DE10316458B4 - Transmission control and method for performing a Reibwertadaption - Google Patents

Abstract

A method for carrying out a friction value adaptation in a clutch, in particular an automated clutch, a transmission, wherein the coefficient of friction is changed depending on at least one clutch size, wherein the coefficient of friction is changed by a predetermined value, if the slip time is too short, characterized in that the coefficient of friction is increased absolutely by a certain amount and / or relatively by a certain percentage, if the friction value adaptation is not carried out due to a short slip duration in a shift and / or the slip time for the specific shift is too short for a given speed jump.

Description

The present invention relates to a method with the features according to the preamble of claim 1 and a transmission control with the features according to the preamble of claim 11.

For gearbox control, the coefficient of friction can be used as the adjustment factor. The coefficient of friction can be z. B. be adapted for approaches and circuits. On this short-term and long-term changes of the physical torque characteristic can be detected and taken into account in the control of the clutch.

The pamphlets DE 102 13 946 A1 . WO 2002/23066 A1 . DE 101 14 880 A1 . DE 195 04 847 A1 . DE 42 41 995 A1 such as DE 26 28 486 A1 disclose method for performing a Reibwertadaption an automated clutch. However, none of the documents has a method which corrects by means of the Reibwertadion an urtümlicherweise very low estimated coefficient of friction in a short time upwards.

In a known from practice Reibwertadaption in an automated transmission, a comparison of the clutch torque with the engine torque can be performed during the slip phases on the basis of the engine torque equilibrium. Depending on this, then an adjustment of the clutch characteristic can be carried out, in which the coefficient of friction is corrected accordingly.

In the known Reibwertadion an extremely low coefficient of friction can be corrected due to a short slip phases disadvantageously only slowly upwards. As a result, for example, during commissioning in the vehicle plant, for example, due to a not completely carried out ventilation, an excessive touch point can be learned. This behavior is compensated by the friction value adaptation, in which the coefficient of friction is lowered considerably. During the first circuits the air can escape from the system. The touch point can be corrected relatively quickly via the touch point adaptation. However, the coefficient of friction can not follow the fast sensing point adaptation and it can continue to assume the low values. As a result, the driver may experience an extremely uncomfortable shift, d. H. a hard disengagement and engagement can occur.

Furthermore, it has been shown that the coefficient of friction is strongly temperature-dependent. The physical coefficient of friction of the clutch increases with the temperature and drops sharply again at a coating-specific temperature. This is also called fading.

It may happen that the lining temperature changes very quickly during the clutch, for example during many hill climbs, and thus the coefficient of friction changes very quickly. It has been shown that in the known Reibwertadaption sometimes very long slip phases occur during approaches and circuits. Due to the long slip phases, the clutch can be heated even faster, so that the physical coefficient of friction changes even faster.

The present invention has for its object to propose a transmission control and a method for performing a Reibwertadaption to further improve the control of the clutch.

The object is achieved by a method with the features of claim 1 and by a transmission control with the features of claim 11.

In this way, the Reibwertadaption invention is modified so that the known from the prior art disadvantages are avoided. For example, depending on a preferred clutch size or engine or transmission size of the process, such. As the beginning and / or duration, the adaptation can be influenced.

According to the invention, it is advantageously provided that the change in the coefficient of friction is carried out as a function of the coupling temperature or the like. Accordingly, according to the invention, the Reibwertadaption z. B. are accelerated at high clutch temperatures, so that the rapid changes in the physical coefficient of friction are absorbed at the clutch. It is also possible for the course of the friction value adaptation to be changed as a function of the clutch temperature gradient.

A next development of the invention may provide that in the context of Reibwertadaption z. B. during a slip phase of the clutch detected in the clutch control clutch actual torque is compared with the physical, actual value of the clutch torque. Preferably, the physical clutch torque from the engine torque and the engine acceleration can be determined.

It is particularly advantageous according to an embodiment of the invention that, if the two moments assume different values, the software friction coefficient, ie the value stored in the transmission or clutch control, is preferably changed in small steps or increments. For example, 15 increments per circuit or per approach can be used preferably. The maximum change can be chosen such that a sufficient robustness of the Reibwertadaption is ensured.

In order to further accelerate the friction value adaptation so that rapid changes in the physical coefficient of friction can be adjusted, various possibilities are conceivable.

For example, upon reaching a predetermined clutch temperature and / or upon reaching a predetermined temperature gradient, the maximum value of the increments used can be increased by a certain value.

Furthermore, it is conceivable that upon reaching a predetermined clutch temperature and / or upon reaching a predetermined temperature gradient, the maximum value of an increment used is preferably increased linearly with the clutch temperature.

It is also possible for the maximum increment to be determined in a characteristic diagram, which is dependent on the clutch temperature and / or the temperature gradient, for example, so that the friction coefficient can be changed more quickly at higher temperatures and / or higher temperature gradients.

The measures mentioned make sure that in the normal working range in which the clutch is most of the time, the Reibwertadaption is sufficiently robust, but adapted in an area in which the coefficient of friction can vary greatly, the adaptation speed in an advantageous manner accordingly becomes.

In the context of the invention presented here, it is provided according to the invention that the friction value adaptation corresponds to these properties due to the above-mentioned properties of Reibwertadaption described, namely relatively fast to act at too high friction value, while in the opposite direction, the adaptation much slower or not at all is adjusted. In this way, the shifting comfort can be increased by the proposed Reibwertadaption.

According to the invention, an adaptation of the friction value adaptation takes place if the friction value adaptation is not carried out or evaluated due to an excessively short slip duration in a shift and / or the slip duration for the specific shift is too short for a given speed jump. It is also possible to determine other criteria in order to further improve the friction coefficient adaptation according to the invention.

When checking whether the aforementioned or other criteria are present, it is provided according to the invention that the coefficient of friction is increased either absolutely by a certain proportion and / or relatively by a certain percentage. In this way it can be ensured that the Reibwertadaption from the so-called deadlock situation described above can adapt again. This means that the coefficient of friction can increase accordingly. Due to the increasing coefficient of friction, the usual duration of slippage can be achieved, so that a normal evaluation of the friction value adaptation can take place in an advantageous manner.

In the context of an advantageous development of the present invention, the friction coefficient adaptation according to the invention can also be further simplified. Preferably z. B. below a predetermined Reibwertschwelle or a limit value, for example, at each adaptation phase or the like, without the evaluation of the adaptation, a certain, the friction coefficient increasing proportion be taken into account. If this proportion is smaller than the maximum permissible change in the adaptation, this can also ensure the desired behavior in the control of the clutch.

It is also possible to make other changes in the Reibwertadaption in order to simplify the process of the invention further on. This type of friction value adaptation is preferably to be used in an electronic clutch management (EKM), in an automated transmission (ASG), in a parallel transmission (PSG) and / or in an uninterruptible transmission (USG).

The object underlying the invention can be solved by a device device for a transmission with a clutch, in particular an automated clutch, in particular for carrying out the proposed method, in which at least one device for detecting and evaluating clutch sizes is provided to the course of Adjust friction coefficient accordingly.

Further advantages and advantageous embodiments will become apparent from the dependent claims.

Furthermore, the invention relates to a method, an apparatus and their use for operating a motor vehicle, with a drive motor, a clutch and / or a transmission in the drive train, in particular for controlling the transmission.

According to 1 indicates a vehicle 1 a drive unit 2 such as a motor or an internal combustion engine. Furthermore, in the drive train of the vehicle 1 a torque transmission system 3 and a gearbox 4 arranged. In this embodiment, the torque transmission system 3 arranged in the power flow between the engine and transmission, wherein a drive torque of the engine via the torque transmission system 3 to the gearbox 4 and from the transmission 4 on the output side to an output shaft 5 and to a downstream axis 6 as well as to the wheels 6a is transmitted.

The torque transmission system 3 is as a clutch, such as. Example, as a friction clutch, multi-plate clutch, magnetic powder clutch or torque converter lock-up clutch, configured, wherein the clutch may be a self-adjusting or a wear-compensating coupling. The gear 4 is an uninterruptible gearbox (USG). According to the inventive concept, the transmission can also be an automated transmission (ASG), which can be switched automatically by means of at least one actuator. As an automated manual transmission is further to be understood an automated transmission, which is connected to a traction interruption and in which the switching operation of the transmission ratio is controlled by at least one actuator.

Furthermore, as USG also find an automatic transmission use, wherein an automatic transmission, a transmission is substantially without interruption of traction in the switching operations and which is usually constructed by planetary gear stages.

Furthermore, a continuously variable transmission, such as conical-pulley transmission can be used. The automatic transmission can also be equipped with a torque transmission system arranged on the output side 3 , be designed as a clutch or a friction clutch. The torque transmission system 3 can continue to be designed as a starting clutch and / or reversible clutch for reversing the direction of rotation and / or safety clutch with a selectively controllable transmissible torque. The torque transmission system 3 may be a dry friction clutch or a wet friction clutch, for example, running in a fluid. It can also be a torque converter.

The torque transmission system 3 has a drive side 7 and an output side 8th on, with a torque from the drive side 7 on the output side 8th is transferred by z. B. the clutch disc 3a by means of the pressure plate 3b , the plate spring 3c and the release bearing 3e as well as the flywheel 3d is charged with force. To this admission, the release lever 20 by means of an actuating device, for. As an actuator operated.

The activation of the torque transmission system 3 takes place by means of a control unit 13 , such as B. a control unit, which the control electronics 13a and the actor 13b may include. In another advantageous embodiment, the actuator 13b and the control electronics 13a also in two different units, such. As housings, be arranged.

The control unit 13 can the control and power electronics for controlling the drive motor 12 of the actor 13b contain. This can be advantageously achieved, for example, that the system as a single space, the space for the actuator 13b needed with electronics. The actor 13b consists of the drive motor 12 , such as B. an electric motor, wherein the electric motor 12 via a transmission, such. As a worm gear, a spur gear, a crank gear or a threaded spindle gear, on a master cylinder 11 acts. This effect on the master cylinder 11 can be done directly or via a linkage.

Movement of the initial part of the actuator 13b , such as B. the master cylinder piston 11a , comes with a clutch travel sensor 14 detected, which detects the position or position or the speed or the acceleration of a large, which is proportional to the position or engagement position or the speed or acceleration of the clutch. The master cylinder 11 is via a pressure medium line 9 , such as B. a hydraulic line, with the slave cylinder 10 connected. The starting element 10a the slave cylinder is with the release means 20 , z. B. a release lever, operatively connected, so that a movement of the output part 10a the slave cylinder 10 causes the release means 20 also moved or tilted to that of the clutch 3 to control transmissible torque.

The actor 13b for controlling the transmittable torque of the torque transmission system 3 can be druckmittelbetätigbar, ie, it may have a Druckmittelgeber- and slave cylinder. The pressure medium may be, for example, a hydraulic fluid or a pneumatic medium. The actuation of the Druckmittelgeberzylinders can be done by an electric motor, wherein the drive element as 12 provided electric motor can be controlled electronically. The drive element 12 of the actor 13b may be in addition to an electric motor drive element and another, for example, pressure-medium-actuated drive element. Furthermore, magnetic actuators can be used to adjust a position of an element.

In a friction clutch, the control of the transmittable torque takes place in that the contact pressure of the friction linings of the clutch disc between the flywheel 3d and the printing plate 3b targeted. About the position of the release 20 , such as As a release fork or a Zentralausrückers, the application of force to the pressure plate 3b respectively the friction linings are selectively controlled, the pressure plate 3b can be moved between two end positions and set and fixed as desired. The one end position corresponds to a fully engaged clutch position and the other end position of a fully disengaged clutch position. To control a transmittable torque, which is, for example, less than the momentarily applied engine torque, for example, a position of the pressure plate 3b be driven, which lies in an intermediate region between the two end positions. The coupling can by means of targeted activation of the release 20 be fixed in this position. However, it is also possible to control transmittable clutch torques which are defined above the currently pending engine torques. In such a case, the currently pending engine torques can be transmitted, wherein the torque irregularities in the drive train are damped and / or isolated in the form of, for example, torque peaks.

For controlling the torque transmission system 3 Furthermore, sensors are used which at least temporarily monitor the relevant quantities of the entire system and supply the control variables necessary for the state variables, signals and measured values which are processed by the control unit, wherein a signal connection to other electronic units, such as to an engine electronics or electronics of a Antilock braking system (ABS) or traction control (ASR) can be provided and can exist. The sensors detect, for example, speeds, such as wheel speeds, engine speeds, the position of the load lever, the throttle position, the gear position of the transmission, a shift intent and other vehicle-specific characteristics.

The 1 shows that a throttle sensor 15 , an engine speed sensor 16 and a tachometer sensor 17 Can be used and readings or information to the control unit 13 hand off. The electronics unit, such as. B. a computer unit, the control electronics 13a processes the system inputs and gives control signals to the actuator 13b further.

The transmission is as z. B. stepped change gear, wherein the gear ratios by means of a shift lever 18 be changed or the gearbox by means of this lever 18 is operated or operated. Furthermore, on the shift lever 18 of the manual transmission at least one sensor 19b arranged, which detects the shift intention and / or the gear position and to the control unit 13 forwards. The sensor 19a is hinged to the transmission and detects the current gear position and / or a shift intention. The shift intent detection using at least one of the two sensors 19a . 19b can be done by the sensor is a force sensor, which on the shift lever 18 acting force detected. Furthermore, however, the sensor can also be configured as a displacement or position sensor, wherein the control unit recognizes a switching intent from the time change of the position signal.

The control unit 13 is at least temporarily in signal communication with all sensors and evaluates the sensor signals and system input variables in such a way that depending on the current operating point, the control unit control or regulation commands to the at least one actuator 13b outputs. The drive motor 12 of the actor 13b , z. As an electric motor, receives from the control unit, which controls the clutch actuation, a manipulated variable in response to measured values and / or system input variables and / or signals of the connected sensors. This is in the control unit 13 a control program implemented as hardware and / or software that evaluates the incoming signals and calculates or determines the output variables based on comparisons and / or functions and / or maps.

The control unit 13 Advantageously, it has implemented or is in signal communication with at least one of these units, a torque determination unit, a gear position determination unit, a slip determination unit, and / or an operating condition determination unit. These units can be implemented by control programs as hardware and / or as software, so that by means of the incoming sensor signals the torque of the drive unit 2 of the vehicle 1 , the gear position of the gearbox 4 as well as the slip, which in the area of the torque transmission system 3 prevails and the current operating state of the vehicle 1 can be determined. The gear position determination unit determines based on the signals of the sensors 19a and 19b the currently engaged gear. Here are the sensors 19a . 19b on the shift lever and / or on internal gear actuating means, such as a central shift shaft or shift rod, hinged and detect these, for example, the location and / or speed of these components. Furthermore, a load lever sensor 31 at the load lever 30 , such as B. on an accelerator pedal, which detects the load lever position. Another sensor 32 can act as an idle switch, ie when the load lever is actuated 30 or accelerator pedal is this idle switch 32 switched on and with the load lever not actuated 30 it is off, so that can be detected by this digital information, whether the load lever 30 is pressed. The load lever sensor 31 detects the degree of operation of the load lever 30 ,

The 1 shows next to the load lever 30 and the associated sensors, a brake actuator 40 for actuating the service brake or the parking brake, such. As a brake pedal, a hand brake lever or a hand or foot operated actuator of the parking brake. At least one sensor 41 is on the actuator 40 arranged and monitors its operation. The sensor 41 is for example as a digital sensor, such. B. as a switch, designed, which detects that the brake actuator 40 is pressed or not operated. With the sensor 41 can a signaling device, such. As a brake light, are in signal communication, which signals that the brake is actuated. This can be done both for the service brake and for the parking brake. The sensor 41 However, it can also be designed as an analog sensor, wherein such a sensor, such as a potentiometer, the degree of actuation of the brake actuator 41 determined. This sensor can also be in signal connection with a signaling device.

Hereinafter, a possible embodiment of the present invention will be described, in which a suitable Reibwertadaption is provided to increase in particular the shifting comfort.

In the case of a possible friction value adaptation, in particular in the case of an electronic clutch management (EKM) and in an automated manual transmission (ASG), during the slip phases on the basis of the engine torque equilibrium J · ω. _Mot = M _Mot - M _Kupp a comparison of the clutch torque with the engine torque can be performed. Depending on this comparison result, an adaptation of the clutch characteristic can then be carried out, in which the friction coefficient is suitably corrected. As slip phases, the start-up phase and the Einkuppelphase or other phases can be evaluated after the circuits. In addition, additional constraints may be provided, such as | ω. _Mot | <ω. _Max . It is possible that other constraints may be used.

• a) Ein extrem zu hoher Reibwert kann relativ schnell nach unten korrigiert werden, da hier die Voraussetzung, wie zum Beispiel lange Schlupfphasen oder dergleichen, gegeben sind.
• b) Ein extrem zu niedriger Reibwert kann nur langsam nach oben korrigiert werden, da hier die Voraussetzung bei Einkuppelphasen aufgrund der kurzen Schlupfphasen prinzipiell nicht mehr gegeben sind.

This results in the following properties of the intended Reibwertadaption:

• a) An extremely high coefficient of friction can be corrected relatively quickly downwards, since the prerequisite, such as long slippage phases or the like, are given here.
• b) An extremely low coefficient of friction can be corrected only slowly upwards, since the prerequisite for engagement phases due to the short slip phases are no longer given here in principle.

As a result, in particular during commissioning in the vehicle plant, for example because of a not completely carried out ventilation, an excessive touch point can be taught. This behavior compensates for the friction value adaptation in which the coefficient of friction drops sharply; especially partly on the lower one Border. During the first circuits the air can escape from the system. The touch point can be corrected relatively quickly via the touch point adaptation. The coefficient of friction can then not follow the fast touch point adaptation and this can continue to assume the low values. As a result, the driver may experience an extremely uncomfortable shift, ie a hard disengagement and engagement is present.

Due to the characteristics of the above-described Reibwertadaption, namely to act relatively fast at too high friction value, while in the opposite direction, the adaptation much slower or not responding, can be provided according to the invention presented here, that the Reibwertadaption suitably adapted to these properties becomes.

• – Die Reibwertadaption konnte aufgrund zu kurzer Schlupfdauer bei einer Schaltung nicht durchgeführt bzw. ausgewertet werden, und/oder
• – die Schlupfdauer ist für die spezielle Schaltung bei gegebenem Drehzahlsprung zu kurz.

For this purpose, certain criteria can be established, such as the following:

• - The coefficient of friction adaptation could not be performed or evaluated due to too short a slip time in a circuit, and / or
• - The slip time is too short for the specific circuit given a speed jump.

It is also possible to determine other criteria in order to further improve the friction coefficient adaptation according to the invention. When checking whether the aforementioned or other criteria are present, it can be provided that the coefficient of friction is increased either absolutely by a certain proportion and / or relatively by a certain percentage. In this way it can be ensured that the Reibwertadaption from the so-called deadlock situation described above can adapt again. This means that the coefficient of friction can increase accordingly. Due to the increasing coefficient of friction, the usual duration of slippage can be achieved so that a normal evaluation of the friction value adaptation can take place in an advantageous manner.

In the context of an advantageous development of the present invention, the friction coefficient adaptation according to the invention can also be further simplified. Preferably z. B. below a predetermined Reibwertschwelle, for example, at each adaptation phase, without the evaluation of the adaptation, a certain, fixed increasing proportion are taken into account. If this proportion is smaller than the maximum permissible change in the adaptation, this would also ensure the desired behavior. It is also possible to make other changes in the Reibwertadaption in order to simplify the process of the invention further on. This type of friction value adaptation is preferably to be used in an electronic clutch management (EKM), in an automated manual transmission (ASG), in a parallel transmission (PSG) and / or in an uninterruptible transmission (USG).

Hereinafter, another possible embodiment of the present invention will be described, in which a suitable displacement measurement, especially for clutch plates, is proposed.

It has been shown that an incremental displacement measurement relative to an absolute displacement sensor does not have the limitation of a mechanically limited measuring range. Accordingly, the present invention has the object, preferably to propose an absolute displacement measurement in clutch plates, which does not have the aforementioned disadvantage.

Accordingly, for example, an absolute angle sensor, for example, inductive, be provided such that over a revolution, for example, the armature of the electric clutch actuator, the entire display area is swept over. After one revolution, the sensor signal can then return to the initial state. This is for example in 2 shown.

In this way, an absolute displacement measurement, for example with an absolute angle sensor, be provided, without resulting in the disadvantage of a limited maximum path or a too low resolution of the sensor signal. Thus, by using an absolute angle sensor, a purely incremental displacement measurement can be avoided.

In 2 is a sensor signal for measuring distance in a clutch actuator shown. The sensor signal corresponds to a counting absolute displacement measurement, as proposed by the embodiment of the present invention presented here.

Hereinafter, another embodiment of the present invention will be described, in which an initialization of the position of the clutch actuator at characteristic points, preferably the clutch spring force characteristic is proposed.

It is possible that an initialization of the position of the clutch actuator is performed by means of an incremental displacement measurement; this in particular to compensate for mechanical changes to the respective clutch.

According to the invention can be provided that instead of a stop or a detent on the actuator or on the release system as a reference point at least one characteristic point of the force curve of the clutch spring is used, for example, a plate spring or the like.

As a reference point can be used in the force curve, for example, a kink or the like. This kink occurs in a single clutch, for example, when the diaphragm spring is relieved so that the release system stands out from the plate spring. For example, in a combination clutch, in particular for an uninterruptible gearbox (USG), a similar kink may be provided in the course of the force curve, for example in a central position at the transition between the two characteristics of the individual clutches. It is also possible that other reference points are used. For example, force zero passes can also be used as reference points.

In an advantageous manner, it can be provided that the search for and determination of such a reference point simultaneously represents so-called software snooping. In this case, mechanical shifts, as they can occur due to wear or by thermal effects, are compensated in the control of the actuator in an advantageous manner. On the other hand, by measuring the mechanical changes, for example when sniffing, it may also be possible to draw conclusions about the thermal state or degree of wear of the coupling. It is essential that the reference point is not exactly determined, but a method is to be specified which identifies a position on the characteristic curve in a reproducible manner.

To carry out the so-called software snooping, the following may be considered; For the electric motor of the actuator, a change in the disengagement force can be represented as a change in the friction torque in the almost self-locking release gear. This friction torque may also depend on the sign of the external force. A movement against the force results in a much greater friction than a movement in the direction of the force. This can result in a virtually linear relationship between the force and the friction torque for not too small forces.

The implementation of this so-called sniffing and the position adjustment can be realized in different ways. For example, a speed measurement at constant motor voltage can be provided. When moving the actuator with a constant voltage, the speed results from the friction torque, which results in self-locking actuator gears. A change in the clutch spring force can cause a change in the friction torque and thus a speed change of the actuator. The measurement can preferably take place in a movement against the clutch spring force and with increasing spring constant (.DELTA.F / .DELTA.x). The change of the speed over the location (Δv (Δx) can be independent of the magnitude of the speed, neglecting the inertia (ṅJ> MR) In each case almost linear force characteristics before and after the kink, one results only from the spring constant ( (.DELTA.F / .DELTA.x) and the engine characteristics dependent value for .DELTA.v / .DELTA.x Since the speed is the most accurate signal of the actuator in a incremental path measurement, this speed change can be advantageously easily measured and detected.

In 3 is a possible load curve and a resulting theoretical course of the actuator speed shown, wherein the direction of movement of the actuator is indicated during the measurement by an arrow above the two diagrams. Faults which can occur periodically during the engine revolution are suppressed in the simplest way in this presented measurement process. It is only necessary that exactly one motor revolution is to be selected as the basis for the determination of Δv / Δx as Δx. For example, the reference point can then be determined by the intersection of speed ramps. It is also possible to use other reference point.

The speed measurement can initially take place only at the same angular position or phase position of the actuator, where xi with i is to be seen as an index for the engine revolution. From the velocities vi determined in this way, it can then be recognized at which engine revolution the system is in the area of the hard spring. The result is the following limit: abs (vi - vi-1)> limit

If it is detected from which motor revolution the system is in the range of the hard spring, this means that the measurements of vi and vi-1 limit the temporally first interval from xi to xi-1 on the spring load. The measurements of vi-2 and vi-3 can limit the last time interval (xi-2 to xi-3) before the spring load. Between these intervals, there may be the position interval (xi-1 to xi-2) in which the reference point is provided.

From these intervals and the velocities measured at their boundaries, the point at which the actuator speed changes can advantageously be determined. This corresponds to the reference point. This determination of the reference point is graphically in 4 indicated, wherein the determined position, ie the reference point is indicated as an intersection between the two lines shown.

Any existing noise can be minimized by evaluating a measurement process at different phase angles with respect to the engine revolution and by subsequent averaging. Another possible improvement in the reference point determination may result from the suppression of implausible results (for example, if the reference position is outside the interval xi to xi-3).

Known problems, such as a strong dispersion of velocities at certain phase angles, can also be taken into account in the determination of the reference point or in the averaging according to a development of the present invention.

In 5 is a possible algorithm tabulated. In this case, it can be provided that a set of data is used for each phase position in which the reference position is searched. First, under the keyword "phase angle", it is checked to which phase the data belong. Under the keyword "revolution", a counter for the engine revolutions since the start of the measurement can be used, which serves the position determination. Afterwards, under the keyword "Vbuffer" a 4-value array can be used, in which the last speeds are stored. Finally, under the keyword "Position" the determined position can be output, which serves the flow control.

In 6 is a possible flowchart shown in an overview. Initially, the initialization of the process variables takes place. Thereafter, the measurement and setting of the considered phase angles is started. Then the measurement is continued with the determination of the reference points for the individual phase positions. Finally, the final results are determined by averaging from the found reference positions of the individual phase positions. In 7 this is shown in a flow chart, which includes the start of the measurement and the definition of the considered phase angles. In 8th is also shown a flow chart in which the reference point for each phase is determined.

It is possible that the so-called "software snooping" in the same driving conditions is performed, in which also the hydraulic sniffing is performed. This is possible, in particular when the engine is at a standstill and while driving, when not switched. As a result, a regular adjustment of the position is realized. In particular, in an uninterruptible transmission (USG), this approach corresponds to a sniffing for the starting clutch, so that can be sniffed here to the same driving conditions, if not driven straight with the load shift gear.

This proposed method can be used in particular for electrical Zentralausrückern or for the operation of a single or two coupled couplings. Overall, it can be stated that this method can be used with any actuator with a mechanical release system as so-called "software snooping", whereby the type of displacement sensor used in the actuator system is irrelevant.

In addition, it is conceivable that this proposed initialization of the position of the actuator can also be used in mechanical end stops. For example, the position adjustment can also be replaced by a suitable touch point adaptation or supplemented.

Particularly advantageous is the so-called "software snooping" with respect to the reference point search, since no touch point adaptation can be used as a replacement strategy especially for long highway trips for a long time.

wobei

n:

Motordrehzahl

n₀:

Leerlaufdrehzahl

M_R:

Das auf den Motor wirkende Reibmoment

M_A(U):

Anlaufmoment (Moment bei n = 0) Dabei gilt: M_A(U) ∝ U

In particular, the proposed position initialization is based on the following considerations:
The properties of permanently excited DC motors cause a given speed U sets a speed that is in a linear relationship to the braking torque.

in which

n:

Engine speed

n ₀ :

Idle speed

M _R :

The friction torque acting on the engine

M _A (U):

Starting torque (moment at n = 0) where: M _A (U) α U

Therefore, in a system in which the inertia of the motor has a negligible influence (ṅJ <M _R ) from the rotational speed of the motor can be closed to the load. A position-dependent load thus leads to a position-dependent speed.

In a linear load curve, this also leads to a linear velocity course:

• • Keine integrierte Kompensationsfeder
• • Inkrementalwegsensor (aus einem Wählmotor) auf der Ankerwelle des Kupplungsaktors mit 40 Inkrementen pro Umdrehung; daraus folgt ca. 40 Inkremente pro mm am Geberzylinder, daraus folgt wiederum ca. 12,5 μm/Inkremente an der (theoretischen) Tellerfeder.

It can z. B. Measurements with a modified clutch actuator (ASG) are performed on a special test setup. As possible modifications of the clutch actuator, the following can be provided, the enumeration being not exhaustive:

• • No integrated compensation spring
• • Incremental travel sensor (from a selector motor) on the armature shaft of the clutch actuator with 40 increments per revolution; This results in approx. 40 increments per mm on the master cylinder, which in turn is followed by approx. 12.5 μm / increments on the (theoretical) diaphragm spring.

It can z. B. different loads (springs, detent, stops or combinations thereof) are given to the clutch plate. The structure may preferably consist of a skeleton of aluminum rails, on which the controller is attached to the clutch actuator. Different load modules can be positioned on the aluminum profiles. The plunger of the clutch actuator is firmly connected to a rod which is slidably mounted on the skeleton. The Rastierungsprofil and the stop discs (also for the springs) are freely positioned on this rod and also solvable. Thus, the positions of the different loads can be varied and also multiple loads combined can be used. The test setup is schematic in 9 shown.

• • Die (ehemalige) Kompensationsfeder mit 5.9 N/mm; sie dient der Simulation einer kleinen Last, wie sie auch im realen System benötigt wird, um ein sicheres Funktionieren des Ausrücksystems zu gewährleisten.
• • Die „Kupplungsfeder” mit 19.9 N/mm; sie simuliert die Kraft der Tellerfeder.

For the investigations presented here two spring loads are used, whereby both springs are loaded in the same direction.

• • The (former) compensation spring with 5.9 N / mm; It is used to simulate a small load, as is required in the real system, to ensure the safe functioning of the release system.
• • The "clutch spring" with 19.9 N / mm; it simulates the force of the diaphragm spring.

With the test setup presented here, the in 10 shown load profile can be generated. The theoretical load profile includes the force curves of spring 1 and 2 as well as the summed course. Due to the (almost) self-locking actuator gear, the position-dependent load results in a position-dependent (and direction-dependent) friction torque on the motor. The friction torque for movements against the force is almost proportional to this force.

In the following, increments are preferably used for the position information. The conversion between increments and the true position is only approximately linear due to the actuator operation. In addition, the sign rotates around. In the measurements shown here, the weak spring starts from a position of 0 increments (with increasing force with negative position) and the strong spring starts at -100 increments. The motor voltages are given by the pulse width ratio. A value of 255 corresponds to approx. 10 V. For the speed, the values given in the control unit are preferably used.

It turns out that when using an incremental travel sensor, the actuator speed is the most accurate signal for the state of the actuator. Even a (externally mounted) measurement of the motor current is significantly inaccurate. First, the result of a series of measurements will be shown here, below the above The actuator is driven in both directions with different voltages (PWM ratios ± 30, ± 35. ± 40, ± 45, ± 50, ± 55 and ± 60). In 11 Measured velocities are shown across the position for different motor voltages.

The effect of the load on the speed of the actuator is much better with movements against the load (here at negative speeds) than with the load. This behavior can be explained by the properties of the self-locking gearbox. It shows a significant disturbance with a period of 40 increments = 1 engine revolution. When running towards the load, a constant speed change over the location (Δν / Δx) is detected (under the periodic disturbance) regardless of the magnitude of the speed. The reproducibility of the measurements is very high.

In 12 the reproducibility of the speed measurements is shown. In this measurement, the actuator is reciprocated eight times in both directions with always the same voltage (PWM ratio = ± 35). The curves are very consistent. Even the small velocity variations with a period of 3 to 4 increments are reproducible. Presumably, these are due to the magnetic Ankererrastierung. When evaluating the speed measurements, first the calculation of the speed change per revolution is taken into account. The first consideration is to calculate, for each position in a measurement, the difference of the velocity to the exactly 40 increments, which is performed according to the following equation: B (x) = ν (x) - ν (x + 40) (for movements with negative velocity.

Since the velocity is not measured at every position due to the temporal sampling of the signals, the velocity profile should be suitably interpolated. If the measurement data shown above are evaluated in this way, the result is the 13 , Again, the movement is from right to left. The signal is recognizable by the noise but so bad that it can not be used in this form for the position adjustment. Averaging advantageously reduces the noise. However, the position resolution may be degraded.

In the following, the calculation of the buckling position by linear interpolation of the speeds is considered. If the speed during a measurement is always considered at the same phase position of the motor revolution (Pos mod 40 = const.), A signal released from the modulation by the motor rotation may result. Depending on the selected phase position, a slightly different velocity pattern is shown here. However, apart from a bit of noise, the speed curves differ only by a speed offset. In 14 this is shown for two phases.

Two velocity profiles are shown at different phase angles. The next step in the evaluation is to know from which engine revolution you are on the hard spring. This is z. B. with a simple limit value violation (| ν _i - ν _i-1 |> C) to recognize. The limit value C can be dependent on the properties of the engine and the hardness of the spring. In 15 a rough position reference can be seen from the speed changes. This means that the buckling position of the load characteristic curve can be exactly found to within ± 40 increments.

A better position indication can be obtained if the last velocity measurements are used as interpolation points for straight lines and their intersection point is calculated. The measurements of ν _i and ν _i-1 limit the first interval on the hard spring, which is the position interval found above. The measurements of ν _i-2 and ν _i-3 limit the last interval before the hard spring. In 16 the determination of the exact reference position is shown graphically by way of example.

The position determined in this way is still not very accurate for each individual phase position (in the best case, it is 20 increments); but it is possible to apply this method for several phases of a measurement series. However, a simple averaging of all possible positions that can be determined in this way is not enough to significantly improve the accuracy. Errors caused by noise can be partially detected and suppressed. Thus it can happen that the determined intersection of the two straight lines outside the interval [x _i ; x _i-3 ]. Such a case can be easily recognized and the result for averaging suppressed.

It has also been found that the speed determination for phase positions which are located on the flanks of the periodic disturbance is subject to very great fluctuations as a result of the interpolation. For this reason, care is taken right at the beginning of the measurement that these phase angles are not correct consider. This is done by evaluating the acceleration (Δν / Δi). Phasing that borders on excessive acceleration in the first motor revolution will not be considered further. Averaging from the thus selected values for reference point determination then achieves the accuracy of about 10 to 12 increments (or less than 5.5 increments of standard deviation).

The quality of the proposed method is tested by measurements with different starting positions for the hard spring. The specified positions are only approximate because they are adjusted by hand. Important here is the reproducibility of recognition of the same position. This is in 17 represented, wherein the determined positions of the reference point are indicated. Apart from too small (<40) or too large (> 55) motor voltages, the position where the kink of the force curve is found is independent of the motor voltage. The scatters of the determined positions are each in a range of ± 3 to ± 5.5 increments. The individual measurements lasted between 3 s (for U = -35) and 1.6 s (U = -55). If the measurements are limited to the necessary range of motion, measurement times between 1.6 s (U = -35) and 0.6 s (U = -55) are obtained. These can be further reduced.

It can be assumed that the hardness of the spring, whose starting position is sought, enters into the accuracy of the position determination. However, there is the possibility that a harder spring allows a more accurate determination of the reference position. The influence of tension on the detected position should also be considered. Furthermore, the effects of the starting position (relative to the reference point) and the duration of the measurement should be taken into account. Since the actuator temperature influences in particular the idling speed and the starting torque of the electric motor, these should also be taken into account.

In summary, it can be stated that a method is proposed with which the actuator speed can be detected, in particular at the time at which the release system touches the diaphragm spring. In this case, it is particularly advantageous that, apart from the incremental travel sensor, no additional sensors or mechanical parts are required. Preferably, this method can be used in actuators without compensation spring and with a mechanical release system (eg USG with EZA). The achieved reproducibility for the position adjustment corresponds approximately to the standard deviation in the intended test. In this system, this would correspond to a positioning accuracy of about 0.15 mm on the diaphragm spring.

Hereinafter, a further embodiment of the invention presented here will be described, in which a possible monitoring of an automated clutch is proposed in particular during creeping.

Especially with a permanent crawl without vehicle movement, z. B. due to an obstacle or on a slope, a high Reibenergieanteil can arise in particular on the clutch disc on the slipping clutch. This is associated with increased wear of the clutch documents. Due to the heating or the forming temperature gradients in the clutch unit, the transmission properties of the clutch change. Under certain circumstances, it is possible that there may be a sharp increase in the torque transmitted by the clutch. This can lead to unwanted vehicle movements and associated safety-critical situations.

Accordingly, it can be provided according to the invention that for monitoring the clutch during creep several monitoring methods are proposed, which can be used alternatively or in combination.

It is possible, for example, that the coupling temperature preferably with the aid of a temperature sensor or a suitable model z. B. from the friction energy at the clutch disc and cooling effects, such as speed-dependent heat transfer and heat conduction is determined. Preferably, when a predetermined limit value is exceeded, the driver warning can be activated, i. H. the driver can be warned acoustically, visually and / or mechanically, for example.

Furthermore, it is possible that the time duration is determined with the aid of a time measurement, in which a predetermined slip speed threshold is exceeded when the creep condition is activated. If preferably the time exceeds a predetermined limit, the driver warning can also be activated.

In the context of another development, it can be provided that when the creep condition is activated, the engine torque and its time derivative are checked or monitored. For both sizes can appropriate limits are defined. Since the engine torque due to consumers, such as an air conditioner or a power steering, can briefly increase sharply, two limit conditions should be met to distinguish between influence of consumers and temperature-related change in the coupling transmission characteristics. Other conditions are also possible to recognize the distinction between the different influences. If the engine torque exceeds the preset limit while the derivative remains below its limit, the driver warning and / or adjustment of the clutch transmission characteristics may be activated.

Preferably, as an alternative to the aforementioned strategy, it can also be provided that the torque transmitted by the clutch is determined directly with the aid of a torque sensor. If this torque exceeds a predetermined limit, the driver warning and / or adjustment of the clutch transmission characteristics may also be activated.

As a driver warning, various options can be used, which are used alternatively or in combination with each other. It is conceivable that an acoustic warning, for example by a beeper or an announcement, is provided. Furthermore, an optical warning can also be provided by a flashing warning light or by a text message in a display. In addition, a mechanical warning, for example by jerking the vehicle is possible. The jerking can be caused for example by an oscillatory movement of the clutch actuator. Here, the main frequency band should be in the range of the vehicle's own frequency, so that with small torque amplitudes a clearly noticeable jerking is caused. Preferably, the jerking can be limited in time, for example, for a few seconds, and / or repeated with a certain time offset, for example, 10 seconds, until a driver reaction, for example by pressing the accelerator pedal or the brake takes place.

The software adaptation of the coupling transmission properties can be done in principle in two ways. For example, by the correction of the touch point and / or by the correction of the coefficient of friction of the clutch characteristic.

In the monitoring method proposed according to the invention, however, provision is made for the change in the coupling transmission characteristics to be analyzed for the relevant coupling unit, and the corresponding procedure preferably be selected with corresponding weighting factors.

It is conceivable that the monitoring method according to the invention is also used if the clutch temperature rises above a first limit value during creep. Then the torque transmitted by the clutch can be lowered. From a second limit, the clutch can be fully opened. However, if the vehicle is to be stopped with a slipping clutch on an incline, opening the clutch will cause the vehicle to roll back. If the clutch temperature exceeds a limit value, a warning jerk can be activated when starting depending on the gear.

The inventively proposed monitoring method can be used in transmission concepts with automated coupling, in particular in EKM, ASG, USG, PSG or the like, preferably when a creep function is integrated.

Hereinafter, a further embodiment of the invention presented here will be described, in which a variable current limiting preferably a clutch actuator for example, an XSG transmission is proposed.

It has been shown that an optimization of the actuation of a clutch actuator is necessary in view of the different demands on dynamics and comfort.

A central requirement for the actuators of an automated clutch is u. a. a high dynamic; This means the least possible adjustment. To meet this requirement, the actuator has a high power consumption in known couplings, which is reflected in electromechanical actuations in relatively high currents. In this way, very high currents (50 to 100 A) occur in the short term when starting electric motors, if no suitable measures for limiting the current are provided. However, these high start-up currents are accepted only to a limited extent, as these inadmissible on-board network loads and comfort impairments, such as light flares, actuator noise or the like, can cause.

By changing the design of the electric motor, for example with regard to a higher resistance of the winding or the leads, higher currents can be avoided. On the one hand, however, this can also increase the energy dissipation and, on the other hand, this can adversely affect the dynamics of the engine. Consequently, for example, limiting the motor current to a fixed value is more advantageous.

According to the invention, the maximum permissible motor current can be varied depending on the driving situation, in particular by the control software. As a result, in certain situations, the dynamics can be increased in an advantageous manner, while in other situations, the burden on the electrical system and the loss of comfort for the driver are avoided.

• – Der Fahrzeuggeschwindigkeit und/oder der Fahrzeugbeschleunigung; beispielsweise kann beim Ausrollen vor der Ampel bei niedriger Geschwindigkeit v und schwacher negativer Beschleunigung a der Maximalstrom I_max niedrig gewählt werden (zum Beispiel 30 A), da hier die Ansprüche des Fahrers an den Komfort großer als an die Dynamik sind.
• – Die Motordrehzahl; daraus folgt ein höherer Maximalstrom I_max bei höherer Motordrehzahl. Andererseits könnte unterhalb der Leerlaufdrehzahl ein höherer Maximalstrom I_max einen verbesserten Abwürgeschutz sicherstellen.
• – Die Fahrpedalstellung und/oder der Gradient der Fahrpedalstellung; eine hohe Fahrpedalstellung zeigt einen hohen Anspruch an Dynamik, daraus folgt ein hoher Maximalstrom I_max, wobei höchste Dynamikanforderungen bei erkanntem Kickdown vorgesehen sind.
• – Die Wählhebelstellung bzw. der gewählte Fahrmodus; eine Anpassung des Maximalstromes I_max an den Fahrmodus, wie zum Beispiel Automatik, Tipp-Modus, Sport-Modus oder dergleichen, ist denkbar. Es können zum Beispiel während des Automatikmodus 40 A als Maximalstrom und zum Beispiel im Sportmodus 50 A als Maximalstrom gewählt werden.
• – Der Fahrertyp; bei einer vorhandenen Fahrererkennung oder bei einer Fahrererkennungsadaption kann einem sportlichen Fahrer ein höherer Maximalstrom I_max zugeordnet werden als zum Beispiel einem komfortorientierten Fahrer.
• – Der Ladezustand der Batterie; falls der Ladezustand von einem XSG-Steuergerät erfasst werden kann, beispielsweise über eine CAN-Verbindung zu einem Batteriemanagement-SG, könnte bei gut geladener Batterie ein höherer Maximalstrom I_max als zum Beispiel bei einer nahezu leeren Batterie eingestellt werden.
• – Die Temperatur, wie zum Beispiel Motortemperatur, Außenlufttemperatur, Getriebetemperatur, Aktortemperatur oder dergleichen; bei der Messung oder der Modulierung dieser Temperaturen ist eine Anpassung von dem Maximalstrom I_max möglich. Falls die Aktorik hohe Temperaturen aufweist, und eine zusätzliche Erwärmung vermieden werden soll, kann vorzugsweise der Maximalstrom I_max in Abhängigkeit zum Beispiel der Aktor- und/oder Getriebetemperatur entsprechend verringert werden.
• – Die Bremspedalstellung bzw. der Bremsdruck; es ist denkbar, dass zum Beispiel bei einer Vollbremsung zum Vermeiden des Abwürgens des Motors ein sehr hoher maximaler Strom I_max vorgesehen ist.
• – Die Fahrzeugsituation; bei einer für den Fahrer nachvollziehbaren Situation, wie zum Beispiel bei einer Schaltung oder dergleichen, kann ein höherer Maximalstrom I_max eingestellt werden, als in einer für den Fahrer nicht spürbaren Situation, wie zum Beispiel bei Adaptionen, Momentennachführungen oder dergleichen.

In the following, the possible influencing variables on the maximum current and the method for calculating the maximum current are given, whereby the enumeration of the possible influencing variables should not be regarded as conclusive. The driving situation-dependent maximum current I _max may depend on one or more of the following influencing variables:

• The vehicle speed and / or the vehicle acceleration; For example, when coasting in front of the traffic light at low speed v and low negative acceleration a, the maximum current I _{max can be set} low (for example, 30 A), since here the driver's demands on comfort are greater than on dynamics.
• - The engine speed; this results in a higher maximum current I _max at higher engine speed. On the other hand, below the idle speed, a higher maximum current I _max could ensure improved stalling protection.
• - The accelerator pedal position and / or the gradient of the accelerator pedal position; a high accelerator pedal position shows a high demand on dynamics, this results in a high maximum current I _max , with the highest dynamic requirements are provided with recognized kickdown.
• - The selector lever position or the selected driving mode; an adaptation of the maximum current I _max to the driving mode, such as automatic, jog mode, sports mode or the like, is conceivable. For example, during the automatic mode, 40 A can be selected as the maximum current and, for example, in sports mode 50 A as the maximum current.
• - the driver type; In the case of an existing driver recognition or a driver recognition adaptation, a higher maximum current I _{max can be} assigned to a sporty driver than, for example, a comfort-oriented driver.
• - The state of charge of the battery; if the state of charge can be detected by an XSG control unit, for example via a CAN connection to a battery management SG, a higher maximum current I _max could be set with a well-charged battery than, for example, an almost empty battery.
• The temperature, such as engine temperature, outside air temperature, transmission temperature, actuator temperature or the like; in the measurement or the modulation of these temperatures, an adaptation of the maximum current I _{max is} possible. If the actuator has high temperatures, and an additional heating is to be avoided, preferably the maximum current I _max can be reduced correspondingly, for example, to the actuator and / or transmission temperature.
• - The brake pedal position or the brake pressure; It is conceivable that, for example, in a full braking to avoid the stalling of the engine, a very high maximum current I _{max is} provided.
• - The vehicle situation; in a driver-comprehensible situation, such as in a circuit or the like, a higher maximum current I _max can be set, as in an unnoticeable situation for the driver, such as in adaptations, torque tracking or the like.

A further development of the present invention can provide that a limitation of the time duration is preferably additionally provided, while the maximum current I _max can be set. This is particularly advantageous for the use of an electrical central release.

• 1) Es ist möglich, dass in vorbestimmten Zeitabständen vorzugsweise etwa alle 10 ms die möglichen Einflussgrößen bzw. mindestens eine Einflussgröße ermittelt werden;
• 2) An Hand der verwendeten Einflussgroßen kann dann der Maximalstrom I_max in vorbestimmten Abständen ermittelt werden. Daraus ergibt sich eine Funktion I_max = f(Einflussgrößen);
• 3) Der Maximalstrom I_max kann an die Begrenzungseinheit ausgegeben werden. Dies kann zum Beispiel der Lageregler bei einer sogenannten Software(SW)-Realisierung oder der Stromregler bei einer sogenannten Hardware(HW)-Realisierung sein.

The limitation of the current according to the invention can be effected, for example, by an analog current control, in which preferably the desired value is predetermined by the controller, and the actual value is detected, for example, by suitable sensors, such as a Hall sensor or the like. It is also possible that a model-based limitation of the manipulated variable, for example the voltage, is realized in the digital position controller. In a model-based limitation, a simultaneous online identification of the engine parameters may be advantageous. A possible expiration of the limitation can be provided as follows:

• 1) It is possible for the possible influencing variables or at least one influencing variable to be determined approximately every 10 ms at predetermined time intervals;
• 2) Based on the influencing variables used, the maximum current I _max can then be determined at predetermined intervals. This results in a function I _max = f (influencing variables);
• 3) The maximum current I _max can be output to the limitation unit. This can be, for example, the position controller in a so-called software (SW) implementation or the current controller in a so-called hardware (HW) implementation.

It is also conceivable that the sequence of the limitation according to the invention is carried out in a different manner.

When determining the maximum current I _max as a function of the influencing variables used, characteristic curves or two-dimensional characteristic maps can preferably be used. In 18 is tabulated an example of a map for determining the maximum current I _max on the basis of the pedal value and the vehicle speed shown. In the illustrated example, the maximum current I _max can be determined by linear interpolation between the influencing variables.

In the following, a further possible embodiment of the invention presented here is described, in which an increase in the reproducibility of the reel cycle in a vehicle plant is proposed.

It is known that in the vehicle plant in an EKM or an ASG system, the clutch is put into operation. It is u. a. determines the touch point and made on the basis of the detected touch point adjustment or adjustment of the coefficient of friction. In addition, the entry of the friction energy into the clutch (model), and the correct startup of the clutch line or the like can be checked by means of various parameters, such as the difference of the clutch temperature before and after the role of the roller cycle. However, this assumes that the test is run on the reel cycle under exactly reproducible boundary conditions.

In particular, the problem may arise that the reel cycle can not be driven in kickdown mode, since 4-5 upshifting can not be driven here due to the shift point selection and moreover a defined pedal value setting below the "kickdown" state in the vehicle plant on the reel cycle can not be met.

Accordingly, it is the object of the present invention to develop a suitable strategy for carrying out a reproducible roll test.

Accordingly, it is provided in the present invention that a so-called acceleration acceleration is activated during the rolling test in order to effect a faster convergence of Reibwertadaption. As long as the adaptation acceleration is activated, the kickdown detection provided by the transmission control is deactivated, ie. h., Even if a kickdown state is signaled by an external control unit or switch, the kickdown state can not be detected in the transmission control unit.

Particularly advantageous in this strategy is that the simplest possible instruction for the test on the reel cycle for the corresponding employees in the vehicle plant can be given (for example, full pedal pass) and beyond thus a simple reproducibility is given. Furthermore, it is advantageous that the switching points can be selected as in 100% full load circuits; Thus, the 4-5 upshift is still possible in the acceptable speed range.

As part of a development of the present invention can be provided that the role recognition is also coupled to other suitable signals than the adaptation acceleration; For example, an automatic role recognition can be provided.

It is also possible that in addition to the deactivation of the kickdown signal, the pedal value can be limited to the top. This especially if only small maximum speeds on the roll are possible, d. H. the 4-5 upshift should still be shifted to lower speeds.

This strategy according to the invention can preferably be used in automated manual transmissions (ASG systems).

In the following, a further possible embodiment of the invention presented here will be described, in which a correction of the engine torque is proposed in order to preferably achieve a convergence of the coefficient of friction adaptation.

In particular, in EKM and ASG systems, the coefficient of friction of the clutch and other influences can be adapted by means of the so-called friction coefficient adaptation. In this adaptation, the physical clutch torque can be calculated from the engine torque and the engine acceleration, and then the coefficient of friction is adapted on the basis of the deviation from the current actual value of the engine torque. The adaptation can only be carried out in slip phases, ie during startup and during shifts. The accuracy of the motor torque has a great influence on the adaptation. For technical reasons, it may happen that the accuracy of the engine torque is different for circuits and when driving. As a result, the coefficient of friction converges to different values during switching and when approaching.

The object of the present invention is to correct the engine torque in such a way that the coefficient of friction preferably converges to the same values for approaches and circuits.

According to the invention, it can be provided that the correction is carried out, for example, in a condition-dependent manner, that is, carried out differently in the states driving and starting. To carry out the correction, the following equation can be used: MMotKorr = MMot · Factor wherein the factor can be determined by one of the following conditions or in any combination thereof. With MMotKor the corrected engine torque is designated, and with MMot the engine torque is marked.

As conditions z. B. be provided that the factor is constant. Further, the factor may be a function of pedal value, engine speed, engine speed gradient, engine torque gradient, clutch temperature, engine temperature, and / or engine run time. It is possible that other conditions may also be provided which may then be used individually or in combination with the other conditions. The coefficient of friction adaptation may preferably use the corrected engine torque in this procedure.

Preferably, this engine torque correction strategy may be applied to electronic clutch management, automated transmission, or the like.

Hereinafter, another possible embodiment of the invention will be described, in which a suitable continuous slip control is preferably proposed as a function of the coupling load.

It is envisaged that in the case of an EKM and an ASG control a so-called continuous slip control is implemented. In this case, the clutch can be closed by way of example ramped, for example, if the following conditions for a certain time of usually 4 sec are met.

A first condition may be that the transmission speed is greater than the idle speed and the accelerator pedal is being depressed. If the accelerator pedal is operated very slightly, for example pedal value <10%, the ramping of the clutch may cause the vehicle to slow down so that the transmission speed falls below the idling speed. In this case, the clutch should be reopened to avoid stalling the vehicle. This opening of the clutch can be very uncomfortable for the driver. Furthermore, this can lead to vibrational excitations.

Accordingly, it is an object of the invention to avoid this jerk or this Anfahrschwingungen particular by a continuous slip control.

According to the invention, this can be achieved by suitably expanding or supplementing the conditions for activating the continuous-slip control. Here, the term continuous slip control is to be understood that a continuous slip on the clutch should be avoided in order to avoid damage or unnecessary heating of the clutch.

At pedal values of about 10%, however, the clutch torque and the differential speed are so low that it can hardly lead to heating and damage to the clutch. In this case, the continuous slip control may be disabled.

Accordingly, it is provided according to the invention that the continuous slip control is activated only if in addition one or more of the following conditions are met.

One possible condition may be that the clutch temperature T _clutch should be greater than the threshold clutch temperature, ie the clutch temperature exceeds the limit. It is also possible to provide as a condition that the clutch power loss is larger than a limit value of the clutch power loss, that is, the clutch power loss exceeds a predetermined threshold. In addition, it may be provided as a condition that the pedal value is greater than a limit value of the pedal value, that is, the pedal value should be greater than a predetermined limit value. Another condition may relate to the clutch torque. The clutch torque should be greater than a limit value of the clutch torque. Finally, it can be provided that the engine torque is greater than a predetermined limit value of the engine torque. Other conditions are also possible in which the endurance control should definitely be activated.

In this context, it is particularly useful to provide as a condition the query of the coupling power. For example, the continuous slip control may only be activated if the clutch power loss P _{clutch is} greater than the limit value of the clutch power loss and the transmission speed is greater than the idling speed and in addition the accelerator pedal is actuated.

This inventive strategy for the long-term slip control can be preferably used in electronic clutch management and automated manual transmission systems.

Hereinafter, a further embodiment of the invention will be described, in which a possible detection of a faulty clutch filling is preferably proposed during gear engagement.

It has been shown that the clutch should be open so that a gear can be engaged while the engine is running. In the case of a service or even at the end of the assembly line, it can happen that a serious error has occurred during filling and the hydraulic section is not filled correctly. This is then the fact that it is not possible to insert a synchronized gear while the engine is running. The maximum achievable position is the synchronous position in a synchronized gear. It is known that in such a case, only the error "gear actuator" is set to "overtemperature", but only if the gear actuator has exceeded the limit temperature.

Accordingly, a strategy according to the invention is proposed, in which a detection of a faulty clutch filling is provided. The maximum position reached when the gear is engaged can be evaluated. If this is in the range of the synchronization, in particular taking into account the elasticity of switching, and the engine is turned on, an error code can be generated, which indicates the possible defect or the incorrect filling in the hydraulic path of the clutch.

In this case, the rotation of the engine is required to produce a differential speed between the transmission input and the transmission output, whereby engagement of the gear is not possible with the clutch closed.

It has been shown that a similar blocking behavior is to be expected even when the engine is stationary, in particular by adhesive synchronizer rings, which prevent the engagement of the gear. Therefore, an engine speed may be present to avoid falsely setting this error.

This strategy can not be applied to non-synchronized gears, as non-synchronized gears can always be expected to have a tooth-tooth position and thus the error can be set incorrectly. To avoid this, for example, a suitable counter can be provided, which sets the error only after, for example, three unsuccessful gear engagements.

The inventive strategy for detecting a faulty clutch filling can preferably be used in all automated manual transmissions, in uninterruptible manual transmissions, in dual-clutch transmissions or the like.

Hereinafter, a further embodiment of the invention presented here will be described, in which a possible Notadaption of the touch point is preferably proposed when crawling the vehicle.

It is possible to provide emergency adaptation in the creep state of the vehicle in the case of the EKM / ASG software in which the software touch point is corrected only downwards under certain conditions. An important condition may be that the creep _torque is built up to a very high clutch _torque within approx. 6 seconds (0.75 to 1.0 · M _{creep_max} ). Too high a target creep torque indicates that the software touch point is far too high. Therefore, it may be useful to reduce the touch point in this situation. A detection that the software touch point is too low, so that a Tastpunkt-Notadaption is provided in the direction of higher values is not provided in the known Notadaption.

Accordingly, it is proposed according to the invention that it is preferably detected during creep by secondary effects that the software touch point is too low, and that a touch point emergency adaptation in the direction of higher touch point values can thus be carried out. The following explains how the creep strategy works in clutch control.

Before the start of creep, the engine torque can be determined for about 400 ms, and thus the average idling engine torque can be detected. The creep torque build-up can then preferably take place in two stages in ramps with different gradients. Until the moment of contact, the nominal clutch _torque M _{RSoll is built up} as quickly as possible, and then the structure is reduced. The creep torque buildup is in 19 exemplified.

By increasing the clutch torque, the engine torque can also be increased because the idle controller is trying to keep the idle speed LL speed. Now, if the difference between the current engine torque and the idle engine torque is greater than the nominal creep torque, for example, about 10 Nm (vehicle-dependent), so the target clutch torque can not be further built, but it can be maintained. Creep can then be continued with this moment. Since the actual clutch torque with which the vehicle is moved, is determined by the change in the engine torque, it can be detected on the basis of the torque build-up, if the software touch point is too low.

• – Die Erkennung eines zu tiefen Software-Tastpunktes aufgrund der Höhe des Kriechmomentenaufbaus; wenn der Software-Tastpunkt zu tief ist, kann bei dem in 19 dargestellten Momentenaufbau das physikalische Kupplungsmoment sehr schnell aufgebaut werden. Dementsprechend erhöht sich auch das Motormoment relativ schnell. Somit kann dann die Differenz zwischen dem Motormoment und dem Leerlaufmotormoment das nominelle Kriechmoment früher als bei einem normalen Software-Tastpunkt übersteigen. Demzufolge und aufgrund der festgelegten Steigungen der Rampen wird das Kriechmoment bei einem sehr kleinen Kupplungsmoment, zum Beispiel bei ca. 5 Nm, eingefroren. Über die Differenz des aktuellen Kupplungsmomentes und des nominellen Kriechmomentes kann dann auf einen zu niedrigen Software-Tastpunkt zurückgerechnet werden, und der Software-Tastpunkt dementsprechend geeignet verändert werden.
• – Ebenso wie die Höhe des eingefrorenen Kupplungsmomentes kann auch die Zeit des Momentenaufbaus als Erkennung eines zu tiefen Software-Tastpunktes vorgesehen werden. Bei einem zu tiefen Software-Tastpunkt übersteigt die Differenz von Motormoment und Leerlaufmotormoment das nominellen Kriechmoment früher als bei einem normalen Software-Tastpunkt. Dementsprechend kann die Zeit, welche für den Aufbau erforderlich ist, als Maß des zu tiefen Software-Tastpunktes verwendet werden. Dabei bedeutet eine sehr kurze Aufbauzeit, dass der Software-Tastpunkt sehr tief ist.
• – Ein weiteres Indiz, dass der Software-Tastpunkt zu tief ist, kann der Einbruch der Motordrehzahl sein. Wenn das physikalische Kupplungsmoment beim Kriechen sehr schnell aufgebaut wird, welches bei einem zu tiefen Software-Tastpunkt der Fall ist, kann der Leerlaufsregler die Motordrehzahl meist nicht schnell genug wieder einregeln, so dass die Motordrehzahl n_Mot gesenkt wird. Wenn jedoch die Motordrehzahl unter eine bestimmte Grenze sinkt, kann davon ausgegangen werden, dass der Software-Tastpunkt viel zu tief ist. Somit kann diese Erkennung als Notadaption eines zu tiefen Software-Tastpunktes herangezogen werden. Allerdings sollte diese Erkennung auf den Leerlaufregler angepasst werden, da der Leerlaufregler bei verschiedenen Fahrzeugtypen unterschiedlich reagiert.
• – Eine weitere Möglichkeit liegt darin, dass bei einem zu tiefen Software-Tastpunkt mit einem daraus folgenden sehr schnellen Aufbau des physikalischen Kupplungsmomentes die Last für den Leerlaufregler sehr schnell ansteigt. Diese Last kann mit dem Leerlaufregler durch geeignete Algorithmen erkannt werden. Wenn diese Last sehr groß ist, und der Leerlaufregler von der Kupplungssteuerung die Information „Fahrzeug kriech” erhält, kann der Leerlaufregler die Information „Last ist sehr groß” zurücksenden, oder auch wenn es möglich ist, kann dieser einen Wert über die Höhe der Last angeben. Mit dieser Information kann dann der Software-Tastpunkt entsprechend erhöhte werden.
• – Eine weitere mögliche Ausgestaltung kann vorsehen, dass die Fahrzeugbeschleunigung für die Erkennung eines zu tiefen Software-Tastpunktes herangezogen wird. Wenn das physikalische Kupplungsmoment aufgrund eines zu tiefen Software-Tastpunktes sehr schnell aufgebaut wird, kann auch das Fahrzeug sehr schnell beschleunigt werden. Somit kann der Software-Tastpunkt erhöht werden, wenn die Fahrzeugbeschleunigung einen bestimmten Wert überschreitet. Wenn die Fahrzeugbeschleunigung nicht gemessen wird, kann diese aus der Rad- oder der Tachodrehzahl sowie der Drehzahl der Getriebeeingangs- oder Getriebeausgangswelle durch Ableitung ermittelt werden. Dabei ist jedoch zu beachten, dass die Fahrzeugbeschleunigung nur bedingt für die genannte Erkennung geeignet ist, da z. B. beim Ankriechen an einem steilen Berg ebenfalls starke Fahrzeugbeschleunigungen entstehen können. Hierzu ist dann entweder eine Bergabrollerkennung notwendig oder die Fahrzeugbeschleunigung sollte als zusätzliche Erkennung zu den anderen genannten Erkennungen vorgesehen sein.

There are preferably the following recognition possibilities:

• The detection of a too low software touch point due to the amount of creep torque build-up; if the software touch point is too deep, the in 19 shown torque build the physical clutch torque are built up very quickly. Accordingly, the engine torque increases relatively quickly. Thus, then, the difference between the engine torque and the idle engine torque may exceed the nominal creep torque earlier than a normal software touchpoint. As a result, and due to the fixed slopes of the ramps, the creep torque is frozen at a very small clutch torque, for example, at about 5 Nm. The difference between the current clutch torque and the nominal creep torque can then be used to calculate back to a software soft-touch point that is too low, and the software touch point can be suitably changed accordingly.
• - As well as the height of the frozen clutch torque and the time of the torque build-up can be provided as detection of a too low software touch point. If the software probe point is too low, the difference in engine torque and idler torque exceeds the nominal creep torque earlier than a normal software touch point. Accordingly, the time required for the setup can be used as a measure of the too low software touch point. A very short setup time means that the software touch point is very low.
• - Another indication that the software touch point is too deep may be the drop in engine speed. If the physical clutch torque builds up very quickly during creep, which is the case with a too soft software touch point, the idle speed control usually can not adjust the engine speed fast enough so that the engine speed n _{Mot is} lowered. However, if the engine speed drops below a certain limit, it can be assumed that the software touch point is far too low. Thus, this detection can be used as emergency adaptation of a too low software touch point. However, this detection should be adjusted to the idle speed control because the idle speed control reacts differently for different vehicle types.
• - Another possibility is that at too low a software touch point with a consequent very rapid construction of the physical clutch torque, the load for the idle controller increases very quickly. This load can be detected with the idle controller by suitable algorithms. If this load is very large, and the idle controller receives the information "vehicle crawl" from the clutch controller, the idle controller may return the information "load is very large", or although it is possible, this can give a value about the height of the load. With this information, the software touch point can then be increased accordingly.
• - Another possible embodiment may provide that the vehicle acceleration is used for the detection of a too low software touch point. If the physical clutch torque is built up very quickly due to a too low software touch point, the vehicle can also be accelerated very quickly. Thus, the software touch point can be increased if the vehicle acceleration exceeds a certain value. If the vehicle acceleration is not measured, this can be determined from the wheel speed or the speedometer speed as well as the rotational speed of the transmission input or output shaft by derivation. It should be noted, however, that the vehicle acceleration is only partially suitable for the mentioned detection because z. B. when Ankriechen on a steep mountain also strong vehicle acceleration can arise. For this purpose, either a Bergabrollerkennung is necessary or the vehicle acceleration should be provided as additional detection to the other mentioned detections.

The manner of increasing the software touch point can be provided in various ways. The proposed emergency adaptation of the touch point, in particular in the creep state of the vehicle, can make the adaptation precise and very robust in an advantageous manner. The adaptation according to the invention can be used in all vehicles with automatic clutch.

Hereinafter, another possible embodiment of the invention presented here will be described, in which a Tastpunktermittlung is preferably provided at the belt end of a vehicle plant.

It is possible that the Tastpunktermittlung is performed at the end of the tape, in particular without test device. According to the invention, certain operating parameters can be assigned default values in the production of an ASG or an EKM control unit. Among other things, the touch point can be described with a value outside the tolerance range. A touch point determination in the factory without request by a test device can only take place if the touch point is outside the range. The test can then be carried out in the control unit. It has been shown that one bit should be present for each probe point determination. If a probe point determination succeeds, the corresponding bit can be set. The next touch point determination can only be performed if the previous one was successful. Thus, it is advantageously ensured that all three touch point determinations are performed in the correct order.

In particular, on the role of the manufacturing plant at the end of the tape normally only be driven and under no circumstances the handbrake be operated. For this reason, the handbrake can be used as a signal for the Tastpunktermittlung in an advantageous manner. This can be used to perform a volume compensation and a probe point determination if the conditions specified as examples are met:
Handbrake actuated; serves as the actual start signal
Foot brake actuated;
in an ASG system, the selector lever should be in the "N" or "D"position; if this is in the "D" position, the gear should be designed;
in an EKM system, the shift lever should be in "N";
the idle switch should be activated;
Vehicle is stationary, that is, the transmission speed is at the value "zero";
Previous probe point determination should succeed, which is recognizable by the bit.

If these mentioned conditions are preferably fulfilled, a volume compensation can first be carried out and the next step can be continued. It should be noted, however, that the third touch point determination is performed only at the end of the roll. In order to carry out a further plausibility check, it may be useful to start the third determination only when the vehicle speed is greater than a predetermined limit value, for example 100 km / h or the like. In this way it can be ensured that the third touch point determination, which represents a fine adjustment, is provided at the end.

If the above conditions are met, it can be provided that a gear is engaged. For an ASG system with selector lever in position "D" this can be done automatically. In an EKM system or an ASG system where the selector lever is in the "N" position, the driver should be alerted that the first gear is now engaged. This can be done either by one optical signal, such as flashing of the gear indicator in the vehicle display, or be provided by an acoustic signal, such as beeps with a specific Beep sequence.

The next step is the touch point determination. If the result of the probe-point determination lies within the tolerance range, the bit for the probe-point determination can be set. After correct touch point determination may preferably be given a corresponding signal and the second touch point determination can be done afterwards. The signal may be, for example, a visual signal, which is indicated by flashing in the display. If the second touch point determination fails or is not completed, a signal may appear in the display. The driver can be informed about the result of the touch point determination, for example by flashing the gear display in the display.

In the case of the touch point determination according to the invention, the abovementioned steps can be combined as desired with one another, and it is also possible for further steps to be combined therewith. In this way, a touch point determination can be performed without diagnosis.

Hereinafter, another possible embodiment of the present invention will be described, in which a strategy is proposed, which allows a start with an automated manual transmission without clutch as emergency in a Kupplungsaktorausfall.

In a vehicle, the clutch, also called the starting clutch, serves to equalize the rotational speed between the motor shaft and the transmission input shaft in order to enable starting or shifting. During the starting process, the clutch may separate the engine from the drive train during engine idling or vehicle standstill, the coupling between the engine and the drive train during the starting process (clutch slips) and the coupling of the engine with the drive train after starting (usually liable the coupling). Accordingly, the engine torque is not transmitted from the engine to the wheels at all (with the clutch fully open), partially (with a sliding clutch) or completely (with the clutch engaged). The energy dissipation, which is the difference between the engine and the transmission input shaft speed with sliding clutch, is converted into heat in the clutch.

In vehicles with an automated manual transmission, the clutch is actuated by the clutch actuator by means of an electric motor. If the clutch actuator has failed, the clutch is always closed, so that starting is no longer possible.

A specially modified automated manual transmission (ASG3) has a synchronization that allows start-up in the event of a failure of the clutch actuator. According to the invention can then be provided that the separation and the coupling of the engine is carried out by or with the drive train by two devices on the drive train. Preferably, this can be provided by the clutch and the synchronizer in the transmission. In this way, the clutch can be replaced by the synchronization in case of emergency clutch failure, if the synchronization is designed for high energy loss.

In an automated manual transmission (ASG), the synchronization is usually designed only for low energy losses. Therefore, the object is to provide the speed synchronization of the transmission output shaft against the transmission input shaft. The energy lost during the synchronization process is only the kinetic energy of the transmission input shaft when the clutch is open.

The synchronization in an automated transmission can also be called power shift clutch, which are designed for the circuits without traction interruption. This is the speed synchronization of the transmission output shaft against the transmission input shaft, however, with closed or sliding starting clutch. This means that the loss energy that occurs during synchronization must not only absorb the kinetic energy of the transmission input shafts due to their inertia, but in addition the drive energy from the engine. Consequently, the synchronizers of an automated manual transmission are designed for high performance.

As a result, the synchronization of an automated manual transmission meets all the conditions that a normal clutch should have at startup. Therefore, starting up as emergency running with the synchronization in the event of clutch actuator failure is possible in an advantageous manner.

Hereinafter, another possible embodiment of the present invention will be described, in which a strategy for avoiding inadvertent closing of the clutch in case of failure is proposed.

When using a Schraubaktors as a clutch actuator, such as an electrically operated Zentralausrücker (EZA), a mechanically operated Zentralausrücker (MZA) or the like, in the case of failure of the release bearing or blocking, an unintentional and spontaneous closing of the clutch can be done, which then possible uncontrolled rotations of the helical gear unit of the screw actuator can not be ruled out.

Accordingly, according to the invention, a correlation of the direction of rotation of the internal combustion engine and the direction of the pitch of the screw gear unit in Schraubaktoren depending on the type of clutch used to prevent inadvertent closing of the clutch in case of failure are provided. Especially with the use of Schraubaktoren with firmly connected to the release bearing disc springs (LAC, push-pull) may come in any position of the releaser to the failure of the release bearing. Consequently, it may then come to blocking the clutch actuator, since the screw gear unit of the screw actuator is driven by the plate spring or ultimately by the internal combustion engine and set in uncontrolled rotation. Even if fasteners of Schraubaktors are finally demolished, is actuated by the screw-nut principle of the releaser, which can in the worst case unintentional and uncontrolled closing of the clutch by itself.

To avoid this in an advantageous manner, it may be proposed that the direction of rotation of the motor, the pitch direction of the Schraubgetriebeeinheit the Schraubaktors and / or the coupling type must correlate to such that in the case of failure of the release bearing of the actuator is actuated such that the clutch does not close, but is opened.

The screw actuator is in 20 shown schematically, the situation is shown with blocking release bearing.

Hereinafter, another possible embodiment of the present invention is described, in which a height approach aid is preferably influenced by the coupling temperature.

For ASG and EKM controls, a so-called height approach help (HAH) is implemented. If the pedal value is above a high limit, for example 75%, and the vehicle does not move within a certain time with the gear engaged, the clutch can be easily opened to shift the engine speed to a range where the engine is more Can deliver moment. The clutch can then be completely closed again when the engine speed is greater than a predetermined threshold and the vehicle is moved and a predetermined time t _{1 has} passed. If the vehicle still does not stop at a predetermined time about t ₂ = 6 s or the engine speed is still less than a predetermined threshold, the clutch can still be closed quickly to avoid permanent slip. As a result, however, a start attempt is not possible.

Accordingly, it can be provided in the present invention that the height approach help is modified so that as possible under all circumstances, a starting operation is still possible. It is conceivable that the time threshold at which the clutch is engaged or closed, is suitably increased. In the above example, this would mean that the time t ₂ would have to be increased. If, for example, the time t _{2 is raised} to a value which is greater than 6 s, there is the possibility that the starting operation is terminated. However, there is also the danger that the clutch can be damaged by an excessive coupling temperature.

In order to solve this problem as well, the time threshold t ₂ can only be increased if the clutch temperature is less than a predetermined threshold of, for example, about 250 ° C. In this way, the coupling is not damaged in an advantageous manner and the starting process can u. U. still be terminated.

Hereinafter, a further embodiment of the present invention will be described, in which a suitable structure of the creep torque is proposed for variations in the engine torque, which is improved in terms of its comfort.

In 21 a measurement is shown in various diagrams, which indicate the structure of the creep torque in several stages due to the breakdown of the engine torque. This Kriechmomentenaufbau acts uncomfortable on the driver, because it is crucial in a Kriechmomentenaufbau that the clutch torque is built steadily, so that results in an approximately uniform vehicle acceleration. The cause of the in 21 shown Kriechmomentenaufbau lies in the fluctuations of the engine torque. The creep torque strategy in clutch control may provide that, prior to creep, the engine torque is averaged for approximately 400 ms (M _{Mot_Filt} ). If the creep starts, the creep torque with a constant ramp gradient depending on the respective vehicle at the beginning first slowly, then faster, but each with a constant pitch can be built. If the difference between the current engine torque M _{Mot_actual} and the averaged engine torque M _{Mot_Filt exceeds} a fixed nominal creep _torque M _{creep_nom} , the creep _{torque buildup} can be terminated and the creep _torque can be kept constant. For example, as in 21 illustrated, vibrations are generated in the course of the engine torque, for example by a poorly tuned idle controller, the torque build-up can be stopped when the above condition is met. If the difference of the current engine torque and the averaged engine torque again falls below the threshold of the fixed nominal creep torque, the creep torque can be further built up according to the known strategy. This process can last for a maximum of 6 s, after which the creep torque can be frozen. With this creep torque build-up in multiple stages, the driver can perceive vehicle acceleration as two thrusts during creep. The vehicle crawls and suddenly accelerates for a short time; then the vehicle crawls on. This is uncomfortable for the driver.

Accordingly, the invention proposes that the structure of the creeping torque is suitably modified in order to avoid these vehicle accelerations, and thus to make the creep torque build-up more comfortable for the driver.

It is within the scope of a development of the present invention possible that in the creep torque structure according to the invention, the gradient of the current engine torque M _{Mot_aktuell is} observed. When the gradient is normal, the clutch control expects a steadily increasing gradient. However, if the gradient of the current engine torque is negative or falls below a threshold to be determined, the engine torque can advantageously be frozen immediately if the difference of the current engine torque and the average engine torque is greater than the fixed nominal creep torque. Thereafter, this can not change in the current creep process. This inventive structure of the creep torque is in 22 shown. There, the courses of the old Kupplungssollmomentverlaufes and the new clutch desired torque curve over time are shown.

It is also possible that the idle controller is prone to severe vibration, and the condition that the difference between the current engine torque and the averaged engine torque is greater than the nominal creep torque is satisfied so fast that only a very small creep torque is built up the vehicle is moved insufficiently. In this case it can be provided that also the gradient of the engine torque is observed. If the gradient of the current engine torque is negative or falls below a threshold to be determined after the aforementioned condition has been met, it can be provided according to the invention that the desired clutch torque is built up more slowly but steadily. This speed in building the clutch desired torque can then be maintained, even if the current engine torque is a second time smaller and the creep torque would be further built in the current strategy. This procedure is in 23 indicated, where again the old and the new course of the clutch desired torque are shown over time.

As a further variant, it may be provided that the body speed is selected so as to be able to switch back to the normal body speed when the current engine torque drops, whereby the difference to the situation in the current creep strategy then is that the creeping torque structure according to the invention is not stopped but slowly increases, so that the acceleration differences are very low. This is in 24 indicated, wherein the differences of the different gradients are not too large in the new course of the desired clutch torque.

With this option, due to the continuously increasing creep torque, demolition conditions must be defined for the further development of the creep torque, so that the creep torque is not built up too high. The following conditions may be provided as termination conditions, whereby suitable combinations of these or other conditions are also possible:
Termination of the rising creep torque at a fixed maximum creep torque;
Abort after a fixed time;
Stop if the vehicle acceleration exceeds a predetermined value;
Abort when synchronization between engine speed and transmission speed is reached.

There are also other demolition conditions possible, which can then possibly be combined with the conditions already mentioned suitable.

Another strategy for establishing the creep torque may provide that the gradient of the current engine torque be observed and if the condition that the difference of the current engine torque and the averaged engine torque is greater than the nominal creep torque is satisfied at least once, and the gradient of the current one Motor torque is negative or falls below a threshold to be determined, the creep can be built from this point only very slowly, so that the change in acceleration by the driver is advantageously barely perceptible. This is in 25 indicated by the new and the old course of the clutch desired torque.

These strategies presented for. Structure of the creep torque are basically applicable to all vehicles with automatic clutch and with a torque strategy in creep torque build-up.

Hereinafter, a further embodiment of the invention presented here will be described, in which a suitable limitation of the nominal clutch torque is preferably proposed when pulling the clutch in the jerk function.

The jerk function of the clutch control is a driver warning when a driver wants to start in a too high gear, for example gear 3, 4 or 5. It has been shown that it can sometimes come when pulling the clutch to the route of the engine. The cause may be that the clutch torque is limited to the height of the engine torque due to the risk of stalling. By means of the closing function, the clutch can be closed further from the current clutch torque, in particular via a travel control with approximately 0.5 mm / s, up to the maximum clutch torque. Thus, always the higher moment is set. Accordingly, the clutch torque is at a maximum value as a function of the torque control and the travel control. By pulling it can thus be given a small Überanpressung that increases over time. If the software coefficient of friction is erroneously too high, it may happen that the overpressure due to the path control is too low and thus the motor is moving.

• 1) Wartezeit bis zum Ruckeln;
• 2) Ruckelrampen durchführen;
• 3) Zuziehen der Kupplung, wobei die Kupplung über die Wegsteuerung mit etwa 0,5 mm/s zugezogen werden kann.

In this case, the clutch torque is modulated so that the driver experiences a strong change in acceleration and is thus advised to switch back to the first gear. The jerk function can consist, for example, of three phases:

• 1) Waiting time to jerk;
• 2) perform jerking ramps;
• 3) Tightening the coupling, whereby the coupling can be tightened via the travel control with about 0.5 mm / s.

According to the invention, the situation is first recognized that the engine is moving. This is the case when the engine has an engine speed to be set or when the slip exceeds a specified limit. Similarly, the engine speed during jerking (Phase 2) can be heavily filtered or averaged and this speed stored. If the engine speed during pull-in (phase 3) exceeds this stored speed or exceeds this speed with an offset to be determined, a route of the engine can be detected.

A possible reaction to this situation may be that the limitation of the clutch torque is increased as a function of the engine speed. This can be realized with a factor set via a following map. Limitation = factor · motor torque

A possible map is in 26 shown. The factor over the engine speed is shown.

Furthermore, it is possible that the slip speed between the engine and the transmission is used instead of the engine speed. Otherwise, it is possible to proceed as in the aforementioned embodiment.

Another variant may provide that the introduced energy is detected in the clutch or the power converted to the clutch. With these detected magnitudes, a Überanpressungsfaktor can then be built using a similar map and then can proceed as in the aforementioned embodiments.

A further possibility for modifying the limitation of the nominal clutch torque may provide that after the detection of the path of the engine, preferably the clutch is pulled much faster. This closing speed can be constant, for example. Furthermore, it is possible that the closing speed may also depend on the engine speed and / or the slip. It is possible that the introduced energy is detected in the clutch and the closing speed is suitably changed depending on this energy and / or depending on the converted power at the clutch.

The aforementioned embodiments can also be suitably combined with one another in order to prevent an unintended route of the engine. The proposed strategy can preferably be used in all vehicles with automatic coupling and with a provided bucking function, ie in an EKM, ASG, USG, PSG or the like system.

Hereinafter, a further embodiment of the present invention will be described, in which an improved torque build-up after a gear change is proposed.

In one possible strategy for torque build-up after a gearshift, it may happen that the transmittable torque of the still slipping clutch exceeds that moment which transmits the clutch a short time later, if it sticks. This means that vehicle acceleration may become greater during re-engagement than immediately thereafter. This acceleration increase can be perceived by the occupants of the vehicle as unpleasant and comfort-reducing.

Accordingly, it is desirable to allow a steady build up of acceleration from zero to the target acceleration without local extremes.

• – Das Motormoment kann linear von einem Anfangswert auf das Fahrerwunschmoment gebracht werden. Der Anfangswert hängt dabei vom Gang und der Gaspedalstellung ab;
• – Parallel dazu kann das Kupplungsmoment T_cl erhöht werden. Dabei bleibt es um einen gewissen Betrag ΔT über dem Motormoment. Dieser Betrag nimmt mit abnehmendem Schlupf ab.

A possible torque build-up after a gear change, as implemented in an automated manual transmission (ASG), works essentially as follows:

• - The engine torque can be linearly brought from an initial value to the driver's desired torque. The initial value depends on the gear and the accelerator pedal position;
• - In parallel, the clutch torque T _{cl can be} increased. It remains at a certain amount .DELTA.T above the engine torque. This amount decreases with decreasing slip.

This potential strategy can cause unnecessary loss of comfort. For example, at the beginning of the torque buildup, the slip is very large. It follows that the ΔT is greater than zero and thus also greater than the clutch torque T _cl . The torque transmitted by the clutch does not grow steadily but abruptly. This in turn can stimulate vibrations in the drive train, which can make comfort noticeable noticeable.

Furthermore, it may happen that the clutch torque T _cl still rises during the slip phase above this value, which later transmits the adhesive coupling. In such a case, just before the transition from slipping to sticking, a higher torque is transmitted to the wheels than afterwards. This means that an increase in the vehicle acceleration occurs, which can be noticed by the driver comfort-reducing.

Thus, according to the present invention, the physical behavior of the drive train is represented by a simple model, which is described in US Pat 27 is indicated. The engine has the moment of inertia Je and can deliver the adjustable torque T _e , which is the torque of the internal combustion engine. The rotational speed of the motor is ω _e . The coupling can transmit a moment whose amount is less than or equal to T _cl . With slipping clutch, this amount is exactly equal to T _cl ; with an adhesive coupling this is at most T _cl . The transmittable torque of the clutch T _cl is adjustable. Between the clutch and the vehicle a translation i is provided. The rotational speed of the left side of the translation, ie the transmission _input shaft, is _denoted by ω _gbi . The vehicle has the moment of inertia J _v . From the outside, the driving resistance T _r acts on the vehicle. which is composed of the rolling resistance, the air resistance and the slope force.

The strategy of the invention may, for. B. be performed as follows. The clutch torque can preferably be brought from 0 linearly to that torque T _cl2 , which transfers the clutch immediately after the transition. At the same time, the engine torque can be controlled in such a way that the slip is reduced precisely when the clutch _torque reaches the value T _cl2 and / or the engine torque again corresponds to the driver's desired torque from this point in time.

If the engine torque can not be controlled accurately enough to implement the proposed strategy, z. B. the engine speed can also be controlled via a modulation of the clutch torque. However, care should be taken to ensure that the clutch torque increases monotonously during reconnection. This type of control strategy is a so-called feedforward control.

wobei T_e,d das vom Fahrer angeforderte Motormoment ist. Um die rechte Seite der Gleichung 1 auflösen zu können, muss T_r bekannt sein. Bei haftender Kupplung sind die Bewegungsgleichungen für den Motor und das Fahrzeug:

wobei T * / cl das von der haftenden Kupplung übertragene Moment ist. Das Eliminieren dieser Größe aus den Gleichungen 2 und 3 sowie das Auflösen nach T_r führt zu der Gleichung 4:

The torque T _cl2 can be calculated in the proposed strategy by the following equation 1:

where T _{e, d is} the requested engine torque by the driver. To resolve the right side of Equation 1, T _r must be known. For a stuck coupling, the equations of motion for the engine and the vehicle are:

in which T * / cl is the moment transmitted from the adhesive coupling. Eliminating this quantity from Equations 2 and 3 and resolving it to T _r leads to Equation 4:

From this equation we can calculate T _r , where ω. _{e is} determined by the numerical derivative of ω _e .

In the strategy _{according to the} invention, the clutch torque is increased linearly from 0 to T _cl2 during torque _build-up : T _cl = s · t Eq.5 where s is the rate of momentum construction. The duration τ is given by:

wobei T_r im betrachteten Zeitraum nicht signifikant sich ändert, d. h. nicht von t abhängt. Die rechte Seite von Gleichung 9 lässt sich durch Einsetzen der Gleichungen für T_cl(t) auflösen.The further steps or promotions of the strategy can be formulated as follows: ω _e (τ) = ω (τ) Eq. 7 T _e (τ) = T _{e, d} (τ) Eq. 8th where T _{e, d} (τ) is the driver's desired torque at the end of the torque _build-up . Here, the assumption is made that this is approximately constant; from this follows T _{e, d} (τ) = Te _{, d} (t) = Te _{, d} . The rotational speeds of The left and right sides of Equation 7 can be determined by integrating equations of motion 2 and 3:

where T _r does not change significantly in the considered period, ie does not depend on t. The right-hand side of Equation 9 can be resolved by substituting the equations for T _cl (t).

With equations 8 and 9, two conditions are associated with the time course of T _e . For T _e (t) an arbitrary approach can now be chosen that depends on exactly two parameters a and b to be determined. Preferably, a time course can be selected in which t occurs up to the second power. Additionally it can be demanded that T _e is continuous at t = 0: T _e (t) = T _e (0) + a * t + b * t ² Eq. 11

The right side of Equation 11 is inserted into the left side of Equation 8 and resolved to a: a = 2 / τ (T _{e, d} - T _e (0) - b · τ ²⁾ Eq. 12

The right-hand sides of Equations 5, 10 and 11 are inserted into Equation 9, integrated and resolved into b:

Accordingly, the clutch and the engine torque can be suitably controlled by means of equations 5 and 11. The quantities contained in the equations can be determined by consecutive evaluation of the equations 4, 1, 10, 13 and 12.

It was assumed, as already mentioned, that the driver's desired torque and the driving resistance remain approximately constant during the re-engagement. However, this does not necessarily have to be accepted. The control code is continuously updated with the current values of these variables. In this way, the target torque T _{cl2 of} the clutch, which is required to calculate the actual clutch _torque according to Equation 5, can be adjusted for each interrupt corresponding to the changed values of T _{e, d} and T _r . The same can be done with the parameters a and b, which are used to calculate the engine torque. However, with a change of T _{e, d} and / or T _r it can be assumed that the clutch slip is not reduced exactly at the expected time τ, but before or after it. This then results in the jerk in the vehicle acceleration, which, however, is so low that it does not have a comfort-reducing effect. However, the change in T _r is so small that the air resistance, rolling resistance and pitch angle of the road do not change appreciably during the short re-engagement phase.

In 27a FIG. 12 shows a flowchart of the proposed strategy, which procedure may preferably be called once per clock cycle.

Hereinafter, another embodiment of the present invention will be described, in which a suitable voltage control for closing the clutch under slippage is proposed.

When closing the clutch, it may happen, for example with a so-called slow roll-out slip, that the driver perceives at least one jerk in the long-term slip control, for example. The jerk can result from small jumps in the clutch torque curve. If the clutch is slowly closed by Wegvorgabe, due to the Abschalthysteresen in the clutch position control at the actual actual value, a step-shaped course can arise. This jerk is uncomfortable for the driver and should therefore be prevented.

Accordingly, the invention provides that for tightening the clutch preferably under slip instead of the position control voltage control is provided.

It is possible that a reduction of the turn-off hysteresis is provided in the above situations. In the clutch plate used usually very strong cogging moments are provided. The switch-off hysteresis is introduced due to these cogging moments. If now the switch-off hystereses for the above-mentioned situation are reduced, it is possible that a similar behavior is provided by the locking of the electric motor as in the staircase with large Abschaltthysteresen. However, this course is less pronounced and the flanks have a lower slope. As a result, the momentary jumps are substantially reduced, so that they are no longer perceived by the driver in an advantageous manner. In this way, the comfort when tightening the clutch for the driver is increased.

It is also possible, according to an advantageous development, for the clutch position control to be switched to a voltage control. If in these situations, in which the clutch is closed by a travel control in the slip phase, a constant voltage curve is provided, a largely continuous course of the actual position can be made possible.

There may, however, be two limitations; because the clutch plate used has the above-mentioned strong cogging moments. If the clutch actuator is moved with a very small voltage, these cogging torques can ensure that the actual position is not continuous, but also runs in small steps. However, these stages are advantageously not as pronounced as in the Abschalthysteresen. According to a further development, it can be provided that a minimum voltage is set which should not be undershot so as not to make the intended steps too large.

Another limitation may be that the clutch plate depending on the clutch used at constant voltage due to the different release force of the clutches can move at different speeds. In order to avoid this, it can be provided according to the invention that the change in the actual position is suitably observed. If the change is too fast, the voltage can be reduced accordingly, and if the change is too slow, the voltage can be increased accordingly. Furthermore, it can be provided that the required voltage is learned by, for example, when starting the parking lock when switching the clutch actuator, which takes place, for example, by path control, the voltage waveform records and an average voltage is stored. It is also possible that other possible ways of communicating the voltage may be provided instead of a record. The learned voltage can then be converted to the corresponding ramp slope and thus any slope of the ramp of a path control can be adapted to the respective clutch.

It can be provided that shortly before the zero point, the voltage control is switched back to the position control so that the clutch actuator can not drive against the stop. The position of the re-switching may be, for example, dependent on the slope of the ramps. Since the switch is provided at almost full moment of a clutch, it can be assumed that the slippage is reduced, and the possibly caused by the position control momentary steps are not perceived by the driver.

This clutch engagement strategy of the present invention may be applied to all vehicles having a clutch control.

A further embodiment of the present invention will be described below, in which a suitable friction coefficient adaptation is preferably proposed at high clutch temperatures and clutch temperature gradients.

The physical actuating torque characteristic can be mapped to a nominal characteristic in the software. An adjustment factor can be the software coefficient of friction, also called SW-RW. The software coefficient of friction can be adapted for approaches and circuits. With the software coefficient of friction and its adaptation, short-term and long-term changes of the physical torque characteristic can be detected and adapted to the software. One factor that can greatly change the physical moment in the short term is the physical coefficient of friction. This is strongly temperature-dependent. The physical coefficient of friction increases with the temperature and drops sharply at a specific coating temperature. This is also called fading.

It can happen that the pad temperature changes very quickly during the clutch, for example during many mountain runs, and thus the software friction coefficient changes very quickly. It has been shown that the software friction coefficient adaptation is partially adapted too slowly and so it can come to very long slip phases in drives and circuits. Due to the long slip phases, the clutch can be heated even faster, so that the physical coefficient of friction changes even faster.

Accordingly, it can be provided according to the invention to accelerate the friction value adaptation at high clutch temperatures, so that the rapid changes in the physical coefficient of friction can be absorbed by the software.

In the friction coefficient adaptation, the clutch actual torque determined in the clutch control can be compared with the physical value in slip phases. The physical moment is determined from the engine torque and the engine acceleration. If the two moments assume different values, the software coefficient of friction can preferably be changed in small steps. For example, preferably a maximum of 15 increments per circuit or per journey can be used. The maximum change can be selected such that a sufficient robustness of the coefficient of friction adaptation is ensured. To speed up the friction coefficient adaptation so that rapid changes in the physical coefficient of friction can be adjusted, various possibilities are conceivable.

It is possible that upon reaching a predetermined clutch temperature and / or upon reaching a predetermined temperature gradient, the maximum value of the increments used is increased by a predetermined value.

Furthermore, it is conceivable that upon reaching a predetermined clutch temperature and / or upon reaching a predetermined temperature gradient, the maximum value of the increments used is preferably increased linearly with the clutch temperature.

It is also possible for the maximum increment to be determined in a characteristic diagram, which is dependent on the clutch temperature and / or the temperature gradient, for example, so that the friction coefficient can be changed more quickly at higher temperatures and / or higher temperature gradients.

The measures mentioned make sure that in the normal working range, in which the clutch is most of the time, the coefficient of friction adaptation is sufficiently robust, but in a range in which the coefficient of friction can vary greatly, the adaptation speed in an advantageous manner is adjusted.

The proposed Reibwertadaption can be used in particular in ASG, EKM, PSG, USG systems or the like.

Hereinafter, another possible embodiment of the present invention will be described, in which a Tastpunktkorrektur is preferably proposed using the clutch temperature.

In particular, in an electronic clutch management (EKM) or an automated manual transmission (ASG) with hydraulic control of the clutch displacements of the touch point due to temperature changes can be compensated by regular sniffing. If as a result of the driving maneuvers, such as permanent crawl or stop and go, no so-called sniffing is possible, this can be compensated for by a medium-term adaptation of the touch point. At the next snoop, which is provided at the latest when the vehicle is switched off, this change in the medium-term touch point can be withdrawn by resetting the touch point to the long-term value. By sniffing when parking the vehicle and shifts of the touch point can be compensated, so that when you turn the ignition of a defined, balanced system can be assumed.

If a system without a hydraulic link is used, as for example with the electrically operated central release (EZA), this adjustment is not possible when the ignition is switched off. Accordingly, it can be provided according to the invention that when switching off the ignition, the current clutch temperature is stored with, in particular when in the coupling, the behavior between the temperature and the displacement of the touch point is known. When switching on the ignition then the current gripping point can be corrected according to the temperature during storage. As a relationship between the coupling temperature and the Greifpunktverschiebung can be stored in the control unit preferably any mathematical function.

Another possibility according to the invention can provide that the physical control torque characteristic of the clutch is mapped to the nominal control torque characteristic stored in the software.

The working range of the control torque characteristic, that is, the distance between the touch point and the maximum torque of the clutch, is also temperature-dependent. During driving operation, this working range can be determined precisely via the touch point adaptation and the friction coefficient adaptation or more generally via a characteristic adaptation. Furthermore, the change of the working range on the temperature of the coupling is known. If this behavior is not sufficiently known, it can be recorded accordingly during driving. For example, this can be provided on the basis of the touch point and the coefficient of friction and the position of the torque maximum.

If the vehicle is turned off with a warm coupling, the information about the clutch temperature and the behavior of the working range above the clutch temperature when the ignition is switched on again with a cold clutch can be used to determine the position relative to the locking point. As a relationship between the clutch temperature and the change in the working range of the clutch, any mathematical function can be preferably stored in the control unit.

Since the position of the touch point is known, all the necessary information of the working range of the clutch together with the coefficient of friction are known in order to map the physical control torque characteristic of the clutch to the nominal actuating torque characteristic stored in the software.

The proposed touch point correction can be used in particular for each electrical Zentralausrücker for EKM, ASG, USG, PSG systems or the like.

Hereinafter, another embodiment of the present invention will be described, in which a suitable coarse positioning of a clutch actuator with incremental displacement measurement is proposed.

Especially with an uninterruptible gearbox with central release (USG-EZA) only one incremental travel sensor is provided.

Thus, only relative movements of the actuator can be detected. Thus, there may be situations where nothing is known about the position of the actuator, e.g. B. at startup or at the end of the tape.

Accordingly, a possibility is to be found, with the at least a rough information (± 1 mm) can be obtained on the current actuator position without the risk of driving the actuator against a mechanical stop so that the self-locking gear clamps afterwards.

One way of obtaining information about the current position of a clutch actuator is to measure the engagement and disengagement forces. The properties of the self-locking transmission (spring band transmission) are used. Thus, for the actuator motor, the load (engagement / disengagement) is a friction torque in the spring band transmission.

The relationship between the load on the actuator and the friction torque in the spring band transmission is utilized, in which the speed of the actuator is evaluated under the constant motor voltage. Since no large movements are possible for the first position reference (unknown position of the stops), this method is unsuitable here.

• 1. Einem echten lastabhängigen Reibmoment im Federbandgetriebe (richtungsunabhängig); und
• 2. dem aus der Last übersetzten Moment (richtungsabhängig).

However, the frictional torque in the spring band transmission also depends on the sign of the moment (or the movement) of the actuator motor. This friction torque consists of at least two components:

• 1. A true load-dependent friction torque in the spring band transmission (independent of direction); and
• 2. the torque translated from the load (directional).

Since the gearbox is self-locking, the load torque can never be greater than the friction torque. Overall, this results in a dependent of the (intended) direction of movement friction torque (Mr + and Mr-), where the difference between the values for the two directions corresponds to twice the load torque, which is proportional to the load.

In 28 is the theoretical relationship between the external force and the friction torque on the self-locking gear of the uninterruptible transmission (USG) shown, with Mr + motion against the load and Mr-: motion with the load are designated.

In addition to the possibility of triggering a movement of the actuator by defined short voltage / current surges and to evaluate their width (or in extreme cases by "shaking" to find the force equilibrium position), the friction moments on the required for starting in the respective direction of the actuator voltages roughly determined. The latter has the advantage that the measurement results obtained in this way are proportional to the load torque, ie also proportional to the external force.

When the motor is stationary, a voltage leads to a proportional current and thus also to an engine torque proportional to this voltage. Electrical voltages are achieved in this system via a pulsed full voltage (12 V). The effective output voltage is set via the ratio "voltage on" to "voltage off"; This is called pulse width modulation (PWM).

For force measurement, the motor voltage is slowly increased until movement of the actuator is detected. The actuator starts to move when the motor torque M _M α I α U becomes greater than the friction torque Mr This is done in both directions so that Mr + and Mr- can be measured. The difference between the minimum required voltages then gives a measure of the external force on the actuator.

In 29 a single force measurement is shown, with the actuator position at the top and the output voltage at the bottom as the PWM ratio.

At the in 29 The force measurement shown above requires a larger voltage (about a PWM of 32) than the movement down (about 22). This difference now gives a statement about the external force. The applied after the voltage ramps, the movement counteracting small motor voltages, serve to slow down the actuator motor. This voltage is so small (PWM = 6) that it alone can not maintain or even trigger any movement of the actuator.

In 30 is a possible sequence of force measurement shown in a flow chart. As can also be seen, the movement is not recognized immediately by a change in position by one increment. The interaction between the anchor latch and the positions where the increments are detected requires at least one movement to wait for two increments. However, a limit of four increments further reduces the uncertainty. This measurement is not very accurate. However, when a large number of measurements are taken, the force history can be detected. In 31 the results of 10 measurement series are compiled in which the force is measured at the same 21 positions. With the solid line is the course of the disengagement or

Engagement force clarified. The coordinate system for the position is set in these measurements in such a way to achieve the goal of the complete position reference.

In 31 The results of the force measurements on the coupling system are shown. In addition, the position ranges used for coarse positioning are already marked.

• (a) Werte unter –3 sind nur im stark negativen Bereich (< –4 mm) zu finden;
• (b) Im Bereich zwischen –4 mm und 0 mm treten nur kleine Messwerte zwischen –3 und +3 auf;
• (c) Werte über +3 sind nur bei positiven Positionen zu finden;
• (d) Werte über +10 sind nur bei Positionen oberhalb von +1,5 mm zu finden.

An evaluation of the measurements for coarse positioning clearly shows the strong variance in the measurement results. Nevertheless, a rough conclusion can be drawn from the measurement results on the position. Hereinafter, the ranges (a) to (d) in FIG 31 explained.

• (a) Values below -3 are found only in the strongly negative range (<-4 mm);
• (b) In the range between -4 mm and 0 mm, only small readings between -3 and +3 occur;
• (c) Values above +3 can only be found in positive positions;
• (d) Values over +10 can only be found at positions above +1.5 mm.

Using these results, a coarse positioning of the actuator can be performed. In this, the actuator scans through the system until it detects the transition from negative to positive position values. It is important to find a compromise between the speed and accuracy of this positioning. The low accuracy of the force measurement allows only a rough approach to the reference position. Only with clearly large measured values can a larger position step be carried out in order to keep the duration of the coarse positioning low.

In 32 a flow chart is shown in which the limits are specified as they are also z. B. in the uninterruptible transmission (USG) are implemented. This method finds a position that is slightly (about 1 mm) above the reference position. However, this can be compensated by a position correction at the end of the coarse positioning.

The following are examples of the course of coarse positioning.

A series of measurements are carried out to check the coarse positioning. Coarse positioning was started at different positions in the adjusted state (reference position at position 0). In 33 is a measurement of a coarse positioning with start at -6 mm shown graphically. By resetting the incremental travel sensor, the movements always start at position 0 mm and then move in the direction of the negative start position. The actual force measurement process is visible on the pairwise forward and backward movements of the actuator.

In 34 a measurement of a coarse positioning with start at +6 mm is shown graphically. The measurements shown here are extreme cases. The closer the start position of the reference run is to the reference point, the fewer steps are required to find it.

Due to the noise in the force measurement, the gradients are different for different coarse positions of the same starting position. In rare cases (less than 1 in 25 in these examinations), noise also leads to misrecognition of the position. In these cases, the subsequent fine positioning does not work. But this can be detected by the software, whereupon a new coarse positioning is started.

Overall, this method of coarse positioning is quite robust. It can preferably be used in three different combinations of clutch and EZA, if the mechanical characteristic of the clutches is changed by a mechanical defect (flatter increase of the force in the direction of positive positions), the coarse positioning functions unchanged.

It is conceivable that the actuator can lift more than 1 mm from the diaphragm spring. This force-free millimeter can be found with this coarse positioning. In addition, then the fine positioning has sufficient Wegreserven.

In push-pull couplings, this method can be used advantageously. Here a clear difference between positive and negative coupling force occurs; In addition, the proportion increases with characteristic release forces by about one millimeter. The used method of reference point search must of course be adapted to the circumstances of the respective coupling system.

Overall, with the described embodiment of the present invention, a method is given, with which a coarse determination of its reference position can be carried out in particular in a USG clutch actuator at an unknown position, without using a mechanical stop. This inventive method can be used in particular for a basic initialization, as used for example at the end of the tape. It is provided over a rough measurement of the training or engagement force on the minimum voltage for movements of the actuator motor. The reference position can initially be determined to be about +/- 0.5 mm. Afterwards, the reference position can be determined exactly by a suitable fine adjustment. This method according to the invention can also be used for other coupling systems.

In the method according to the invention can be dispensed with additional elements, such as force sensors in the system. This particular because the actuator motor itself is used for position determination. For this purpose, the motor voltage of the actuator motor can be increased twice with opposite signs in a time ramp until each movement of the actuator is detected or a maximum voltage is reached. The difference between the two voltages at which the motor begins to move is, apart from the noise, proportional to the extraction or engagement force. The course of a single force measurement is in 35 shown in a flow chart. To save time in the process, the voltage ramps can not at zero, but z. B. be started at a predetermined starting value.

In order to minimize interactions of the magnetic Ankerrastierungen and the position of the position increments, the movement of the actuator can be detected only at a minimum number of position increments, which is preferably greater than 1. For example, the limit may be 4. The braking voltage U_Brems is chosen so small that it does not hold the actuator in a rectified movement or even cause the actuator to start.

It has been found that although this force measurement is subject to heavy noise, sufficient information about the current position of the actuator is provided. The force curve in the coupling system should be sufficiently structured. There must be large proportions of about <1/6 of the range of motion with distinct levels of force. Particularly advantageous here next to a region with very little load, a sign change of the load. From this it can be decided which position change is to be made next in order to obtain more information about the absolute position of the actuator.

A further embodiment of the invention presented here will be described below, in which a plausibility check of the actuator position is preferably proposed when the control unit is switched on without the use of stops.

For clutch actuators with incremental travel measurement, such as the EZA, there can be a problem with not knowing where the clutch is when the controller is turned on. There are possibilities to store the position of the controller when the control unit is switched off, so that it is possible to continue from this position when the control unit is switched on. This assumes, however, that in the time in which the controller is turned off, no change in the position of the actuator has been made. This requirement is not met if the clutch plate has been emergency opened, for example, and the position has been changed manually.

It is possible, in the above-mentioned situation in the "ignition on" phase, always to first approach the stop in the direction "clutch close", so that a reference point for the physical coordinate system of the clutch is given.

According to the invention it can be provided that the stop for the plausibility of the position is not required. Here, the procedure defined in the EKM / ASG control is to be changed from the zero position (parking lock) to the position HUB, ie over the entire working range of the clutch control, so that a plausibility check is provided at the same time, whether a rough change of zero Position in the "ignition off" phase. It is assumed that the working range of the clutch control with the limits zero and HUB from the last phase "ignition on" is known.

For example, when the controller is turned on, the actuator may be in the park lock position (corresponding to the zero point of the clutch control workspace). This is the starting position of the travel path. It is driven a short distance in the direction "clutch close" (delta_HUB), as well as by the arranged below the diagram arrows in 36 is shown. In this case, a maximum speed to be determined is not exceeded. If the actuator is still in the same position, which corresponds to the switching off of the control unit, the actuator can not encounter a stop in this process. Consequently, the travel distance should be selected accordingly.

In 36 is a comparison of the physical coordinate system of the clutch and the coordinate system of the clutch control indicated, is indicated by the arrows arranged below the diagram of the proposed travel of the actuator for plausibility of the emergency position.

Afterwards, the actuator can be moved to the position HUB + delta_HUB, since in neutral mode the starter is actuated. This procedure should only be carried out with a limited speed, since it is possible to move against the stop, especially if the actuator position has been adjusted in the "Ignition off" state. If the actuator was in the same position when switching on the ignition as when switching off the ignition, which characterizes the normal case, the actuator is not moved against the stop. As a result, delta_HUB should be selected accordingly. Thereafter, it is possible to approach the position HUB and the normal driving mode is present. Overall, the time to proceed is not significantly extended.

A prerequisite for the method according to the invention is that the distances between the zero-point drift and the distance between stop and the position HUB are greater than all the tolerances of the hard goods, as the travel paths delta_NP and delta-HUB.

A further advantage of the method according to the invention is that more information is obtained, in particular about the position of the stops.

A further development of the method according to the invention is that during the opening of the clutch in the state "ignition on" takes place a suitable monitoring whether a stop has been approached. In this way, the process can be dispensed with in the other direction (delta_NP and delta_Hub are equal to zero). This development has the advantage that the paths below the zero point and above the position HUB need not be maintained.

Overall, it can be stated that it can be provided in the inventive method for checking the plausibility of the actuator position that, if corresponding paths are present, it is checked in both directions as to whether attacks can be approached. Furthermore, it can be provided that, if there are no paths available or if the search takes too long, it is groped open.

Possible errors can z. B. an emergency opening, improper coupling exchange, a reset by undervoltage or a reset by ISM. An adjustment when an error is detected can provide that the plausibility checks are also used.

Hereinafter, a further embodiment of the present invention will be described, in which a switchable freewheel is proposed as a load torque lock.

In particular, in self-locking gearboxes with an efficiency of over 50% could be found that the self-adjusting coefficient of friction is often too large, so then it comes to a self-locking second order.

It can then be operated with existing self-locking only against this, with no more can be done in the direction of the load. This behavior corresponds to a classic freewheel, d. H. a load torque lock, wherein the rotational movement in only one device is possible. The self-locking is essentially provided by two gears, which can be formed internally, externally, diagonally or arrow-toothed. In addition, the training with parallel, crossed and / or cut axes is possible.

The behavior described above can not be limited to one direction only. This is especially in 37 shown. Thus, in the application of a directed moment 2 to the gear of the Hemmgliedes 3, the direction of the freewheel or the self-locking can be determined. The application of the moment can be done in all conceivable forms, for example, magnetic, mechanical, electro-mechanical, piezo-mechanical or the like. Another advantage here is that low power is required for conversion and activation.

The inhibition drive described above allows by not applying the moment a normal drive train without self-locking. The so-called freewheeling cramp can be integrated in every position of a transmission power train.

In this way, by using an originally unwanted effect, namely the self-locking second order, in a self-locking gear with an efficiency over 50%, a suitable freewheel provided by an inhibiting gear.

This freewheel according to the invention can be used for example in starters, starter-generator drives, Fußkurbelstarter, actuators, clutch adjustments or the like.

Hereinafter, a further embodiment of the invention will be described in which at least one suitable cage for securing the functionality of ramp mechanisms is proposed with rolling elements.

There are various ramp release systems with rolling elements, z. B. balls in a mechanically operated Zentralausrücker (MZA) provided. Due to the non-existent form-locking forced-operation safety device, after a vibration of the entire system or after a large number of short actuating movements, the rolling elements may migrate out of the desired position. Due to the special shape of the ramps itself, it can then lead to malfunction. For example, a stiffness or a shortening of the stroke length is possible.

The operation of a mechanically actuated Zentralausrücker (MZA) is based on a wedge effect between two rotating parts with corresponding ramps. At the contact points rolling elements are preferably used. A wedge gearbox of the MZA with rolling elements is in 38 shown, wherein a contact point is unwound on the plane. In 39 is a possible basic structure of a MZA indicated.

The aforementioned problem is caused by the 40 and 41 shown graphically. In 40 different states of movement of the rolling elements between the ramps are indicated, wherein a) shows the ramp mechanism in the starting position and b) the ramp mechanism in the lifting position (operated). In c) blocking of the ramp mechanism is indicated by the rolling elements, which represents a disadvantageous Hubverkürzung.

In 41 is shown a modified ramp shape. The changed ramp shape (without stop) or a waiver of a cage between the rolling elements can not solve the above problem.

According to the invention, provision may be made for several rolling bodies to be provided at each ramp. The rolling elements can preferably be guided in a suitable cage. In this way, the aforementioned disadvantages are avoided.

The shape of the openings in the cage, which are preferably formed as axial slots or the like, can keep the distance of the rolling elements in the advancing direction in an advantageous manner. However, a movement of the rolling elements in the stroke direction is possible. Thus, the rolling elements can adjust the distance of the ramp tracks suitable. Since the distance of the rolling elements is preferably less than their rolling distance, it can be ensured that always at least one rolling element lies between the respective ramps. If the ramps continue to have no travel limit, the cage u. U. may wander as far as you like. Even if the rolling elements move, the function is not adversely affected.

In 42 the ramp mechanism is shown with the cage according to the invention, wherein in a) the starting position and in b) the maximum stroke position are indicated. It is particularly advantageous that with appropriate ramp training Huberzeugung in both directions of actuation is possible. This property can be exploited to realize a longer life of the ramp mechanism.

Claims (11)

A method for carrying out a friction value adaptation in a clutch, in particular an automated clutch, a transmission, wherein the coefficient of friction is changed depending on at least one clutch size, wherein the coefficient of friction is changed by a predetermined value, if the slip time is too short, characterized in that the coefficient of friction is increased absolutely by a certain amount and / or relatively by a certain percentage, if the friction value adaptation is not carried out due to a short slip duration in a shift and / or the slip time for the specific shift is too short for a given speed jump. A method according to claim 1, characterized in that the change in the coefficient of friction is performed in dependence on the coupling temperature. A method according to claim 2, characterized in that the speed of change of the coefficient of friction is increased at high coupling temperatures. A method according to claim 2 or 3, characterized in that in a slip phase, the determined in the clutch control clutch actual torque is compared with the actual physical value of the clutch torque. A method according to claim 4, characterized in that the actual physical value of the clutch torque from the engine torque and the engine acceleration is determined. A method according to claim 4 or 5, characterized in that the coefficient of friction is changed in a predetermined number of increments, when the actual clutch torque and the actual physical clutch torque take different values. A method according to claim 6, characterized in that upon reaching a predetermined coupling temperature, the maximum increment used is increased. A method according to claim 6 or 7, characterized in that the maximum value of the increment is changed in dependence of a characteristic field. A method according to claim 8, characterized in that the increment is increased linearly with the coupling temperature. A method according to claim 1, characterized in that below a predetermined limit value of the coefficient of friction without the evaluation of the adaptation, a specific value is taken into account. Transmission control for a transmission with an automated clutch for carrying out the method according to one of claims 1 to 10, characterized in that at least one means for detecting and evaluating clutch sizes is provided to adjust the course of Reibwertadaption accordingly.

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