the output from sensor  601  provides a relative measure of the load on off-going friction element D. When this embodiment is employed to measure relative load exerted onto an off-going friction element such as when torque sensor  130  is used to measure the load on friction element D, it is readily understood that optimal friction element release timing is identified when the sensor output level approaches to L 0  corresponding to zero load level. 
         [0070]      FIGS. 17A and 17B  depict another sensor assembly  611  which has a similar structure to assembly  601  in  FIG. 16A . Assembly  611  includes a grounded core  612 , a force sensor  613  and a sleeve  614 . However, as illustrated in  FIG. 17A , assembly  611  is placed against a pin  615  that is fixed to an end plate  616  of off-going friction element D. Sensor  613  is preloaded against pin  615 , providing non-zero output in the absence of torque load on off-going friction element end plate  616  ( FIG. 17B ). When a torque load is exerted on off-going friction element D, pin  615  is pressed with a force  617  against sensor  613  across sleeve  614 . This enables sensor  613  to provide the relative measure of torque load on off-going friction element D.  FIG. 17B  shows a view of sensor assembly  611  and off-going friction element end plate  616  with pin  615  in a transmission case  618 . 
         [0071]      FIG. 18  shows another potential embodiment of this invention wherein a thin film-type force sensor  621  is directly attached to a tooth  622  of a friction element plate  623 , covered with a protective sleeve  624 . Sleeve  624  is preferably made from rubber, plastic or metal. When plate  623  is installed into a transmission case  625 , sensor  621  directly measures contact load  626  between friction element tooth  622  and a spline  627  through sleeve layer  624 , providing a relative measure of the load exerted onto off-going friction element D. 
         [0072]      FIG. 19  shows another preferred embodiment of the invention wherein a metal beam  631 , which is securely anchored to a transmission case  632 , is installed and positioned between two teeth  633 ,  634  of an off-going friction element plate  635 . As a load level  636  exerted on plate  635  varies, a strain level of beam  631  changes. The level of the strain is detected through a strain sensor  637 , providing a relative measure of is torque load exerted on off-going friction element D. Optionally, a cover may be added to protect strain sensor  637 . 
         [0073]      FIGS. 20 ,  21 A,  21 B,  21 C,  22 A,  22 B and  23 - 29  show various preferred embodiments of the invention relating to directly measuring torque in a friction element. More specifically,  FIG. 20  shows a partial view of a band brake system  700  with a load sensing assembly  731 . Brake system  700  includes an anchor end of a band strap  732 , a pin or a hook  733 , and an anchor bracket  734 . Band strap  732  is preferably either a single-wrap or double-wrap type. Load sensor assembly  731  includes an assembly core  735 , a load sensing unit  736  and a protective sleeve or cover  737 . Assembly core  735  is made of a metal and securely mounted to a transmission case  738  with a bolt  739  or any other means. Cover  737  may be made of metal, rubber, plastic or any other materials. Cover  737  protects sensor unit  736  from direct contact with pin or hook  733  for reduced sensor material wear. Cover  737  may be made of a thermally-insulated material to protect sensor  736  from heat. Cover  737  also acts as a protective shield against any other hostile conditions that include electro-chemical reaction with transmission oil. Load sensing unit  736 , which may be a pressure resistive film-type, is positioned between core  735  and cover  737 . The tip of sensor  736  is positioned against pin  733  across cover  737 . When a band engagement is commanded, strap  732  is pulled by a hydraulic servo (which is described below) in the direction shown with an arrow  740 . Band strap  732  stretches slightly, pushing pin or hook  733  against load sensor  736 . Load sensor  736  generates an electrical signal according to a magnitude of the contact force. That is, sensor  736  provides a relative measure of band tension at the location of pin  733 . The electrical signal is transmitted to a data acquisition unit (not shown) and then to controller  4  through an electrical cable  741 . 
         [0074]      FIGS. 21A ,  21 B and  21 C depict band strap designs in detail. In  FIG. 21A , a band strap  732  has a part punched out and bent to form a pin or a hook  753  and a hole  752 . Hole  752  also acts as an oil drain during band engagement. In  FIG. 21B , a small pin or a block  754  is riveted, screwed or welded to strap  732 . Alternatively, a pin or a hook  755  can be formed as a part of an anchor bracket  734  as shown in  FIG. 21C . A pin  755  is attached to a band anchor bracket  734  instead of a strap  732 . Sensor assembly  731  is positioned against the pin  755 . Since bracket  732  is stiffer than the strap  732 , its strain is smaller under loaded conditions during both holding and engagement. Thus, a level of force exerted onto a load sensor  736  through a micro displacement of pin  755  is reduced significantly. The lower stress level improves the life of the sensor assembly  731  while enabling the use of a sensor  736  rated for a lower maximum force. 
         [0075]      FIG. 22A  illustrates sensor functions during a band engagement process. When the engagement is initiated, transmission controller  4  sends an electrical signal I(t) to raise and regulate a hydraulic force  761  applied to a servo piston  762 . As servo piston  762  is stroked, a servo rod  763  pulls one end  764  of band strap  732 . Tension around strap  732  builds up, squeezing out lubrication oil  766  from a band-drum interface. During the engagement, brake torque from strap  732  to a drum  767  is partly transmitted through viscous shear across oil  766 . The brake torque is transmitted through a mechanical frictional force once strap  732  makes physical contact with drum  767 . According to a conventional analysis, the relationships between engagement torque T eng , band tension at a pin F pin    733  and band tension at a servo F servo    769  can be written as follows, assuming a Coulomb friction model as a primary torque transfer mechanism between band strap  732  and drum  767 : 
         [0000]        T   eng   =F   servo   R ( e   μβ −1)  Eq. (4)
 
         [0000]        F   pin   =F   servo   e   μβ   Eq. (5)
 
         [0000]    where R=drum radius, m=a Coulomb friction coefficient, b=a band wrap angle  770  assuming that pin  733  is positioned sufficiently close to an anchor  734 . Drum  767  rotates in the same direction  772  as the hydraulic force  761 . Substituting Eq. (5) into Eq. (4) yields: 
         [0000]    
       
         
           
             
               
                 
                   
                     T 
                     eng 
                   
                   = 
                   
                     
                       
                         F 
                         pin 
                       
                        
                       
                         R 
                          
                         
                           ( 
                           
                             1 
                             - 
                             
                                
                               
                                 - 
                                 μβ 
                               
                             
                           
                           ) 
                         
                       
                        
                       
                           
                       
                        
                       or 
                        
                       
                           
                       
                        
                       
                         F 
                         pin 
                       
                     
                     = 
                     
                       
                         T 
                         eng 
                       
                       
                         R 
                          
                         
                           ( 
                           
                             1 
                             - 
                             
                                
                               
                                 - 
                                 μβ 
                               
                             
                           
                           ) 
                         
                       
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     6 
                     ) 
                   
                 
               
             
           
         
       
     
         [0076]    Since the electrical output signal S pin  from the sensor is approximately linear with band tension F pin : 
         [0000]        S   pin   =kF   pin   Eq. (7)
 
         [0000]    where k is a proportional constant. Substituting Eq. (7) into Eq. (6) yields: 
         [0000]    
       
         
           
             
               
                 
                   
                     
                       S 
                       pin 
                     
                     = 
                     
                       
                         
                           k 
                           
                             R 
                              
                             
                               ( 
                               
                                 1 
                                 - 
                                 
                                    
                                   
                                     - 
                                     μβ 
                                   
                                 
                               
                               ) 
                             
                           
                         
                          
                         
                           T 
                           eng 
                         
                       
                       = 
                       
                         
                           
                             k 
                             ′ 
                           
                            
                           
                             T 
                             eng 
                           
                            
                           
                               
                           
                            
                           or 
                            
                           
                               
                           
                            
                           
                             
                                
                               
                                 S 
                                 pin 
                               
                             
                             
                                
                               t 
                             
                           
                         
                         = 
                         
                           
                             k 
                             ′ 
                           
                            
                           
                             
                                
                               
                                 T 
                                 eng 
                               
                             
                             
                                
                               t 
                             
                           
                         
                       
                     
                   
                    
                   
                     
 
                   
                    
                   where 
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     8 
                     ) 
                   
                 
               
             
             
               
                 
                   
                     k 
                     ′ 
                   
                   = 
                   
                     k 
                     
                       R 
                        
                       
                         ( 
                         
                           1 
                           - 
                           
                              
                             
                               - 
                               μβ 
                             
                           
                         
                         ) 
                       
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     9 
                     ) 
                   
                 
               
             
           
         
       
     
         [0000]    According to Eq. (8), the sensor output S pin  provides a relative measure of band brake engagement torque T eng . 
         [0077]    This embodiment provides a relative measure of T eng  and its derivative (dT eng /dt) that enables a closed loop control of on-coming friction element engagement process during torque phase  32 . It significantly improves band engagement control, mitigating a sudden rise of band brake torque known as “grabbing” behaviors. Alternatively, the sensor signals may be utilized to adaptively optimize open-loop calibration parameters such as a rate of pressure rise as a function of oil temperature in order to achieve a consistent (dT eng /dt). The similar analysis can be applied to the so-called “de-energized” band engagement where the drum spins in the opposite direction of the servo. 
         [0078]      FIG. 22B  illustrates sensor functions while band strap  732  is securely engaged around drum  767  under a holding condition without any slippage. In this case, the band tension F pin  at pin  733  reflects both the level of the band tension F servo    784  at the servo and the level of torque load T load    785  exerted onto band  732  and drum  767  from the adjoining components (not shown). It is important that one should clearly differentiate T load  from T eng  which is brake torque exerted from the band to the drum under slipping conditions. 
         [0079]    According to a conventional analysis, the relationships between F pin , F servo  and T load  can be algebraically written as: 
         [0000]    
       
         
           
             
               
                 
                   
                     F 
                     pin 
                   
                   = 
                   
                     
                       
                         F 
                         servo 
                       
                       + 
                       
                         
                           
                             T 
                             load 
                           
                           R 
                         
                          
                         
                             
                         
                          
                         or 
                          
                         
                             
                         
                          
                         
                           T 
                           load 
                         
                       
                     
                     = 
                     
                       R 
                        
                       
                         ( 
                         
                           
                             F 
                             pin 
                           
                           - 
                           
                             F 
                             servo 
                           
                         
                         ) 
                       
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     10 
                     ) 
                   
                 
               
             
           
         
       
     
         [0000]    Substituting Eq. (10) into Eq. (7), the sensor output S pin  can be described as a function of F servo  and T load  as: 
         [0000]    
       
         
           
             
               
                 
                   
                     S 
                     pin 
                   
                   = 
                   
                     
                       kF 
                       pin 
                     
                     = 
                     
                       
                         kF 
                         servo 
                       
                       + 
                       
                         
                           k 
                           R 
                         
                          
                         
                           T 
                           load 
                         
                       
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     11 
                     ) 
                   
                 
               
             
           
         
       
     
         [0000]    Note that F servo  is a function of an electrical signal I commanded to a hydraulic control system from a transmission controller. That is: 
         [0000]        F   servo   =F   servo ( I )  Eq. (12)
 
         [0000]    Substituting Eq. (12) into Eq. (11) results in: 
         [0000]    
       
         
           
             
               
                 
                   
                     S 
                     pin 
                   
                   = 
                   
                     
                       kF 
                       pin 
                     
                     = 
                     
                       
                         
                           kF 
                           servo 
                         
                          
                         
                           ( 
                           I 
                           ) 
                         
                       
                       + 
                       
                         
                           k 
                           R 
                         
                          
                         
                           T 
                           load 
                         
                       
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     13 
                     ) 
                   
                 
               
             
           
         
       
     
         [0000]    In the absence of T load , Eq. (13) becomes: 
         [0000]        S   pin   =kF   servo ( I )≡ S   pin   noload ( I )  Eq. (14)
 
         [0000]    where S pin   noload  is defined as the sensor output measured under no load condition for a given level of I. In practice S pin   noload  can be readily obtained, as required, by sweeping the servo actuator with a varying level of I while a vehicle is in a stationary condition. Substituting Eq. (14) into Eq. (13) yields: 
         [0000]    
       
         
           
             
               
                 
                   
                     
                       S 
                       pin 
                     
                     - 
                     
                       
                         S 
                         pin 
                         noload 
                       
                        
                       
                         ( 
                         I 
                         ) 
                       
                     
                   
                   = 
                   
                     
                       k 
                       R 
                     
                      
                     
                       T 
                       load 
                     
                   
                 
               
               
                 
                   Eq 
                   . 
                   
                       
                   
                    
                   
                     ( 
                     15 
                     ) 
                   
                 
               
             
           
         
       
     
         [0000]    Thus, S pin −S pin   noload (I) provides a relative measure of torque load T load  for a given electrical input I. The optimal timing to release off-going friction element during a synchronous shift is when the load exerted onto off-going friction element or T load  becomes zero. This can be readily determined by sampling S pin  and evaluating S pin −S pin   noload (I) for a given electrical signal I. The use of the load sensor assembly according to this embodiment significantly improves band release controllability during a synchronous shift under all the operating conditions. 
         [0080]      FIG. 23  shows a cross-sectional view of another sensor assembly  811  including a cushion element  812  inserted between a load sensor  813  and a pin or a block  814  that is attached to a band strap or an anchor bracket. Cushion element  812  is preferably made of a rubber. Alternatively, cushion element  812  may be made of a metal in the form of a spring such as a disk spring or a conical spring. A protective cover  815  is preferably positioned between cushion element  812  and block  814 . Cover  815  is readily slidable at a nominal force under loaded conditions. The loading force is transmitted from block  814  to load sensor  813  by deformation of cushion element  812 . Accordingly, cushion element stiffness is used to specify a force range at sensor  813  for a given range of loading force at block  814 . The force transmitted to load sensor  813  becomes limited once the cushion element surface becomes flush with surface  817  of the assembly core. This non-linear characteristic indicated at  818  enables high resolution force measurement for a targeted load range  819  as shown in  FIG. 24  while protecting sensor  813  from excessive loading. 
         [0081]      FIG. 25  shows an alternative embodiment of this invention. In this design, a load sensor  821  is placed at the bottom of a band anchor pin  822  inside a transmission case  823 . Electrical cable  824  attached to sensor  821  is routed outside through case  823 . The tip of pin  822  is inserted into an anchor bracket  826 , which is attached to band strap  825 . When the band brake system is actuated, strap  825  is hydraulically or mechanically tightened around a drum such that anchor bracket  826  pulls pin  822  in the direction of anchor load  828  as represented by an arrow. Accordingly, load sensor  821  directly measures an anchor load  828  exerted onto pin  822  from the anchor bracket  826 . A cushion element  831  is preferably placed between the bottom of an anchor pin  822  and load sensor  821 . Note that the sensing area of sensor  821  is smaller than the surface area of cushion element  831 . The anchor load supported by pin  822  is distributed over the surface of cushion element  831 . Accordingly, only part of the anchor load is transmitted to load sensor  821 . This enables the use of a sensor rated for a lower maximum force. 
         [0082]    In  FIG. 26 , a strut  841  is inserted between an anchor bracket  826  and an anchor pin  843 . Strut  841  enables the flexible placement of anchor pin  843  with respect to band strap  825  and transmission case  823 . Also, an angle  845  between strut  841  and pin  843  may be adjusted to optimize a level of the axial loading that bracket  876  exerts onto pin  843  through strut  841 . Cushion element  831  and the reduced axial loading allow the use of a sensor  821  rated for a lower maximum force. Alternatively, angle  845  may be adjusted to reduce the side loading onto pin  843  to minimize sensor output hysteresis caused by sticky pin displacement under the loaded conditions. 
         [0083]    The embodiment of the invention in  FIG. 27  shares many of the same features described in connection with the embodiment in  FIG. 26 . First, anchor pin  853  is inserted into an unthreaded hole  852  inside transmission case  823 . Its large head  854  prevents pin  853  from falling through hole  852 . A cushion element  836  and a load sensor  821  are placed against pin head  854 . Cushion element  836  may be made of a rubber and act as a seal to protect the sensor  821  from transmission oil. Behind sensor  821  and cushion element  836  is a sensor support dish  857 , which may be made of a metal. Sensor support dish  857  is backed by a large plug  858  inserted into a threaded hole  859 . The position of plug  858  may be adjusted and locked with a nut  860  in order to set anchor pin  853  to a desirable position with respect to anchor bracket  826  and strut  841 . 
         [0084]    The embodiment of the invention shown in  FIG. 28  shares features with the embodiment for  FIG. 27 . Specifically, a load sensor  821  is placed behind a cushion element  872  inside support dish  874  with a raised retaining wall  873 . Cushion element  872  is preferably made of rubber. Alternatively, cushion element  872  may be made of metal in the form of a spring such as a disk or a conical spring. Under a no load condition, the surface of cushion element  872  is in contact with that of a pin  875 , while the end of retaining wall  873  is away from the surface of pin  875 . When the anchor load is below a predetermined level, the entire load is transmitted to sensor  821  through the elastic deformation of cushion element  872 . As the anchor load increases, cushion element  872  becomes compressed. Once the surface level of cushion element  872  becomes flush with the end of retaining wall  873 , retaining wall  873  starts supporting the load exerted on pin  875 , limiting the load on sensor  821 . 
         [0085]    As shown in  FIG. 29 , cushion element stiffness determines where the sensor output starts leveling off at  876 . This embodiment of the invention enables the sensor performance to be targeted for a specific load range, maximizing a measurement resolution  877 . In addition, sensor output voltage at limiting load level  876  and at zero load level  878  can be used to auto-calibrate sensor  821  for enabling absolute load measurements. That is when the sensor output reaches its maximum plateau, a transfer function between sensor output voltage and load level can be mapped based on two point calibration. This feature is extremely useful, especially if sensor characteristics drift over time or vary under different operating conditions. This load-limiting feature also protects the sensor from overloading, preventing its failure. 
         [0086]    Based on the above, it should be readily apparent that the present invention provides numerous advantages over prior friction element control during a torque phase of gear-ratio changing. The preferred embodiments provide a consistent output shaft torque profile for a powertrain system with a step-ratio automatic transmission system during a synchronous friction element-to-friction element upshift, which reduces shift shock. Also, there is a significant reduction in shift feel variability for a powertrain system with a step-ratio automatic transmission system during a synchronous friction element-to-friction element upshift. The preferred embodiments of the invention permit the use of either absolute or relative load levels which are directly measured or estimated. The use of a relative load profile, instead of absolute load levels, eliminates the need of full-sensor calibration, while the use of a relative load profile only requires one point sensor calibration that corresponds to zero load level and improves robustness against sensor drift over time. The preferred embodiments also provide for reduced output shaft torque oscillation at the beginning of the inertia phase due to the release of the off-going friction element at or near the ideal release timing where its load level is zero or close to zero and robustness against the variability of off-going friction element breakaway friction coefficient by means of a quick release of the off-going friction element at the ideal synchronization timing. 
         [0087]    Further advantages include a consistent output shaft torque profile and significant reduction in shift feel variability for a powertrain system with a step-ratio system during a torque phase of a synchronous friction element-to-friction element upshift and during a torque phase of a non-synchronous upshift with an overrunning coupling element. Further, the system provides robustness against the variability of off-going friction element breakaway friction coefficient by means of a quick release of an off-going friction element at an ideal synchronization timing during a synchronous shift and against the variability of a friction element actuation system for both synchronous and non-synchronous shifts. 
         [0088]    A clutch load sensor assembly provides a relative measure of torque load exerted to the clutch while it is engaged. A band brake load sensor assembly provides a relative measure of engagement torque (brake torque) and its derivative during an engagement process while a band slips against a drum and a relative measure of torque load exerted onto a band and a drum while the band is securely engaged to the drum without slippage. Sensor output may be calibrated with respect to a command signal to a band servo actuator while torque load is zero. Use of a protective cover in the sensor assembly prevents a direct contact between a load sensing material and the pin for reduced sensor material wear; and shields the sensor from hostile conditions that include heat and electro-chemical interaction, such as with transmission oil. 
         [0089]    Although described with reference to preferred embodiments of the invention, it should be understood that various changes and/or modifications can be made to the invention without departing from the spirit thereof. For example, the invention could be extended to a double-wrap band brake system. In general, the invention is only intended to be limited by the scope of the following claims.