Abstract:
A method of calibrating an equilibrium position of an actuator driven by an electric motor with the assistance of a compensating spring, wherein the actuator moves a mechanism and the mechanism opposes the movement with an elastic force, and wherein in the equilibrium position the compensating spring counterbalances the elastic force, includes the following steps: 
     applying a large-amplitude, high-frequency alternating position signal to energize the electric motor, wherein the position signal is biased so that alternating extremes of the position signal straddle the equilibrium position, 
     measuring an actual dynamically variable position of the actuator by means of a position sensor, and 
     determining the equilibrium position as that position where the actuator settles.

Description:
CROSS-REFERENCE TO RELATED APPLICATION 
     This application is a continuation of International Patent Application Ser. No. PCT/DE02/01300, filed Apr. 9, 2002, published in German, which is hereby incorporated by reference in its entirety. 
    
    
     BACKGROUND OF THE INVENTION 
     The invention relates to a method of calibrating an equilibrium position of an electrically operated actuator in a motor vehicle clutch or in a gear-selecting mechanism, where the electric motor driving the actuator is assisted by a compensating spring. 
     Electrically operated actuators of the kind that the present invention relates to are disclosed for example in GB 2325036, GB 2313885, or GB 2309761. The disclosures of these documents are expressly incorporated herein by reference. Such actuators include an electric motor, which drives a hydraulic master cylinder that communicates with a slave cylinder which, in turn, actuates a vehicle clutch or a gear-ratio selecting mechanism. The electric motor in these actuators can work through an appropriate gear mechanism such as a worm-drive mechanism to drive a push rod. One end of the push rod is connected to a crank that is tied to the gear wheel of the worm-drive mechanism, while the other end of the push rod is connected to a piston that slides in a master cylinder, so that the rotary movement of the gear wheel is converted into a linear movement of the piston. The electric motor, the gear mechanism, and the master cylinder are preferably arranged together in a common housing. 
     The master cylinder of the electric actuator described above is typically connected to a slave cylinder of a clutch. When pressure is applied to the slave cylinder, a clutch release fork is actuated which acts on a clutch release bearing to generate a force that disengages the clutch. The release bearing typically acts on a diaphragm spring which in its normal (i.e., non-actuated) state holds the discs of the clutch in frictional engagement. When the diaphragm spring is depressed by the release bearing, the clutch discs move apart, so that the clutch becomes disengaged. The force generated by the electric motor therefore has to be large enough to depress the diaphragm spring to an extent that is sufficient to release the engagement of the clutch. The force required to disengage the clutch is typically of the order of 450 N. 
     In order to reduce the size of the electric motor required for such actuators, it has been proposed to include a compensating spring in the electric actuator, so that the compensating spring counteracts the opposing force of the diaphragm spring. This may be accomplished for example with an arrangement where the compensating spring is fully compressed in the completely engaged state of the clutch, whereby-the compensating spring generates a force of, e.g., 250 N in the electric actuator in the direction of disengagement of the clutch. In the process of disengaging the clutch, the initial amount of force to depress the diaphragm spring will now be supplied by the compensating spring. Although the force of the compensating spring decreases over the phase where the compensating spring and the diaphragm spring counteract each other, the electric motor only needs to generate a force of the order of 200 N to fully disengage the clutch. Thus, the requirement for the electric motor to produce 450 to 500 N, the amount of force that would be required without the compensating spring, can be reduced to 250 to 300 N through the use of a compensating spring. 
     In electric actuators of the type disclosed in the aforementioned references, a high level of static friction between the worm and the gear wheel provides a self-holding effect. However, in the interest of optimizing the efficiency of the actuator, it may be desirable if the internal static friction of the actuator is smaller than would be required to keep the actuator immobilized. In this case, it is possible that the force exerted by the diaphragm spring in the disengaged state of the clutch will force the actuator back, or that the force exerted by the compensating spring in the engaged state of the clutch will push the actuator forward, i.e., in the direction of disengagement. If this causes the actual position of the actuator to deviate from the required position by more than a predetermined tolerance, the controller will reactivate the actuator motor. 
     The German Patent Application DE 10062456.1, which is hereby incorporated by reference in the present disclosure, proposes the following concept to counteract the force of the diaphragm spring which could cause an unintended re-engagement of the clutch: When the actuator is in its rest position, a voltage of typically 7 percent of the maximum PWM voltage (Pulse Width Modulation voltage) is applied to the electric motor in the direction where the motor will support the compensating spring to counteract the force of the diaphragm spring. This voltage generates a force in the actuator which prevents the actuator from moving backward because of the reactive force of the diaphragm spring. However, in order to prevent a forward movement of the actuator in the engaged state of the clutch, the 7 percent voltage is applied only when the force generated by the compensating spring is smaller than the force produced by the diaphragm spring. 
     Consequently, this system requires that the equilibrium position is known where the forces of the compensating spring and the diaphragm spring keep each other in balance. In clutches with self-adjusters that adjust the position of the pressure plate to compensate for the wear on the friction surfaces, the equilibrium position remains substantially unchanged over the life of the clutch. In clutches of this type, the equilibrium position may be precalibrated. However, in a clutch without the self-adjusting feature, the equilibrium position changes significantly as the components of the clutch wear down in use. In clutches of this latter type, the concept of energizing the electric motor with 7 percent of the maximum PWM voltage cannot be used, and a much more complex adaptive strategy with a continuous current supply has to be used. 
     OBJECT OF THE INVENTION 
     The present invention therefore has the objective to provide a method of calibrating the equilibrium position where the forces of the diaphragm spring and the compensating spring counterbalance each other, so that an initial calibration can be performed at the end of the assembly line after the system has been installed in the vehicle, and recalibrations can be performed at regular time intervals, for example with the routine maintenance services, so that the concept of using the assistance of the electric motor at 7 percent of the maximum PWM voltage can be used in clutches without self-adjusters. 
     SUMMARY OF THE INVENTION 
     A method according to the present invention serves to calibrate the equilibrium position of a clutch actuator that is driven by an electric motor and assisted by a compensating spring, where the latter is used to compensate an opposing elastic force of the mechanism that is operated by the actuator. The method includes the steps of: 
     applying a large-amplitude, high-frequency alternating position signal to energize the electric motor, where the position signal alternates between extremes that correspond to actuator positions spanning across the equilibrium position, 
     measuring the actual position of the actuator by means of a position sensor associated with the actuator, and 
     determining the equilibrium position as the position into which the actuator settles as long as the alternating position is applied. 
     Due to the combined force/displacement characteristic of the compensating spring and the counteracting elastic force of the mechanism, the actuator with the assistance of the compensating spring moves rapidly into the equilibrium position, while the speed of the actuator movement is significantly reduced when the motor is working against the opposing elastic force of the mechanism. Likewise, when the motor is energized in the reverse direction, the actuator returns rapidly to the equilibrium position and then continues slowly beyond the equilibrium position as the compensating spring is being compressed. As a result, the range of the movement of the actuator is centered on the equilibrium position. The higher the frequency of the alternating signal, the shorter the distance by which the actuator moves in either direction beyond the equilibrium position. Consequently, the equilibrium position can be determined more precisely by using a higher frequency. According to a preferred embodiment, the frequency of the alternating position signal is 25 Hz or higher. Particularly preferred is a position signal with a frequency of about 50 Hz. 
     The closer the midpoint of the alternating position signal is to the equilibrium position, the more accurate will be the result of the equilibrium determination. Consequently, it is possible to use an iterative technique of successive determinations of the equilibrium position where in each iteration the midpoint of the alternating position signal is positioned on the equilibrium position determined in the previous step until the equilibrium position coincides with the midpoint of the alternating position signal. Initially, the midpoint of the alternating position signal may be set so that it coincides with a theoretical equilibrium position calculated from the design characteristics of the actuator and mechanism or with the last known equilibrium position at which the system of actuator and mechanism was recalibrated. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     An embodiment of the invention will be described below with reference to the attached drawings, wherein: 
     FIG. 1 schematically represents a vehicle with a clutch actuator driven by an electric motor; 
     FIG. 2 gives a more detailed view of the clutch actuator driven by the electric motor and the clutch of the vehicle of FIG. 1; 
     FIG. 3 shows graphic plots of force vs. actuator travel for the clutch diaphragm spring and for the actuator compensation spring illustrated in FIG. 2; 
     FIG. 4 shows a graph of the actual position taken by the actuator when alternating position signals of high amplitude are applied to the actuator at varying frequencies; 
     FIG. 5 shows a graph of the actual position taken by the actuator when an alternating high-frequency position signal is applied to the actuator at varying amplitudes; and 
     FIG. 6 shows a graph of the actual position taken by the actuator when an alternating position signal of large amplitude and high frequency is applied to the actuator with a varying midpoint of the alternating signal moving progressively nearer to the equilibrium position of the actuator. 
    
    
     DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS 
     The vehicle  10  illustrated in FIG. 1 has an internal combustion engine  12  that is connected to a shiftable transmission  14  by way of a clutch  16 . The transmission  14  is connected by way of the drive shaft  18  and the rear axle  20  to the driven rear wheels  22  of the vehicle  10 . 
     A gear-shifting lever  24  is connected mechanically to the transmission  14  in a conventional manner for a manual selection of the gear ratio. The engagement and disengagement of the clutch  16  is controlled by a clutch actuator  30  that is driven by an electric motor. A sensor  32  at the gear-shifting lever  24  sends a signal to a control unit  34  which directs the clutch actuator  30  to disengage and re-engage the clutch  16  as needed when a gear shift is initiated by a movement of the gear-shifting lever  24 . 
     As shown in FIG. 2, the clutch actuator  30  has a direct current motor  40  (for example a brushless, electronically commutated DC motor) mounted on the housing  42 . The electric motor  40  is either directly connected to a push rod  48  or through a fixed-ratio gear mechanism with a worm  44  and a worm gear wheel  46  connected to a crank  50  driving the push rod  48 , so that the rotary movement of the worm gear wheel  46  is converted into a linear movement of the pushrod  48 . Instead of the worm gear drive  44 ,  46 , one could also use other arrangements for transmitting the rotation of the electric motor  40  to the push rod  48 . Possible alternatives include, e.g., a planetary gear mechanism, a spur gear mechanism, a cam disk mechanism, or a spindle drive mechanism. 
     The free end of the pushrod  48  is connected to a piston  54  of a hydraulic master cylinder  52  that is formed integrally in the housing  42  of the electric motor. The push rod  4  is connected to the piston  52  through a ball joint  56  that is snap-fitted into a spherical cavity  58  formed on the axis of the piston  54 . A helical compression spring  60  acts between the housing  42  and a ring collar  62  on the push rod  48  to urge the push rod  48  towards the closed end  64  of the master cylinder  52 . A hydraulic port  66  is arranged at the end  64  of the master cylinder  52 . 
     A position sensor  68  in the form of a linear potentiometer is arranged at the push rod  48  to deliver a signal indicating the position of the push rod  48 . 
     The port  66  of the master cylinder  52  is connected by way of a hydraulic conduit  72  to a slave cylinder  70  of the clutch  16 . The slave cylinder  70  is connected to a clutch release fork which acts on the clutch release bearing  76  to move the clutch  16  into and out of engagement in a conventional manner. 
     The clutch  16  includes a friction disc  80  that is connected to the input shaft  82  of the transmission  14 . The friction disc  80  is mounted coaxially between a flywheel  84  that is connected to the engine and a pressure plate  86  that is connected to the flywheel  84  by way of a clutch housing  88  so that the pressure plate  86  is constrained to share the rotation of the flywheel  84  but is axially movable relative to the flywheel  84 . The pressure plate  86  is pushed towards the flywheel  84  by a diaphragm spring  90 , so that the friction disc  80  is clamped between the pressure plate and the flywheel in order to transmit torque between the engine  12  and the transmission  14 . The clutch  16  is released (taken out of engagement) when a force in the direction towards the flywheel is applied through the release fork  74  and the release bearing  76  to a radially inner portion of the diaphragm spring. 
     Instead of a hydraulic link between the clutch actuator  30  and the release fork  74 , one could also use a pneumatic or mechanical connection where the push rod  48  could act, e.g., either directly or through a mechanical linkage or cable on the release fork. 
     When the clutch  16  is fully engaged, the clutch actuator  30  will be in the position shown in FIG. 2, where the push rod  48  is moved hard to the left, so that the piston  54  of the master cylinder  52  is at the limit of its movement on the far side from the end  64  of the master cylinder and the spring  60  is fully compressed. 
     When the electric motor  40  is energized to disengage the clutch  16 , the push rod  48  is moved to the right, so that the piston  54  moves towards the end  64  of the master cylinder  52 . As a result, fluid is displaced from the master cylinder  52  to the slave cylinder  70 , causing the piston of the slave cylinder  70  to exert a force on the release fork  74 , so that the release bearing  76  will move in the direction towards the flywheel  84  and exert a force on the inner circumference of the diaphragm spring  90 . This reduces the force exerted by the diaphragm spring on the pressure plate  86  and thus releases the clamping pressure on the friction disc  80 . As shown in FIG. 3, the force generated by the fully compressed spring  60  is initially larger than the reactive force of the diaphragm spring  90 . Consequently, the movement of the push rod  48  and the piston  54  are driven initially by the force that is generated by the spring  60 , so that the load on the electric motor  40  will be very small, as the motor has to deliver only enough power to permit the movement of the push rod  48  under the action of the spring  60 . 
     In the equilibrium position where the force generated by the spring  60  is in balance with the reactive force of the diaphragm spring, the force required to continue the disengagement of the clutch  16  will be supplied the electric motor  42 . As shown FIG. 3, the force applied to the diaphragm spring to fully disengage the clutch  16  is typically of the order of 430 N. The rating of the compensating spring  60  is such that a force of the order of 250 N is generated when the clutch is fully engaged and the reactive force of the diaphragm spring is substantially zero. From the fully engaged position to the equilibrium position, the force of the compensating spring decreases to about 210 N. Consequently, the electric motor  40  needs to be capable of generating a force of sufficient magnitude to depress the diaphragm spring  90  from the equilibrium position to the fully disengaged position of the clutch, i.e., the difference from 210 N to 430 N, and to fully compress the spring  60  from the equilibrium position to the completely engaged position of the clutch. Thus, an electric motor  40  capable of generating a force of 220 N to 250 N will be adequate instead of a motor  40  capable of more than 430 N which would be required in the absence of the compensating spring. As illustrated in the graph of FIG. 3, the equilibrium position BP dis  in a disengagement phase of the clutch  16  differs from the equilibrium position BP eng  in an engagement phase because of the hysteresis of the diaphragm spring  90 . The equilibrium positions BP dis  and BP eng  are at 4.5 mm and 5.7 mm, respectively, measured from the fully engaged position of the clutch. 
     In electric actuators of the type disclosed herein, when the electric motor is switched off, the force exerted by the compensating spring  60  in the fully engaged clutch position or the force of the diaphragm spring  90  in the fully disengaged position will cause the electric motor to turn back, so that the actuator  30  moves out of its required position, unless there is a significant amount of friction in the mechanism. If during a gear shift, the actual position of the actuator  30  differs from the required position by more than a predetermined amount, the electric motor  40  is energized again to return the actuator to the required position. In order to avoid this condition when the actuator  30  is at rest, it has been proposed to apply a current of sufficient magnitude to the electric motor  40  to hold the motor  40  in position, but not large enough to cause the actuator  30  to move. Typically, a voltage of 7% of the full PWM voltage is applied to the electric motor  40  for this purpose. Preferably, the value of 7 percent represents the percentage of “on” time (also called duty cycle) of the PWM voltage, in which case the pulse height is constant, but alternatively it is also possible to use a variable DC voltage. The method of applying a PWM voltage level of 7% to the electric motor is used only if the actuator  30  is in its rest position during a gear shift, where the actuator  30  is between the equilibrium position and the fully disengaged position of the clutch  16 . To use this method, it is therefore necessary to know the exact equilibrium position of the actuator/clutch system. 
     As the equilibrium position of the actuator/clutch system changes as a result of wear on the friction surfaces of the clutch  16 , it will be necessary to calibrate the actuator/clutch system at some points in time during the life of the vehicle. 
     FIG. 4 illustrates the effect of applying an alternating position signal to the electric motor  40  of the actuator  30  at different frequencies, i.e., 5 Hz, 10 Hz, 25 Hz, and 50 Hz. The position signal has an amplitude of 6 mm and a midpoint MP of the oscillation lying in the vicinity of the predicted equilibrium point of the system composed of the actuator  30  and clutch  16 . If the position signal is applied when the clutch  16  is in the fully engaged position, the electric motor  40  with the assistance of the compensating spring  60  will move the actuator  30  rapidly to the equilibrium position. At this point, the actuator  30  continues to move at a slower speed because the motor itself will now have to contribute a part of the force to overcome the reactive force of the diaphragm spring  90 . Consequently, even at a frequency of 5 Hz the actuator will not completely follow the position signal to the required position before the position signal is reversed. Upon reversal of the position signal, the electric motor  40  with the assistance of the diaphragm spring  90  rapidly returns the actuator  30  to the equilibrium position and then continues to move at a slower speed because the motor will now have to contribute a part of the force to compress the compensating spring  60 . The higher the frequency of the position signal, the less the actuator will overshoot the equilibrium position, and at the frequencies of 25 Hz and 50 Hz shown in FIG. 4, the actuator settles at the equilibrium position. 
     FIG. 5 shows the behavior of an actuator  30  at alternating position signals with a uniform frequency of 50 Hz and different amplitudes. With a smaller amplitude of the position signals, the actuator  30  settles near the midpoint of the alternating position signal, and the larger the amplitude, the closer the actuator  30  will settle to a position corresponding to the equilibrium state of the actuator/clutch system. 
     Finally, FIG. 6 illustrates the effect of varying the midpoint of a position signal with a frequency of 50 Hz and an amplitude of 6 mm. As the graph shows, the accuracy of determining the equilibrium position is improved as the midpoint of the position signal approaches the equilibrium position. 
     According to an embodiment of the invention, the equilibrium position of an actuator  30  is calibrated by applying an alternating position signal to the actuator with a frequency of 50 Hz and an amplitude of 6 mm, while the actual position of the actuator  30  is determined, e.g., by a position sensor  68 . Initially, the midpoint of the alternating position signal is set to coincide with a calculated or previously determined equilibrium position. 
     The foregoing cycle is repeated with the midpoint repositioned to coincide with the equilibrium position that has been newly determined in the foregoing cycle. Further iterations are performed until the detected equilibrium position coincides with the midpoint of the position signal. 
     Various modifications are possible without departing from the scope of the invention. Although an alternating position signal of 50 Hz is used in the foregoing preferred embodiment of the invention, one could use position signals alternating at some other frequency above 25 Hz. Furthermore, while an amplitude of 6 mm is used in the preceding embodiment, it is self-evident that the amplitude used in an actual use of the invention depends on the travel range of the actuator between the fully engaged and fully disengaged positions of the clutch and on the location of the equilibrium position relative to the endpoints of the travel range. 
     While the invention has been described with reference to a clutch actuator, it is equally applicable to other electronically controlled, motor-driven actuators that are equipped with a compensating spring and are used to actuate the movement of a mechanism that produces an elastic reactive force. Examples of other possible uses of the inventive concept include actuators used in gear selector mechanisms. Actuators according to the present invention may also be used in automatic or semi-automatic transmissions.