Abstract:
A transmission clutch control method includes defining a transfer function relating clutch torque to a control signal under transmission operating conditions; determining a target clutch torque for current operating conditions; determining the target control signal from the transfer function to produce target torque at the clutch; correcting clutch torque on the basis of a difference between the target clutch torque and the actual torque at the clutch by adjusting the control signal; calculating actual clutch torque with reference to transmission input torque and transmission output torque; computing a clutch torque error as a difference between calculated clutch torque and the target clutch torque; and repetitively adjusting the transfer function on the basis of the clutch torque error.

Description:
BACKGROUND OF THE INVENTION 
       [0001]    1. Field of the Invention 
         [0002]    This invention relates generally to a method for controlling a transmission clutch during a clutch control event. 
         [0003]    2. Description of the Prior Art 
         [0004]    A clutch transfer function is defined as a relationship between clutch torque transmitted through frictional interfaces and a clutch actuator control signal, which may be electric current to an electric clutch actuator, hydro-electric actuator pressure, clutch piston position or other variables. 
         [0005]    Clutch torque is affected by various uncontrolled noise factors, such as actuator system variability and thermal sensitivity of hydrodynamic torque, during the clutch actuation process. For example, for a hydraulically-actuated wet clutch, clutch torque may be relatively linear with respect to a given control signal profile at a certain transmission fluid temperature. However, at a different transmission fluid temperature, clutch torque may exhibit significant nonlinearity and its value may be considerably different for the same given control signal. 
         [0006]    The clutch transfer function varies from unit to unit due to hardware variability and also changes over the life of a vehicle due to degradation and wear of system components, including friction material, transmission fluid additives, hydraulic valves, etc. In practice, it is not easily possible to capture the changing transfer function behaviors in volume production applications using prior art technologies. Accordingly, a conventional clutch control methodology primarily relies on a clutch transfer function, which may be obtained a priori based on limited vehicle tests or bench tests. 
         [0007]    The clutch transfer function may be adjusted based on indirect observations such as increased transmission shift duration. However, such an approach cannot directly and accurately map a detailed functional relationship between clutch torque and actuator control signal under all drive conditions. 
       SUMMARY OF THE INVENTION 
       [0008]    A transmission clutch control method includes defining a transfer function relating clutch torque to a control signal under transmission operating conditions; determining a target clutch torque for current operating conditions; determining the control signal from the transfer function to produce target torque at the clutch; correcting clutch torque on the basis of a difference between the target clutch torque and the actual torque at the clutch by adjusting the control signal; calculating actual clutch torque with reference to transmission input torque and transmission output torque; computing a clutch torque error as a difference between calculated clutch torque and the target clutch torque; and repetitively adjusting the transfer function on the basis of the clutch torque error. 
         [0009]    The clutch transfer function constructed through the method provides valuable tool for controlling clutch behaviors, collectively accounting for all the noise factors which are difficult to characterize individually. The method also provides a systematic means to account for unit-to-unit variability or a characteristics change over time. The transfer function can be utilized to back calculate the control signal required to achieve a desired clutch torque. 
         [0010]    The scope of applicability of the preferred embodiment will become apparent from the following detailed description, claims and drawings. It should be understood, that the description and specific examples, although indicating preferred embodiments of the invention, are given by way of illustration only. Various changes and modifications to the described embodiments and examples will become apparent to those skilled in the art. 
     
    
     
       DESCRIPTION OF THE DRAWINGS 
         [0011]    The invention will be more readily understood by reference to the following description, taken with the accompanying drawings, in which: 
           [0012]      FIG. 1  is a schematic diagram of a gearing arrangement for an automatic transmission; 
           [0013]      FIG. 2  is chart showing the engaged and disengaged state of each of the clutches and brakes of the transmission of  FIG. 1  for each of the forward gears and reverse gear; 
           [0014]      FIG. 3  is a graph of a general process of a synchronous friction element-to-friction element upshift event from a low gear to a higher gear for the transmission of  FIG. 1 ; 
           [0015]      FIG. 4  is an example of clutch behaviors as a function of transmission oil temperature; 
           [0016]      FIG. 5  is a graph illustration of varying shift behaviors under different operating conditions; 
           [0017]      FIG. 6  is logic flow diagram of a method for constructing a transmission friction element transfer function; 
           [0018]      FIG. 7  illustrates with reference to a graph the method for updating the clutch transfer function; 
           [0019]      FIG. 8  is a logic flow diagram of a method for controlling a transmission clutch using the updated, adapted transfer function as described with reference to  FIG. 6 ; and 
           [0020]      FIG. 9  is a graph that illustrates the variation of the control signal and clutch torque with time during a clutch event controlled according to the method. 
       
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
       [0021]    Transmission  2  schematically illustrated in  FIG. 1  is an example of a multiple-ratio transmission having a controller  4 , wherein speed ratio changes are controlled by friction elements acting on individual gear elements. Engine torque from vehicle engine  5  is carried to torque input element  10  of hydrokinetic torque converter  12 . An impeller  14  of torque converter  12  develops turbine torque on a turbine  16 . Turbine torque is transmitted on a turbine shaft, which is also transmission input shaft  18 . Transmission  2  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  and meshing with sun gear S 1  and ring gear R 1 . Transmission input shaft  18  is drivably connected to ring gear R 1 . 
         [0022]    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 mesh with short planetary pinions P 3  and torque output ring gear R 3 . Short planetary pinions P 3  further mesh with 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). 
         [0023]    Gearset  20  is an underdrive ratio gearset arranged in series with respect to  15  compound gearset  21 . Torque converter  12  preferably includes a lockup or torque converter bypass clutch  25 , which directly connects transmission input shaft  18  to engine  5  after a torque converter torque multiplication mode is completed and a hydrokinetic coupling mode begins. 
         [0024]      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 . 
         [0025]    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 clutch A. During operation in the third ratio, fifth ratio and reverse, direct clutch 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 clutch E connects turbine shaft  18  to compound carrier  23  through shaft  28 . Friction brake C acts as a torsion reaction brake for sun gear S 2  during operation in the second and sixth forward driving ratios. During operation of the third forward driving ratio, direct friction brake B is applied together with forward friction clutch A. The elements of gearset  21  then are locked together to effect a direct driving connection between shaft  28  and output shaft  26 . 
         [0026]    Forward friction clutch A is connected through torque transfer element  29 , torque transfer element  29  to direct friction clutch B during forward drive. 
         [0027]    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 clutch B concurrently. 
         [0028]    For the purpose of illustrating one example of a synchronous ratio 1-2 upshift produced by transmission  2 , it will be assumed that the upshift occurs between the first ratio and the second ratio. During such a 1-2 upshift, forward friction clutch A remains engaged, friction brake C starts in the disengaged state before the upshift and is engaged during the upshift, and low/reverse friction brake D starts in the engaged position before the upshift and is released during the upshift. Friction clutch B and overdrive friction clutch E remain disengaged throughout the upshift. 
         [0029]    Friction brake C is referred to as the oncoming element, oncoming clutch or oncoming element (OCE) during the synchronous 1-2 upshift. Friction brake D is referred to as the offgoing element, offgoing clutch or offgoing element (OGE). 
         [0030]      FIG. 3  depicts a general process of a synchronous friction element-to-friction element upshift from a low gear to a higher gear for the automatic transmission of  FIG. 1 . For example, the process has been described in relation to a 1-2 synchronous ratio upshift 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. 
         [0031]    The upshift 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 . 
         [0032]    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  drop 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 oncoming friction element C completes engagement or when its slip speed becomes zero at  43 , inertia phase  33  ends, completing the shift event. 
         [0033]    A clutch transfer function is defined as a relationship between clutch toque (Tcl), transmitted through frictional interfaces and a clutch actuator control signal (Ucom), which may be electric current to an electric clutch actuator, hydro-electric actuator pressure, clutch piston position or other variables. Clutch torque is affected by various uncontrolled noise factors, such as actuator system variability and thermal sensitivity of hydrodynamic torque, during the clutch actuation process. For example, for a hydraulically-actuated wet clutch system illustrated in  FIG. 4 , clutch torque Tcl may be relatively linear with respect to a control signal Ucom at Toil=200° F.  401 , where Ucom may be a commanded pressure of hydraulic transmission fluid. However, at 30° F., clutch torque may exhibit significant nonlinearity  402  and its value may be considerably lower at a given commanded signal level  403 . The clutch transfer function or the relationship between Tcl and Ucom may vary from unit to unit and also change over the life of a vehicular system due to the degradation and wear of system components, including friction material, transmission fluid additives, hydraulic valves, etc. In practice, it is not easily possible to capture the changing transfer function behaviors in volume production applications using prior art technologies. Accordingly, a conventional clutch control methodology primarily relies on a clutch transfer function, which may be obtained a priori based on limited vehicle tests or bench tests. Clutch transfer function may be adjusted based on indirect observations such as increased transmission shift duration. However, such an approach cannot accurately map a detailed functional relationship between Tcl and Ucom under all drive conditions. 
         [0034]    A clutch system, which includes an actuator and frictional elements, exhibits widely varying behaviors under different operating conditions. In the case of a hydraulic actuator system, its performance is very sensitive to hydraulic fluid conditions inside the hydraulic circuits. A wet clutch pack whose frictional interface is lubricated with transmission fluid also exhibits sensitivity to a number of factors such as slip velocity, fluid additives, oil temperature, etc. In the case of a dry clutch system, it is known that its frictional torque is sensitive to interface temperature conditions. Accordingly, even if the same actuator force profile is commanded, torque transmitted through frictional interfaces may differ significantly. Clutch torque variability generally degrades transmission output torque consistency or shift quality during a shift event. 
         [0035]    For example,  FIG. 5  shows varying output shaft torque behaviors under two temperature conditions, resulting in inconsistent shift feel that may be negatively perceived by vehicle occupants. The transmission output shaft torque has a pronounced peak  501  at the beginning of the shift at cold temperature while the peak moves toward the end  502  at high temperature, even though the commanded pressure profiles are nearly identical  503 . The performance of the clutch actuator and frictional components changes during the initial break-in phase and over time due to component wear, affecting shift quality. The behaviors of the clutch system also vary from unit to unit due to manufacturing and assembly variability. There is a need to accurately characterize a clutch transfer function or a relationship between a commanded control signal and actual torque transmitted across clutch hardware in order to accurately control clutch torque for improved shift quality, throughout the life of a transmission system. 
         [0036]      FIG. 6  illustrates is a logic flow diagram  600  of a method for constructing and adaptively improving clutch transfer function. Clutch control starts at step  601 . Torque transmitted on transmission input shaft  18  and output shaft  24  is preferably determined at steps  602  and  603  according to the methods described in U.S. Patent Application Publication No. U.S. 2010/0318269 at paragraphs [0050] through [0058] using driveline torque sensing or other means. The entire disclosure of U.S. Publication No. 2010/0318269 is incorporated herein by reference. 
         [0037]    At step  606  clutch torque T*cl of either the oncoming friction control element or offgoing friction control element is determined based on Tin and Tout according to the methodology described with reference to Eq. (3) and Eq. (6) of U.S. Patent Publication No. U.S. 2010/0318269, respectively. 
         [0038]    A clutch transfer function F(.)is defined in a functional form as Tcl=F(Ucom, Xk), wherein Ucom is a commanded clutch actuator control signal determined at step  604 , and Xk, determined at step  605 , are the corresponding transmission operating conditions. 
         [0039]    F(.) may be defined using any suitable base function, such as a multi-variable polynomial with multiple coefficients or a neural network. Alternatively, F(.) may be defined as a look-up table with multiple dimensions for Xk, representing key operating conditions such as transmission oil temperature. F(.) is stored in a control system memory and utilized to compute Tcl under given operating conditions Xk at step  607 . 
         [0040]    At each time interval or time step (ti) during a clutch control event, the coefficients in F(.) or lookup table entries from which F(.) is determined are updated at step  609 , based on T*cl, Tcl and ΔTcl, for a given Ucom and Xk. ΔTcl, is calculated at step  608 . A conventional optimization method, such as a least square optimization method, may be employed for updating the coefficients of F(.) or the lookup table entries from which F(.) is determined. The optimization method reduces the magnitude of the difference ΔTcl between the clutch torque magnitude of T*cl computed from Tin and Tout at step  606  and the clutch torque magnitude computed from F(.) at step  607 . 
         [0041]    Each change to F(.) that occurs during successive executions of the control method  600  is stored in the control system memory at step  610 . 
         [0042]    A change or a change rate of the transfer function F(.) over time can be computed and stored in a powertrain control module (PCM) at step  616 . 
         [0043]    At step  617  a test is performed to determine whether the level of clutch system performance degradation is sufficient to warrant issuing at step  618  an early service warning before a system failure occurs. 
         [0044]    At step  611  a test is performed to determine whether to end at step  613  execution of the clutch control. 
         [0045]      FIG. 7  illustrates graphically the method described with reference to  FIG. 6  for updating clutch transfer function. Under the given powertrain or transmission operating condition Xk, Tcl  701  is computed from the transfer function F(Ucom, Xk)  702  stored in PCM for the control signal Ucom=U1  703  at time step (ti). T*cl  704  is obtained from Tin and Tout, independently from the transfer function F(.) by employing the methodologies described in U.S. Publication No. 2010/0318269. 
         [0046]    As shown in  FIG. 7 , T*cl(ti)  704  is larger than Tcl(U1, Xk)  701 . Following the systematic methodology described with reference to  FIG. 6 , the coefficients of F(Ucom, Xk) are adjusted to move the function upward to  705 , thereby reducing ΔTcl, based on Tcl and T*cl. 
         [0047]      FIG. 8  is a logic flow diagram of a method for controlling a transmission clutch using the updated, adapted transfer function as described with reference to  FIGS. 6 and 7 . The clutch control event may include vehicle launch, transmission upshifting, downshifting or any other drivability control actions. 
         [0048]    Clutch control starts at  801 , where all the relevant powertrain and transmission variables are initialized. 
         [0049]    At  802 , clutch operating conditions Xk are determined based on measured data available in the transmission system  2 . 
         [0050]    At  803 , a target clutch torque profile Ttar is determined for current clutch operating conditions Xk for a given drivability control event. 
         [0051]    At step  804 , based on the inverse of the clutch transfer function F(.) −1  stored at step  610 , control signal profile Ucom is determined to realize the target magnitude of clutch torque Ttar for the current operating conditions Xk. 
         [0052]    At step  805 , Ttar is corrected based on ΔT as a feedback signal, where G is a control gain. Note that ΔT is set to 0 for i=0. 
         [0053]    At step  806 , Ucom is commanded for clutch control. 
         [0054]    At step  807 , T*cl is computed based on Tin and Tout. 
         [0055]    At step  808 , a clutch torque error ΔT is computed. 
         [0056]    If end of clutch control is reached, then the control process ends at step  811 . 
         [0057]    If further clutch control is required, the iterative process  800  returns from step  809  to step  802  after incrementing (i) by 1 before re-executing step  802 . 
         [0058]      FIG. 9  is a graph that illustrates the variation of the control signal and clutch torque with time during a clutch event controlled according to the method illustrated with reference to  FIG. 8 . At the first time step (ti)  901 , actual clutch torque T*cl  902 , which is determined from Tin and Tout, is lower than the target magnitude of clutch torque Ttar (ti)  903 , which is a point on the target profile Ttar(t)  904 . Ucom is corrected to a higher magnitude  905  for the time steps that follow ti, based on Ttar and ΔT and inverse clutch transfer function F −1  to reduce ΔT. 
         [0059]    The corrected Ucom  905  results in T*cl  906 , which closely follows Ttar  904 . In comparison,  FIG. 9  illustrates T*cl  907 , which differs significantly from Ttar  904  when neither clutch transfer function F(.) nor torque feedback ΔT is available to correct the control signal Ucom  908 . 
         [0060]    In accordance with the provisions of the patent statutes, the preferred embodiment has been described. However, it should be noted that the alternate embodiments can be practiced otherwise than as specifically illustrated and described.