DE10316446B4 - Control device and method for controlling an engine torque and a clutch torque in a drive train - Google Patents

Abstract

Method for controlling an engine torque and a clutch torque in a drive train of a vehicle with a transmission and an engine after a gear change, the clutch torque (T _cl ) being increased to a predetermined value, and at the same time the engine torque (T _e ) being controlled in such a way that The fact that the slip on the clutch is reduced is characterized in that the slip on the clutch is reduced by controlling the engine torque (T _e ) when the clutch _torque reaches the value (T cl2) and the engine torque (T _e ) from this onwards Time again corresponds to the torque requested by the driver, the engine speed for controlling the engine torque (T _e ) being regulated by changing the clutch _{torque (T cl} ).

Description

The present invention relates to a control device and a method for controlling an engine torque and a clutch torque in a drive train of a vehicle with a transmission and an engine after a gear change.

document DE 195 24 412 A1 discloses a control of the clutch torque of a separating clutch.

document EP 1 310 401 A1 discloses a strategy for controlling the engine and clutch during ASG downshifts. Speed specification motor quick engagement in the first step, then sanding

In the case of an automated manual transmission, a torque build-up can be carried out after a gear change, in which the engine torque is increased linearly from an initial value to the torque requested by the driver. The initial value depends on the gear and the position of the accelerator pedal. At the same time, the clutch torque can be increased so that it is above the engine torque, whereby it is reduced as the slip decreases. It can happen that the transferable clutch torque of the clutch that is still slipping exceeds the torque that the clutch transfers a short time later when the clutch is in the adhering state. This means that the vehicle acceleration can be greater during the re-engagement than immediately afterwards. This increase in acceleration can be perceived by the occupants of the vehicle as unpleasant and reducing comfort.

At the beginning of the torque build-up, the slip on the clutch can be very large. It follows that the clutch torque does not increase steadily, but is increased abruptly. This in turn can generate vibrations in the drive train, which can also reduce comfort.

Furthermore, it can happen that the clutch torque still rises during the slipping phase above the value which is later transmitted by the adhering clutch. In such a case, a higher torque is transmitted to the wheels shortly before the transition from slipping to sticking than afterwards. This means that the vehicle acceleration is excessive, which can also reduce driving comfort.

The invention is based on the object of proposing a control device and a method for controlling an engine torque and a clutch torque in a drive train of a vehicle in order to enable an improved torque build-up, in particular after a gear change.

In terms of the method, this object is achieved by a method for controlling an engine torque and a clutch torque in a drive train of a vehicle with a gearbox and an engine after a gear change, in which the clutch torque is increased to a predetermined value and at the same time the engine torque is controlled in such a way that that the slip on the clutch is reduced.

Accordingly, the method according to the invention realizes a torque build-up that is as constant as possible, especially after a gear change, so that the acceleration of the vehicle can be built up from zero to the target acceleration without local extremes.

wobei T_e,d das vom Fahrer angeforderte Motormoment bzw. Fahrerwunschmoment, J_v das Trägheitsmoment des Fahrzeuges, i die Übersetzung, J_e das Trägheitsmoment des Motors und T_r die Summe sämtlicher auf das Fahrzeug wirkender Momente sind.In the strategy according to the invention, the clutch torque can for example be increased from the value 0, preferably linearly, to a value which corresponds to the clutch torque immediately after the transition to sticking of the clutch. It is also possible to use other values. The value of the clutch _torque T cl2 can preferably be calculated using the following equation: $$T_{cl2}\frac{J_V{T}_{e,d}i{J}_e{T}_r}{i^2{J}_e{J}_v}$$

where T _{e, d is} the engine torque or torque requested by the driver, J _{v is} the vehicle's moment of inertia, i is the gear ratio, J _{e is} the engine's moment of inertia and T _{r is} the sum of all the moments acting on the vehicle.

According to the invention, the engine torque is controlled in such a way that the slip is reduced precisely when the clutch torque reaches the aforementioned value at the transition to the clutch sticking. From this point in time, the engine torque again corresponds to the torque requested by the driver.

According to the invention it is provided that if, for. B. the engine torque can not be controlled accurately enough, the engine speed is also controlled via a modulation of the clutch torque. However, care should be taken to ensure that the clutch torque increases monotonically during re-engagement. This type of control strategy can also be used as a so-called pre-control.

und die Bewegungsgleichung des Fahrzeuges durch folgende Gleichung

angegeben werden, wobei mit $T_{cl}^{*}$

das von der haftenden Kupplung übertragene Kupplungsmoment und mit gbi die Drehgeschwindigkeit der Getriebeeingangswelle bezeichnet sind.If the coupling is sticking, the equation of motion of the motor can be given by the following equation

and the equation of motion of the vehicle by the following equation

can be specified, with $T_{cl}^{*}$

the clutch torque transmitted by the adhesive clutch and the rotational speed of the transmission input shaft with gbi.

wobei mit

die Drehzahlbeschleunigung des Motor bezeichnet ist.Furthermore, the sum of all moments Tr acting on the drive train can be calculated using the following equation:

with

the speed acceleration of the engine is designated.

The object on which the invention is based can be achieved in terms of apparatus by a control device for a drive train of a vehicle with an engine and a transmission coupled via at least one clutch, in particular for performing the proposed method, in which at least one engine control unit and at least one transmission control unit are provided, which control the engine torque and the clutch torque in such a way that the most constant possible torque build-up is provided after a gear change.

The invention also relates to a method, a device 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 1a instructs a vehicle 1 a drive unit 2 , such as an engine or an internal combustion engine. Furthermore are in the drive train of the vehicle 1 a torque transmission system 3 and a gearbox 4th arranged. In this embodiment, the torque transmission system 3 arranged in the power flow between the engine and transmission, with a drive torque of the engine via the torque transmission system 3 to the gearbox 4th and from the gearbox 4th on the output side to an output shaft 5 and to a subordinate axis 6th as well as to the wheels 6a is transmitted.

The torque transmission system 3 is as a clutch, such. B. designed as a friction clutch, multi-plate clutch, magnetic particle clutch or torque converter lock-up clutch, wherein the clutch can be a self-adjusting or a wear-compensating clutch. The gear 4th is a uninterruptible gearbox (USG). According to the concept of the invention, the transmission can also be an automated manual transmission (ASG), which can be switched automatically by means of at least one actuator. An automated transmission is to be understood below as an automated transmission which is shifted with an interruption in tractive force and in which the shifting process of the transmission ratio is carried out in a controlled manner by means of at least one actuator.

Furthermore, an automatic transmission can also be used as the USG, an automatic transmission being a transmission with essentially no interruption of tractive force during the shifting operations and which is usually built up by planetary gear stages.

Furthermore, a continuously adjustable transmission, such as, for example, bevel pulley belt transmission, can be used. The automatic transmission can also have a torque transmission system arranged on the output side 3 , be designed like a clutch or a friction clutch. The torque transmission system 3 can also be designed as a starting clutch and / or reversing set clutch for reversing the direction of rotation and / or a safety clutch with a specifically controllable transmittable torque. The torque transmission system 3 can be a dry friction clutch or a wet friction clutch that runs in a fluid, for example. It can also be a torque converter.

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

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

The control unit 13th can control and power electronics to control the drive motor 12th of the actuator 13b contain. As a result, it can advantageously be achieved, for example, that the system is the only installation space for the actuator 13b with electronics needed. The actuator 13b consists of the drive motor 12th , such as an electric motor, the electric motor 12th Via a gear, such as a worm gear, a spur gear, a crank gear or a threaded spindle gear, to a master cylinder 11 works. This effect on the master cylinder 11 can be done directly or via a linkage.

The movement of the output part of the actuator 13b such as B. the master cylinder piston 11a , comes with a clutch travel sensor 14th detects, which detects the position or position or the speed or the acceleration of a variable 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 a hydraulic line, with the slave cylinder 10 connected. The starting element 10a of the slave cylinder is with the release means 20th , for example a release lever, operatively connected, so that a movement of the output part 10a of the slave cylinder 10 causes the release agent 20th is also moved or tilted by the clutch 3 to control transmittable torque.

The actuator 13b for controlling the transmittable torque of the torque transmission system 3 can be actuated by pressure medium, ie it can have a pressure medium transmitter and slave cylinder. The pressure medium can be, for example, a hydraulic fluid or a pneumatic medium. The actuation of the pressure medium transmitter cylinder can be carried out by an electric motor, with the as the drive element 12th provided electric motor can be controlled electronically. The drive element 12th of the actuator 13b In addition to an electromotive drive element, it can also be another drive element, for example a pressure medium-actuated drive element. Magnetic actuators can also be used to set a position of an element.

In the case of a friction clutch, the transferable torque is controlled by the friction linings of the clutch disc being pressed between the flywheel 3d and the printing plate 3b takes place in a targeted manner. About the position of the release mechanism 20th , such as a release fork or a central release, the application of force to the pressure plate 3b or the friction linings are specifically controlled, the pressure plate 3b moved between two end positions and set and fixed as required can be. One end position corresponds to a fully engaged clutch position and the other end position corresponds to a fully disengaged clutch position. A position of the pressure plate, for example, can be used to control a transferable torque, which is, for example, less than the currently applied engine torque 3b are controlled, which lies in an intermediate area between the two end positions. The clutch can by means of the targeted control of the release means 20th be fixed in this position. However, it is also possible to control transferable clutch torques that are defined above the currently pending engine torques. In such a case, the currently pending engine torques can be transmitted, with the torque irregularities in the drive train being dampened and / or isolated in the form of, for example, torque peaks.

For controlling the torque transmission system 3 sensors are also used that at least temporarily monitor the relevant variables of the entire system and deliver the state variables, signals and measured values required for control that are processed by the control unit, with a signal connection to other electronic units, such as engine electronics or electronics Anti-lock braking system (ABS) or anti-slip regulation (ASR) can be provided and can exist. The sensors detect, for example, rotational speeds such as wheel speeds, engine speeds, the position of the load lever, the throttle valve position, the gear position of the transmission, an intention to shift and other vehicle-specific parameters.

The 1a shows that a throttle position sensor 15th , an engine speed sensor 16 as well as a speedometer sensor 17th Can be used and measured values or information to the control unit 13th hand off. The electronics unit, such as a computer unit, of the control electronics 13a processes the system input variables and sends control signals to the actuator 13b further.

The transmission is designed as a step change transmission, for example, with the transmission steps by means of a shift lever 18th can be changed or the transmission by means of this shift lever 18th is actuated or operated. Furthermore is on the shift lever 18th 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 13th forwards. The sensor 19a is linked to the transmission and detects the current gear position and / or an intention to shift. The detection of the intention to change gear using at least one of the two sensors 19a , 19b can be done in that the sensor is a force sensor that acts on the shift lever 18th acting force detected. Furthermore, the sensor can also be designed as a displacement or position sensor, the control unit recognizing an intention to switch from the change in the position signal over time.

The control unit 13th is at least temporarily in signal connection 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 sends control or regulation commands to the at least one actuator 13b issues. The drive motor 12th of the actuator 13b , for example an electric motor, receives from the control unit, which controls the clutch actuation, a manipulated variable as a function of measured values and / or system input variables and / or signals from the connected sensors. This is in the control unit 13th 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 characteristic diagrams.

The control unit 13th has advantageously implemented a torque determination unit, a gear position determination unit, a slip determination unit and / or an operating state determination unit or it is in signal connection with at least one of these units. These units can be implemented as hardware and / or software by means of control programs, so that the torque of the drive unit can be determined by means of the incoming sensor signals 2 of the vehicle 1 , the gear position of the transmission 4th as well as the slip, which is in the area of the torque transmission system 3 prevails and the current operating status of the vehicle 1 can be determined. The gear position determination unit determines on the basis of the signals from the sensors 19a and 19b the currently engaged gear. There are the sensors 19a , 19b articulated to the shift lever and / or to internal actuating means, such as a central shift shaft or shift rod, and detect these, for example the position and / or the speed of these components. Furthermore, a load lever sensor 31 on the load lever 30th , such as on an accelerator pedal, which detects the load lever position. Another sensor 32 can function as an idle switch, ie when the load lever is actuated 30th or the accelerator pedal is this idle switch 32 switched on and with the load lever not actuated 30th it is switched off, so that this digital information can be used to identify whether the load lever 30th is operated. The load lever sensor 31 detects the degree of actuation of the load lever 30th .

The 1a shows next to the load lever 30th and the sensors associated therewith, a brake actuator 40 for actuating the service brake or the parking brake, such as a brake pedal, a handbrake lever or a hand- or foot-operated actuating element 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 as. B. configured as a switch, which detects that the brake actuator 40 operated or not operated. With the sensor 41 A signaling device, such as a brake light, can be in signal connection, which signals that the brake is actuated. This can be done both for the service brake and for the parking brake. The sensor 41 can, however, also be designed as an analog sensor, with such a sensor, such as a potentiometer, the degree of actuation of the brake actuation element 41 determined. This sensor can also be in signal connection with a signaling device.

A possible embodiment of the invention presented here is described below, in which a suitable coefficient of friction adaptation is provided in order to increase shift comfort in particular.

ein Vergleich des Kupplungsmomentes mit dem Motormoment durchgeführt werden. In Abhängigkeit dieses Vergleichsergebnisses kann dann eine Anpassung der Kupplungskennlinie durchgeführt werden, bei der der Reibwerte geeignet korrigiert wird. Als Schlupfphasen können die Anfahrphase sowie die Einkuppelphase oder auch andere Phasen nach den Schaltungen ausgewertet werden. Außerdem können zusätzliche Randbedingungen vorgesehen sein, wie zum Beispiel

Es ist möglich, dass auch noch andere Randbedingungen verwendet werden.In the case of a possible adaptation of the coefficient of friction, in particular in the case of an electronic clutch management (EKM) and an automated gearbox (ASG), the engine torque balance can be used during the slip phases

a comparison of the clutch torque with the engine torque can be carried out. Depending on this comparison result, the clutch characteristic curve can then be adapted, in which case the coefficient of friction is appropriately corrected. The start-up phase as well as the coupling phase or other phases after the shifts can be evaluated as slip phases. In addition, additional boundary conditions can be provided, such as, for example

It is possible that other boundary conditions are also used.

1. 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.
2. 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 coefficient of friction adaptation:

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

As a result, a tactile point that is too high can be taught-in, particularly during commissioning in the vehicle factory, for example due to an incomplete venting. This behavior compensates for the coefficient of friction adaptation, in which the coefficient of friction drops sharply; especially partially to the lower limit. Air can escape from the system during the first few switching operations. The touch point can be corrected relatively quickly using 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. This can result in an extremely uncomfortable shifting process for the driver, i.e. hard disengaging and engaging.

Due to the aforementioned properties of the described coefficient of friction adaptation, namely to act relatively quickly when the coefficient of friction is too high, while the adaptation reacts much more slowly or not at all in the opposite direction, the invention presented here can provide that the coefficient of friction adaptation is 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.

Certain criteria can be established for this purpose, such as the following:

• - The coefficient of friction adaptation could not be carried out or evaluated due to a too short slip period during a shift, and / or
• - The slip period is too short for the specific gear shift at a given speed jump.

It is also possible to determine other criteria in order to further improve the coefficient of friction 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 coefficient of friction adaptation can adapt itself out of the so-called deadlock situation described at the beginning. This means that the coefficient of friction can increase accordingly. The usual slip duration can be achieved due to the increasing coefficient of friction, so that a normal evaluation of the coefficient of friction adaptation can advantageously take place.

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

Another possible embodiment of the present invention is described below, in which a suitable path measurement, in particular in the case of clutch actuators, is proposed.

It has been shown that an incremental distance measurement does not have the restriction of a mechanically limited measuring range compared to an absolute distance sensor. Accordingly, the present invention is based on the object of proposing, preferably, an absolute path measurement for clutch actuators that does not have the aforementioned disadvantages.

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

In this way, an absolute path measurement can be provided, for example with an absolute angle sensor, without the disadvantage of a limited maximum path or an insufficiently low resolution of the sensor signal. A purely incremental distance measurement can thus be avoided by using an absolute angle sensor.

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

A further embodiment of the present invention is described below 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 for the position of the clutch actuator to be initialized by means of an incremental path measurement; this in particular to compensate for mechanical changes to the respective coupling.

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

For example, a kink or the like in the force characteristic can be used as a reference point. This kink occurs in a single clutch, for example, when the disc spring is relieved in such a way that the release system lifts off the disc spring. For example, in the case of a combination clutch, in particular for an uninterruptible gearbox (USG), a similar kink can be provided in the course of the force characteristic, for example in the case of a central position at the transition between the two characteristics of the individual clutches. It is also possible to use other reference points. For example, force zero crossings can also be used as reference points.

It can advantageously be provided that the search for and the determination of such a reference point simultaneously represent what is known as software sniffing. Mechanical displacements, such as can occur due to wear and tear or also due to thermal effects, can advantageously be compensated for in the control of the actuator. On the other hand, by measuring the mechanical changes, for example when “sniffing”, a conclusion can possibly also be drawn about the thermal condition or the degree of wear of the clutch. It is essential that the reference point is not determined exactly, but that a method is to be specified which appropriately recognizes a position on the characteristic curve in a reproducible manner.

To carry out what is known as software sniffing, the following can be taken into account; for the electric motor of the actuator, a change in the release force can be represented as a change in the frictional torque in the almost self-locking release gear. This frictional torque can also depend on the sign of the external force. A movement against the force results in significantly greater friction than a movement in the direction of the force. For forces that are not too small, there can be an almost linear relationship between the force and the frictional torque.

unabhängig von der Größe der Geschwindigkeit sein. Bei jeweils nahezu linearen Kraftkennlinien vor und nach dem Knick ergibt sich jeweils ein nur von der Federkonstante ((F / x) und den Motoreigenschaften abhängiger Wert für v / x. Da bei einer Inkrementalwegmessung die Geschwindigkeit das genaueste Signal des Aktors ist, kann diese Geschwindigkeitsänderung in vorteilhafter Weise einfach gemessen und erkannt werden.This so-called sniffing and position comparison can be carried out in different ways. For example, a speed measurement with a constant motor voltage can be provided. When the actuator moves with constant voltage, the speed results from the frictional torque that results from self-locking actuator gears. A change in the clutch spring force can cause a change in the frictional torque and thus a change in the speed of the actuator. The measurement can preferably take place in a movement against the clutch spring force and with an increasing spring constant (F / x). The change in speed over the location (v / x) can be achieved if the inertia is neglected $\left({}^{‵}n\text{J}>\text{MR}\right)$

be independent of the size of the speed. With almost linear force characteristics before and after the kink, the result is a value for v / x that only depends on the spring constant ((F / x) and the motor properties can be easily measured and recognized in an advantageous manner.

In 3 a possible load profile and a theoretical profile of the actuator speed resulting therefrom is shown, with the direction of movement of the actuator during the measurement being indicated by an arrow above the two diagrams. Disturbances, which can occur periodically during the motor rotation, are suppressed in the simplest way with this presented measuring method. It is only necessary that exactly one motor revolution is to be selected as the basis for determining v / x as x. For example, the reference point can then be determined by the intersection of speed ramps. It is also possible to use other reference points.

The speed measurement can initially only take place with the same angular position or phase position of the actuator, where xi with i is to be seen as the index for the motor revolution. From the speeds vi determined in this way, it can then be recognized from which engine revolution the system is in the area of the hard spring. This results in the following limit value: $$\text{Section}\left(\text{vi vi}-\text{1}\right)>\text{limit}$$

If it is recognized from which engine revolution the system is in the area of the hard spring, this means that the measurements of vi and vi-1 limit the first time 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. The position interval (xi-1 to xi-2) in which the reference point is provided can be located between these intervals.

From these intervals and the speeds measured at their limits, the point at which the actuator speed changes can be determined in an advantageous manner. This corresponds to the reference point. This determination of the reference point is graphically in 4th indicated, the determined position, that is to say the reference point, being indicated as the point of intersection between the two straight lines shown.

Any noise that may be present can be minimized by evaluating a measurement process with different phase positions with regard to the motor rotation and then by calculating the mean value. Another possible improvement in the determination of the reference point can result from the fact that Implausible results (for example if the reference position is outside the interval xi to xi-3) are suppressed.

Known problems, such as a strong spread of the speeds at certain phase positions, can also be taken into account in a further development of the present invention when determining the reference point or when calculating the mean value.

In 5 a possible algorithm is shown in a table. In this case, it can be provided that a set of data is used for each phase position in which the reference position is sought. First, under the keyword "phase position", it is checked which phase position the data belong to. Under the keyword "revolution", a counter can be used for the motor revolutions since the start of the measurement, which is used to determine the position. Then a 4-value array can be used under the keyword “Vbuffer”, in which the last speeds are saved. Finally, under the keyword "Position", the determined position can be output, which is used for the sequence control.

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

It is possible that the so-called "software sniffing" is carried out in the same driving states in which the hydraulic sniffing is carried out. This is possible in particular when the engine is at a standstill and while driving, when no gear changes are made. This ensures that the position is regularly compared. In the case of an uninterruptible gearbox in particular, this procedure corresponds to sniffing for the starting clutch, so that the same driving conditions can be sniffed here if the powershift gear is not being used.

This presented method can be used in particular for electrical central release devices or for operating a single or two coupled clutches. Overall, it can be stated that this method can be used as so-called “software sniffing” for every actuator with a mechanical release system, 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 the case of mechanical end stops. For example, the position adjustment can also be replaced or supplemented by a suitable touch point adaptation.

The so-called “software sniffing” is particularly advantageous with regard to the search for reference points, since touch point adaptation cannot be used as a substitute strategy, especially on long motorway journeys over a long period of time.

• Die Eigenschaften permanent erregter Gleichstrommotoren bewirkt, dass sich bei gegebener Spannung U eine Drehzahl einstellt, die in einem linearen Zusammenhang zum Bremsmoment steht.

$$n{n}_0\frac{n_0}{M_A\left(U\right)}{M}_R\text{ f}\text{r }{M}_A\left(U\right)M_R$$

wobei

n:

Motordrehzahl

n0:

Leerlaufdrehzahl

MR:

Das auf den Motor wirkende Reibmoment

MA(U):

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

The proposed position initialization is based on the following considerations:

• The properties of permanently excited DC motors mean that a given voltage U sets a speed that is linearly related to the braking torque.

$$n{n}_0\frac{{\mathrm{<figure-callout\; id="0"\; label="n"\; filenames="DE000010316446B4\_0050.png,DE000010316446B4\_0052.png"\; state="\{\{state\}\}">n</figure-callout>}}_0}{{\mathrm{M.}}_{\mathrm{A.}}\left(U\right)}{\mathrm{M.}}_{\mathrm{R.}}\text{f}\text{r}{\mathrm{M.}}_{\mathrm{A.}}\left(U\right){\mathrm{M.}}_{\mathrm{R.}}$$

in which

n:

Engine speed

n0:

Idle speed

MR:

The frictional torque acting on the motor

MA (U):

Starting torque (torque at n 0) The following applies: M _A (U) U

Therefore, in a system in which the inertia of the motor has a negligible influence (ńJ M _R ), conclusions can be drawn from the speed of the motor about the load. A position-dependent load thus leads to a position-dependent speed.

With a linear load profile, this also leads to a linear speed profile: $$\frac{{\mathrm{M.}}_{\mathrm{R.}}}{x}\text{c}Onst.\frac{v}{x}cOnst.$$

• Keine integrierte Kompensationsfeder

For example, measurements with a modified clutch actuator (ASG) can be carried out on a special test setup. The following can be provided as possible modifications of the clutch actuator, although the list is 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; this again results in approx. 12.5 µm / increment at the (theoretical) disc spring.

For example, different loads (springs, detents, stops or combinations thereof) can be placed on the clutch actuator. The structure can preferably consist of a basic frame made of aluminum profile rails, to which the control unit with the clutch actuator is attached. Different load modules can be positioned on the aluminum profiles. The tappet of the clutch actuator is firmly connected to a rod which is mounted so that it can be moved over the basic structure. The locking profile and the stop disks (also for the springs) can be freely positioned and also detached on this rod. This means that the positions of the different loads can be varied and several loads can be combined with one another. The test setup is shown schematically in 9 shown.

For the investigations presented here, two spring loads are used in each case, with both springs being loaded in the same direction.

The (former) compensation spring with 5.9 N / mm; it is used to simulate a small load, as it is also required in the real system in order to ensure the safe functioning of the release system.

The clutch spring with 19.9N / mm; it simulates the force of the disc spring.

With the test setup presented here, the in 10 load curve shown can be generated. The theoretical load profile includes the force distribution of the spring 1 and 2 as well as the accumulated course. Due to the (almost) self-locking actuator gear, the position-dependent load results in a position-dependent (and direction-of-movement-dependent) frictional torque on the motor. The frictional 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 gear. In addition, the sign is reversed. In the measurements shown here, the weak spring starts at a position of 0 increments (with increasing force with a negative position) and the strong spring starts at 100 increments. The motor voltages are specified via the pulse width ratio. A value of 255 corresponds to approx. 10V. The values given in the control unit are preferably used for the speed.

It turns out that when using an incremental displacement sensor, the actuator speed is the most precise signal for the state of the actuator. An (externally attached) measurement of the motor current is also significantly less accurate. First of all, the result of a series of measurements is to be shown here with the actuator with different voltages (PWM ratios ± 30, ± 35, ± 40, ± 45, ± 50, ± 55 and ± 60) in both directions under the load shown above is driven. In 11 Measured speeds are shown over the position for different motor voltages.

The effect of the load on the speed of the actuator can be seen much better with movements against the load (here with negative speeds) than with the load. This behavior can be explained by the properties of the self-locking gear. There is a clear disorder with a period of 40 increments = 1 motor revolution. When running towards the load (under the periodic disturbance) a constant change in speed over the location (v / x) is recognized regardless of the size of the speed. The reproducibility of the measurements is very high.

In 12th the reproducibility of the speed measurements is shown. In the measurement shown, the actuator is moved back and forth eight times in both directions with the same voltage (PWM ratio = ± 35). The measurement curves agree very well. Even the small speed fluctuations with a period of 3 to 4 increments are reproducible. Presumably these are due to the magnetic armature detent. When evaluating the speed measurements, the calculation of the speed change per revolution is first taken into account. The first consideration is to calculate the difference in speed to the exact 40 increments beforehand for each position during a measurement, which is carried out according to the following equation: $$\mathrm{B.}\left(x\right)v\left(x\right)v\left(x40\right)(\text{f}\text{r movements at negative speed}.$$

Since the speed is not measured at every position due to the temporal sampling of the signals, the speed profile should be suitably interpolated. If the measurement data shown above are evaluated in this way, the result is the 13th . Here, too, the movement is from right to left. The signal can be recognized by the noise but is so bad that it cannot be used in this form for position adjustment. The noise is advantageously reduced by averaging. However, the position resolution may deteriorate.

In the following, the calculation of the knee position by linear interpolation of the speeds is considered. If the speed is always considered at the same phase position of the motor rotation during a measurement (Pos mod 40 = const.), A signal that is freed from the modulation by the motor rotation can result. Depending on the selected phase position, there is a slightly different speed curve here. Apart from some noise, the only difference between the speed curves is a speed offset. In 14th this is shown for two phases.

Two speed profiles are shown with different phase positions. The next step for the evaluation is to recognize from which engine revolution you are on the hard spring. This can be recognized, for example, by simply exceeding the limit value (| v _i v _i 1 | C).

The limit value C can depend on the properties of the motor and the hardness of the spring. In 15th a rough position reference can be seen from the changes in speed. This means that the kink position of the load characteristic can be found with an accuracy of ± 40 increments.

A better position specification can be obtained if the last speed measurements are used as support points for straight lines and their intersection points are calculated. The measurements of v _i and v _i limit the first interval on the hard spring, which is the position interval found above. The measurements of v _{i 2} and v _{i 3} limit the last interval before the hard spring. In 16 the determination of the exact reference position is shown graphically as an example.

The position determined in this way is still not very precise for each individual phase position (in the best case to 20 increments); but it is possible to use this method for several phase positions in a series of measurements. However, a simple averaging of all possible positions that can be determined in this way is not yet sufficient to significantly improve the accuracy. Errors caused by noise can partially be recognized and suppressed. It can happen that the determined intersection of the two straight lines lies _{outside the interval x i} ; x _{i 3.} Such a case can easily be recognized and the result for averaging can be suppressed.

It has also been shown that the determination of the speed for phase positions that are on the flanks of the periodic disturbance is subject to very large fluctuations due to the interpolation. For this reason, care is taken right at the start of the measurement not to consider these phase positions. This is done by evaluating the acceleration (v / t). Phases that border too high an acceleration in the first motor revolution are not considered any further. The mean value formation from the values selected in this way for the reference point determination then achieves an accuracy of approx. 10 to 12 increments (or less than 5.5 increments 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 approximately correct because they are set by hand. What is important here is the reproducibility of the detection of the same position. This is in 17th shown, the determined positions of the reference point being indicated. Apart from motor voltages that are too small (<40) or too high (> 55), the position at which the bend in the force characteristic is found is independent of the motor voltage. The spread of the determined positions is in a range of ± 3 to ± 5.5 increments. The individual measurements took between 3s (for U = -35) and 1.6s (U = -55). If the measurements are restricted to the necessary range of motion, measurement times between 1.6s (U = -35) and 0.6s (U = -55) are obtained. These can be further reduced.

It can be assumed that the hardness of the spring, the starting position of which is being sought, affects the accuracy of the position determination. However, there is the possibility that a harder spring enables a more precise determination of the reference position. The influence of the voltage on the recognized position should also be taken into account. Furthermore, the effects of the start position (relative to the reference point) and the duration of the measurement should be taken into account. Since the actuator temperature particularly influences the idling speed and the starting torque of the electric motor, this should also be taken into account.

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

A further embodiment of the invention presented here is described below, in which a possible monitoring of an automated clutch, in particular during the crawling process, is proposed.

Particularly in the case of continuous crawling without vehicle movement, e.g. B. due to an obstacle or on an incline, a high proportion of frictional energy can arise on the slipping clutch, in particular on the clutch disc. This is associated with increased wear on the clutch linings. As a result of the heating or the temperature gradients that develop in the clutch unit, the transmission properties of the clutch also change. Under certain circumstances it is possible that the torque transmitted by the clutch can increase sharply. This can lead to undesired vehicle movements and the associated safety-critical situations.

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

It is possible, for example, that the clutch temperature is preferably determined using a temperature sensor or a suitable model, e.g. B. from the frictional energy on the clutch disc and cooling effects, such as speed-dependent heat transfer and heat conduction, is determined. The driver warning can preferably be activated when a predetermined limit value is exceeded, i.e. the driver can be warned accordingly acoustically, visually and / or mechanically, for example.

It is also possible, with the aid of a time measurement, to determine the period of time in which a predefined slip speed threshold is exceeded when the creep state is activated. If the duration preferably exceeds a predetermined limit value, the driver warning can also be activated.

Within the scope of another development, it can be provided that when the creep state is activated, the engine torque and its time derivative are checked or observed. Suitable limit values can be defined for both quantities. Since the engine torque can briefly rise sharply due to loads such as an air conditioning system or power steering, two limit value conditions should be met to distinguish between the influence of loads and temperature-related changes in the clutch transmission properties. Further conditions are also possible in order to recognize the distinction between the different influences. If the engine torque exceeds the predetermined limit value while the derivation remains below its limit value, the driver warning and / or the adaptation of the clutch transmission properties can 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 value, the driver warning and / or the adaptation of the clutch transmission properties can also be activated.

Various options can be used as driver warnings, which can be used alternatively or in combination with one another. It is conceivable that an acoustic warning, for example by a beeper or an announcement, is provided. Furthermore, a visual warning can also be provided by a flashing warning light or by a text message in a display. A mechanical warning, for example by jerking the vehicle, is also possible. The jerk can be caused, for example, by an oscillatory movement of the clutch actuator. The main frequency band should be in the range of the vehicle's natural frequency so that a clearly noticeable jolt is caused with small torque amplitudes. The jerking can preferably 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 occurs, for example by pressing the accelerator pedal or the brake.

The software-based adaptation of the clutch transmission properties can in principle be carried out in two ways. For example, by correcting the touch point and / or correcting 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 clutch transmission properties to be analyzed for the clutch unit in question, and for the corresponding procedure to be selected, preferably 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 creeping. The torque transmitted by the clutch can then be reduced. The clutch can be opened completely from a second limit value. If the vehicle is to be kept at a standstill with the clutch slipping on an incline, however, opening the clutch causes the vehicle to roll back. If the clutch temperature exceeds a limit value, a warning jolt can be activated depending on the gear when starting off.

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

A further embodiment of the invention presented here is described below, in which a variable current limitation, preferably of a clutch actuator for an XSG transmission, for example, is proposed.

It has been shown that an optimization of the control 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, among other things, high dynamics; this means the shortest possible adjustment times. In order to meet this requirement, the actuators in known clutches have a high power requirement, which is reflected in relatively high currents in electromechanical actuation. In this way, when electric motors start up, very high currents (50 to 100 A) sometimes occur briefly if no suitable measures are taken to limit the current. However, these high starting currents are only accepted to a limited extent, since they can cause inadmissible loads on the vehicle electrical system and impairment of comfort, such as flickering lights, actuator noises or the like.

By changing the design of the electric motor, for example with regard to a higher resistance of the winding or the supply lines, higher currents can be avoided. However, on the one hand, this can also increase the energy loss and, on the other hand, the dynamics of the motor can be negative to be influenced. As a result, for example, limiting the motor current to a fixed value is more advantageous.

According to the invention, it is provided that the maximum permissible motor current can be varied depending on the driving situation, in particular by the control software. In this way, the dynamics can advantageously be increased in certain situations, while in other situations the load on the vehicle 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 Imax niedrig gewählt werden (zum Beispiel 30 A), da hier die Ansprüche des Fahrers an den Komfort größer als an die Dynamik sind.
• - Die Motordrehzahl; daraus folgt ein höherer Maximalstrom Imax bei höherer Motordrehzahl. Andererseits könnte unterhalb der Leerlaufdrehzahl ein höherer Maximalstrom Imax 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 Imax, wobei höchste Dynamikanforderungen bei erkanntem Kickdown vorgesehen sind.
• - Die Wählhebelstellung bzw. der gewählte Fahrmodus; eine Anpassung des Maximalstromes Imax 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 Imax 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 Imax 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 Imax möglich. Falls die Aktorik hohe Temperaturen aufweist, und eine zusätzliche Erwärmung vermieden werden soll, kann vorzugsweise der Maximalstrom Imax 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 Imax vorgesehen ist.
• - Die Fahrzeugsituation; bei einer für den Fahrer nachvollziehbaren Situation, wie zum Beispiel bei einer Schaltung oder dergleichen, kann ein höherer Maximalstrom Imax eingestellt werden, als in einer für den Fahrer nicht spürbaren Situation, wie zum Beispiel bei Adaptionen, Momentennachführungen oder dergleichen.

The possible influencing variables on the maximum current and the method for calculating the maximum current are specified below, although the list of possible influencing variables is not to be regarded as conclusive. The driving situation-dependent maximum current Imax can 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 lights at low speed v and weak negative acceleration a, the maximum current Imax can be selected to be low (for example 30 A), since the driver's demands on comfort are greater here than on dynamics.
• - The engine speed; this results in a higher maximum current Imax at a higher motor speed. On the other hand, below the idle speed, a higher maximum current Imax could ensure improved stall protection.
• - The accelerator pedal position and / or the gradient of the accelerator pedal position; a high accelerator pedal position indicates a high demand for dynamics, resulting in a high maximum current Imax, with the highest dynamics requirements being provided when a kickdown is detected.
• - The selector lever position or the selected driving mode; an adaptation of the maximum current Imax to the driving mode, such as automatic, inching mode, sport mode or the like, is conceivable. For example, 40 A can be selected as the maximum current in automatic mode and 50 A as the maximum current in sport mode, for example.
• - The type of driver; With an existing driver recognition or with a driver recognition adaptation, a sporty driver can be assigned a higher maximum current Imax 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 Imax could be set when the battery is well charged than, for example, when the battery is almost empty.
• The temperature, such as engine temperature, outside air temperature, transmission temperature, actuator temperature or the like; When measuring or modulating these temperatures, it is possible to adapt the maximum current Imax. If the actuator system has high temperatures and additional heating is to be avoided, the maximum current Imax can preferably be correspondingly reduced as a function, for example, of the actuator and / or transmission temperature.
• - The brake pedal position or the brake pressure; it is conceivable that, for example, in the event of emergency braking, a very high maximum current Imax is provided to prevent the motor from stalling.
• - The vehicle situation; in a situation that is understandable for the driver, such as a shift or the like, a higher maximum current Imax can be set than in a situation that is not perceptible to the driver, such as for example adaptations, torque adjustments or the like.

A further development of the present invention can provide that a limitation of the time period is also preferably provided during which the maximum current Imax can be set. This is particularly advantageous for the use of an electric central slave cylinder.

• 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 Einflussgrößen kann dann der Maximalstrom Imax in vorbestimmten Abständen ermittelt werden. Daraus ergibt sich eine Funktion Imax = f(Einflussgrößen);
• 3) Der Maximalstrom Imax 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 current can be limited according to the invention, for example, by analog current regulation, in which the setpoint is preferably specified by the controller and the actual value is recorded, for example, via 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 implemented in the digital position controller. In the case of a model-based limitation, simultaneous online identification of the motor parameters can be advantageous. A possible expiry 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 at predetermined time intervals, preferably approximately every 10 ms;
• 2) The maximum current Imax can then be determined at predetermined intervals on the basis of the influencing variables used. This results in a function Imax = f (influencing variables);
• 3) The maximum current Imax 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 limitation according to the invention is carried out in a different manner.

When determining the maximum current Imax as a function of the influencing variables used, characteristic curves or two-dimensional characteristic diagrams can preferably be used. In 18th an example of a characteristic diagram for determining the maximum current Imax on the basis of the pedal value and the vehicle speed is shown in a table. In the example shown, the maximum current Imax can be determined by linear interpolation between the influencing variables.

Another possible embodiment of the invention presented here is described below, in which an increase in the reproducibility of the roller cycle in a vehicle plant is proposed.

It is known that the clutch is put into operation in an EKM or ASG system in the vehicle plant. Among other things, the touch point is determined and the coefficient of friction is adjusted or adjusted on the basis of the determined touch point. In addition, various parameters, such as the difference in clutch temperature before and after the role of the roller cycle, the entry of frictional energy into the clutch (model), and the correct start-up of the clutch section or the like can be checked. However, this assumes that the test is run on the roller cycle under exactly reproducible boundary conditions.

In particular, the problem can arise that the roller cycle cannot be run in kickdown mode, since the 4-5 upshift cannot be operated here due to the selection of the shift point and, moreover, a defined pedal value position below the kickdown state in the vehicle plant is not maintained on the roller cycle can be.

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

Accordingly, it is provided in the present invention that a so-called adaptation acceleration is activated during the roller test in order to bring about a faster convergence of the coefficient of friction adaptation. As long as the adaptation acceleration is now activated, the kickdown detection provided by the transmission control is deactivated, i.e. even if a kickdown state is signaled by an external control unit or switch, the kickdown state cannot be detected in the transmission control unit.

It is particularly advantageous with this strategy that the simplest possible instruction for the test on the roller cycle can be given to the corresponding employees in the vehicle plant (for example, fully depress the pedal) and, moreover, simple reproducibility is ensured. It is also advantageous that the switching points can be selected as in 100% full load shifts; thus the 4-5 upshift is still possible in the acceptable speed range.

In the context of a further development of the present invention, it can be provided that the role recognition is also coupled to other suitable signals than the adaptation acceleration; for example, automatic role recognition can be provided.

It is also possible that, in addition to deactivating the kickdown signal, the pedal value can also be limited upwards. This is particularly the case when only small maximum speeds are possible on the roll, i. H. the 4-5 upshift should be shifted to lower speeds.

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

Another possible embodiment of the invention presented here is described below, 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 with EKM and ASG systems, the coefficient of friction of the clutch and other influences can be adapted by means of the so-called coefficient of friction adaptation. With this adaptation, the physical clutch torque can be calculated from the engine torque and the engine acceleration, and the coefficient of friction is then adapted based on the deviation from the current actual value of the engine torque. The adaptation can only be carried out in slip phases, i.e. when starting up and when shifting. The accuracy of the motor torque has a major influence on the adaptation. For technical reasons it can happen that the accuracy of the motor torque is different when switching and when starting. The consequence of this is that the coefficient of friction converges to different values during gear changes and when driving off.

The present invention is based on the object of correcting the engine torque in such a way that the coefficient of friction preferably converges to the same values when starting and shifting.

wobei der Faktor an Hand einer der nachfolgenden Bedingungen bzw. in beliebigen Kombinationen daraus bestimmt werden kann. Mit MMotKorr ist das korrigierte Motormoment bezeichnet, und mit MMot ist das Motormoment gekennzeichnet.According to the invention, it can be provided that the correction is carried out, for example, as a function of the state, ie it is carried out differently in the driving and starting states. The following equation can be used to perform the correction: $$\text{MMotKorr}=\text{MMot}*\text{factor}$$

whereby the factor can be determined on the basis of one of the following conditions or in any combination thereof. The corrected engine torque is identified with MMotKorr, and the engine torque is identified with MMot.

As conditions, for. B. be provided that the factor is constant. Furthermore, the factor can be a function of the pedal value, the engine speed, the engine speed gradient, the gradient of the engine torque, the temperature at the clutch, the temperature in the engine and / or a function of the engine running time. It is possible that other conditions are also provided, which can then be used individually or in combination with the other conditions. In this procedure, the coefficient of friction adaptation can preferably use the corrected engine torque.

This strategy can preferably be used for correcting the engine torque in an electronic clutch management system, in an automated transmission or the like.

Another possible embodiment of the invention is described below, in which a suitable continuous slip control is proposed, preferably as a function of the clutch load.

It is provided that a so-called continuous slip control is implemented in an EKM and an ASG control. The clutch can be closed in a ramp-shaped manner by specifying the path, for example, if the following conditions are met for a certain time, usually 4 seconds.

A first condition can be that the transmission speed is greater than the idle speed and the accelerator pedal is depressed. If the accelerator pedal is only pressed very slightly, for example pedal value <10%, the ramp-shaped closing of the clutch can cause the vehicle to slow down, so that the transmission speed falls below the idling speed. In this case, the clutch should be opened again to avoid stalling the vehicle. This opening of the clutch can be very uncomfortable for the driver. Furthermore, this can lead to vibrations.

Accordingly, it is an object of the invention to avoid this jolt or these starting vibrations, in particular by means of 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. The term continuous slip control is to be understood as meaning that constant slip on the clutch should be avoided in order to avoid damage or unnecessary heating of the clutch.

At pedal values of around 10%, the clutch torque and the differential speed are so low that the clutch can hardly be heated up or damaged. In this case, the continuous slip control can be deactivated under certain circumstances.

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

One possible condition can be that the clutch temperature T _clutch should be greater than the limit value clutch temperature, ie the clutch temperature exceeds the limit value. It is also possible that the condition provided is that the clutch power loss is greater than a limit value for the clutch power loss, ie the clutch power loss exceeds a predetermined limit value. In addition, it can be provided as a condition that the pedal value is greater than a limit value of the pedal value, ie the pedal value should be greater than a predetermined limit value. Another condition can relate to the clutch torque. The clutch torque should be greater than a limit value for 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 continuous slip control should definitely be activated.

In this context, it is particularly useful to include the query of the clutch performance as a condition. 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 the accelerator pedal is also actuated.

This strategy according to the invention for continuous slip control can preferably be used in electronic clutch management and automated manual transmission systems.

Another embodiment of the invention is described below, in which a possible detection of faulty clutch filling, preferably when shifting into gear, is proposed.

It has been shown that the clutch should be open so that a gear can be engaged with the engine running. During service or at the end of the assembly line, it can happen that a serious error was made during filling and that the hydraulic section was not filled correctly. This manifests itself in the fact that it is not possible to engage a synchronized gear while the engine is running. With a synchronized gear, the maximum achievable position is the synchronized position. It is known that in such a case only the "gear actuator" "overtemperature" error is set, 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 faulty clutch filling is provided. The maximum position reached when the gear is engaged can be evaluated. If this is within the range of synchronization, especially taking into account the switching elasticity, and the motor is switched on, an error code can be generated that indicates the possible defect or incorrect filling in the hydraulic section of the clutch.

It is necessary to turn the engine in order to produce a differential speed between the transmission input and the transmission output, which means that it is not possible to engage the gear when the clutch is closed.

It has been shown that a similar locking behavior can be expected even when the engine is not running, in particular due to sticky synchronizer rings which prevent the gear from being engaged. Therefore, in order to prevent this error from being set incorrectly, there may be an engine speed.

This strategy cannot be used on non-synchronized gears, since a tooth-tooth position must always be expected with non-synchronized gears and the error can therefore be set incorrectly. To avoid this, a suitable counter can be provided, for example, which only sets the error after, for example, three unsuccessful gear engagement.

The strategy according to the invention for recognizing faulty clutch filling can preferably be used in all automated manual transmissions, in uninterruptible manual transmissions, in double clutch transmissions or the like.

A further embodiment of the invention presented here is described below, in which a possible emergency adaptation of the touch point is proposed, preferably when the vehicle crawls.

It is possible to provide an emergency adaptation with the EKM / ASG software when the vehicle is crawling, in which the software touch point is only corrected downwards under certain conditions. An important condition can be that the creep torque is built up to a very high clutch torque within approx. 6 seconds (0.75 to 1.0 * Mcriech_max). If the target creep torque is too high, this indicates that the software touch point is much too high. It can therefore make sense to reduce the tactile point in this situation. A detection that the software touch point is too low, so that a touch point emergency adaptation in the direction of higher values is provided, is not provided in the known emergency adaptation.

Accordingly, it is proposed according to the invention that, preferably when crawling, secondary effects recognize 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 the clutch control.

Before the beginning of the creep, the engine torque can be determined for about 400 ms, and thus the mean idle engine torque can be recorded. The build-up of creep torque can then preferably take place in two stages in ramps with different gradients. _{The clutch setpoint torque M RSoll is built up} as quickly as possible up to the touch torque, and then the buildup is reduced. The creep torque build-up is in 19th shown as an example.

The engine torque can also be increased by increasing the clutch torque, since the idle controller tries to maintain the idle speed LL speed. If the difference between the current engine torque and the idling engine torque is greater than the nominal creep torque, for example approx. 10 Nm (depending on the vehicle), the target clutch torque can no longer be built up, but can be maintained. With this moment, the creeping can be carried out further. Since the actual clutch torque with which the vehicle is moved is determined via the change in the engine torque, the torque build-up can be used to determine whether 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.

The following detection options are preferably available:

• - The detection of a software touch point that is too deep due to the level of the creep torque build-up; if the software touch point is too deep, the in 19th The physical clutch torque can be built up very quickly. Accordingly, the engine torque also increases relatively quickly. Thus, the difference between the engine torque and the idle engine torque can exceed the nominal creep torque earlier than at a normal software touch point. As a result, and due to the defined gradients of the ramps, the creep torque is frozen at a very low clutch torque, for example at approx. 5 Nm. The difference between the current clutch torque and the nominal creep torque can then be used to calculate back to a software touch point that is too low, and the software touch point can be appropriately changed accordingly.
• - Just like the level of the frozen clutch torque, the time of the torque build-up can also be provided as a detection of a software touch point that is too low. If the software touch point is too low, the difference between the engine torque and the idling motor torque exceeds the nominal creep torque earlier than with a normal software touch point. Accordingly, the time required for construction can be used as a measure of the software tactile point that is too deep. A very short set-up time means that the software touch point is very deep.
• - Another indication that the software touch point is too low can be the drop in the engine speed. If the physical clutch torque is built up very quickly when crawling, which is the case when the software touch point is too low, the idle speed controller can usually not regulate the engine speed again quickly enough, so that the engine speed n _{Mot is} reduced. However, if the engine speed drops below a certain limit, it can be assumed that the software touch point is far too low. This detection can thus be used as an emergency adaptation of a software touch point that is too low. However, this detection should be adapted to the idle controller, since the idle controller reacts differently in different vehicle types.
• - Another possibility is that if the software touch point is too low, the physical clutch torque builds up very quickly as a result, the load on the idle speed controller increases very quickly. This load can be detected with the idle controller using suitable algorithms. If this load is very large and the idle controller receives the information vehicle is creeping from the clutch control, the idle controller can send back the information Load is very large, or, if possible, it can indicate a value about the level of the load. With this information, the software touch point can then be increased accordingly.
• Another possible embodiment can provide that the vehicle acceleration is used to detect a software touch point that is too low. If the physical clutch torque is built up very quickly due to a software touch point that is too low, the vehicle can also be accelerated very quickly. Thus, the software touch point can be increased when the vehicle acceleration exceeds a certain value. If the vehicle acceleration is not measured, it can be derived from the wheel or tachometer speed and the speed of the transmission input or output shaft. It should be noted, however, that the vehicle acceleration is only conditionally suitable for the detection mentioned, since z. B. when crawling up a steep mountain can also cause strong vehicle accelerations. For this purpose, either rolling downhill detection is necessary or the vehicle acceleration should be provided as an additional detection to the other detections mentioned.

The type of elevation of the software touch point can be provided in various ways. The proposed emergency adaptation of the touch point, especially when the vehicle is crawling, can advantageously make the adaptation precise and very robust. The adaptation according to the invention can be used in all vehicles with an automatic clutch.

Another possible embodiment of the invention presented here is described below, in which a detection point is preferably provided at the end of the line of a vehicle plant.

It is possible that the detection of the touch point is carried out at the end of the line, in particular without a test device. According to the invention, certain operating parameters can be assigned standard values during the manufacture of an ASG or an EKM control device. Among other things, the touch point can be described with a value outside the tolerance range. A touch point determination in the factory without a 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 a bit should be present for each detection of the touch point. If a touch point determination is successful, the corresponding bit can be set. The next contact point determination can only be carried out if the previous one was successful. This advantageously ensures that all three touch point determinations are carried out in the correct sequence.

• Handbremse betätigt; dient als eigentliches Startsignal Fußbremse betätigt;
• bei einem ASG-System sollte der Wählhebel in der Stellung N oder D sein; falls dieser in der Stellung D ist, sollte der Gang ausgelegt sein;
• bei einem EKM-System sollte der Schalthebel in N sein;
• der Leerlaufschalter sollte betätigt sein;
• Fahrzeug steht, das heißt die Getriebedrehzahl ist auf dem Wert Null ;
• vorherige Tastpunktermittlung sollte erfolgreich sein, welches an dem Bit erkennbar ist.

In particular, it is normally only possible to drive on the roll of the manufacturing plant at the end of the line and in no case the handbrake can be operated. For this reason, the handbrake can be used in an advantageous manner as a signal for determining the touch point. A volume compensation and a contact point determination can thus be carried out if the conditions given by way of example are met:

• Handbrake applied; serves as the actual start signal foot brake operated;
• with an ASG system, the selector lever should be in position N or D; if this is in position D, the gear should be disengaged;
• with an EKM system, the gearshift lever should be in N;
• the idle switch should be activated;
• The vehicle is stationary, i.e. the transmission speed is zero;
• Previous contact point determination should be successful, which can be recognized by the bit.

If these conditions are preferably met, a volume equalization can first be carried out and the process can be continued with a next step. It should be noted, however, that the third contact point determination is only carried out at the end of the roll. In order to carry out a further plausibility check, it can be useful to begin 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 contact point determination, which represents a fine adjustment, is provided at the end.

If the aforementioned conditions are met, provision can be made for a gear to be engaged. In the case of an ASG system with the selector lever in position D, this can be done automatically. In the case of an EKM system or an ASG system in which the selector lever is in the N position, the driver should be made aware that first gear is now being engaged. This can be provided either by an optical signal, such as for example flashing of the gear indicator in the vehicle display, or by an acoustic signal, such as for example beeping with a specific beeping sequence.

The next step is the determination of the touch point. If the result of the contact point determination is within the tolerance range, the bit for the contact point determination can be set. After the touch point has been correctly determined, a corresponding signal can preferably be given and the second touch point determination can then take place. The signal can be a visual signal, for example, which is indicated by flashing in the display. If the second contact point determination is negative or has not yet been completed, a signal can appear in the display. The driver can be informed about the result of the tactile point determination, for example by the flashing of the gear indicator in the display.

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

A further possible embodiment of the invention presented here is described below, in which a strategy is proposed which enables starting with an automated manual transmission without a clutch as an emergency run in the event of a clutch actuator failure.

In a vehicle, the clutch, also called the starting clutch, serves to equalize the speed between the motor shaft and the transmission input shaft in order to enable starting or shifting. During the start-up process, the clutch can separate the engine from the drive train when the engine is idling or when the vehicle is at a standstill, the coupling between the engine and the drive train during the start-up process (clutch slips) and the coupling of the engine to the drive train after start-up (usually stuck the clutch). Correspondingly, the engine torque is not transmitted at all (when the clutch is fully disengaged), partially (when the clutch is sliding) or completely (when the clutch is sticking) from the engine to the wheels. The energy loss that arises from the difference between the engine speed and the transmission input shaft speed when the clutch is sliding 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 gearbox (ASG3) has a synchronization system that enables an emergency start-up if the clutch actuator fails. According to the invention it can then be provided that the separation and coupling of the motor from or with the drive train is carried out by two devices on the drive train. This can preferably be provided by the clutch and the synchronizing device in the transmission. In this way, the clutch can be replaced by the synchronization in an emergency in the event of a clutch actuator failure, if the synchronization is designed for high energy losses.

In an automated gearbox (ASG), the synchronization is usually only designed for small energy losses. Therefore, the task 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 gearbox can also be called a power shift clutch, which is designed for shifting without interruption of tractive power. This involves synchronizing the speed of the transmission output shaft with the transmission input shaft, but with the starting clutch closed or sliding. This means that the energy loss that occurs during synchronization must not only absorb the kinetic energy of the transmission input shafts due to their inertia, but also the drive energy from the motor. As a result, the synchronizers of an automated manual transmission are designed for high performance.

As a result, the synchronization of an automated manual transmission fulfills all the conditions that a normal clutch should have when starting off. Therefore, starting as an emergency run with synchronization in the event of a clutch actuator failure is possible in an advantageous manner.

Another possible embodiment of the invention presented here is described below, in which a strategy for avoiding unintentional closing of the clutch in the event of a fault is proposed.

When using a screw actuator as a clutch actuator, such as an electrically operated central clutch release (EZA), a mechanically operated central clutch release (MZA) or the like, in the event of failure of the release bearing or jamming, the clutch can close unintentionally and spontaneously, which then possible uncontrolled rotations of the screw gear unit of the screw actuator cannot 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 screw actuators depending on the type of coupling used should be provided to avoid unintentional engagement of the coupling in the event of a fault. Particularly when screw actuators are used with disc springs firmly connected to the release bearing (LAC, push-pull), the release bearing can fail in any position of the release. As a result, the clutch actuator can block because the screw gear unit of the screw actuator is driven by the disk spring or ultimately by the internal combustion engine and is set in uncontrolled rotation. Even if the fastening elements of the screw actuator are finally torn off, the screw-nut principle actuates the release mechanism, which in the worst case can result in unintentional and uncontrolled closing of the clutch.

In order to avoid this in an advantageous manner, it can be proposed that the direction of rotation of the motor, the pitch direction of the helical gear unit of the screw actuator and / or the clutch type must correlate so that in the event of failure of the release bearing, the actuator is actuated in such a way that the clutch does not close, but is opened.

The screw actuator is in 20th shown schematically, showing the situation when the release bearing is blocked.

Another possible embodiment of the present invention is described below, in which a height starting aid is preferably influenced by the clutch temperature.

A so-called height starting aid (HAH) is implemented for ASG and EKM controls. If the pedal value is above a high limit value, for example 75%, and the vehicle does not move within a certain time with a gear engaged, the clutch can be easily opened in order to shift the engine speed into a range in which the engine can operate more Moment can deliver. The clutch can then be fully closed again when the engine speed is greater than a predetermined threshold and the vehicle is being moved and a predetermined time t _{1 has} passed. If the vehicle is still not stationary for a predetermined time, for example t ₂ = 6s, or the engine speed is still less than a predetermined threshold, the clutch can still be closed quickly in order to avoid permanent slippage. However, this means that a start-up attempt is not possible.

Accordingly, it can be provided in the present invention that the height start-up aid is modified in such a way that, if possible, a start-up process is still possible under all circumstances. It is conceivable that for this purpose the time threshold at which the clutch is pulled or closed is appropriately increased. In the aforementioned example, this would mean that time t ₂ would have to be increased. If, for example, the point in time t _{2 is raised} to a value which is greater than 6 s, there is the possibility that the start-up process will be terminated. However, there is also the risk that the coupling can be damaged by too high a 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 limit value of, for example, approximately 250 ° C. In this way, the clutch is advantageously not damaged and the start-up process can possibly still be ended.

A further embodiment of the present invention is described below, in which a suitable structure of the creep torque in the event of fluctuations in the engine torque is proposed, which is improved in terms of its comfort.

In 21 a measurement is shown in various diagrams, which indicate the build-up of the creep torque in several stages due to the collapse of the engine torque. This build-up of creep torque has an uncomfortable effect on the driver, because it is crucial with a creep torque build-up that the clutch torque is built up steadily, so that an approximately uniform vehicle acceleration results. The cause of the in 21 The creep torque shown here lies in the fluctuations in the engine torque. The creep torque strategy in the clutch control can provide that the engine torque is averaged for about 400 ms before creeping (M _{Mot_Filt} ). When the creep begins, the creep torque can be built up slowly at the beginning with a constant ramp slope, depending on the respective vehicle, but then more quickly, but with a constant slope in each case. If the difference between the current engine torque M _{Mot_aktuell} and the averaged engine torque M _{Mot_Filt exceeds} a fixed nominal creep torque MKriech_nom, the creep torque build-up can be ended and the creep torque can be kept constant. For example, if, as in 21 shown, 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-mentioned condition is met. If the difference between the current engine torque and the averaged engine torque falls again below the threshold of the fixed nominal creep torque, the creep torque can be built up further according to the known strategy. This process can take a maximum of 6 s, after which the creep torque can be frozen. With this creep torque build-up in several stages, the driver can perceive the vehicle acceleration as two thrusts when crawling. The vehicle crawls off and suddenly there is acceleration for a short time; then the vehicle crawls on. This is inconvenient for the driver.

Accordingly, it is proposed according to the invention that the build-up of the creep torque be modified in a suitable manner in order to avoid these vehicle accelerations and thus make the creep torque build-up more comfortable for the driver.

In the context of a further development of the present invention, it is possible for the gradient of the current engine torque M _{Mot_aktuell to} be observed in the case of the creep torque build-up according to the invention. If 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 specified threshold, the engine torque can advantageously be frozen immediately if the difference between the current engine torque and the averaged engine torque is greater than the fixed nominal creep torque. After that, this can no longer change with the current creeping process. This structure of the creep torque according to the invention is shown in FIG 22nd shown. There the courses of the old clutch target torque curve and the new clutch target torque curve are shown over time.

It is also possible that the idle speed controller tends to oscillate strongly, and the condition that the difference between the current engine torque and the averaged engine torque is greater than the nominal creep torque is met so quickly that only a very small creep torque is built up, as a result of which the vehicle is only being moved inadequately. In this case it can be provided that the gradient of the engine torque is also observed. If the gradient of the current engine torque is negative or falls below a threshold to be specified after the aforementioned condition has been met, it can be provided according to the invention that the desired clutch torque is further built up more slowly but steadily. This speed when building up the clutch target torque can then be maintained, even if the current engine torque becomes smaller a second time and the creep torque would continue to build up in the current strategy. This approach is in 23 indicated, the old and the new course of the desired clutch torque being shown there again over time.

As a further variant, it can be provided that the build-up speed is selected in such a way that the normal build-up speed can be switched back to when the current engine torque drops, whereby the difference to the situation in the current creep strategy is that the creep torque build-up according to the invention is not stopped but increases slowly so that the differences in acceleration are very small. This is in 24 indicated, whereby the differences in the different gradients in the new course of the desired clutch torque are not chosen to be too large.

• Abbruch des steigendes Kriechmomentes bei einem fest vorgegebenen maximalen Kriechmoment;
• Abbruch nach einer fest vorgegebenen Zeit;
• Abbruch, wenn die Fahrzeugbeschleunigung einen vorbestimmten Wert überschreitet;
• Abbruch, wenn zwischen Motordrehzahl und der Getriebedrehzahl Synchronisierung erreicht ist.

With this option, due to the steadily increasing creep torque, termination conditions must be defined for the further build-up of the creep torque so that the creep torque is not too high is being built. The following conditions can be provided as termination conditions, whereby suitable combinations of these or other conditions are also possible:

• Cancellation of the increasing creep torque at a fixed maximum creep torque;
• Termination after a fixed time;
• Termination if the vehicle acceleration exceeds a predetermined value;
• Abort if synchronization is achieved between the engine speed and the gearbox speed.

Further termination conditions are also possible, which can then possibly be suitably combined with the conditions already mentioned.

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

The strategies presented for building up the creep torque can be used in principle for all vehicles with an automatic clutch and with a torque strategy for building up the creep torque.

A further embodiment of the invention presented here is described below, in which a suitable limitation of the clutch setpoint torque is suggested, preferably when the clutch is tightened in the jerk function.

The jerk function of the clutch control is a driver warning if a driver wants to start in a gear that is too high, for example gear 3, 4 or 5. It has been shown that the motor can sometimes run away when the clutch is closed. The cause can be that the clutch torque is limited to the level of the engine torque due to the risk of stalling. Using the tightening function, the clutch can be closed from the current clutch torque, in particular via a path control at about 0.5 mm / s, up to the maximum clutch torque. This means that the higher torque is always set. Accordingly, the clutch torque is at a maximum value depending on the torque control and the path control. As a result of the tightening, there can therefore be a slight overpressure, which increases over time. If the software coefficient of friction is incorrectly too high, it can happen that the overpressure is too low due to the position control, and the motor turns away.

• 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 in such a way that the driver experiences a sharp change in acceleration and is thus advised to shift down to first gear. The jerk function can, for example, consist of three phases:

• 1) waiting time until jerking;
• 2) perform jerk ramps;
• 3) Closing the coupling, whereby the coupling can be closed with about 0.5 mm / s via the travel control.

According to the invention, the situation is first recognized that the engine is turning away. This is the case when the engine has a specified engine speed or when the slip exceeds a specified limit. The motor speed can also be heavily filtered or averaged during the jerking (phase 2) and this speed can be saved. If the motor speed exceeds this stored speed when pulling in (phase 3) or exceeds this speed with an offset to be defined, the motor can be recognized as turning.

A possible reaction to this situation can be that the limitation of the clutch torque is increased as a function of the engine speed. This can be achieved with a factor that is set using a following map. $$\text{Limitation}=\text{factor}*\text{Engine torque}$$

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

It is also possible that the slip speed between the engine and the transmission is used instead of the engine speed. Otherwise, you can proceed as in the aforementioned embodiment.

Another variant can provide that the energy introduced into the clutch or the power converted at the clutch is recorded. With these recorded variables, an overpressure factor can then be built up with the aid of a similar characteristic map, and the procedure can then continue as in the aforementioned embodiments.

Another possibility for modifying the limitation of the clutch setpoint torque can provide that after the detection of the turning away of the engine, the clutch is preferably applied much more quickly. This closing speed can be constant, for example. It is also possible that the closing speed can also depend on the engine speed and / or on the slip. It is possible for the energy introduced into the clutch to be recorded and for the closing speed to be suitably changed as a function of this energy and / or as a function of the power converted at the clutch.

The aforementioned configurations can also be combined with one another in a suitable manner in order to prevent the motor from unintentionally turning away. The proposed strategy can preferably be used in all vehicles with an automatic clutch and with an intended jerk function, that is to say in an EKM, ASG, USG, PSG or similar system.

Another embodiment of the present invention is described below, in which an improved torque build-up is proposed after a gear change.

In a possible strategy for building up torque after a gear change, it can happen that the torque that can be transmitted from the clutch that is still slipping exceeds the torque that the clutch transmits a short time later when it is stuck. This means that the vehicle acceleration can be greater during the re-engagement than immediately afterwards. This increase in acceleration can be perceived by the occupants of the vehicle as unpleasant and reducing comfort.

Accordingly, it is desirable to enable the acceleration to build up steadily 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 T über dem Motormoment. Dieser Betrag nimmt mit abnehmendem Schlupf ab.

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

• The engine torque can be brought linearly from an initial value to the torque requested by the driver. The initial value depends on the gear and the position of the accelerator pedal;
• - At the same time, the clutch torque T _{cl can be} increased. It remains a certain amount T above the engine torque. This amount decreases with decreasing slip.

This possible strategy can cause unnecessary loss of convenience. For example, at the beginning of the torque build-up, the slip is very large. It follows from this that the T is greater than zero and is therefore also greater than the clutch torque T _cl . The torque transmitted by the clutch does not grow steadily, but rather abruptly. This in turn can stimulate vibrations in the drive train, which can reduce comfort.

Furthermore, it can happen that the clutch torque T _cl rises above this value during the slipping phase, which is later transmitted by the adhering clutch. In such a case, a higher torque is transmitted to the wheels shortly before the transition from slipping to sticking than afterwards. This means that there is an excessive acceleration of the vehicle, which the driver can notice as reducing comfort.

Accordingly, according to the present invention, the physical behavior of the drive train is represented by a simple model, which is shown in 27 is indicated, described. The motor has the moment of inertia Je and can deliver the adjustable moment Te, which is the torque of the internal combustion engine. The speed of rotation of the motor is e. The clutch can transmit a torque, the amount of which is smaller or is equal to T _cl . When the clutch is slipping, this amount is exactly equal to T _cl ; in the case of an adhesive coupling, this is at most T _cl . The transmittable torque of the clutch T _cl is adjustable. A gear ratio i is provided between the clutch and the vehicle. The speed of rotation of the left-hand side of the transmission, i.e. the transmission input shaft, is denoted by gbi. The vehicle has the moment of inertia Jv. The driving resistance Tr., Which is composed of the rolling resistance, the air resistance and the downhill force, acts on the vehicle from the outside.

The strategy of the invention can e.g. B. be carried out as follows. The clutch torque can be brought from 0, preferably linearly, to that torque T _cl2 which the clutch transmits immediately after the transition to sticking. 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 from this point in time corresponds again to the torque requested by the driver.

If the engine torque cannot be controlled precisely enough to implement the proposed strategy, z. B. the engine speed can also be controlled by modulating the clutch torque. However, care should be taken to ensure that the clutch torque increases monotonically during re-engagement. This type of control strategy is a so-called pre-control.

wobei T_e,d das vom Fahrer angeforderte Motormoment ist. Um die rechte Seite der Gleichung 1 auflösen zu können, muss Tr 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:

With the proposed strategy, the moment T _cl2 can be calculated using the following equation 1: $$T_{cl2}\frac{J_V{T}_{e,d}i{J}_e{T}_r}{i^2{J}_e{J}_v}$$

where T _{e, d is} the engine torque requested by the driver. To solve the right hand side of equation 1, Tr must be known. With a sticky coupling, the equations of motion for the engine and the vehicle are:

in which $T_{cl}^{*}$

is the torque transmitted by the adhesive coupling. Eliminating this quantity from equations 2 and 3 and solving for _Tr leads to equation 4:

durch die numerische Ableitung von _e bestimmt wird.From this equation, T _r can be calculated, where

is determined by the numerical derivative of _e .

wobei s die Rate des Momentaufbaus ist. Die Dauer ergibt sich aus: $$\frac{T_{cl2}}{s}$$

_{In the strategy according to the} invention, the clutch torque is increased linearly from 0 to T cl2 during the torque build-up: $$T_{cl}st$$

where s is the rate of momentum build-up. The duration results from: $$\frac{T_{cl2}}{s}$$

$$T_e\left(\right)T_{e,d}\left(\right)$$

wobei T_e,d() das Fahrerwunschmoment zum Ende des Momentenaufbaus ist. Hier wird die Annahme getroffen, dass dieses annähernd konstant ist; daraus folgt T_e,d () T_e,d(t) T_e,d . Die Drehgeschwindigkeiten der linken und rechten Seite von Gleichung 7 lassen sich durch Integration der Bewegungsgleichungen 2 und 3 bestimmen: $${}_e\left(0\right)\frac{1}{J_{e0}}\left(T_e\left(t\right)T_d\left(t\right)\right)dt{}_{gbi}\left(0\right)\frac{i}{J_{\text{v }0}}\left(T_{r}i{T}_{cl}\left(t\right)\right)dt$$

wobei Tr 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. $${}_{gbi}\left(\right)\frac{i}{J_{\text{v}}}\left(T_r\frac{i}{2}{T}_{cl2}\right){}_{gbi}\left(0\right)$$

The further steps or requirements of the strategy can be formulated as follows: $${}_e\left(\right){}_{Gbi}\left(\right)$$

$$T_e\left(\right)T_{e,d}\left(\right)$$

where T _{e, d} () is the torque requested by the driver at the end of the torque build-up. Here the assumption is made that this is approximately constant; from this follows T _{e, d} () T _{e, d} (t) T _{e, d} . The rotational speeds of the left and right sides of equation 7 can be determined by integrating equations of motion 2 and 3: $${}_e\left(0\right)\frac{1}{J_{e0}}\left(T_e\left(t\right)T_d\left(t\right)\right)dt{}_{Gbi}\left(0\right)\frac{i}{J_{\text{v}0}}\left(T_{r}i{T}_{cl}\left(t\right)\right)dt$$

where Tr does not change significantly in the period under consideration, ie does not depend on t. The right hand side of equation 9 can be solved by substituting the equations for T _cl (t). $${}_{Gbi}\left(\right)\frac{i}{J_{\text{v}}}\left(T_r\frac{i}{2}{T}_{cl2}\right){}_{Gbi}\left(0\right)$$

Equations 8 and 9 involve two conditions on the timing of Te. Any approach can now be chosen for Te (t), which depends on exactly two parameters a and b that have yet to be determined. A time course can preferably be selected in which t occurs up to the second power. In addition, it can be required that Te is continuous at t = 0: $$T_e\left(t\right)T_e\left(0\right)a\mathrm{<figure-callout\; id="2"\; label="t\; b\; t"\; filenames="DE000010316446B4\_0051.png,DE000010316446B4\_0057.png"\; state="\{\{state\}\}">t</figure-callout>}b{t}^{}{}^2$$

The right side of equation 11 is inserted into the left side of equation 8 and solved for a: $$a\frac{2}{}\left(T_{e,d}{T}_e\left(0\right)b{}^2\right)$$

The right-hand sides of equations 5, 10 and 11 are inserted into equation 9, integrated and solved for b: $$\mathrm{<figure-callout\; id="3"\; label="b\; 6th"\; filenames="DE000010316446B4\_0053.png,DE000010316446B4\_0060.png"\; state="\{\{state\}\}">b</figure-callout>}\frac{\mathrm{6th}}{}\frac{\mathrm{}}{{}^3}\frac{1}{2}\left(T_{e,d}{T}_e\left(0\right)T_{cl2}\right)J_e\left({}_{Gbi}\left(\right){}_e\left(0\right)\right)$$

Accordingly, the clutch torque and the engine torque can be controlled appropriately with the aid of equations 5 and 11. The quantities contained in the equations can be determined by evaluating equations 4, 1, 10, 13 and 12 consecutively.

As already mentioned, it was assumed that the torque requested by the driver and the driving resistance remain approximately constant during the re-engagement. However, this need not necessarily be assumed. The current values of these variables are continuously accessible to the control code. In this way, the target torque T _{cl2 of} the clutch, which is required to calculate the current clutch _{torque according to equation 5, can be adapted for each interrupt in accordance with the changed values of T e, d} and Tr. The same can be done with parameters a and b, which are used to calculate the engine torque. However, if T _{e, d} and / or Tr change, it is to be expected that the clutch slip will not be reduced exactly at the expected point in time, but rather before or after. This then has the jolt in the vehicle acceleration, which is so low that it does not reduce comfort. However, the change in T _r is so small that the air resistance, the rolling resistance and the angle of incline of the roadway do not change noticeably during the short re-engagement phase.

In 27a a flowchart of the proposed strategy is shown, wherein the procedure can preferably be called once per clock cycle.

A further embodiment of the present invention is described below, in which a suitable voltage control for closing the clutch while slipping is proposed.

When the clutch is closed, it can happen, for example with a so-called slow path ramp with slip, that the driver perceives at least one jolt, for example during the continuous slip control. The jerk can result from small jumps in the clutch torque curve. If the clutch is closed slowly by means of the path specification, a step-shaped curve can arise due to the switch-off hysteresis in the clutch position control at the actual actual value. This jolt is uncomfortable for the driver and should therefore be prevented.

Accordingly, it is provided according to the invention that a voltage control is provided for closing the clutch, preferably with slip, instead of the position control.

It is possible to reduce the switch-off hysteresis in the above-mentioned situations. In the case of the clutch actuator used, very strong latching torques are usually provided. The switch-off hysteresis is introduced on the basis of these cogging torques. If the switch-off hystereses for the above-mentioned situation are now reduced, it is possible that the latching of the electric motor provides a behavior similar to that of the staircase with large switch-off hystereses. However, this course is less pronounced and the flanks have a lower slope. This has the result that the torque jumps are significantly reduced, so that they are advantageously no longer perceived by the driver. This increases the comfort for the driver when pulling the clutch.

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

However, there can be two limitations; because the clutch actuator used has the strong cogging torques mentioned above. If the clutch actuator is moved with a very low voltage, these detent moments can ensure that the actual position does not run continuously, but also in small steps. However, these stages are advantageously not as pronounced as in the case of the switch-off hysteresis. According to a further development, it can be provided that a minimum voltage is specified, which should not be undershot, in order not to let the intended steps become too large.

Another restriction can be that the clutch actuator can move at different speeds at constant tension due to the different release force of the clutches, depending on the clutch used. In order to avoid this, it can be provided according to the invention that the change in the actual position is appropriately observed. If the change is too rapid, the voltage can be decreased 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 in that, for example, when starting the parking lock when shifting the clutch actuator, which is done for example by way of travel control, the voltage curve is recorded and an average voltage is stored. It is also possible that instead of a recording, other possible procedures for reporting the voltage are provided. The taught-in 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 coupling.

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 cannot move against the stop. The position of the switchover can be dependent on the slope of the ramps, for example. Since the switchover is provided when a clutch is almost at full torque, it can be assumed that the slip is dismantled, and the torque stairs that may arise again due to the position control are not perceived by the driver.

This strategy according to the invention when applying the clutch can be provided in all vehicles with a clutch control.

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

The physical actuating torque characteristic can be mapped to a nominal characteristic in the software. The software coefficient of friction, also called SW-RW, can be an adjustment factor. The software coefficient of friction can be adapted when starting up and shifting. With the software coefficient of friction and its adaptation, short-term and long-term changes in 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 temperature and drops sharply at a temperature that is specific to the surface. This is also known as fading.

It can happen that the lining temperature on the clutch changes very quickly, for example when driving uphill many times in a row, and thus the software coefficient of friction changes very quickly. It has been shown that the software coefficient of friction adaptation is adapted too slowly in some cases, which can lead to very long slip phases when starting up and shifting gears. Due to the long slip phases, the clutch can be heated up even faster, so that the physical coefficient of friction changes even faster.

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

With the coefficient of friction adaptation, the actual clutch torque determined in the clutch control can be compared with the physical value in slip phases. The physical torque is determined from the engine torque and the engine acceleration. If the two torques assume different values, the software coefficient of friction can preferably be changed in small steps. For example, a maximum of 15 increments can preferably be used per switching or per approach. The maximum change can be selected in such a way that sufficient robustness of the coefficient of friction adaptation is guaranteed. In order to accelerate the adaptation of the coefficient of friction so that rapid changes in the physical coefficient of friction can be adapted, various possibilities are conceivable.

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

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

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

The measures mentioned ensure that in the normal working range in which the clutch is located most of the time, the coefficient of friction adaptation is sufficiently robust, but the adaptation speed is advantageous in an area in which the coefficient of friction can change significantly is adjusted.

The proposed coefficient of friction adaptation can be used in particular in ASG, EKM, PSG, USG systems or the like.

Another possible embodiment of the present invention is described below, in which a touch point correction is proposed, preferably using the clutch temperature.

In particular with an electronic clutch management (EKM) or with an automated gearbox (ASG) with hydraulic control of the clutch, shifts in the touch point due to temperature changes can be compensated for by regular sniffing. If so-called sniffing is not possible as a result of the driving maneuvers, such as continuous crawling or stop and go, this can be compensated for by a medium-term adaptation of the touch point. With the next sniffing, which is provided at the latest when the vehicle is switched off, this change in the medium-term touch point can be canceled by resetting the touch point to the long-term value. By sniffing when the vehicle is parked, shifts in the touch point can also be compensated for, so that a defined, balanced system can be assumed when the ignition is switched on again.

If a system without a hydraulic system is used, such as the electrically operated central slave cylinder (EZA), this adjustment is not possible when the ignition is switched off. Accordingly, it can be provided according to the invention that the current clutch temperature is also stored when the ignition is switched off, in particular if the behavior between the temperature and the shift of the touch point is known for the clutch. When the ignition is switched on, the current gripping point can then be corrected according to the temperature when saving. Any mathematical function can preferably be stored in the control device as the relationship between the clutch temperature and the shifting of the grip point.

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

The operating range of the actuating 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 preferably be precisely determined via the touch point adaptation and the friction coefficient adaptation or, more generally, via a characteristic curve adaptation. Furthermore, the change in the working range via the temperature of the clutch is known. If this behavior is not sufficiently known, it can be recorded accordingly while driving. For example, this can be provided on the basis of the touch point and the coefficient of friction as well as the position of the maximum torque.

If the vehicle is parked with a warm clutch, the information on the clutch temperature and the behavior of the working area 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 detent point. Any mathematical function can preferably also be stored in the control unit as a relationship between the clutch temperature and the change in the operating range of the clutch.

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

The proposed touch point correction can be used in particular with every electrical central release for an EKM, ASG, USG, PSG system or the like.

A further embodiment of the present invention is described below, in which a suitable rough positioning of a clutch actuator with incremental travel measurement is proposed.

In particular in the case of an uninterruptible gearbox with a central release mechanism (USG-EZA), only one incremental travel sensor is provided.

This means that only relative movements of the actuator can be recorded. Thus, there may be situations in which nothing is known about the position of the actuator, e.g. B. during commissioning or at the end of the line.

Accordingly, a possibility is to be found with which at least rough information (± 1 mm) can be obtained about the current actuator position without the risk of moving the actuator against a mechanical stop in such a way that the self-locking gear then jams.

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) used. For the actuator motor, the load (engagement / disengagement force) is thus a frictional torque in the spring band transmission.

The relationship between the load on the actuator and the frictional torque in the spring band transmission is used here, 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. 1. Einem echten lastabhängigen Reibmoment im Federbandgetriebe (richtungsunabhängig); und
2. 2. dem aus der Last übersetzten Moment (richtungsabhängig).

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

1. 1. A real load-dependent friction torque in the spring band transmission (independent of direction); and
2. 2. the torque translated from the load (depending on the direction).

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

In 28 the theoretical relationship between the external force and the frictional torque on the self-locking gear of the uninterruptible gearbox (USG) is shown, with Mr + movement against the load and Mr-: movement with the load.

In addition to the possibility of triggering a movement of the actuator with defined short voltage / current surges and evaluating its width (or, in extreme cases, to find the force equilibrium position by trembling), the frictional torques can also be roughly determined using the voltages required for moving in the respective direction of the actuator become. The latter has the advantage that the measurement results obtained in this way are proportional to the load torque, i.e. also proportional to the external force.

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

To measure the force, the motor voltage is slowly increased until a movement of the actuator is detected. The actuator begins to move when the engine torque M _M IU becomes greater than the friction torque Mr. This is done in both directions so that Mr + and Mr- can be measured. The difference in 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 indicated above and the output voltage indicated as a PWM ratio below.

At the in 29 The force measurement shown requires a higher voltage (approx. a PWM of 32) for the upward movement than for the downward movement (approx. 22). This difference now gives a statement about the external force. The small motor voltages applied after the voltage ramps and counteracting the movement are used to brake the actuator motor. This voltage is so small (PWM = 6) that it cannot maintain or even trigger any movement of the actuator on its own.

In 30th a possible sequence of a force measurement is shown in a flow chart. As can also be seen, the movement is not recognized immediately when the position changes by one increment. The interaction between the armature locking and the positions at which the increments are recognized 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 carried out, the force profile can be recognized. In 31 the results of 10 series of measurements are compiled, in which the force is measured at the same 21 positions. The course of the disengagement or engagement force is illustrated by the solid line. The coordinate system for the position is set in these measurements in such a way as to achieve the goal of the complete position reference.

In 31 the results of the force measurements on the clutch system are shown. In addition, the position areas used for rough positioning are already marked.

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

An evaluation of the measurements for the rough positioning clearly shows the strong scatter in the measurement results. Nevertheless, a rough conclusion about the position can be drawn from the measurement results. The areas (a) to (d) in 31 explained.

1. (a) Values below 3 can only be found in the strongly negative range (<-4 mm);
2. (b) In the range between 4 mm and 0 mm, only small measured values between 3 and +3 occur;
3. (c) Values above +3 can only be found in positive positions;
4. (d) Values above +10 can only be found for positions above +1.5 mm.

With the help of these results, a rough positioning of the actuator can be carried out. In this, the actuator feels its way 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 the accuracy of this positioning. The low accuracy of the force measurement allows only a rough approach to the reference position. A larger position step can only be carried out in the case of clearly large measured values in order to keep the duration of the rough positioning short.

In 32 a flowchart is shown in which the limit values are specified as they are also implemented, for example, in the uninterruptible gearbox (USG). This procedure finds a position that is slightly (approx. 1 mm) above the reference position. However, this can be compensated for by a position correction at the end of the rough positioning.

Examples of the rough positioning sequence are given below.

A series of measurements are carried out to check the rough positioning. Rough positioning was started at different positions in the adjusted state (reference position at position 0). In 33 a measurement of a coarse positioning starting at 6 mm is shown graphically. By resetting the incremental position sensor, the movements always start at position 0 mm and then move in the direction of the negative start position. The actual force measuring process can be seen from the back and forth movements of the actuator in pairs.

In 34 a measurement of a coarse positioning starting 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 during the force measurement, the courses are different for different coarse positioning from the same starting position. In rare cases (less than 1 in 25 in these examinations), the noise also leads to incorrect detection of the position. In these cases, the subsequent fine positioning does not work. However, this can be recognized by the software, whereupon a new rough positioning is started.

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

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

This method can be used to advantage with push-pull clutches. Here there is a clear difference between positive and negative clutch force; In addition, the proportion with characteristic release forces increases by around one millimeter. The method used to search for reference points must of course be adapted to the conditions of the respective coupling system.

Overall, the described embodiment of the present invention specifies a method with which, in particular in the case of a USG clutch actuator, when the position is unknown Rough determination of its reference position can be carried out without using a mechanical stop. This method according to the invention can be used in particular for a basic initialization, such as is used, for example, at the end of the tape. A rough measurement of the disengagement or engagement force is provided via the minimum voltage for movements of the actuator motor. The reference position can initially be determined to about +/- 0.5 mm. The reference position can then be precisely determined by means of a suitable fine adjustment. This method according to the invention can also be used for other coupling systems.

With the method according to the invention, additional elements, such as force sensors in the system, can be dispensed with. This is particularly because the actuator motor itself is used to determine the position. For this purpose, the motor voltage of the actuator motor can be increased twice with opposite signs in a time ramp until the movement of the actuator is detected in each case 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 disengagement and engagement force. The process of a single force measurement is shown in 35 shown in a flow chart. To save time in the process, the voltage ramps can not be at zero, but z. B. be started at a predetermined starting value.

In order to minimize interactions between the magnetic armature detents and the position of the position increments, the movement of the actuator can only be detected when there is a minimum number of position increments, which is preferably greater than 1. For example, the limit can be 4. The brake voltage U_Brems is selected to be so small that it does not keep the actuator moving in the same direction or even cause the actuator to start up.

It has been shown that although this force measurement is subject to strong noise, sufficient information is provided about the current position of the actuator. The force curve in the clutch system should be structured sufficiently. There must be large proportions of approx. <1/6 of the range of motion with clear strength levels. In addition to an area with a very small load, a change in the sign of the load is particularly advantageous here. From this it can be decided which position change is to be carried out next in order to obtain more information about the absolute position of the actuator.

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

In the case of clutch actuators with incremental travel measurement, such as the EZA, the problem can arise that when the control unit is switched on, it is not clear where the clutch is located. There are options to save the actuator position when the control unit is switched off, so that it can continue to be driven from this position when the control unit is switched on. However, this assumes that the actuator position has not been changed while the control unit is switched off. This requirement does not apply if, for example, the clutch actuator has been opened in an emergency and the actuator position has been changed by hand.

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

According to the invention it can be provided that the stop is not required for the plausibility check of the position. The procedure specified in the EKM / ASG control is to be changed from the zero position (parking lock) to the HUB position, i.e. over the entire working range of the clutch control, in such a way that a plausibility check is provided at the same time as to whether there is a rough change in the zero Position in the "ignition off" phase. A prerequisite is that the working range of the clutch control with the limits zero point and stroke from the last "ignition on" phase is known.

When the control unit is switched on, the actuator can, for example, be in the parking lock position (corresponds to the zero point of the working range of the clutch control). This is the starting position of the travel path. A short distance is driven in the direction of “close clutch” (delta_HUB), as indicated by the arrows located below the diagram in 36 is shown. A specified maximum speed is not exceeded. If the actuator is still in the same position as when the control unit was switched off, the actuator cannot hit a stop with this method. The travel distance should therefore be selected accordingly.

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

The actuator can then be moved to the HUB + delta_HUB position, as the starter is actuated in neutral gear. This procedure should only be carried out with a limited speed, as it is possible that the vehicle will hit the stop, especially if the actuator position has been adjusted in the "ignition off" state. If the actuator was in the same position when the ignition was switched on as when the ignition was switched off, which characterizes the normal case, the actuator is not moved against the stop. Therefore delta_HUB should be selected accordingly. The HUB position can then be approached and the normal driving mode is present. Overall, the time to the procedure is not significantly increased.

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

Another 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 consists in that during the opening of the clutch in the “ignition on” state, suitable monitoring takes place to determine whether a stop has been hit. In this way, the process in the other direction can be dispensed with (delta_NP and delta_Hub are equal to zero). This development has the advantage that the paths below the zero point and above the HUB position do not have to be kept.

Overall, it can be stated that in the method according to the invention for checking the plausibility of the actuator position, it can be provided that, if corresponding paths are available, a check is made in both directions to determine whether stops can be approached. Furthermore, it can be provided that if there are no paths or if the search takes too long, a tentative opening is made.

Possible errors can e.g. B. an emergency opening, an improper coupling exchange, a reset due to undervoltage or a reset by ISM. A comparison when an error is detected can provide that the stops are also used for the plausibility check.

Another embodiment of the present invention is described below, in which a switchable freewheel is proposed as a load torque lock.

Particularly in the case of self-locking gears with an efficiency of over 50%, it was found that the coefficient of friction that occurs is often too high, so that second-order self-locking then occurs.

In the case of existing self-locking, it can then only be actuated against this, with no further actuation being possible in the direction of the load. This behavior corresponds to a classic freewheel, i.e. a load torque lock, whereby the rotary movement is only possible in one device. The self-locking is essentially provided by two gears, which can be designed with internal, external, helical or herringbone teeth. In addition, the design with parallel, crossed and / or cut axes is possible.

The behavior described above cannot be restricted to just one direction. This is especially true in 37 shown. Thus, when a directed torque 2 is applied to the gear wheel of the inhibiting member 3, the direction of freewheeling or self-locking can be determined. The moment can be applied in all conceivable forms, for example magnetically, mechanically, electromechanically, piezomechanically or the like. Another advantage here is that low performance is required for conversion and activation.

The inhibition drive described above enables a normal drive train without self-locking by not applying the torque. The so-called freewheeling inhibitor can be integrated in any position of a transmission power train.

In this way, by using an originally undesired effect, namely second-order self-locking, a suitable free-wheeling mechanism is provided in a self-locking gear with an efficiency of more than 50%.

This freewheel according to the invention can be used, for example, in starters, starter-generator drives, foot crank starters, actuating gears, clutch adjustments or the like.

A further embodiment of the invention is described below in which at least one suitable cage is proposed to ensure the functionality of ramp mechanisms with rolling elements.

There are different Rampenausrücksysteme with rolling elements, z. B. balls in a mechanically operated central slave cylinder (MZA) are provided. Due to the non-existent positive safety device, the rolling elements may drift out of the target position after vibrations of the overall system or after a large number of short adjusting movements. The special shape of the ramps themselves can lead to malfunctions. For example, stiffness or a shortening of the stroke length is possible.

The functionality of a mechanically operated central slave cylinder (MZA) is based on a wedge effect between two rotating parts with corresponding ramps. Rolling bodies are preferably used at the contact points. A wedge gear of the MZA with rolling elements is in 38 shown, with a contact point being developed on the plane. In 39 a possible basic structure of an MZA is indicated.

The aforementioned problem is caused by the 40 and 41 graphically represented. In 40 various states of motion of the rolling elements between the ramps are indicated, with a) showing the ramp mechanism in the starting position and b) showing the ramp mechanism in the lift position (actuated). In c) a blocking of the ramp mechanism by the rolling elements is indicated, which represents a disadvantageous shortening of the stroke.

In 41 a modified ramp shape is shown. The changed shape of the ramp (without a stop) or the omission of a cage between the rolling elements cannot solve the aforementioned problem either.

According to the invention it can be provided that a plurality of rolling elements are preferably provided for 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 designed as axial slots or the like, can advantageously keep the distance between the rolling elements in the advancing direction. However, the rolling elements can move in the stroke direction. In this way, the rolling elements can suitably adjust to the distance between the ramp tracks. Since the distance between the rolling elements is preferably less than their rolling distance, it can be ensured that there is always at least one rolling element between the respective ramps. If the ramps still have no limit, the cage can possibly move as far as it wants. Even if the rolling elements move, the function is not negatively affected.

In 42 the ramp mechanism with the cage according to the invention is shown, the starting position being indicated in a) and the maximum lifting position being indicated in b). It is particularly advantageous that with a corresponding ramp design, the stroke generation is possible in both directions of actuation. This property can be used to achieve a longer service life for the ramp mechanism.

Further advantages and advantageous configurations emerge from the subclaims and the drawings described below.

• 1 eine schematische Ansicht eines Antriebsstranges mit einer Kupplung, und
• 2 ein Flussdiagram einer möglichen Ausgestaltung des erfindungsgemäßen Verfahrens.

Show it:

• 1 a schematic view of a drive train with a clutch, and
• 2 a flow diagram of a possible embodiment of the method according to the invention.

In 1 a model of a drive train is indicated schematically, on the basis of which the physical behavior of the drive train is described. The motor has the moment of inertia J _e and can deliver the adjustable motor torque T _e , which is the torque of the internal combustion engine. The The rotational speed of the motor is denoted by e. The clutch can transmit a torque, the amount of which is less than or equal to T _cl . When the clutch is slipping, this amount is exactly the same as the value of T _cl ; if the coupling is adhesive, this is at most T _cl . The transmittable torque of the clutch T _cl is adjustable. A gear ratio i is provided between the clutch and the vehicle. The speed of rotation of the left-hand side of the transmission, i.e. the transmission input shaft, is denoted by gbi. The vehicle has the moment of inertia J _v . The vehicle's driving resistance, which is composed of rolling resistance, air resistance and downhill force, acts on the vehicle from outside. The sum of all the moments acting on the vehicle is denoted _{by T r}

In 2 shows a flow chart of a possible embodiment of the method according to the invention. The individual process steps can be seen from the flowchart.

wobei mit T_e,d das vom Fahrer angeforderte Motormoment T_e bezeichnet ist.At the beginning of the process, it is checked whether the clutch is sticking in the drive train. When the new gear is to be engaged, the clutch _torque T cl2 can be calculated using the following equation 1: $$T_{cl2}\frac{J_V{T}_{e,d}i{J}_e{T}_r}{i^2{J}_e{J}_v}$$

where T _{e, d} denotes the engine torque T _e requested by the driver.

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 folgenden Gleichung 4:

To solve the right hand side of equation 1, T _r must be known. With a sticky coupling, the equations of motion for the engine and the vehicle are given by the following equations:

in which $T_{cl}^{*}$

is the torque transmitted by the adhesive coupling. Eliminating this quantity from equations 2 and 3 and solving for T _r leads to the following equation 4:

durch die numerische Ableitung von _e bestimmt wird.From this equation, T _r can be calculated, where

is determined by the numerical derivative of _e .

wobei s die Rate des Momentaufbaus ist. Die Dauer ergibt sich aus der Gleichnung 6: $$\frac{T_{cl2}}{s}$$

In the inventive method, the clutch torque T _cl during the torque formation is increased linearly from 0 to T _cl2: $$T_{cl}st$$

where s is the rate of momentum build-up. The duration results from equation 6: $$\frac{T_{cl2}}{s}$$

$$T_e\left(\right)T_{e,d}\left(\right)$$

wobei T_e,d () das Fahrerwunschmoment zum Ende des Momentenaufbaus ist. Hier wird die Annahme getroffen, dass dieses annähernd konstant ist; daraus folgt T_e,d() T_e,d(t) T_e,d. Die Drehgeschwindigkeiten der linken und rechten Seite aus der Gleichung 7 lassen sich durch Integration der Bewegungsgleichungen 2 und 3 bestimmen: $${}_e\left(0\right)\frac{1}{J_{e0}}\left(T_e\left(t\right)T_{cl}\left(t\right)\right)dt{}_{gbi}\left(0\right)\frac{i}{J_{v0}}\left(T_{r}i{T}_{cl}\left(t\right)\right)dt$$

wobei T_r im betrachteten Zeitraum nicht signifikant verändert wird, d. h. nicht von t abhängt. Die rechte Seite von Gleichung 9 lässt sich durch Einsetzen der Gleichungen für T_cl(t) auflösen. $${}_{gbi}\left(\right)\frac{i}{J_v}\left(T_r\frac{i}{2}{T}_{cl2}\right){}_{bbi}\left(0\right)$$

The further steps or requirements of the strategy can be formulated as follows: $${}_e\left(\right){}_{Gbi}\left(\right)$$

$$T_e\left(\right)T_{e,d}\left(\right)$$

where T _{e, d} () is the torque requested by the driver at the end of the torque build-up. Here the assumption is made that this is approximately constant; from this follows T _{e, d} () T _{e, d} (t) T _{e, d} . The rotational speeds of the left and right side from equation 7 can be determined by integrating equations of motion 2 and 3: $${}_e\left(0\right)\frac{1}{J_{e0}}\left(T_e\left(t\right)T_{cl}\left(t\right)\right)dt{}_{Gbi}\left(0\right)\frac{i}{J_{v0}}\left(T_{r}i{T}_{cl}\left(t\right)\right)dt$$

where T _{r is} not changed significantly in the period under consideration, ie does not depend on t. The right hand side of equation 9 can be solved by substituting the equations for T _cl (t). $${}_{Gbi}\left(\right)\frac{i}{J_v}\left(T_r\frac{i}{2}{T}_{cl2}\right){}_{bbi}\left(0\right)$$

With the equations 8 and 9, two conditions are connected with the time course of the engine torque Te. Any approach can now be chosen for Te (t), which depends on exactly two parameters a and b that have yet to be determined. A time course can preferably be selected in which t occurs up to the second power. In addition, it can be required that T _e is continuous at t = 0: $$T_e\left(t\right)T_e\left(0\right)a\mathrm{<figure-callout\; id="2"\; label="t\; b\; t"\; filenames="DE000010316446B4\_0051.png,DE000010316446B4\_0057.png"\; state="\{\{state\}\}">t</figure-callout>}b{t}^{}{}^2$$

The right side of equation 11 is inserted into the left side of equation 8 and solved for a: $$a\frac{2}{}\left(T_{e,d}{T}_e\left(0\right)b{}^2\right)$$

The right-hand sides of equations 5, 10 and 11 are inserted into equation 9, integrated and solved for b: $$\mathrm{<figure-callout\; id="3"\; label="b\; 6th"\; filenames="DE000010316446B4\_0053.png,DE000010316446B4\_0060.png"\; state="\{\{state\}\}">b</figure-callout>}\frac{\mathrm{6th}}{}\frac{\mathrm{}}{{}^3}\frac{1}{2}\left(T_{e,d}{T}_e\left(0\right)T_{cl2}\right)J_e\left({}_{Gbi}\left(\right){}_e\left(0\right)\right)$$

Accordingly, the clutch _{torque T cl} and the engine torque T _{e can} be controlled appropriately with the aid of equations 5 and 11. The quantities contained in the equations can be determined by evaluating equations 4, 1, 10, 13 and 12 consecutively.

It can be assumed that the driver's desired torque T _{e, d} and the driving resistance remain approximately constant during the re-engagement. However, this need not necessarily be assumed. The current values of these variables are continuously accessible to the control code. In this way, the target torque T _{cl2 of} the clutch, which is required to calculate the current clutch torque according to equation 5, can be adapted for each interrupt according to the changed values of the driver's desired torque T _{e, d} and the sum of all the torques T _r acting on the drive train. The same can be done with the parameters a and b, which are used to calculate the engine torque T _e .

In the proposed method for controlling the engine torque T _e and the clutch _{torque T cl} in a drive train of a vehicle, the method described above and in 2 The procedure shown is preferably called once per clock cycle.

The patent claims submitted with the application are formulation proposals without prejudice for obtaining further patent protection. The applicant reserves the right to claim further combinations of features that have so far only been disclosed in the description and / or drawings.

References back used in subclaims indicate the further development of the subject matter of the main claim through the features of the respective subclaim; they are not to be understood as a waiver of the achievement of an independent, objective protection for the combinations of features of the referenced subclaims.

Since the subjects of the subclaims with regard to the state of the art can form their own and independent inventions on the priority date, the applicant reserves the right to make them the subject of independent claims or declarations of division. They can also contain independent inventions that have a design that is independent of the subject matter of the preceding subclaims.

The exemplary embodiments are not to be understood as limiting the invention. Rather, numerous changes and modifications are possible within the scope of the present disclosure, in particular those variants, elements and combinations and / or materials that can be obtained, for example, by combining or modifying individual elements in connection with those described in the general description and embodiments as well as the claims and in Features or elements or process steps contained in the drawings can be inferred by the person skilled in the art with regard to the solution of the task and, through combinable features, lead to a new object or to new process steps or process step sequences, also insofar as they relate to manufacturing, testing and working processes.

Claims (8)

Method for controlling an engine torque and a clutch torque in a drive train of a vehicle with a transmission and an engine after a gear change, the clutch torque (T _cl ) being increased to a predetermined value, and at the same time the engine torque (T _e ) being controlled in such a way that The fact that the slip on the clutch is reduced is characterized in that the slip on the clutch is reduced by controlling the engine torque (T _e ) when the clutch _torque reaches the value (T cl2) and the engine torque (T _e ) from this onwards Time again corresponds to the torque requested by the driver, the engine speed for controlling the engine torque (T _e ) being regulated by changing the clutch _{torque (T cl} ). Procedure according to Claim 1 , characterized in that the clutch torque (T _{cl ) increases monotonically from a clutch torque directly after engaging the gear} (value 0) to a predetermined value which corresponds to the clutch torque (T cl2) immediately after the transition to sticking. Procedure according to Claim 2 , characterized in that the value of the clutch _torque (T cl2) is calculated by the following equation $$T_{cl2}=\frac{J_V\cdot T_{e,d}-i\cdot J_e\cdot T_r}{i^2\cdot J_e+J_v}$$

Method according to one of the preceding claims, characterized in that the clutch torque (T _cl ) is increased monotonically during the re-engagement. Method according to one of the preceding claims, characterized in that when the clutch is sticking, the equation of motion of the motor is given by the following equation: $$J_e\cdot {\dot{\omega }}_e=T_e-T_{cl}^{*}$$

Method according to one of the preceding claims, characterized in that when the clutch is sticking, the equation of motion of the vehicle is given by the following equation: $$J_v\cdot \frac{{\dot{\omega }}_{Gbi}}{i}=T_r+i\cdot T_{cl}^{*}$$

Method according to one of the preceding claims, characterized in that the sum of all the torques (T _r ) acting on the drive train is calculated by the following equation: $$T_r=\left(i\cdot J_e+\frac{J_v}{i}\right)\cdot {\dot{\omega }}_e-i\cdot T_e$$

Control device for a drive train of a vehicle with an engine and a transmission coupled via at least one clutch, for performing the method according to one of the Claims 1 to 7th , characterized in that at least one engine control unit and at least one transmission control unit are provided which control the engine torque (T _e ) and the clutch _{torque (T cl} ) in such a way that a steady build-up of torque is provided after a gear change.

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