DE10236539A1 - Method of controlling motor vehicle clutch involves adjusting clutch take up point dependent on operating temperature - Google Patents

Abstract

The method of controlling a motor vehicle clutch involves adjusting the take up point of the clutch dependent on the clutch temperature. The temperature change is compensated for as the clutch is operated. The value of the take up point of the clutch is reduced as the clutch temperature increases. Claims include a clutch control using the method

Description

The invention relates to a method, a device and the use thereof for Operation of a motor vehicle, in particular for controlling a electronic clutch management and / or an automated gearbox, preferably with a drive motor, a clutch and / or a gear in the drive train.

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

The torque transmission system 3 is a clutch, such as. B. designed as a friction clutch, multi-plate clutch, magnetic powder clutch or torque converter clutch, the clutch can be a self-adjusting or a wear-compensating clutch. The transmission 4 is an uninterruptible manual transmission (USG). According to the idea of the invention, the transmission can also be an automated manual transmission (ASG), which can be shifted automatically by means of at least one actuator. In the following, an automated manual transmission is to be understood as an automated transmission which is shifted with an interruption in tractive force and in which the shifting operation of the transmission ratio is carried out in a controlled manner by means of at least one actuator.

An automatic transmission can also be used as the USG, whereby an automatic transmission a transmission essentially without interruption of tractive power in the switching operations and this is usually through planetary gear stages is constructed.

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

The torque transmission system 3 has a drive side 7 and an output side 8 , wherein a torque is transmitted from the drive side 7 to the output side 8 by z. B. the clutch disc 3 a by means of the pressure plate 3 b, the plate spring 3 c and the release bearing 3 e and the flywheel 3 d is subjected to force. For this application, the release lever 20 is actuated by means of an actuator, for. B. an actuator operated.

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

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

The movement of the output part of the actuator 13 b, such as. B. the master cylinder piston 11 a is detected with a clutch travel sensor 14 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. B. a hydraulic line, connected to the slave cylinder 10 . The output element 10 a of the slave cylinder is with the disengaging means 20 , for. B. a release lever, operatively connected so that a movement of the output part 10 a of the slave cylinder 10 causes the disengaging means 20 is also moved or tilted to control the torque transmitted by the clutch 3 .

The actuator 13 b for controlling the transmissible torque of the torque transmission system 3 can be actuatable 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 take place by means of an electric motor, wherein the electric motor provided as the drive element 12 can be controlled electronically. The drive element 12 of the actuator 13 b, besides an electromotive drive element, another, for example, be pressure-actuated drive member. Magnetic actuators can also be used to adjust a position of an element.

In the case of a friction clutch, the transferable torque is controlled in that the friction linings of the clutch disc are pressed in a targeted manner between the flywheel 3 d and the pressure plate 3 b. About the position of the disengagement means 20 , such as. B. a release fork or a central release, the application of force to the pressure plate 3 b or the friction linings can be controlled in a targeted manner, the pressure plate 3 b being moved between two end positions and being able to be set and fixed as desired. One end position corresponds to a fully engaged clutch position and the other end position corresponds to a fully disengaged clutch position. To control a transmissible torque, which is, for example, less than the engine torque currently present, a position of the pressure plate 3 b can be controlled, for example, which lies in an intermediate region between the two end positions. The clutch can be fixed in this position by means of the targeted actuation of the disengaging means 20 . However, it is also possible to control transmissible clutch torques that are defined above the engine torques currently pending. In such a case, the currently occurring engine torques can be transmitted, the torque irregularities in the drive train being damped and / or isolated in the form of, for example, torque peaks.

To control the torque transmission system 3 , sensors are also used which at least temporarily monitor the relevant variables of the entire system and deliver the state variables, signals and measured values necessary for control, which are processed by the control unit, with a signal connection to other electronic units, such as, for example Engine electronics or an electronics of an anti-lock braking system (ABS) or an anti-slip control (ASR) can be provided and can exist. For example, the sensors detect 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.

Fig. 1 shows that a throttle sensor 15, an engine speed sensor 16 and a speedometer sensor may find use 17 and transmit measured values or information to the controller 13. The electronics unit, such as. B. a computer unit, the control electronics 13 a processes the system input variables and transmits control signals to the actuator 13 b.

The gear is as z. B. stage change gear designed, the gear ratios are changed by means of a shift lever 18 or the transmission is operated or operated by means of this shift lever 18 . Furthermore, at least one sensor 19 b is arranged on the shift lever 18 of the manual transmission, which detects the intention to shift and / or the gear position and forwards it to the control unit 13 . The sensor 19 a is articulated on the transmission and detects the current gear position and / or an intention to shift. The intention to shift is detected using at least one of the two sensors 19 a, 19 b in that the sensor is a force sensor which detects the force acting on the shift lever 18 . 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 13 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 issues control or regulation commands to the at least one actuator 13 b. The drive motor 12 of the actuator 13 b, z. B. 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 of the connected sensors. For this purpose, a control program is implemented in the control unit 13 as hardware and / or as software, which evaluates the incoming signals and calculates or determines the output variables on the basis of comparisons and / or functions and / or characteristic maps.

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

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

The following is a possible embodiment of the invention presented here described in which an EKM and / or ASG control is provided, which in particular enables maximum engine torque adaptation.

It has been shown that especially at extreme heights (e.g. 3000 m) the Engine of the vehicle u. U. only delivers about 60% of its nominal torque.

In particular with an EKM start-up control via a suitable characteristic map M _K (n _Mot ), lower start-up speeds result. In extreme cases, the engine torque may drop again at these low speeds. A corresponding map is shown schematically in FIG. 2. The control is such. B. depending on the motorization and essentially from the starting map shown in Fig. 2. For detection, e.g. B. a CAN signal "vehicle code" can be read.

According to an embodiment of the presented invention, it is conceivable that the Control of the electronic clutch management or automated gearbox (ASG) has a suitable engine torque adaptation. This can advantageously be the starting behavior at extreme heights be stabilized. The control is more independent of the present motorization of the vehicle.

The height detection can e.g. B. directly via a designated edition of Motor control take place, which preferably measures the intake pressure. It is it is also conceivable that the height detection indirectly through the EKM / ASG Control via a determination of the output from the engine control Moment is provided for example at a maximum load. The The specified engine torque can then be corrected accordingly.

The result of this adaptation can e.g. B. stored and preferably to be updated in the medium term. The time constant required here can e.g. B. assume the value of several minutes.

The EKM / ASG control can correct the starting map M _K (n _Mot ) in such a way that, for. B. the adapted actual engine maximum torque is present at a target starting speed. Of course, other suitable adaptations are also possible.

With this adaptation described above, the motorization of the vehicle are automatically recognized and the starting map is suitably adapted become. It can also be provided that the type of engine is recognized of the vehicle (diesel engine / petrol engine) is provided.

Another development of the invention presented here can preferably prevent the overrun of a transmission input shaft. It has been shown when the clutch is closed in the neutral position of a transmission or z. B. a drag torque transmits that, for example, to follow the Transmission input shaft can come. The transmission input shaft can only are slowed down by the losses in the transmission. Especially with one warm gearbox can be run on for up to 4 seconds. It is possible, that when inserting z. B. an unsynchronized reverse gear unpleasant noises are caused. Especially with an automated one Shift gearboxes (ASG) should avoid this.

• 1. Sobald der Befehl zum Einlegen des Rückwärtsganges kommt, kann der vorherige Gang herausgenommen werden. Somit ist bei stehendem Fahrzeug sichergestellt, dass die Getriebeeingangsdrehzahl den Wert Null oder bei einem Restmoment der Kupplung einen möglichst kleinen Wert annimmt.
• 2. Vor Einlegen des Rückwärtsgangs wird bei stehendem Fahrzeug kurz ein Gang eingelegt. Alternmativ kann auch vorgesehen sein, dass mittels der Synchronisierung eines Ganges die Getriebeeingangswelle bis und/oder fast bis zum Stillstand abgebremst wird. Dies kann vorzugsweise mittels eines Getriebeeingangsdrehzahlsensors und/oder durch das Feststellen des Synchronisierens vorzugsweise durch eine Wegmessung an der Schiebemuffe und oder an einem mit ihr verbundenen Bauteil detektiert werden. Anschließend kann z. B. der Rückwärtsgang eingelegt werden.

According to the embodiment of the invention, this can be avoided in that the transmission input speed is preferably braked to a standstill or to a predetermined speed. This can e.g. B. can be achieved by at least one of the strategies described below:

• 1. As soon as the command to engage the reverse gear comes, the previous gear can be removed. This ensures when the vehicle is at a standstill that the transmission input speed assumes the value zero or the smallest possible value for a residual torque of the clutch.
• 2. Before engaging reverse gear, a gear is briefly engaged when the vehicle is stationary. Alternatively, provision can also be made for the transmission input shaft to be braked to and / or almost to a standstill by means of the synchronization of a gear. This can preferably be detected by means of a transmission input speed sensor and / or by determining the synchronization, preferably by measuring the distance on the sliding sleeve and or on a component connected to it. Then z. B. the reverse gear.

A further embodiment of the invention presented here follows described, which a suitable strategy for gear engagement preferably at cold gearbox concerns.

It has been shown that especially at extremely low temperatures high forces have to be applied to engage a gear. It is too note that twisting the gears and gear shafts against each other, which is required for engagement, extremely high forces are required are. In extreme cases, it is therefore possible that no gear is engaged can.

Especially in manual transmissions with a non-automated clutch the driver therefore briefly close the clutch or when inserting the clutch z. B. let it grind. The shafts can thus be rotated against each other and the next time you try to insert a cheaper one-track position is present or the drag torque can also occur after the cold oil has been displaced Gearbox be correspondingly lower.

With an automated clutch and with an automated manual transmission the clutch is usually kept open in neutral to a to avoid so-called idle rattling. Preferably with the so-called Sniffing function, the clutch can be closed briefly. When engaging the gear however, the clutch must be open to prevent ratcheting avoid.

According to the z. B. be provided that at temperatures below one predetermined limit, the clutch is closed in the neutral position (neutral) becomes. This can be achieved in an advantageous manner that the gears Each insertion attempt is in a different position, which is for the engagement is cheaper.

It is also conceivable to take different temperatures into account. For example, the transmission temperature, the engine temperature, the Outside temperature and / or any other suitable temperature can be used. A suitable combination of different temperatures may also be possible be beneficial.

According to a development of the invention presented here, a suitable model can be used to calculate the temperature appropriately. A determined limit temperature can preferably vary as a function of time and / or in predetermined situations, such as. B. before inserting the Reverse gear can be changed so that the clutch then remains open. Other suitable measures are also conceivable, the one Further improve gear engagement, especially when the transmission is cold.

A further embodiment of the present invention can be a suitable one Control unit function, by means of which preferably a seat belt Control is provided.

It has been shown that, for security reasons, it is advantageous if the Drivers are prevented from doing this by taking the appropriate measures Driving seat belt.

It is possible that a suitable sensor e.g. B. in the buckle of the seat belt is provided. Of course, a switch or the like can also be used be provided. The control of the EKM / ASG system, the motor and / or the Immobilizer can send a corresponding signal, which is generated by the sensor or switch is sent, read in and then initiate suitable measures. For example, certain driving functions can be prevented.

It is also possible for the motor to be stopped at a predetermined level Speed threshold is carried out. Furthermore, z. B. by the Control of the ASG system predetermined gears, such as. B. gears 4 and 5 or the like. There is also a possibility that a Jerking when starting is provided by the clutch control. As another Measure can also z. B. at high gears by closing the clutch the clutch control can be prevented. In addition, of course also suitable warning elements, such as. B. beeper or flashing devices be provided.

With sufficient reliability of the sensor or switch can also further measures are provided, such as. B. locking the Starter and / or a suitable control of the clutch control and / or the ASG control so that the vehicle can no longer be moved.

The use of a suitable seat belt contact is included crucial. Once this contact, switch, sensor or the like is a suitable one Any measures can be taken. For example it should also be provided that the driveability and maneuverability of the Vehicle is still guaranteed. However, at higher speeds, larger ones Gears or greater performances can drive the z. B. be blocked.

Furthermore, the invention presented here is the entire content of the based on British patent application GB 2 088 608 A.

Another embodiment of the present invention follows described, which in particular a suitable modification z. B. the Torque characteristic and / or the clutch control concerns.

It is possible that after an energy input into the clutch Change in the torque characteristic of the clutch occurs: This can negatively affect the Function and comfort z. B. an automatic clutch control impact.

Accordingly, it may be advantageous to carry out a suitable correction, preferably of Actuating torque characteristic to be provided after an energy input into the clutch. It is particularly advantageous if preferably a calculation and / or compensation of thermal effects in an automated Clutch control is provided.

For example, it is conceivable that a calculation of the displacement and / or Change in the torque characteristic and compensation of the shift and / or change in the clutch control is carried out.

Furthermore, a similar effect can be seen in shifting the position of the Belleville spring tongues occur after an energy input. In particular, this can with the help of measurement technology in the vehicle and / or a test bench, in particular an EKM functional test bench.

The influence of the coupling temperature on the Gripping point described.

The gripping point is preferably for the control of an automatic clutch a crucial size. The gripping point corresponds to the path of the actuator a predetermined point, which is preferably a clutch torque of corresponds to about 9 Nm. The gripping point is not constant, but can e.g. B. also change by introducing energy into the clutch.

The investigation described below is intended to provide basic information for the Control of the clutch supply to the quality of the clutch control continue to improve.

• 1. Verschiebung der Position der Tellerfederzungen bei geschlossener Kupplung. Daraus folgt eine Änderung des Greifpunktes am Steller, wenn die sogenannte Schnüffelfunktion nicht aktiviert wird.
• 2. Änderung der Momentenkennlinie der Kupplung. Daraus folgt eine Änderung des Greifpunktes am Steller.

First, the displacement of the gripping point when the vehicle is crawling is considered as an example. By shifting the gripping point by an energy input, it can, for. For example, it may be possible for the vehicle to crawl at about 30 Nm instead of 10 Nm. In the case of energy input into the clutch, two effects described in the following act in particular:

• 1. Shift the position of the disc spring tongues when the clutch is closed. This results in a change of the gripping point on the actuator if the so-called sniffing function is not activated.
• 2. Change in the torque characteristic of the clutch. This results in a change in the gripping point on the actuator.

These two effects can also overlap, especially if none Sniff function is activated. When activating the sniffing function can preferably only the effect mentioned under 2.

Furthermore, there may be a short-term dependence of the clutch temperature grip point. It is possible that based on an initial state within a short period of time, such as. B. <3 min. the change in the gripping point is calculated and taken into account in the control of the automated clutch. The coupling temperature can be e.g. B. short term, e.g. B. at a crawl, with several approaches in a row and / or during a traffic jam on the mountain. It can be important here whether the sniffing function is activated or not, as is also indicated in FIG. 3 by the different trend lines. In Fig. 3 a calculated Steller gripping point thus is shown as a function of the clutch temperature.

It has been shown that u. May not be a long-term or absolute dependency of the gripping point from the coupling temperature. Others can Influences on the gripping point can be of importance, such as. B. the temperature distribution within the clutch, especially a self-adjusting clutch (SAC), the adjustment of the clutch, in particular a SAC clutch, a setting the main actuator spring and / or a setting or embedding of the pad suspension.

It when the displacement of the gripping point at the Energy input in the EKM control is included, so that one Improve the functionality and comfort of automatic Clutch control is possible.

• 1. Durch die aufeinanderfolgenden Anfahrten an einer leichten Steigung nach folgendem Programm:
  10 Anfahrten im 1. Gang, 5 min Pause
  5 Anfahrten im 2. Gang, 5 min Pause
  15 Anfahrten im 1. Gang, 5 min Pause
  10 Anfahrten im 2. Gang, 5 min Pause
• 2. Durch das Kriechen des Fahrzeuges an einem Bordstein wie folgt:
  4 min Kriechen, 2 min Pause
  2,5 min Kriechen, 10 min Pause

The energy input into the coupling can preferably be carried out using the methods described below:

• 1. By successive approaches on a slight slope according to the following program:
  10 starts in 1st gear, 5 min break
  5 starts in 2nd gear, 5 min break
  15 starts in 1st gear, 5 min break
  10 starts in 2nd gear, 5 min break
• 2. By crawling the vehicle on a curb as follows:
  4 min crawl, 2 min break
  2.5 min crawl, 10 min break

The energy input per journey can be less than 50 kJ. When crawling the friction power was less than 2 kW.

A light stall test was carried out before and after each start-up program, in which the accelerator pedal is applied to the torque characteristic of the Determine clutch.

The gripping point can be determined with the torque characteristics of the coupling determined in the so-called stall test. The clutch temperature can be calculated using an EKM temperature model, whereby the temperature of the pressure plate, the path at the central release and the pressure are used as measured variables. Of course, other measured variables can also be used. The summary of the measurement results is shown in the table below:

When the term gripping point is used in general, it can be preferred the gripping point with regard to the Stellerweg are considered. In the Investigation of the processes on the clutch should be between the clutch gripping point and the actuator gripping point.

The clutch gripping point can be the release path of the clutch in one predetermined point, which is on the torque characteristic of the clutch at 9 Nm is present. The release path begins with the clutch closed, i.e. H. the Clutch zero. The clutch grip point only includes the influence of the Clutch. The actuator gripping point is the path of the actuator at the point at which there is 9 Nm on the torque characteristic of the clutch. The Stellerweg begins in a so-called sniffing position, d. H. the actuator zero point. The Steller Gripping point includes influences of the clutch and the release system.

• 1. Verschiebung der Zungenposition bei geschlossener Kupplung. Dies entspricht einer Parallelverschiebung der Kennlinie des Kupplungsmoments f (Zentralausrücker-Weg) inklusive dem Nullpunkt der Kupplung. Daraus folgt, dass der Nullpunkt der Kupplung ungleich dem Nullpunkt des Stellers ist. Dies bedeutet, dass eine Parallelverschiebung der Kennlinie des Kupplungsmoments f (Stellerweg) ist. Es ändert sich somit der Steller- Greifpunkt. Beim Schnüffeln kann ein Abgleich der beiden Nullpunkte erfolgen, d. h. der Nullpunkt der Kupplung ist gleich dem Nullpunkt des Stellers. Die Kennlinie Kupplungsmoment f (Stellerweg) ist wieder ähnlich wie vorher.
• 2. Änderung des Kupplungs-Greifpunktes, wie in Fig. 4 angedeutet. Der Steller-Greifpunkt ändert sich dabei analog.

The actuator grip point can be influenced in particular by two effects:

• 1. Shift of the tongue position with the clutch closed. This corresponds to a parallel shift of the characteristic curve of the clutch torque f (central release travel) including the zero point of the clutch. It follows that the zero point of the clutch is not equal to the zero point of the actuator. This means that there is a parallel shift in the characteristic of the clutch torque f (Stellerweg). The actuator gripping point thus changes. When sniffing, the two zero points can be compared, ie the zero point of the coupling is equal to the zero point of the actuator. The clutch torque f characteristic curve (Stellerweg) is similar again as before.
• 2. Change the clutch grip point, as indicated in Fig. 4. The actuator gripping point changes analogously.

With the help of the hydraulic translation and the loss path due to the rigidity of the section, the actuator gripping point can be calculated as a function of the pressure, as shown in FIG. 5. Since the so-called sniffing has a significant influence on the actuator gripping point, it can be evaluated once with and once without sniffing, as indicated in FIG. 3.

If there is no sniffing, the tongue position is shifted, d. H. the zero position, the clutch directly on the actuator gripping point.

When using the so-called sniffing function, the body can Gripping point when compensating for the displacement of the tongue position by sniffing be calculated. Only the change of the clutch grip point is effective.

A trend line with an equation and a coefficient of determination R ^{2 is} depicted in the diagrams of FIGS. 3 to 6. The trend line and the coefficient of determination are inserted appropriately. If there is a clear linear dependency, the coefficient of determination is 1.

Coupling temperatures of up to 290 ° C were evaluated. at Temperatures greater than 300 ° C are expected to separate the clutch later and the Gripping point becomes larger again. The cause may be a change in Pad suspension characteristic of the clutch disc.

• 1. Es kann eine kurzfristige Abhängigkeit des Kupplungs-Greifpunktes von der Kupplungstemperatur festgestellt werden. Bezüglich eines Ausgangszustandes kann innerhalb eines kurzen Zeitbereiches von ca. weniger als 3 min die Änderung des Greifpunktes berechnet und in der Ansteuerung berücksichtigt werden. Die Kupplungstemperatur kann sich kurzfristig z. B. bei einem Kriechvorgang, bei mehreren Anfahrten in Folge und/oder bei einer Stauanfahrt am Berg, ändern. Es verschieben sich die Zungen der Tellerfedern, insbesondere der SAC-Kupplung und der Druck im Greifpunkt; dies entspricht der Ausrückkraft, welche sich in Abhängigkeit der Kupplungstemperatur ändert. Mit zunehmender Kupplungstemperatur (bis 300°C), kann folgendes vorgesehen sein:
  • - Der Kupplungs-Greifpunkt kann kleiner werden, wie dies die Werte und Gleichungen in der Fig. 4 andeuten.
  • - Die Zungen können sich in Richtung Motor verschieben, wie dies die Werte und Gleichungen der Fig. 6 zeigen.
  • - Der Druck, welcher der Ausrückkraft entspricht, wird im Greifpunkt kleiner, wie aus der Fig. 5 ersichtlich ist.
• 2. Es existiert eine kurzfristige Abhängigkeit des Steller-Greifpunktes von der Kupplungstemperatur, wie dies in Fig. 3 angedeutet ist. Dabei ist von großer Bedeutung, ob die Schnüffelfunktion aktiviert ist oder nicht. Mit zunehmender Kupplungstemperatur (bis 300°C) kann folgendes festgestellt werden:
  • - Der Steller-Greifpunkt ohne Schnüffeln wird größer, weil die Zungen der Tellerfedern mit zunehmender Temperatur in Richtung Motor wandern (Kupplungs-Nullpunkt verschiebt sich), wie dies durch die Werte und Gleichungen in der Fig. 3 angedeutet ist.
  • - Der Steller-Greifpunkt mit Schnüffeln wird kleiner, weil der Kupplungs- Greifpunkt ebenfalls kleiner wird und durch Schnüffeln ein Abgleich des Kupplungs-Nullpunktes erfolgt, wie dies aus den Werten und Gleichungen in Fig. 3 zu entnehmen ist.
• 3. Es kann keine Aussage über eine langfristige (absolute) Abhängigkeit des Kupplungs-Greifpunktes von der Kupplungstemperatur getroffen werden, da weitere Einflüsse auf den Greifpunkt von Bedeutung sein können, wie z. B.:
  • - Temperaturverteilung innerhalb der Kupplung, insbesondere einer SAC. Kupplung,
  • - Nachstellung der SAC-Kupplung,
  • - Setzen der Haupttellerfeder,
  • - Setzen bzw. Einbetten der Belagfederung und dergleichen.
  Weiterhin beginnt die errechnete Kupplungstemperatur erst bei ca. 90°C, welche insbesondere mit einem geeignetem EKM-Temperaturmodell ermittelt wird.
• 4. Die Verschiebung des Steller-Greifpunktes wird in der EKM-Ansteuerung berücksichtigt, wodurch in vorteilhafter Weise eine Verbesserung der Funktionalität und des Komforts der automatischen Kupplungsansteuerung ermöglicht wird.

Overall, the following findings can be determined:

• 1. A short-term dependence of the clutch gripping point on the clutch temperature can be determined. With regard to an initial state, the change in the gripping point can be calculated within a short time range of approximately less than 3 minutes and taken into account in the control. The coupling temperature can change for a short time, e.g. B. with a crawl, with several approaches in a row and / or with a traffic jam on the mountain, change. The tongues of the disc springs, in particular the SAC clutch, and the pressure at the gripping point shift; this corresponds to the release force, which changes depending on the clutch temperature. With increasing coupling temperature (up to 300 ° C), the following can be provided:
  • The clutch gripping point can become smaller, as indicated by the values and equations in FIG. 4.
  • - The tongues can move towards the motor, as the values and equations of Fig. 6 show.
  • - The pressure, which corresponds to the disengagement force, becomes smaller at the gripping point, as can be seen from FIG. 5.
• 2. There is a short-term dependence of the actuator gripping point on the coupling temperature, as indicated in FIG. 3. It is very important whether the sniff function is activated or not. With increasing coupling temperature (up to 300 ° C) the following can be determined:
  • - The actuator gripping point without sniffing becomes larger because the tongues of the plate springs move towards the motor with increasing temperature (clutch zero point shifts), as indicated by the values and equations in FIG. 3.
  • - The actuator gripping point with sniffing becomes smaller because the coupling gripping point also becomes smaller and the snapping of the coupling zero point takes place, as can be seen from the values and equations in FIG. 3.
• 3. No statement can be made about a long-term (absolute) dependence of the clutch gripping point on the clutch temperature, since further influences on the gripping point can be of importance, such as. B .:
  • - Temperature distribution within the clutch, especially an SAC. Clutch,
  • - adjustment of the SAC clutch,
  • - setting the main disc spring,
  • - Set or embed the pad suspension and the like.
  Furthermore, the calculated coupling temperature only begins at approx. 90 ° C, which is determined in particular with a suitable EKM temperature model.
• 4. The displacement of the actuator gripping point is taken into account in the EKM control, which advantageously enables the functionality and comfort of the automatic clutch control to be improved.

Furthermore, a short-term thermal effect can preferably be added Energy input into the clutch can be considered.

• 1. Eine kurzfristige, reversible Verschiebung der Position der Tellerfederzungen, welche insbesondere durch thermische Verformung infolge von Temperaturgradienten innerhalb der Anpressplatte und des Schwungrades verursacht werden kann. Des weiteren kann dies in Abhängigkeit von der zugeführten Reibleistung sein.
• 2. Eine langfristige, reversible Verschiebung der Position der Tellerfederzungen, welche insbesondere in Abhängigkeit von der Temperatur der Kupplung (Anpressplatte) sein kann.

In particular, the short-term displacement of the plate spring tongues due to z. B. considered pot effects on the flywheel and / or on the pressure plate, preferably to estimate effects on the EKM system. A lot of frictional energy can be supplied to the clutch in a relatively short time, particularly when starting at full load, starting bangs and / or so-called stall tests. Two effects can preferably have an effect here:

• 1. A short-term, reversible shift in the position of the plate spring tongues, which can be caused in particular by thermal deformation due to temperature gradients within the pressure plate and the flywheel. Furthermore, this can be dependent on the friction power supplied.
• 2. A long-term, reversible shift in the position of the plate spring tongues, which can be dependent in particular on the temperature of the clutch (pressure plate).

The long-term shift in the position of the disc spring tongues can are preferably provided in the EKM control.

For the friction power range in the vehicle, preferably the short-term Displacement of the position of the disc spring tongues and the time of regression z. B. be proportionally dependent on the friction power supplied. Of course, other dependencies are also possible.

Depending on the frictional power supplied, a so-called sniffing ban can preferably be provided. The friction power supplied to the clutch can e.g. B. are constantly determined by the EKM control, a suitable EKM clutch temperature model can be used as a basis. The so-called sniffing ban can thus be interpreted in the simplest way, as shown in FIG. 7.

These measures are preferably for a predetermined clutch Use in a vehicle with a petrol engine and in a vehicle with provided a diesel engine. Of course, the above measures can also be used in other vehicles, each with a suitable one Adaptation is possible, the z. B. can be determined from test bench tests.

Furthermore, there is a short-term behavior of the position of the plate spring tongues examined at energy input.

For this purpose, a predetermined clutch is loaded in slipping phases with parameters which are listed in the table below. The clutch can then be closed and the position of the plate spring tongues determined.

So-called potting effects preferably on the pressure plate and on the Flywheels can depend on their geometry.

In FIG. 8, the diaphragm spring fingers and the time is, inter alia, the short-term displacement .DELTA.S _k t _R, this shifting back is formed in the 90% evaluated. This results in a dependency between the short-term shift, the time and their regression on the friction power supplied. The mean values of the evaluated slip phases are shown in the table below. The scatter is approximately +/- 50%.

It can be seen that the greater the friction power supplied, the greater the shift and the time of the regression. The friction power supplied can be a product of the clutch torque and the slip speed and can be calculated using the following equation:

P _to = M _Ku .ns ~ Δs _k ~ t _R

P supplied _to frictional power
M _Ku clutch torque
n _S Slip speed
Δs _k displacement of the TF tongues
t _R time of regression

The data from the table above are shown graphically in FIG. 7. Two different axes are selected there for the displacement of the plate spring tongues Δs _k and the time for the regression t _R. Thus, the displacement of the diaphragm spring fingers and the time of recovery is shown as a function of the frictional power in Fig. 7.

For the area of clutch friction, the vehicle preferably the short-term displacement of the position of the plate spring tongues and the The time of regression is preferably proportional depending on the feed Be frictional power.

The effects on the EKM system are described below. To Slip phases with an applied friction power, which is greater than about 10 kW the position of the disc spring tongues shifts for a short time. The The clutch zero point cannot move during the shift with the set zero point match, so that the release system as a whole can be detuned. This Upset is short-lived and has no major driving behavior Influence.

During this phase, the sniff function should be deactivated, otherwise the The release system is adjusted to an incorrect clutch zero point because by sniffing the clutch zero point set to the set zero point becomes.

If during this short-term displacement of the disc spring tongues is sniffed, a wrong one can revert after regression of the shift Actuating torque characteristic for all driving situations until the next sniffing result. This should be avoided in order to have a positive impact on the Driving comfort of a vehicle with an automated clutch system to reach.

Accordingly, a so-called sniffing ban can be provided. The frictional power supplied to the clutch is continuously determined by the EKM control, the EKM clutch temperature model again preferably being used as the basis. The so-called sniffing ban can thus be interpreted without great difficulty, as is also indicated in FIG. 8.

Furthermore, the results on the temperature behavior of the clutch, especially the SAC clutch for further testing in the EKM Control incorporated in the form of a characteristic curve shift.

Attempts were made during the so-called permanent crawling against one Curb performed. It has been shown that vehicles with warm engine the compensation works correctly. For vehicles with a cold engine can under u. A slight over- or under-compensation, i. H. a Reduction or increase in the actual clutch torque.

For example, a fully loaded vehicle may stop repeatedly on the mountain become. The starting speeds can be too high Torque transmission decrease, so that u. A start on a 30% gradient may not is possible. In addition, drag torque can also engage a gear verhindem. A possible overtemperature can be caused by suitable devices be recognized in time. It has been shown that therefore a suitable Control should be kept or compensated.

It is possible that the displacement of the gripping point is provided, whereby the actuator should be sufficiently taken into account (i.e. coupling effect and Hydraulic effect).

Furthermore, a change in temperature since the last sniff can be detected the hydraulic effect is reset when sniffing.

• 1. Adaption des Greifpunktes auf den aktuell korrekten Wert und/oder die Deaktivierung der Reibwertadaption.
• 2. Sogenanntes Stall mit deutlicher Temperaturerhöhung.
• 3. Adaption des Greifpunktes (Aufholen einer etwaigen Verschiebung).
• 4. Adaption des Greifpunktes (Nachführen der Verschiebung bei Abkühlung).
• 5. Schnüffeln.
• 6. Adaption des Greifpunktes auf den korrekten Wert.

The results without compensation for a predetermined vehicle are described below. The following actions can be taken to control the effect:

• 1. Adaptation of the gripping point to the currently correct value and / or deactivation of the friction value adaptation.
• 2. So-called barn with significant temperature increase.
• 3. Adaptation of the gripping point (catching up any shift).
• 4. Adaptation of the gripping point (tracking the shift when cooling).
• 5. Sniffing.
• 6. Adaptation of the gripping point to the correct value.

In Fig. 9, the course of the temperature and the gripping point during the aforementioned actions is shown schematically. In particular, the temperature effect is made clear by the gripping point adaptation. The characteristic curve is shifted through the barn, which can be recognized by catching the gripping point. When cooling, the characteristic curve can shift back, which can also be recognized by the gripping point.

There is a temperature shift on the master cylinder of approximately 2.9 mm / 100 ° (Heating) or 6.1 mm / 100 ° (cooling). These values lie well above the predicted values of approx. 1 mm / 100 °. These values for the Warming and cooling unexpectedly differ. Since the vehicle before Measurement cooled for an hour, it can be assumed that e.g. B. Cold start effects of the temperature model influence this. When the vehicle is cold the cooling takes place faster and the temperature is lower than in Model reached. This can explain the large value of 6.1 mm / 100 °.

The presence of a temperature effect, which is qualitative with the model matches, is confirmed. The effect was less expected quantitatively.

An implementation of compensation is also provided. The Shift takes place parallel to the gripping point shift of the characteristic at Calculation of NPUNKT (in b_flag.c). This affects the shift on End of the status sequence when determining the new clutch target travel (KUPPLUNGSWEG_NORMIEREN in k_solnom.c) and also when querying the Modulation limits etc. (e.g. DRIVING in s_fahr.c).

So that the path control parts such. B. sniffing and / or towing function, can not be disturbed, the continuity of the clutch target torque query KSOLL⇒-NPUNKT are retained even under NPUNKT shifts. To If necessary, it can be provided that the size KSOLL is changed appropriately becomes.

For the calculation of the shift u. A global variable (short) TEMP_ALT may be required, which is the same when SNOWING (in s_fahr.c) the current temperature is set. This can also be done with the RUNG_INIT (in m_steuer.c) and when initializing the temperature at Engine start (in t_temp.c) can be provided, in particular to jump in temperature avoid or intercept.

In the event of an emergency run or emergency run program, the compensation can preferably slowly reduced to 0 (in_temp.c).

To be able to realize this shift, the first time Use of the model temperature, for example, at a relatively sensitive point in the Control can be intervened. It has been shown that such interventions should be tested in the longer term.

In Fig. 10 the result of a measurement with 15 Nm continuous creep with temperature increase according to the model from 130 ° C to 260 ° C is shown schematically. The compensation causes a delta GP of approx. +1.3 mm. Due to the compensation, the engine torque runs parallel to the clutch torque. Without compensation, an estimated increase in the actually transmitted torque by approx. 20 Nm would have occurred. The self-amplification, ie increased torque, thereby increased energy input and thereby greater temperature rise, thereby increased torque etc., is not taken into account in this estimation.

In Fig. 11 the result of a measurement with 15 Nm continuous creep with temperature increase according to the model from 130 ° C to 200 ° C is shown graphically. The compensation causes a Delta_GP of approx. +1 mm. Despite the compensation, the engine torque increases by approx. 5 Nm, ie it is slightly undercompensated. Without compensation, an estimated increase in the actual torque of approx. 20 Nm would be expected.

The result of a measurement at 10 Nm continuous creep with temperature increase according to the model from 150 ° C. to 240 ° C. is shown graphically in FIG. 12. The compensation causes a Delta_GP of approx. + 1 mm. The motor torque drops by approx. 5 Nm due to the compensation, ie it is slightly overcompensated. Without compensation, an estimated increase in the actual torque of approx. 15 Nm would have occurred.

When cooling down with a temperature reduction according to the model of approx. 240 ° C Creep tests are carried out at 150 ° C. The compensation causes Delta_GP of approximately -0.9 mm (i.e. the clutch is closed further and further). The peak torque of creep dependent on engine torque is despite the Compensation getting bigger, d. H. it's slightly undercompensated (or that Cooling takes place faster than modeled, e.g. B. because engine and transmission are still cold).

Without compensation, the vehicle could probably cool down when it cools down complete even less crawling.

In Fig. 13 the result of measurement at 10 Nm duration creep with temperature increase according to the model of 150 ° C to 270 ° C is shown graphically. The compensation causes a delta GP of approx. +1.3 mm. The motor torque drops by approx. 2 Nm due to the compensation, ie it is minimally overcompensated. Without compensation, an estimated increase in the actually transmitted torque by approx. 20 Nm would have occurred.

Overall, it can be stated that the described compensation can advantageously avoid a dangerous self-amplification of the clutch torque when creeping continuously. The compensation can be used in particular in vehicles that are at operating temperature. The slight over- or under-compensation can possibly be caused by the fact that the temperature model (as the trigger of the shift) is not adapted to the warm-up phase of the vehicle. Of course, an adaptation can therefore be provided in order to also advantageously use the compensation in vehicles that are not at operating temperature. A possible relationship between the actual clutch temperature and the model temperature is shown schematically in FIG. 14.

From this it can be seen that the model is particularly useful when the vehicle is cold too high a temperature of 80-120 ° C provides because the model is cooling against a bell housing at 100 ° C. The model is in this form optimized for this at high temperatures (where measures by the Control) to achieve high accuracy. As a result, it makes sense that with a further development of the temperature model as Transmission bell temperature the known engine temperature is used. Of course, too other measures for optimizing the temperature model are possible.

• - Wenn der Greifpunkt und/oder der Reibwert zu tief ist, ist das tatsächliche Kriechmoment größer als das gesteuerte Kriechmoment, demzufolge ist die tatsächliche Erwärmung größer als die modellierte Erwärmung. Daraus ergibt sich eine Unterkompensation, wodurch die Auswirkung des GP/RW- Fehlers reduziert wird.
• - Wenn der Greifpunkt und/oder der Reibwert zu hoch ist, ist das tatsächliche Kriechmoment kleiner als das gesteuerte Kriechmoment. Demzufolge ist die tatsächliche Erwärmung kleiner als die modellierte Erwärmung. Somit liegt eine Überkompensation vor, wodurch die Auswirkung des GP/RW- Fehlers verstärkt wird.

It is possible that the influence of characteristic fields on the described compensation can be predetermined qualitatively:

• If the gripping point and / or the coefficient of friction is too low, the actual creeping torque is greater than the controlled creeping torque, and consequently the actual heating is greater than the modeled heating. This results in undercompensation, which reduces the impact of the GP / RW error.
• - If the gripping point and / or the coefficient of friction is too high, the actual creep torque is less than the controlled creep torque. As a result, the actual warming is smaller than the modeled warming. Thus there is overcompensation, which increases the impact of the GP / RW error.

The compensation can preferably depend on whether the characteristic correct is. If the above GP / RW error amplification caused instability of the Adaptions leads, it could be checked whether the adaptation result if necessary large compensation is rejected.

A further embodiment of the invention presented here can in particular involve electromechanical sniffing. With an automated Coupling controls become special to the transmission line and the coupling Requirements. If e.g. B. a release travel measurement is dispensed with and only a path measurement is carried out in the actuator, changes should be made the clutch or on the transmission link are constantly corrected. To can e.g. B. defined points of the clutch characteristic can be used. one these points can preferably be the clutch that is closed when no power is applied. Here the fact that the transmission link can also be used becomes powerless. Hydraulic release systems take advantage of this condition Adjustment of the zero point. The connection system open / closed can doing z. B. be established via a path relationship.

For cost reasons, you want automated clutch actuation often do without hydraulic transmission, but then need one another mechanism for detecting the zero point. Accordingly, one should Mechanism can be found, the z. B. on an electromechanical basis a power relationship works.

• 1. Schalter und Elastizität parallel zum Kraftfluss
• 2. Schalter in Reihe zum Kraftfluss und Elastizität raumfest abgestützt.

It can e.g. B. be provided that the zero point is detected via the absence of force in the transmission link by means of an elasticity and / or an electrical switch. For this, e.g. B. the following variants are provided:

• 1. Switch and elasticity parallel to the flow of force
• 2. Switch in series supported for power flow and elasticity.

• - Sehr geringer Zusatzaufwand bei geschickter Integration
• - Mechanismus kann für beliebige Ausrückkraftcharakteristik der Kupplung verwendet werden (Zug und Druck mit mehreren Nulldurchgängen)
• - Keine Beschränkung des Modulationsbereichs
• - Leerweg kann stark reduziert werden
  (Hydraulik: Leerweg bis zu 3,0 mm möglich)
• - Lage des Nullpunkts unabhängig von Verfahrgeschwindigkeit und Temperatur
  (Hydraulik: starke Abhängigkeit von Temperatur und Verfahrgeschwindigkeit)
• - Lage des Nullpunkts konstant über Lebensdauer
  (Hydraulik: Leerweg verändert sich mit Verschleiß der Dichtung etc.)
• - Größe der Vorlast kann eingestellt werden und bleibt über Lebensdauer konstant
  (Hydraulik: Vorlast ändert sich um bis zu 30%)

This has at least the following advantages:

• - Very little additional effort with skillful integration
• - Mechanism can be used for any clutch release force characteristics (pull and push with multiple zero crossings)
• - No restriction of the modulation range
• - Empty travel can be greatly reduced
  (Hydraulics: free travel up to 3.0 mm possible)
• - Position of the zero point independent of travel speed and temperature
  (Hydraulics: strong dependence on temperature and travel speed)
• - Position of the zero point constant over the service life
  (Hydraulics: free travel changes with wear of the seal etc.)
• - The size of the preload can be set and remains constant over the service life
  (Hydraulics: preload changes by up to 30%)

• - geschlossene Kupplung = Übertragungsstrecke kraftfrei = Nullpunkt
• - Greifpunkt = Kupplung überträgt ein definiertes Moment

The transmission link establishes the correlation between the coupling and the actuator. If a disengagement distance measurement is dispensed with, the behavior of the clutch and transmission link should be checked continuously. For this purpose, at least two prominent points of the clutch characteristic can be used by the clutch control. For example:

• - Closed coupling = transmission path without force = zero point
• - Gripping point = clutch transmits a defined torque

There are numerous effects on the coupling and the transmission link that make it necessary to correct the zero point. The table below shows the main effects and an approximate size. Installation tolerances were not taken into account, these can be compensated for by an initial setting.

The quality of the clutch control depends in particular on the accuracy with which the zero point is determined. The feature can e.g. Legs power-free transmission path.

Various mechanisms for detecting the Described zero point. In the case of a hydraulic transmission link, the Hydraulic zero adjustment. By opening the transmission path it is depressurized and the clutch can reach zero to adjust. The length of the transmission line is now hydraulic adjusted (fluid compensation via fluid reserve). This process is "sniffing" called.

It should be noted here that the "release of force" of the transmission link z. B. is path-controlled and inevitably increases the tolerance chain becomes. This can lead to a sniffing game (loss path) of up to 3 mm can be (15% of the useful stroke of the actuator). Furthermore, this is The loss path is not constant across all operating states. An influence can e.g. B. especially the fluid viscosity and the travel speed. Because that Sealing element is destroyed when the direction of movement is reversed Sealing element takes place as long as it is in contact with the sniffer bore and a If there is pressure, the modulation range must be restricted. This can u. U. to a significant loss of comfort in the Lead clutch control.

The clutch is controlled via a mechanical Transmission path, the zero point adjustment can be carried out according to the one presented here Mechanism done in a simple manner.

• - Es ist z. B. denkbar, dass ein elastisches Element zu der Übertragungsstrecke parallel geschaltet ist. Wird der Kraftfluss in der Übertragungsstrecke z. B. unterbrochen, weil der Nullpunkt der Übertragungsstrecke erreicht ist, so wirkt allein die Federkraft. Diese kann über einen Federweg z. B. durch einen elektronischen Schalter exakt detektiert werden.
• - Es ist auch möglich, dass zu der Übertragungsstrecke in Reihe geschaltet wird. Die Vorlastfeder kann z. B. raumfest angebracht werden. Wird der Kraftfluss unterbrochen, so kann der Schalter dadurch ausgelöst werden.

The release system can e.g. B. biased by a preload on the release bearing. The release bearing / release system always follows the movement of the plate spring tongues. This requirement is always met with conventional release systems.

• - It is Z. B. conceivable that an elastic element is connected in parallel to the transmission link. If the power flow in the transmission path z. B. interrupted because the zero point of the transmission path is reached, the spring force alone acts. This can via a travel z. B. can be detected exactly by an electronic switch.
• - It is also possible to connect in series to the transmission link. The preload spring can, for. B. be attached fixed. If the power flow is interrupted, the switch can be triggered.

• - Nullpunkt kann sehr genau bestimmt werden.
• - Mechanismus kann an einer beliebigen Stelle der Übertragungsstrecke angebracht werden. Er kann also auch in den Kupplungsaktor integriert werden.
• - Wesentlich höhere Genauigkeit des Nullpunkts und konstante Lage über alle Betriebsbedingungen im Vergleich zur Hydraulik.
• - Keine Beschränkung des Modulationsbereichs.
• - Mechanismus kann für beliebige Ausrückkraftcharakteristik der Kupplung verwendet werden (Zug und Druck mit mehreren Nulldurchgängen).

This has at least the following advantages:

• - The zero point can be determined very precisely.
• - Mechanism can be placed anywhere on the transmission link. It can also be integrated in the clutch actuator.
• - Much higher accuracy of the zero point and constant position over all operating conditions compared to the hydraulics.
• - No restriction of the modulation range.
• - Mechanism can be used for any clutch release force characteristics (pull and push with multiple zero crossings).

In Fig. 15, two mechanisms of action for electromechanical sniffing are exemplified. Two basic sketches show the necessary system components. The mechanism for the detection of the zero point can be attached at any point in the transmission path. A location near the actuator is advisable. This eliminates the need for complex electrical plug connections and increases the signal quality.

Variant 1 in Fig. 15 shows a preload spring connected in parallel with the power flow, so that a constant preload can be set via the switch (equivalent to the pre-pressure valve)

Variant 2 in FIG. 15 shows a spatially fixed preload spring, this corresponds to an arrangement customary today, here too the absence of force in the transmission link can be detected by a sensor.

In Fig. 16 is schematically shown that the mechanism with a corresponding design can also detect multiple zero crossings of the Ausrückkennlinie. It is conceivable that z. B. a length compensation can be done by a path adjustment on the route and / or on the actuator.

A further embodiment of the present invention follows described, which relates to a transmission which, for. B. in the forward drive a high and Z. B. in the reverse drive has a low efficiency, so that Gearbox self-locking, self-holding and / or when load is applied is self-braking.

The efficiency of drive gearboxes should be increased, e.g. B. the Reduce power requirement or the output power at the same Increase input power. In particular, a gear should be found which is used for an electronic clutch management (EKM) is provided and which has a high efficiency from the drive side and from the load side, as already mentioned, self-braking, is self-locking and / or self-holding. This is intended in particular for the clutch actuator without additional power consumption of the drive in a predetermined Be held in position. Accordingly, a propulsion system is proposed which in the drive direction, d. H. Motor for load for back and forth movement, one has high efficiency and, conversely, less Has efficiency, so that the entire transmission without counteracting Drive power is self-holding, self-locking or self-braking.

According to an advantageous development of the present invention, this can be done a helical gear, a double helical gear or equivalent trained rack (special case of a gear, but with infinite Radius) in an actuator that is used to actuate clutches, be provided. The helical gear can be used both as a single and as Intermediate link can be used in a strand, which is a drive gear and has an actuator gearbox for actuating clutches.

It is possible that in connection with a clutch actuator rotary and / or linear input, intermediate or output for electromechanical and / or hydraulic actuation of a clutch and in combination with a electric drive and an incremental and / or absolute travel or Angle measurement is used.

As a result, in the event of an action of force (load to motor) the gear train no additional energy for self-holding and / or Self-braking can be applied. The clutch actuator and thus also the Clutch is in a predetermined position without power consumption of the drive held.

It is particularly advantageous that the overall efficiency of the transmission of the clutch actuator in the drive direction of the drive forces or moments (Motor to load for back and forth movement) is significantly higher than against the Direction of load forces or moments (load to motor for outward and return Return movement).

It is also possible that e.g. B. a ball screw drive is integrated, which itself characterized by a high power density and high efficiency and can serve for rotation-translation conversion.

Accordingly, the proposed transmission has the property of being external To act self-locking or self-braking. Thus the force of the clutch spring (uncompensated or compensated by a counteracting force, here mostly spring force) or the resulting one Torque that acts on the clutch actuator is suitably compensated. A Further development of the invention can provide that a modified Helical gear is used, which, for. B. in a switching and / or selection actuator Manual transmission or used in a clutch release system.

It is conceivable that the system-related high axial forces on the Gearing z. B. to operate a brake or a mechanical, electrical or electromechanical system can be used.

The proposed drive system can of course also be used for others Use cases as a basis, so that taking into account the the respective properties of each application also in these applications Use is possible.

Further advantages and advantageous refinements can be found in FIGS. 17 to 19, a first exemplary embodiment of the drive system according to the invention being shown in FIG. 17, in which a screw wheel thread or a double screw wheel thread is provided. In Fig. 18, a second embodiment of the drive system of the invention is shown. FIG. 19 shows a third exemplary embodiment of the drive system according to the invention, in which a ball screw drive is provided as an example of the rotation-translation conversion in addition to the double helical gear.

A further embodiment of the present invention follows described in which a method for monitoring the EKM and / or ASG Control is proposed.

It has been shown that in particular when the control unit is subjected to extreme loads by interrupt events, e.g. B. for speed counts with high frequencies Malfunctions, insufficient processing time for complete processing of the activation process is available. It can happen that a so-called error bit is used and individual measures Parts of the process or the entire control process is suspended.

Accordingly, e.g. B. long-term monitoring with EKM and / or ASG Control provided to avoid the situation described above.

It is possible that e.g. B. a so-called bit "Control_Active" is set, which is preferably only at the end of the process is deleted again. If the control is called again and this Bit is still set, there may be a timeout. As Remedial measures can e.g. B. be provided that stored in a so-called error bit becomes. This can be used to bypass the control to computing time and Save storage space. Thus, for example, one can advantageously The cause of the crash can be counteracted by a "memory overflow".

A further embodiment of the present invention follows described, in which a so-called abuse protection in particular so-called "garage shift" is provided. Under a "garage shift" means shifting from R-gear to D-gear in an automatic transmission or with an ASG gearbox. In particular, noise ratings are used (Noise when engaging the gears) and evaluations of the gearshift when A gear is made. This circuit is in an evaluation often repeated in succession. It is also possible for the driver to do this Circuit often makes consecutive for certain reasons. For example, simply by playing on the gear lever or by consciously Try it out to get closer to the noises that often occur with these circuits analyze.

Especially with these shaft operations, e.g. B. if you frequently very quickly in succession, the switching and selector motors of the ASG actuators can be powerful be charged. This leads to increased power consumption and u. U. to one excessive heating of the motors. It is possible that this warming even leads to the ASG being switched off. Accordingly, the present Invention proposed a method by which the movement of the motors beforehand and / or when a predetermined temperature model responds (or Exceeding a measured engine temperature) is prevented.

It can be provided that when reverse gear or forward gear is requested, unless other lock-out conditions exist; this is inserted immediately. According to the present invention, it is initially provided that the garage shift situation is recognized in a suitable manner. This can be done by the method described below, which is shown in FIG. 20 in the form of a block diagram.

First, a counter (Z_G) can be initialized to the value zero. This Counter can record the number of garage shifts. This can be done happen that when the selector lever is moved to the R or D or M position, the Counter is increased by 1. A time counter (Z_Z) can preferably be displayed simultaneously a certain value (Z_0) can be initialized. During normal operations this counter can be decremented until it has the value zero reached. As soon as the counter has reached zero, the process can begin start over. However, if the selector lever during the measured time (Time counter> 0) is brought back into the R or D or M position, the Garage shift counters are incremented by 1 and the time counter returns to its Original value to be reset. From this it can be determined that a garage shift was carried out within the defined period of time.

When the number of garage shifts reaches a predetermined number (e.g.> 5), a variety of garage shifts can be recognized and a corresponding one Action to be taken to prevent another garage shift. It it is also conceivable that another type of time counting, such as e.g. B. by a moving average or the like is used. Of course it can The methods described above can also be modified appropriately in order to achieve a Further optimize abuse protection.

In particular, the conditions under which the shift counter is incremented can be extended by suitable conditions. So it is z. B. possible the Only increase the counter if the hand and foot brakes are activated at the same time becomes. These signals can indicate that the driver has no intention of going rank. In addition, it is also conceivable that the inclusion of the Idle switch and / or the accelerator pedal position or the like is provided.

According to another development of the present invention be provided that in the detection of a variety of garage shifts desired gear is not initially engaged. In this way, the Electric motors of the switch actuator are no longer loaded and the current consumption goes back to zero. It is also conceivable that these measures be provided if (possibly for other reasons) the temperature of the Electric motors of the switch actuator is too high. This temperature can be both measured as well as suitably calculated from a temperature model. To enable the driver to start driving when necessary, the gear should still be inserted if the driver z. B. the accelerator pedal (recognizable from the Accelerator pedal position or the idle switch).

According to another development of the present invention, instead of one complete prevention of the switching process when operating the selector lever too be provided that the switching process is suppressed only as long as the Foot or hand brake are operated. So it is advantageous possible that the gear is engaged when releasing these brakes and the usually provided creep function, if somewhat delayed, is activated. It it is also conceivable that the proposed prevention strategies or Anti-abuse strategies are used in succession or in stages.

If it is assumed that this causes the movement of the switching actuator is triggered that the selector lever from the position R to D or M, from the Position D (M) is moved to R or from position N to R or D (M). However, it is also conceivable that such circuits are triggered by this; that the winter program is selected and deleted periodically or in manual mode in standstill about 1 → 2, 2 → 1 operations can be carried out. The above Strategies for preventing circuits described can also be used this application. It is also possible that the The lock for garage shifts is automatically released after a predetermined Period can follow. The above-described method for Protection against misuse in garage shifts can also be modified appropriately further optimize this process.

A further embodiment of the present invention can be adapted in particular due to the activation of the EKM and / or the ASG system of information or conclusions from the active chassis or from Provide road conditions and / or lateral dynamics.

There are others in the implementation of active chassis in motor vehicles Sensor signals are available, by evaluating this information Sensors on road conditions, longitudinal and lateral dynamics of the Vehicle to be closed. This can advantageously work the EKM and / or the ASG system in special situations be improved.

It is Z. B. possible that due to suitable sensor signals of the active Chassis of the motor vehicle, preferably so-called rough road sections be recognized. However, this should influence the sensitivity of the Gear detection can be avoided to rule out error detection.

Furthermore, it can be provided that, for. B. due to the longitudinal and / or Transverse dynamics of the motor vehicle an existing driving behavior is detected. For example, it can be provided that the engagement strategy and / or the shift line design is adapted appropriately.

A further embodiment of the present invention follows described, in particular taking into account path hysteresis in the Control of the clutch preferably in an electronic Coupling management (EKM) is provided.

With a hydraulic release system, a force hysteresis (pressure hysteresis) intended. Due to the elasticity of the release system, the Hysteresis also as path hysteresis between the master and / or slave cylinder impact. This path hysteresis can depend on the direction of movement Lead inaccuracies in the control of the master cylinder.

For example, the change in pressure hysteresis (d_hyst _{master cylinder} ) can depend on the following factors:
temperature
print
wear
pollution
Travel speed or the like.

The path hysteresis and thus also _{hydraulically} can depend on the elasticity of the actuation path, such as. B .:
Diameter of master and slave cylinders
Other geometries of the cables (e.g. length, diameter, etc.)
Cable materials
temperature
print
Aging of the mechanical elements (e.g. hoses)
Aging of the active medium

Of course, there may be other dependencies on the change the pressure hysteresis or the path hysteresis.

Based on the aforementioned influences, it can be determined that the path hysteresis in particular is dependent on many variable factors. The accuracy with which the path hysteresis is taken into account depends on the knowledge of the influencing factors. The influences, in particular on the d_hyst _{master cylinder} and hydraulically, can be measured and / or calculated. Of course, it can also be determined from suitable characteristics (e.g. elasticity over age).

An advantageous development of the present invention can provide that the path hysteresis is appropriately taken into account in the control of the clutch. There are several options for this, which are described below:
It is Z. B. possible that, for. B. a predetermined characteristic curve for the assignment between the master and slave cylinders is stored in the control for each direction of movement. With this possible variant, an average error in particular can be taken into account.

It is also conceivable that, for. B. a characteristic curve for the assignment between the master and slave cylinders is stored in the control for each direction of movement, the distances between which are preferably made variable. With this option, changes with regard to the d_hyst _{master cylinder} can be taken into account in an advantageous manner.

Another possibility of the present invention can provide that a characteristic curve for the assignment between the master cylinder and the slave cylinder is stored and z. B. a second characteristic curve is calculated via the changes in the pressure hysteresis d_hyst _{master cylinder} . Changes in the pressure hysteresis d_hyst _{master cylinder can be} taken into account.

Furthermore, it is conceivable that a middle characteristic curve for the assignment between the master cylinder and the slave cylinder is stored and z. B. can be calculated proportionately via the changes in the pressure hysteresis d_hyst _{master cylinder} . With this option, changes in the pressure hysteresis d_hyst _{master cylinder} can again be taken into account in an advantageous manner.

According to a development of the present invention, it is also possible that a characteristic curve for the assignment between the master cylinder and the slave cylinder (i _hydraulic ) is calculated and z. B. the second characteristic curve is calculated via d_hyst _{master cylinder} . In this embodiment, changes in the path hysteresis i _hydraulic and in the pressure hysteresis d_hyst _{master cylinder} are taken into account in an advantageous manner.

A further possibility can be that an average characteristic curve for the assignment between the master cylinder and the slave cylinder (i _hydraulic ) is calculated and from this calculated proportionately by changing the pressure hysteresis d_hyst _{master cylinder,} preferably both characteristics. With the possibility described above, changes in the path hysteresis i _hydraulic as well as changes in the pressure hysteresis d_hyst _{master cylinder can be} sufficiently taken into account.

The possibilities described above, preferably the path hysteresis in the Coupling control to be taken into account can be suitably combined become. Other options may also be provided to further improve the present invention.

For a better understanding of the present invention, the following are the Effects of force hysteresis on release paths described.

An automatic clutch can be actuated by a hydrostatic Path transmission path from the master cylinder via a line to Central release take place. Due to elasticity, the hydraulic Change translation depending on the pressure. This can cause the force hysteresis and the resulting pressure hysteresis on the release path.

Due to force hysteresis occurs when the direction of the master cylinder is reversed a release path behavior of the release piston of approx. 0.5 mm. By Due to dynamic influences, the hysteresis can decrease, but the Influences of temperature can in turn increase the hysteresis. The The clutch characteristic curve can have a gradient of approx. 30 Nm / mm. This means, that the hydraulic line with a dead path behavior of approx. 0.5 mm with their current control has an insensitivity of approx. 15 Nm. Therefore is an exact one with the current state of the control Moment tracking is not absolutely possible.

In the case of a rigid system, there are no elasticities in the hydraulic transmission path, so that after the sniffer bore path has been subtracted from the master cylinder path, the release path can be calculated via the hydraulic translation, as is also indicated in FIG. 21. A corresponding equation is given below:

i _{hydraulically stiff} = A _releaser / A _{master cylinder}

s_ausr stiff s_MS / i _{hydraulically stiff} = (s_GZ - s_NP) / i _{hydraulically stiff}

In the case of a rigid system, the hydraulic transmission link is actually designed to be elastic. The pressure acting in the hydraulic line affects all pressurized hydraulic components and leads to elastic deformations. Part of the volume displaced by the master cylinder can be used up by the elastic deformation of the hydraulic components, ie the releaser moves a smaller distance than that in the ideal rigid system. The following formulas result from this:

i _hydraulic <i _{hydraulic rigid} for p> 0

s_ausr <s ausr _stiff - for p> 0

The situation described above is shown in particular in FIG. 22. In Fig. 22 the pressure dependence of the difference from the release path is stiff and from the release path is shown elastically (see also Fig. 22: d elast). This difference as a function of the pressure (d_elast = f [p]) is investigated for each individual hydraulic component and, in the method according to the invention for calculating the release travel to the master cylinder travel, is combined below.

The path deviation of the releaser can depend on the line pressure and thus also on the force hysteresis. The force hysteresis should therefore express itself in the form of path hysteresis, as indicated in the measurement shown in FIG. 23.

In Fig. 24 a so-called Totwegverhalten the releaser is shown. The master cylinder first extends so that the line pressure moves in the upper part of the pressure hysteresis, which is caused by the force hysteresis of the central release and the clutch. The release can move on the lower path hysteresis. If the master cylinder stops, the releaser remains in position a. When the master cylinder is retracted, the line pressure drops first (point b). However, this cannot cause the releaser to move out, so that the releaser remains in its position (point b) until the actuator travel has moved back so far that the releaser can safely move back out of its upper path hysteresis. The same effect occurs again when the master cylinder changes its direction from "retract" to "extend".

The calculation of the dead path behavior is indicated in FIG. 25. From the pressure hysteresis, the path hysteresis of the releaser can be calculated (d_hyst) by assigning pressure to displacement loss at the releaser. The gradient angle α can be calculated using the hydraulic ratio as follows:

tan α = d_ausr / d_st = 1 / i _hydraulic

The dead travel on the master cylinder can be calculated from α and d_hyst as follows:

Dead path = d_hyst / tan α

Dead path = d_hyst.i _hydraulic

If the misalignment of the master cylinder is known for a given pressure hysteresis, the dead travel can be calculated as follows:

Dead travel = d_hyst.i _hydraulic = (d_hyst _{master cylinder} / i _hydraulic ) .i _hydraulic

Dead travel = d_hyst _{master cylinder}

For a predetermined vehicle, the magnitudes of the Dead path behavior can be specified. This occurs in the work area Pressure hysteresis of 3 bar. For 3 bar, measurements (elasticity of the hydraulic path) a misalignment on the master cylinder of 0.6 mm. This results in from the equations above that the dead travel also has the value of Assumes 0.6 mm.

It should also be noted that the pressure hysteresis and thus the Reduce dead travel due to dynamic influences in the vehicle, however Temperature influences can also increase this effect.

In the partial load range of the master cylinder, 0.5 mm travel corresponds to Master cylinder with a torque change of around 15 Nm on the clutch. If that Clutch torque should be adjusted sensitively and the Direction of movement of the master cylinder is to be changed, there is a displacement of controlled torque of approx. 15 Nm. Without appropriate measures the hydraulic route can only have a sensitivity of approx. 15 Nm. This is z. B. the case with a suitable moment tracking.

As a remedy, e.g. B. the storage of two actuating torque characteristics are provided, which are preferably read in depending on the direction. It can be provided that a second actuating torque characteristic is preferably also realized by an off-set, which, for. B. is temperature dependent. In doing so, precise knowledge of the hysteresis behavior should be taken into account the temperature and / or dynamic influences are taken into account.

A further embodiment of the present invention follows described, in which preferably a speed-dependent compensation of Consumer moments is provided.

In particular with EKM and / or ASG controls, it can happen that the friction value adaptation based on the engine torque tends to lead to lower coefficients of friction when starting with the air conditioning system switched on than when the air conditioning system is switched off. This is particularly due to the fact that the control system calculates a lower engine torque when starting up than when the air conditioning system is switched off. The motor torque can be calculated as follows for EKM and / or ASG control:

MM = MM_CAN - (M_KOMP + Offset_im_Leerlauf)

With

MM_CAN = mmoti - mmotsch = indexed engine torque = drag torque. CAN signal

M_KOMP = air conditioning compressor torque (CAN signal)

Offset_im_Leerlauf = Consideration of any other consumers, if MM_CN - M_KOMP> 0 Nm applies during idle.

It has been shown that the compressor torque sent via the CAN M_KOMP is output regardless of engine speed, d. H. while an arrival is the speed-independent term (M_KOMP + Offset_im_Leerlauf) subtracted from the engine torque. The two last summands in the above equation are only correct, if it can be assumed that consumers (air conditioning, Generator etc.) can be operated with constant torques.

Rather, it can be assumed that the main consumer Air conditioning, generator, etc. are operated with approximately constant power, d. H. that the torque decreases hyperbolically with increasing engine speed. Under this assumption would apply to the starting speed according to the previously valid equation too high a consumer torque (M_KOMP + Offset_im_Leerlauf) from Subtracted engine torque, which is only valid for the idle speed. This in turn means that when the air conditioner is switched on it tends to be on smaller engine torque is calculated than when the air conditioning is switched off, which in turn tends to result in a lower coefficient of friction.

Accordingly, assuming that the essential consumers require an approximately constant input power, the above equation for calculating the engine torque can be adapted accordingly.

MM = MM_CAN - (M_KOMP + Offset_im_Leerlauf). Idle speed / n_Motor

Of course, the above equation can be further modified to the to further improve speed-dependent compensation of consumer torques.

A further embodiment of the present invention follows described, in which a possibility is given, in particular the tightening display the parking brake appropriately.

The engine can be switched off, especially when adhering to the mountain the ignition key is removed. The clutch can be in gear then be closed slowly or with a delay, so that the vehicle is at Release the foot brake a little. This can be advantageous to the driver Way that the vehicle does not yet have the parking brake is held. This can prevent the vehicle from using only the Parking lock function of the engine is held and z. B. with a small pulse can roll off.

A further embodiment of the present invention is described below described, in which the clutch characteristic preferably in a 3-parameter Representation z. B. is shown in the so-called EEPROM.

When activated in the EKM and / or ASG system, the Coupling characteristic as a polyline with z. B. 6 support points are stored in the so-called ROM. The clutch characteristic can be particularly with regard to Torque tracking can be crucial for control. For this reason, the Coupling characteristic through the gripping point and the coefficient of friction during operation be adapted. The corresponding shifts in the characteristic are shown in Transformations of the support points converted and realized as a result. The The characteristic curve can be adapted mainly in the area of large (coefficient of friction) and smaller (Gripping point) consider moments. Qualitative changes in the characteristic (e.g. due to wear or the like) in the middle torque range cannot be fully compensated for by these adaptations. This can problems in the long run, especially with low overpressure (KME factor), in particular lead to slippage. Accordingly, e.g. B. be provided that a third adaptive parameter is introduced. However, such an adaptation is only to implement with relatively great effort and the informative value is limited.

Therefore, in particular according to a further development of the present Invention can be provided that the clutch characteristic preferably by analytical function with e.g. B. 3 parameters is represented. It is Z. B. possible that two of these parameters directly through the adaptation (coefficient of friction, gripping point) being affected. The third parameter could e.g. B. not adapted, but z. B. external to be changed.

It follows that the characteristic curve in the form of the 3 parameters z. B. in the so-called EEPROM is stored. Two of these parameters are adaptive variables and the third parameter can e.g. B. by external test equipment (e.g. from customer service) based on the functionality can be changed directly.

It can be provided that based on possible replacement strategies the faulty 2 out of 3 selection a suitable replacement characteristic from the so-called ROM is used.

• - Die Berechnung des Kupplungsmomentes ist durch die analytische Funktion einfacher.
• - Der Einfluss von qualitativen Kennlinienveränderungen auf die Funktionalität des Systems kann dann leicht durch Testgeräte simuliert werden, ohne dass sogenannte Source-Code-Änderungen vorgenommen werden müssen.
• - Der wesentliche Vorteil liegt in der nachträglichen Manipulierbarkeit der Kennlinien, sodass eine Nacharbeit im Produktionswerk oder beim Kundendienst möglich ist.

At least the following advantages can be given in this procedure:

• - The calculation of the clutch torque is easier due to the analytical function.
• - The influence of qualitative changes in the characteristic curve on the functionality of the system can then be easily simulated by test devices without having to make so-called source code changes.
• - The main advantage lies in the fact that the characteristic curves can be subsequently modified, so that reworking in the production plant or at customer service is possible.

With the proposed method, problems in particular with the Customer service on the vehicle can be eliminated relatively easily, taking experience can be used. In addition, production-related spreading of the Characteristic curves are compensated.

Of course there are also other modifications of the invention Procedure possible to further optimize the representation of the coupling line.

A further embodiment of the present invention follows described, in which preferably the starting map in an EKM and / or an ASG control is suitably modified.

It has been shown that maneuvering operations on the mountain are particularly difficult design, so that in particular a suitable dosing function when starting should be provided. The clutch torque can e.g. B. linear as a function the accelerator pedal position can be set, however, if the engine is small Pedal values already provide a significant moment, this can lead to problems.

An engine revs can z. B. be provided, since then Motor torque reduced again. For example, the metering speed should then be up to approx. 4000 rpm to be allowed for moments up to B. 50 Nm with To be able to dose pedal positions up to 40 °. With this procedure, however, the Wear can be increased.

Another possibility can be that the clutch torque is set depending on the engine speed, preferably a suitable one Map is used, which between the idle speed and about 2000 rpm increases parabolically from 0 Nm to the maximum engine torque. It has shown that this parabolic function does not modify as much as desired (deflected), because this could improve the meterability, however, the normal approach would be swinging (steep part of the curve).

• - Die Anfahrt erfolgt aus der Handbremse heraus, d. h., dass die Anfahrt in Abhängigkeit von dem Handbremssignal durchgeführt wird.
• - Kleine Pedalwerte (Schwelle z. B. 30%)
• - Raddrehzahlen (bei stehendem bzw. langsamen Fahrzeug dosieren, sonst anfahren)
• - Pedalwertänderung (dosieren bei langsamen Gasgeben, sonst anfahren)

According to an advantageous development of the present invention, this can preferably be improved in that the starting map in certain situations by the application, for. B. a factor is appropriately modified. The factor can e.g. B. assume the value (<1) so that the starting map is made flatter. It is possible that a suitable transition to the normal starting map (factor 1) is also provided. The following can be provided as possible conditions for the flat map:

• - Approach takes place from the handbrake, ie the approach is carried out depending on the handbrake signal.
• - Small pedal values (threshold e.g. 30%)
• - Wheel speeds (meter when the vehicle is stationary or slow, otherwise drive off)
• - Pedal value change (dose when accelerating slowly, otherwise start off)

These conditions mentioned above may also be suitable be linked together. It is also possible that other conditions be provided.

Another development of the present invention can provide that a Full engine revs with the use of a suitable map can be combined with each other. For example, it can be provided that under under certain conditions a high revs of the engine continuously in suitable map is transferred. As conditions such. B. the already mentioned conditions are used. Of course, others can suitable conditions are provided in this configuration to a to achieve the best possible starting strategy for the vehicle.

A further embodiment of the present invention follows described in which, in particular, blocking the actuator of the clutch a suitable reaction is provided.

In particular with an EKM = system it can e.g. B. to diagnose the condition the coupling, the actuator and / or the hydraulic route may be advantageous, if it is recognized whether the actuator z. B. is blocked. In normal driving operation can the actuator z. B. in the event of a deviation of the actual path from the target path for example, switch off due to excessive load (blocking, stop). That state can only be switched off and on again at any time be withdrawn. Another way to overload the actuator detect, may be that the motor current is monitored appropriately (e.g. PWM signal).

It has been shown that in the stroke detection during diagnosis, for. B. in customer service, switching off the actuator and the related Switching the ignition off an illegal interruption of the diagnostic procedure represents.

According to an advantageous development of the present invention therefore it should be provided that the diagnostic routines in which the danger exists to drive to the stop, the actuator does not warn of overload off. Instead, it can be provided that the target path after a predetermined period of time is adapted to the actual path that the Control deviation reaches an amount that does not thermally overload the actuator and the outputs the actual route.

Preferably, in normal driving, the adaptation of the target path to the actual path can enable a sensible reaction to the detection of the overload of the actuator. In normal driving, other criteria can also be taken into account if the target paths are adjusted. Possible prerequisites are given below:
The gear detection sensor system is ready for operation.
The actuator is ready for operation.
An intelligent workshop test device can facilitate communication with the system.

In the present invention, the following procedure provided in the table below can be used:

A further embodiment of the present invention follows described, in which a so-called bad road detection preferably at an EKM and / or ASG system is provided. In known systems no bad road detection is carried out and therefore neither can appropriate measures are initiated if the vehicle is hit by heavy Railing is moved.

According to a further development of the present invention, it can be provided that the EKM and / or ASG system is used for poor road detection, which can further increase the comfort of the overall system in an advantageous manner. It is conceivable that, for example, the wheel speeds are suitably evaluated. If there is a rough road, the speed ratios shown in FIG. 26 may result. As in FIG. 26, the wheel speeds of the ABS sensors of the driven axle are analyzed, with a suitable speed difference of the wheels being shown on a curve and on a rough road.

It can e.g. B. be provided that all four wheel speeds be evaluated in order to carry out an exact bad road detection (Taking speed differences into account when cornering). Of course you can too fewer and / or also different signals are used for this.

It is Z. B. possible that a digital switchover between the conventional operation and the rough road operation is provided. The description of the poor road condition can e.g. B. by a suitable classification with z. B. a rough road key figure SK from 0 to 100%. The rough road key figure SK can e.g. B. be a function of frequency and / or amplitude of the speed difference. If a rough road is recognized, e.g. B. the following actions are initiated:
EKM / ASG: Closing the clutch (no torque tracking)
Strategy: increasing the kme value
Increase the minimum torque
Transition to the sniffing state
ASG: Increasing the switching hysteresis to prevent switching oscillations

The kme value is primarily dependent on the engine speed (or the transmission input speed) and the engine torque. If the rough road characteristic number SK is increased, the kme value is increased, as shown in FIG. 27, kme_kon corresponding to the conventional kme value, ie without taking into account the rough road detection.

The minimum torque can be _dependent on the engine speed and the current gear in a conventional manner (without taking into account the _{rough road detection} ), as shown in FIG. 28, the minimum _{torque being} denoted by M _{min_kon} . An increased rough road index leads to an increase in the minimum torque, as shown in FIG. 29.

The increase in the minimum torque can be additive, multiplicative or by means of a characteristic curve, e.g. B. with a non-linear approach. A multiplicative factor k for the moment is presented as an example:

M _min = kM _{min_con}

where the range of values is in the range of:


lies and M _{r_max corresponds to} the maximum possible clutch _torque that can be transmitted.

Another possibility is to increase the switching hysteresis, the procedure for determining the increased switching hysteresis being presented using a shift from second to first or from first to second gear (upshift or downshift). In Fig. 30, conventional switch characteristics are shown with a switching hysteresis for the gears 1 and 2 are indicated. The conventional switching characteristics can be seen from FIG

The factor K (lh) is the load lever-dependent quotient from the speed of the upshift to the downshift characteristic. The switching hysteresis K (lh) is increased as the rough road key figure SK increases. An example of a specific load lever (lh = const.) Can be seen in FIG. 31, with variant 1 recognizing increased consideration of the rough road key figure. Accordingly, the dependence of the switching hysteresis on the rough road = characteristic number SK for a defined load lever for variants 1 and 2 is indicated in FIG. 31.

A further embodiment of the present invention follows described in which a suitable starting strategy or starting speed at a EKM and / or ASG system depending on a selected one Driving program is provided.

With a foot-operated clutch, the starting process can be influenced both via the clutch pedal and the accelerator pedal. In contrast to this, in a vehicle with automated clutch actuation, for example by means of an EKM or ASG system, the driver has only the accelerator pedal available to control the starting process. In the starting strategy used in the clutch control, the clutch torque M _Ksoll can be determined essentially via a map as a function of the engine speed n _mot during starting. B. by

M _Ksoll = f (n _mot )

Here, a steady state can arise, which depends on the accelerator pedal position α by a defined starting speed and one on the Drive wheels transmitted engine torque is marked. The functional one The relationship between the starting torque as a function of the starting speed is shown in Hereinafter referred to as the starting characteristic.

According to an advantageous development of the present invention, not only the position of the accelerator pedal, but z. B. additional information from controls can be used. This can lead to an increase or a decrease in the set clutch torque, which in turn causes a change in the engine speed. In order to better take into account the driver's request in the starting strategy, z. B. the controls
Driving program switch (e.g. sport, economy, winter)
Kick-down switch
intended. Of course, other controls are also conceivable.

If e.g. B. the driver selects the sport mode, the clutch torque, which tends to lead to a higher starting speed, can be selected, as is also indicated in FIG. 32. The application depends on the selected engine (maximum engine torque; engine speed-torque characteristic), the vehicle (e.g. very sporty vehicle) and / or the vehicle manufacturer (application requests / customer philosophy).

The starting speed can be at values which, for. B. at least about 500 rpm are lower than the speed at which the maximum torque is reached and can up to 5000 rpm (with highly motorized and very sporty designed vehicles). In diesel vehicles, for. B. the maximum of Engine speed, which is selected in sport mode, by about 1500-2000 rpm be lower (maximum of approx. 3000-3500 rpm).

When the kick-down switch is actuated, the starting speed can reach Values increase, which are still 500 rpm higher than in sport mode.

In the so-called winter mode, a lower coefficient of friction can be used Road surface and thus a lower starting speed be desired. The values can range from idle speed to 1200 Rpm are above the idle speed of the current operating point (Consideration of warm-up phases). The so-called economy mode is called Starting position designated with predetermined starting speeds. For recognition the driver's request (and thus the classification sport, economy etc.) is not an operating element is absolutely necessary. The determination can also be adaptive take place, the driver's past driving behavior used for analysis becomes. Here, for. B. a statistical recording of the pedal values and the Pedal value gradients are formed.

In Fig. 32, the starting speeds are shown in various positions of the drive program switch.

On the one hand, the torque Speed characteristic or the consumption map of gasoline engines, as well the subjective assessment of the starting behavior can be considered. In economy The aim is to keep the vehicle in the range of the specific mode To operate consumption minimums, which in gasoline engines at lower Speeds. Furthermore, a speed curve is desired, which too low wear of the clutch lining.

In contrast to this, in the so-called sport mode or when the kick-down switch is actuated, the focus should not be on minimizing consumption or loss, but on the driver's desire to apply more traction. The subjective assessment of a start-up process by the driver also represents a decisive criterion when determining the start-up speed. The following points must be taken into account:
noise
vibration
acceleration curve
Jerk (change in acceleration over time)

In sport mode or by pressing the kick-down switch can be provided be that the driver wants a high tractive force that z. B. by a increased vibration or played by an increased noise level can. This approach is particularly useful because people only is insufficiently able to estimate an absolute acceleration.

A procedural implementation can be traced on the basis of the flow chart, which is shown in FIG. 33. The influence on the starting strategy as a function of the operating point and the position of the operating elements is shown schematically in FIG. 33.

At the beginning, the release of a route can be based on input information, such as. B. idle switch, brake switch (operating and / or brake application) are checked. If a driver's request to start is recognized, operating point detection can be provided on the basis of operating elements and further input signals (currently selected gear, height or the like), so that a suitable starting strategy is determined. To calculate the clutch torque, the starting characteristic can be factor z. B. be multiplied so that the following equation results:

M _KSoll = Fak.f (n _mot )

The settings of the calculated target clutch torque can be made at presented starting strategy.

A further embodiment of the present invention can provide that preferably the starting speeds depending on the height at which the Vehicle is located, appropriately increased.

At high altitudes, what is available increases due to the low air pressure Engine torque of the vehicle. It is therefore necessary that the Engine in a favorable speed range, which is a sufficient Starting torque provides, should be brought.

According to a further development of the present invention, that this situation is recognized appropriately. For example, this can the evaluation of a standard approach can be used if e.g. B. it is found that despite full almost no adequate acceleration of the vehicle is present. The resulting long slip phase can be avoided if the starting speeds are based on an evaluation of the atmospheric pressure be corrected in advance. Of course, the starting speed can also be done in a different way and appropriately modified to suit the starting process of this situation adapt. The atmospheric pressure in many vehicles is modern Engine management already available.

A further embodiment of the present invention follows described, in which suitable measures are proposed to respond to a React to controller overload.

An overload of the clutch actuator can, for. B. occur when at low Temperatures due to rapid movements of the control piston in the hydraulic System is sucked in liquid through the suction valve. With that the inflated hydraulic route. As a result, z. B. when opening the Coupling large reaction forces on the actuator piston, which leads to a Can overload the actuator.

It is Z. B. possible that with imminent actuator overload a reduction in Maximum position of the clutch actuator is provided. However, in In limit cases, the maximum clutch travel is not reduced across all limits (clutch otherwise transmits torque). In these cases a shutdown can occur of the actuator only by volume compensation of the hydraulic line can be avoided, as the sucked fluid in the route over the Compensating hole can drain. The volume compensation of the hydraulic line cannot be carried out in all driving situations. Thus, the Actuator switch-off cannot always be avoided.

Accordingly, according to a further development of the present invention be provided that an impending actuator overload from the position control of the Clutch actuator recognized and the control of the clutch (Coupling software) is communicated. This condition usually occurs in driving situations, which require the clutch to be fully opened (neutral gear, or at stationary vehicle engaged and brake applied). In particular by Driving mode with the gear engaged cannot equalize the volume as this requires the clutch to be fully closed. If at Impending actuator overload for a short time due to the gearbox actuator system Neutral gear is engaged, volume can be balanced in this neutral phase take place and an imminent switch-off is advantageously avoided become.

A further embodiment of the present invention preferably relates to one Initial start-up of a mechanical disengagement link for an EKM and / or ASG system.

In the case of semi-hydraulic release systems in particular, the adjustment can be carried out between the hydraulic transmission line and the clutch z. B. by that Driving over the sniffer bore takes place, whereby a pressure equalization of the Route is done. This ensures that the maximum Transmission capacity of the clutch can be achieved with the sniffer bore open. This point is therefore stored within the clutch control logic and can be approached reproducibly. With a mechanical The point of the maximum coupling torque depends on the transmission distance Initial adjustment of the route and the coupling. Therefore, when starting up the coupling the connection between the position of the actuator and the set clutch torque.

According to a development of the present invention, an initial comparison of the Coupling onto the transmission line with the coupling fully closed take place, of course, other positions are also conceivable. It is it is also possible for the actuator to be in a defined position during the initial adjustment is driven, which is the sniffer bore position with semi-hydraulic Distance corresponds. Of course, another position can also be defined appropriate position can be used. The first adjustment with a mechanical Route with clutch fully closed is an essential feature for the initial commissioning or recommissioning.

Another embodiment of the present invention relates to a heat-insulated Clutch line. To reduce pressure drops at low temperatures and to reduce interference in the transfer function This can be provided by reducing temperature gradients.

The clutch release systems are preferably operated with brake fluid. Since the viscosity of brake fluid is strongly temperature-dependent, large pressure drops occur, especially at low temperatures. These can be calculated for laminar flows using the following formula:

It is composed of a throttle and an aperture portion, whereby the throttle component is strongly temperature-dependent over the viscosity of the medium. In addition, the throttle is linearly dependent on the volume flow, while this dependency on the aperture is quadratic. Both shares are in Try determined.

In the formula:
≙: volume flow [mm ³ / s]
ρ: density of the fluid [kg / mm ³ ]
ν: kinematic viscosity of the fluid
k _throttle : constant into which the throttle _geometry flows
k _Aperture : Constant into which the aperture geometry flows

Based on the evaluation of a series of tests is possible sought, the resulting pressure drops at low temperatures reduce.

When the system is completely cooled, the pressure losses shown in FIG. 34 can occur. In Fig. 34 are pressure drops in the entire release system in dependence on speed and temperature, especially for constant temperature, shown in the whole system.

The pressure losses in the advance stroke significantly increase the actuator load in the return stroke the system reaches the suction limit and there is a risk of Inflation.

In contrast to other projects, pumping is favored here the fluid temperature does not increase while driving. Only when it is at a standstill Vehicle heats up the fluid through the waste heat of the engine. A Cause this is because the line is made of metal, in addition, the line has an im Large surface area in relation to the volume of fluid. This makes it a very quick one Temperature compensation between line and environment possible.

The temperature gradients in the fluid cause a shift in the Transfer characteristic. The bare metal line in the vehicle is also relative to this vulnerable because the system has a large volume of fluid on the primary side. The Upset can only be compensated for by sniffing regularly.

• - dort häufige schnelle Bewegungen im Ausrücksystem auftreten, die zu einer Erwärmung des Fluids führen und
• - dort eine temperaturbedingte Verstimmung der Übertragungsstrecke auftreten kann. Fußbetätigte Systeme stehen immer im Schnüffeln.

The considerations apply preferably to automated release systems, since only

• - There occur frequent rapid movements in the release system, which lead to heating of the fluid and
• - There may be a temperature-related detuning of the transmission link. Foot-operated systems are always sniffing.

With each stroke, the actuator must take into account the pressure losses in the line overcome. The resulting energy is transferred within the line the fluid friction is converted into heat.

For example, in a vehicle measurement (ramps) at -34 ° C in the preliminary stroke, an average speed of 15.06 mm / s was achieved on the slave cylinder, the pressure drop being 7.81 bar, as indicated in FIG. 35. Fig. 35 shows the pressure and slave speed over the disengagement path (fast ramp). The power loss generated by the viscous fluid in the line can be calculated from the data obtained during the measurement.

The table shows the calculation of the power loss per stroke by the tough Fluid is generated. The pressure loss in the return stroke is significantly lower because that System sucked up.

A fast ramp results in a power loss of 7.41 W in Forward stroke and 1.05 W in the return stroke. This power is called heat to the cold Given the environment because the line is made of metal and also the Surface is large in relation to the fluid volume. Because of the metal pipe can also result in large temperature gradients, which a regular Require sniffing.

If the resulting loss can be used to heat the fluid can, for example by isolating the line, the pressure drops decrease in the system over time. The actuator burden would decrease, that The tendency of the system to suck up could be defused. As well could be susceptible to detuning the transmission link Temperature gradients can be reduced, since an insulated line with the heat Exchange environment much more slowly.

With the formula

Q = cmΔT

the possible heating of the fluid due to a full stroke (forward and return stroke) can be calculated.

With the material values from the table below, a heat quantity of 123.7 J is required to increase the temperature of the line and fluid by 1 K, of which 46% for the fluid. With the existing power loss of 8.46 W in the first stroke, the fluid and line heat up by 0.3 K.

• - Es wird keine Wärme an die Umgebung abgegeben.
• - Es wurden die Stoffwerte für Raumtemperatur verwendet.
• - Die Rechnung gilt für eine Ganzmetallleitung ohne Schlauchanteil.
• - Zudem gilt diese Rechnung nur für den ersten Hub; aufgrund der Erwärmung des Fluids nimmt die Verlustleistung und damit auch die Erwärmung mit jedem zusätzlichen Hub weiter ab.

The following simplifications apply to this invoice:

• - No heat is released to the environment.
• - The material values for room temperature were used.
• - The invoice applies to an all-metal line without a hose component.
• - In addition, this calculation only applies to the first stroke; due to the heating of the fluid, the power loss and thus the heating continue to decrease with each additional stroke.

On the basis of the considerations described, a hydraulic transmission link is set up. This section is insulated with styrofoam against heat loss, as indicated in Fig. 36.

• - 5er Prüfleitung (A2 = 9,62 mm²), Länge 510 mm, damit Volumen 4906,2 mm³.
• - 6er Zuleitung an jeder Seite (A1 = 14,52 mm²), Volumen innerhalb Isolierung auf jeder Seite 4350 mm³.
• - Messleitung mit Drucksensor an jedem Ende der Prüfleitung.
• - Isolierung. Die Leitungen sind allseitig von Styropor umschlossen.

The assembled system consists of the following components:

• - 5 test lead (A2 = 9.62 mm ² ), length 510 mm, thus volume 4906.2 mm ³ .
• - 6-way supply line on each side (A1 = 14.52 mm ² ), volume inside insulation on each side 4350 mm ³ .
• - Measuring line with pressure sensor at each end of the test line.
• - insulation. The lines are enclosed on all sides by polystyrene.

The flow resistance of the test line is about 10% higher than that of the Management of the vehicle.

The fluid column is constantly pushed back and forth for the experiment. It is the displacement path is adjusted so that no cold fluid is out of the area reaches the test lead outside the insulation and falsifies the result.

The shift is carried out by two coupled central releasers. You move a volume of 4235 mm ³ with each stroke, the speed is chosen low due to the high load. It is also temperature-dependent.

Due to the viscous fluid, a speed-dependent pressure difference dp will occur between the two pressure sensors. In Fig. 37 the volume flow and the pressure difference over time for a respective forward and return strokes at -20 ° C are shown.

As a result of the experiment, a diagram is obtained which shows the pressure drop over the actuation speed (analogous to Fig. 34). It is expected that the measured pressure drop will decrease over time, since the pressure drops cause the fluid temperature to rise and the viscosity to decrease at the same time. The measurements are carried out at three different temperatures: -30 ° C, -20 ° C and + 80 ° C. The low temperatures result in large pressure drops. From this, the throttle component of the loss can be determined and the expected effect of fluid heating can be observed. In order to be able to calculate the temperature increases in the line, the aperture portion must also be known. This requires measurement at 80 ° C.

Fig. 38 shows the pressure drops in the line for different fluid temperatures in relation to the flow. The diagram has the volume flow as the x-axis, so that it is valid for all projects regardless of the piston area. A volume flow of 20 cm ³ / s corresponds to a sensor speed of 100 mm / s for a predetermined vehicle and 70 mm / s for MTA / ASM. In Fig. 38, the pressure drops in the test line as a function of volume flow and temperature, is valid for constant temperature throughout the line, is shown. The lifting movements are divided into different blocks, each block contains 23 strokes. The first and then every fourth block is recorded.

In Fig. 39 the first stroke at -30 ° C is shown. A pressure of about 20 bar is set and then the movement back and forth is carried out. The two measured pressures and the speed at the slave cylinder over the slave cylinder path are shown. As in Table 1, the following values result:

Table 3 shows the calculation of the power loss per stroke by the tough Fluid is generated is shown. The measurement carried out here The power loss achieved is much lower than when measuring the vehicle. The reasons for this are, on the one hand, the much lower volume flow, and secondly the 4 ° C higher temperature.

In Fig. 40 the results of the measurement are shown. In Fig. 40 there are pressure losses, shown above the volume flow at -30 ° C. There are 92 strokes between the ramps shown in FIG. 40.

It can be seen that there is a significant improvement especially between curves dp1 and dp2 (the pressure drop in both directions is reduced by 2 bar, that is around 22%). Even after another 92 strokes there is still a noticeable improvement, which is still 0.6 bar or around 9%. As expected, no changes were found after about 500 strokes. The final value is about 5.6 bar, which corresponds to an improvement of about 3.2 bar (see Fig. 43). In a predetermined system, a 3.2 bar reduction in system pressure corresponds to a force of 90 N on the GZ tappet or 200 N in relation to the clutch. It can be seen from Fig. 41 that the pressure loss does not decrease after several strokes. The first 14 strokes over time are shown there. In this short time, an improvement of 0.77 bar in each direction is visible. The decreasing pressure drop within 14 strokes is indicated in FIG. 41.

Using the calculated values for the orifice and throttle component, the temperature increase associated with a specific pressure loss can now be calculated in reverse. This results in FIG. 42. The pressure losses over the volume flow for four different temperatures are entered there. At the start of the experiment, the fluid temperature was -30 ° C. After 92 operations, it has already reached -27.7 ° C. The temperature after 184 strokes is -26.8 ° C. After around 700 strokes, the final value is -25.3 ° C.

The improvement in behavior is very clearly visible. For an opening Movement at 100 mm / s results with the line used here Reduction in pressure drop from 17.4 to 11.2 bar.

Finally, Fig. 43 shows the drop in pressure loss and the temperature rise over the number of operations. It can be seen that the possible warming is limited. Despite insulation, heat is released to the environment, so that after a certain time a state of equilibrium between heating by internal friction and cooling by heat is achieved. The following diagrams show the behavior at -20 ° C.

In Fig. 44, the measured pressure loss is presented to the flow. Here, too, there are 92 strokes between the three curves shown. It is striking that the curves here are much closer together. Due to the higher initial temperature, the viscosity is lower, so that the power required for the movements is lower and the fluid heating accordingly. The improvement between the first two ramps is 0.35 bar or almost 9%. The pressure drop decreases more and more slowly with the following operations.

For this measurement too, the fluid temperature is calculated from the measured pressure losses. This results in the behavior from Fig. 45. Starting from -20 ° C, the situation improves within 92 strokes to -19.1 ° C, the next 92 strokes bring the same increase again. The final value of -16.8 ° C is reached after around 500 operations.

Fig. 46 shows the change in temperature and pressure loss . The behavior is similar, but the pressure losses are far less pronounced. With increasing temperatures, the pressure drops in the line decrease significantly. This also reduces the heat loss that leads to the heating of the fluid. The mechanism described is only effective at low temperatures at which it is also desirable. At high temperatures, there is almost no heating of the fluid.

The release system is almost always in the case of foot-operated systems in the sniffing state. Volume change due to changing Fluid temperatures are thus compensated for via the sniffer bore. Different this is what EKM / ASG vehicles do. Because of the moment tracking the sniffer hole only opened for a short time. All volume changes that occur between two sniffing cycles, lead to a detuning of the Transmission link. An insulated pressure line can help that keep temperature-related detuning low by changing ambient temperature result in a slower fluid temperature change.

For the test carried out, two line sections are provided with temperature sensors and filled. One of the two lines is provided with thick polystyrene insulation, the other is a conventional metal line. For the measurement, the ambient temperature is changed and the fluid temperature response is recorded. As a result, the following diagrams in Figs. 47 and 48 are obtained.

For the measurement in FIG. 47, the measured lines in the climatic cabinet are brought to a uniform temperature of 75 ° C. From there, the ambient temperature is cooled as quickly as possible (initially with the air conditioning cabinet door open).

The ambient temperature drops quickly, especially at the beginning. The maximum gradient reached is briefly 60 K / min. In the metal line, the temperature drops to a maximum of 9 K / min, since the fluid and line have a greater inertia than the air. The temperature in the insulated line drops again much more slowly (2 K / min).

Table 4 shows a comparison of air and fluid temperatures.

Large temperature gradients usually only occur locally in the vehicle. thats why it is not possible from the measurement made to detune the system to determine. The comparison of the gradients achieved in conventional and isolated line, however, allows the conclusion that at most every third Sniffer cycle would be necessary.

The next measurement begins with cooled systems at -2 ° C. From there is brought up to 60 ° C with the maximum heating power, cooled down again -10 ° C and heated again. It is interesting how the two fluid temperatures respond to this cycle.

In Fig. 48 temperature and impact are displayed on fluid temperatures. The non-insulated cable exactly shows the temperature specification with a time delay of about 2 minutes. The same gradients are thus achieved in the fluid. Not so with the insulated cable. The temperature gradients are somewhat below the conventional line. The system is also becoming significantly sluggish. It becomes clear that the insulated line requires less sniffing. On the other hand, it would also be possible to run the line along components that quickly change their temperature without having to expect excessive detuning of the route.

Based on the strong pressure drops in the system, a Possibility sought to reduce the losses in the release system. Because losses in Heat is converted, an attempt is made to use this heat for to make the system usable. For this purpose, a line is completely filled with polystyrene or the like isolated.

The tests carried out have shown that it is actually possible reduce the losses in the system through an insulated cable. Depending on the initial temperature, the fluid warms up by several Degree, which brings a significant improvement in pressure loss. at During the measurements, the pressure losses at -30 ° C between the first and the 93rd actuation by up to 22%.

An additional benefit can arise from an insulated cable if that the fluid cools less quickly from the wind or engine heat or is heated. So there are lower temperature gradients. In order to the transfer function detunes less quickly and the Sniffing frequency to compensate for the upset could be reduced by a factor of 3 become.

A system with a plastic pressure line most likely exhibits behavior that is better than one in terms of the phenomena described conventional metal pipe. The insulating effect is likely to be quite a bit are below special insulation. Have previous plastic lines but the disadvantage that their volume consumption is strongly temperature-dependent. This creates problems for the system, since the transfer function also becomes temperature dependent.

Therefore, according to the current status, metal lines with insulation be provided. This combines the advantages of constant volume absorption with self-heating and reduction of temperature gradients in the fluid. The insulation should be as consistent as possible from the connection of the CSC to the Run GZ to prevent cold bridges from forming. In the future Development of plastic lines could provide thermal insulation may be meaningfully integrated.

• - dort häufige schnelle Bewegungen im Ausrücksystem auftreten, die zu einer Erwärmung des Fluids führen und
• - dort eine temperaturbedingte Verstimmung der Übertragungsstrecke auftreten kann. Fußbetätigte Systeme stehen immer im Schnüffeln.

The considerations described above preferably apply to automated release systems, because

• - There occur frequent rapid movements in the release system, which lead to heating of the fluid and
• - There may be a temperature-related detuning of the transmission link. Foot-operated systems are always sniffing.

A further embodiment of the present invention follows described, in which preferably a path-controlled valve for partial Limitation of the clutch closing speed especially for foot-operated Release systems are proposed.

Many of today's vehicles are difficult to start. It has quite different causes, but all with the improvement of consumption and Exhaust gas behavior. For example, have small volume Motors today have high maximum torques, which means high clutch torques and forces required, but only low moments when idling. Add to that a long time Axle ratios that reduced the gearbox input shaft Increase vehicle mass significantly. Flywheel masses on the engine side are also reduced like the idle speeds.

In order to reduce the disadvantages for the starting behavior, different ways are searched. There will be changes on Disengagement system described from which an improved starting behavior is expected becomes.

The goal of simplifying the coupling ability can be explained on the basis of FIG. 49. There, pressure and clutch torque are plotted over the release path. The torque should be built up on the clutch within a relatively small range without the engine speed falling too much. The interesting area of the moment structure is highlighted by a frame. In this area, the driver has to move the pedal very slowly to reduce the clutch closing speed. It is precisely in this area that the methods described should also work and make it easier for the driver to start off in order to prevent the engine stalling. In Fig. 49 the torque build-up and the diaphragm spring characteristic are shown on disengagement.

In another diagram, the desired behavior is idealized over time ( Fig. 50). The master cylinder extends a full stroke and the central release follows, the torque being reduced to zero by opening the clutch. After a short time (e.g. engaging gear), the driver closes the clutch by taking the path back at the GZ. The slave cylinder follows the backward movement until the torque builds up. There (and only there) the system to be developed should reduce the closing speed and thus the further torque build-up. In Fig. 50 this point can be recognized by the kink in the disengagement path, with Fig. 50 indicating the idealized representation of the target path on the master and slave cylinder as well as the torque build-up over time.

This kink shifts the further closing of the clutch a certain amount, so the clutch is only at this time completely closed after driving over the sniffer hole. Through this Intervention in the release system should not occur a moment jump.

Based on the briefly presented requirements, four solutions can preferably be imagined, which are described below:
The best solution is certainly a suitable EKM system, since the actuator takes over the function there. In addition, by evaluating numerous driving state parameters, the coupling process can be tailored very precisely to the prevailing situation.

• - Volumenstrom, auf den
• - Systemdruck oder auf die
• - GZ-Kolben-Position reagieren.

The other three variants are characterized by the fact that they are purely hydraulic components that do not require any sensors or external connections. These are controls that are based on the

• - volume flow to which
• - system pressure or on the
• - React GZ piston position.

• - Der Widerstand darf nur beim Rückhub (Kupplung schließen) wirken.
• - Er soll die gesamte Einkuppelzeit nicht wesentlich verlängern, es ist also eine Beschränkung auf den Bereich des ersten Momentenaufbaus notwendig.
• - Es sollen nach Möglichkeit keine Abhängigkeiten von Temperatur oder anderen Parametern (z. B. Kupplungsverschleiß) entstehen.
• - Der Fahrer soll möglichst nichts merken.

The basis in any case is a limitation of the closing speed of the clutch by a resistance in the hydraulic line. The following points are important for this resistance:

• - The resistance may only act on the return stroke (close clutch).
• - It should not significantly extend the entire engagement time, so it is necessary to limit it to the area of the first torque build-up.
• - If possible, there should be no dependencies on temperature or other parameters (e.g. coupling wear).
• - The driver should not notice anything if possible.

The solution in all cases consists of one or more orifices that are controlled by valves in the desired manner. Figs. 51 to 53 show the basic structure of all three variants are based. The master cylinder is shown on the left, which pushes fluid into the slave cylinder in the upstroke via the check valve at the top.

In Fig. 51, the path of the fluid is shown in the forward stroke. On the return stroke, the check valve is closed, the fluid flows unhindered back to the GZ via the valve shown ( Fig. 52). When the set criterion is reached, the valve switches to the other position against the spring, the direct path is then blocked. The fluid then flows through the orifice connected in parallel ( FIG. 53). When the set range has been reached, the spring opens the valve again and allows the fluid to flow through ( Fig. 52). By using orifices to reduce the speed, the temperature dependence can be kept low. There applies

The pressure drop created by an ideal orifice is not that Toughness (= f (T)) and quadratic depending on the volume flow. (Because a real one Aperture also has a throttling character, is nevertheless a certain one Temperature dependence exists.)

There is a fundamental difference in the behavior between volume flow or pressure control and path control. Control the first two variants depending on the clutch position, that means that the time in which the Control acts on the system regardless of the driver. However, there is a direct influence of the driver when controlling the route, but the driver has still in full control of his vehicle. Control would be ideal regardless of environmental influences depending on the Pressure plate position works. But since that means a disproportionate effort would be transferred to the position of the master cylinder piston. Since at foot-operated system of pistons usually behind the sniffer bore (SB), environmental influences (especially temperature influences in the clutch and release system) can be compensated via the liquid column. The The hydraulic route is therefore always balanced when actuated. The comparison works very well with the EKM due to the sniffing cycle.

The maximum engagement speed is controlled here depending on the position of the master cylinder piston. In Fig. 54 is the area in which the clutch engagement speed is affected is shown. In Fig. 54, the clutch characteristic curve and the torque build-up are shown above the control travel.

One limits itself to the small area of the first Torque build-up because there is only a small moment on the engine. A higher moment is only on the clutch when the driver accelerates. In this case but there was no danger of stalling.

The function of the path control is explained on the basis of FIG. 58. The path of the fluid during the preliminary stroke has already been explained. It flows from the GZ via the check valve to the ZA.

In Fig. 55 is a block diagram of the position control is shown. On the return stroke, fluid first flows unhindered through the valve (in this case, both ways - over the orifice and past the orifice - are possible). When the GZ piston reaches a certain position, the valve is switched over and the fluid can only flow through the orifice. Depending on the driver's operating speed, the valve opens a certain time later, the clutch closing at normal speed.

In all variants, the suction valve (NSV) influences the Control behavior to be expected. If the route lacks a possibility to Sucking in while the fluid only flows through the orifice is created in the GZ-side line a vacuum (requirement: The GZ is by a spring withdrawn). This vacuum could lead to air at the plug connections penetrate, in addition, with the path-controlled variant described here a moment jump is expected at the time when the opening parallel to the. Aperture is released. Then the fluid flows freely from the Slave cylinder (NZ) and suddenly closes the clutch by a certain amount. The driver can feel a jerk. If an NSV is present, the GZ sucks Fluid as the fluid flows from the NZ through the orifice. Because of this The route remains balanced. A jump only occurs here when the sniffer hole is released.

In Fig. 56, the idealized representation of the time course of the GZ and ZA travel and clutch torque without a suction valve is indicated. In Fig. 57, however, the idealized representation is indicated the time course of GZ and ZA-way clutch torque and with anti-cavitation valve.

The system may be inflated if one Suction valve is present. This danger exists when the driver presses the pedal pressed repeatedly before the clutch is fully closed. To this one To be able to check the considerations presented are of the variants Pressure control and path control each built prototypes. A Flow control may be established later. With the superstructures attention was paid to modularity, to the effort for the construction and the Keep optimization small.

A modular structure means that all components are in separate housings are set up so that changes can be made easily. But that also means that a lot more space is required overall than is absolutely necessary.

An assembly drawing of the path-dependent control is shown in Fig. 58. The additional components required were attached to this cylinder. A housing was constructed that has two separate pressure outlets. One outlet (in Fig. 58, left) can be closed by the additional piston. The additional piston is screwed into the previous valve seat of the suction valve (NSV). The function of the NSV is performed by a check valve which is connected to the right pressure outlet as shown in Fig. 58. This makes it very easy to carry out experiments with and without NSI. In Fig. 58, the construction of path-dependent control is thus (NSV: replenishing valve; RSV: check valve), respectively.

A non-return valve is located in parallel between the two pressure outlets (RSV) and a series connection of several apertures. The RSV will needed to achieve the maximum speed during the forward stroke. Through the Series connection of the orifices can experimentally the required pressure loss be determined. The individual pressure losses add up, so that with several "weak" apertures the effect of a very small aperture can be replicated.

As a result, vehicles can be approached with little effort be improved. By knowing the characteristics of clutch and Release system can be used to develop methods for starting to help the driver in critical situations.

In summary, an approach can be given that the maximum possible engagement speed with hydraulic and semi-hydraulic release systems in the critical area of the first torque build-up the clutch is reduced and thus an unwanted stalling of the engine prevented.

The structure of such a control consists, for. B. in a parallel connection of a check valve as a bypass, an orifice for almost temperature-independent increase in flow resistance and thus to reduce the Closing speed as well as a line, depending on the situation by a Valve can be closed. Are possible for the situation-dependent Valve control z. B. the system pressure, the volume flow and the Piston position of the master cylinder.

Different approaches are conceivable to improve the starting behavior, with which it is possible to reduce the clutch closing speed hydraulic and semi-hydraulic systems in the critical area for starting to limit.

In Fig. 59, the clutch characteristic curve and the torque build-up are shown above the control travel. The goal of simplifying the approachability can be explained in FIG. 59. There, pressure and clutch torque are plotted over the master cylinder travel. In the area between the two lines, the torque on the clutch should be built up at a reduced speed in order to facilitate comfortable starting. This reduction in speed is achieved with an orifice controlled by a valve, which only works on the return stroke.

A basic structure of this variant is indicated in FIG. 60. The master cylinder moves fluid in the preliminary stroke via the check valve into the slave cylinder. Thus, 60, Fig. Is a block diagram of the travel-dependent clutch control.

On the return stroke, the fluid initially flows back unhindered via the valve to the GZ. When the set piston position is reached, the valve switches to the other position, so the direct path is then blocked. The fluid flows through the orifice connected in parallel. Once the set range has been reached, the valve opens again and allows the fluid to flow through. By using orifices to reduce the speed, the temperature dependence can be kept low. There applies

The pressure drop created by an ideal orifice is not that Toughness (= f (T)), but depending on the square of the volume flow. Is an NSV is present, the GZ fluid sucks in, while the fluid from the NZ through the Aperture flows. This sucking in is supposed to close the jerky Coupling at the end of the controlled area can be prevented.

A disadvantage of vacuuming is that the route can be inflated, if the sniffer bore is not released after every clutch actuation becomes. Therefore, when using this structure, a safety device for the release system can be provided, for example the stop of a SAC.

In order to be able to review the considerations briefly repeated here, a first prototype GZ was built. The structure of the GZ is shown in Fig. 61. The right half of the picture ( Fig. 61) shows a plastic GZ. The most important change is that the additional piston has been screwed into the valve seat of the NSV. The NSV functions as a non-return valve, which is externally connected to the right pressure outlet as shown in the figure. In the extension of the GZ, a second housing was screwed on, which has two pressure outlets. The rear one (right) is always open, the front one is closed depending on the position using the extended piston. There are two O-rings in the housing.

The length of the controlled area was chosen to be relatively long so that the results are clearer. In an actual application, the optimum length and position must be determined depending on the parameters from the coupling and the route. A non-return valve (RSV) and several orifices, each with a diameter of 0.3 mm, are located in parallel between the two pressure outlets. For pressure purposes, two pressure sensors were installed in the line, one on the gas side (p_LGZ) and one on the gas side (p_LZA) of the orifices. The positions are indicated in Fig. 61.

The tests were carried out on the EKM hydraulic test facility, in addition to the master cylinder described, the components of the W168 / Otto were used. The GZ area is therefore 197.8 mm ² , the area of the ZA is 380 mm ² . Different speeds were set on the actuator to check the functionality of the structure.

Fig. 62 shows a disengagement curve of the experiment. The following can be seen: the pressure signals of the two sensors are identical in the advance stroke. This corresponds to the requirement that the orifice in the forward stroke should not influence the behavior of the route. In Fig. 62, the Ausrückkennlinie, the pressure before and after release travel through aperture therefore are shown. In contrast, the two signals differ significantly on the return stroke. The pressure p_LZA runs according to the force on the disc spring. The pressure p_LGZ falls at around 7.2 mm from 7 bar below the ambient pressure. In the area from the 3 mm position, the valve opens again and clears the way past the orifice. From there, the pressure profiles in front of and behind the panel are similar again.

In Fig. 63, the displacement signals of the master cylinder and the releaser are plotted versus time. The master cylinder specifies a speed of 50 mm / s. The closing speed on the release mechanism is reduced from the initial 25 mm / s to 7.5 mm / s due to the covers used.

The duration / length of the reduced-speed return stroke is directly dependent on the GZ speed. The slope, i.e. the reduced speed itself, is determined by the resistance in the aperture. The dependencies mentioned can clearly be seen in the curves in FIG. 64. The paths over time are shown again, for comparison on the left with an encoder speed of 35 mm / s, on the right with 100 mm / s. The resulting speed at the clutch is around 7.5 mm / s in both cases. However, there is a different duration of action and thus a smaller disengagement travel range in which the clutch speed is reduced. (left from s_AL 2.95 to 3.22 mm, right from 2.9 to 3.0 mm)

There is an area below each of the master and slave cylinder paths Coupling characteristic and a torque structure marked in which the Control affects the release system. With slow GZ movement is this area is relatively large, at high speed the area is small. How the valve switches very early. The Speed reduction therefore begins in relation to the one shown below Torque build-up too early.

In Fig. 64, the return stroke, the various donor speeds over time are presented with torque build-up in the clutch disengagement. The time during which the aperture is effective is marked.

Contrary to expectations, the release path drops quickly at the end of the reduced-speed range, which means that the clutch closes suddenly by a certain amount. This is because the external suction valve is probably not working properly. In principle, it enables re-suctioning (recognizable in Fig. 65). However, the negative pressure that occurs in the primary space is striking ( Fig. 4: p_LGZ). A NSV normally opens at p_LGZ ≍ 0.

Presumably the system with NSV outside the GZ is with them Speeds too sluggish: the whole thing has to be sucked up Liquid column can be accelerated from the branch of the NSV to the expansion tank. In contrast, with an NSV integrated in the GZ only the moves Piston with the valve. The fluid column itself remains at rest:

The transfer function of the system is shown in Fig. 65. The transfer function and the disengagement path via the encoder path are indicated. The broken line represents the expected function on the return stroke. The forward stroke corresponds to a conventional system as expected. The transfer function is quite linear.

The characteristic curve runs back a bit on the same curve until the aperture is switched on at the encoder position 16 mm. Between the GZ positions 6 and 1.3 mm (end of the diaphragm effect to the sniffer bore), the curve should run parallel to the forward stroke if the NSV is working correctly and then fall off relatively quickly when the sniffer bore is released (the sniffer bore itself has an aperture with a diameter of 0.7 mm The pressure cannot drop suddenly either, this behavior is shown with the broken line in Fig. 65. The experiments have shown that the desired behavior can be achieved with a structure such as that presented here.

It should be noted that in contrast to the variants printing and Volume flow control no arbitrarily low clutch closing speeds can be generated. This is due to the special characteristics of the Up steering. The direct dependence of the function on the driver is typical. There are two limits to the effectiveness of the system. The driver moves that Pedal very slowly, so the orifice remains ineffective, because the volume flow, that the driver sets can flow completely over the panel. Not so with very fast pedal movements. Here the controlled area becomes very fast passed through. Thus, the aperture is also almost ineffective. Between there is a range of an average operating speed in which the Driver is supported when engaging.

The advantage of this characteristic is that the driver always has full Keeps control of his vehicle. He can do that in high gear Build up clutch torque very quickly. Nevertheless, the clutch / start simplified. A stall of the engine is not prevented.

A further embodiment of the present invention follows described, in which a suitable coefficient of friction adaptation is carried out. In the Gripping point adaptation can be between the short and long-term gripping point be distinguished. It is Z. B. possible that only directly after a Volume compensation the assignment between master and slave cylinder position exactly is defined. The deviations in the gripping point are in this case directly to one Change due to the clutch. The longer the last one However, the volume equalization lies behind, the more a determined gripping point error can on a temperature-related contraction or volume expansion of the hydraulic Transmission path between the master cylinder and the slave cylinder to be attributed. Consequently, only within a predetermined period of time Connection to a volume compensation changes the long-term gripping point are used, while the short-term, in the actuating torque characteristic, Gripping point is changed with every possible adaptation. With everyone Volume equalization becomes the short-term used in the adaptation process Gripping point reset to the value of the long-term gripping point. of course is a different coefficient of friction adaptation is also possible.

Analogous to the gripping point adaptation, deviations in the coefficient of friction can occur increasing distance from the last volume adjustment with decreasing Security on a physical change in the transmission behavior of the clutch to be led back. As with gripping point adaptation, therefore, between one a short-term and a long-term coefficient of friction can be distinguished. The short term Coefficient of friction, which the Adaptation procedure is used, with every adaptation possibility (starting, Switching) can be updated. The influence of the coefficient of friction adaptation on the Long-term coefficient of friction can increase with distance from the last volume compensation decline. For example, the change determined for the short-term coefficient of friction can be entered directly in the Connection to a volume compensation z. B. weighted with a factor of 1 in the long-term coefficient of friction. This weighting factor could then be about the time interval for volume compensation z. B. to zero within 60 s be dismantled.

With this distinction between short and long-term coefficient of friction could can advantageously be excluded that other coupling effects (e.g. Stuffing due to short-term, massive energy input), which is the Change the transmission behavior of the clutch, but without changing its physical To influence coefficient of friction, reflected in the long-term coefficient of friction. It is therefore possible that z. B. above a predetermined coupling temperature (e.g. approx. 300 ° C) or only the short-term coefficient of friction after high energy input is changed. If the temperature falls below or after a The short-term coefficient of friction could be based on the time gap to the massive energy input the value of the long-term coefficient of friction is reset. Of course you can this adaptation method can be modified even further to the present Improve invention further.

A further embodiment of the present invention follows described, in which preferably a force-dependent engagement threshold is provided. Thus, a reliable gear recognition and or a Form-fit detection also implemented with overpressed switching. Based on information The engagement threshold is shifted appropriately via the current shifting force.

In particular with an EKM system and also with an ASG system, this can Engaging after a gear change at the earliest when the Form locking on the sliding sleeve. With a path or angle sensor be tried to sense the position of the sliding sleeve and from one predetermined signal value, the so-called engagement threshold, is then the Reconnection permitted.

The sensors for position sensing can, however, for various reasons not be applied directly to the sliding sleeve. There are often between the sensor system (e.g. on the central control shaft or on the gear motor at the ASG) and the sliding sleeve more or less elastic components, e.g. B. Shift forks, shift fingers, central shift shafts or the like.

If a force now acts on the circuit, the parts of the Deform the internal circuit, the greater the switching force. This is what happens Deviations between the actual position of the sliding sleeve and the sensed Signal as soon as power is switched. However, some effort is required absolutely necessary to overcome the friction of the internal circuit. In addition, the synchronizer ring can be rotated when the synchronization is released, which takes a considerable amount of force when the synchronizer ring sticks leads. In addition, the sliding sleeve or the Claw teeth are rotated. It may also be in the tip-to-tip position come between the claw teeth and the sliding sleeve. Then you can do almost anything high forces are introduced into the circuit.

If the engagement threshold is exceeded in the signal before the sliding sleeve has reached sufficient positive locking, the clutch can be closed too early ratchet. The internal circuit exists predominantly made of metallic materials. The are subject to Stiffness properties only minor component fluctuations and minor changes over their lifetime. The rigidity of the internal circuit can therefore be known be assumed. With known force on the external circuit, the Deformation of the internal circuit can be calculated in the control unit. Thus, the Coupling threshold in the signal curve then by the calculated amount of Deformation of the internal circuit can be shifted.

Advantageously, the sensing of the switching force can be via or several force sensors or also about the existing stretch of the External switching via two position sensors or the like (differential path). As a result, early clutch engagement can be prevented if the shift is suppressed become.

A further embodiment of the present invention follows described in which a suitable control strategy for friction-independent Torque control of the powershift clutch, especially with USG gearshifts is provided.

The coefficient of friction of a dry clutch is short due to operational reasons subject to long-term fluctuations. A decisive influencing factor on the The current friction coefficient represents the clutch temperature. The temperature of the Power shift clutch of the USG changes significantly during one Circuit, which means that with the temperature, the coefficient of friction varied.

Since the LSK (power shift clutch) generally only transmits torque during a power shift (with the exception of driving in the gear ratio of the power shift clutch), the coefficient of friction can only be adapted in this phase. In Fig. 66 the temperature dependence of the friction lining is currently used in the USG load clutch is shown.

This means that at the start of a power shift, the coefficient of friction of the Powershift clutch is not exactly known (the temperature of the clutch and thus it The coefficient of friction can change since the last adaptation process, namely the last Circuit, have changed significantly). As long as the short-term, temperature-dependent fluctuations are small, this should not be viewed as critical. Shows however, the friction lining has a strongly temperature-dependent coefficient of friction behavior, such as such as the friction lining of the powershift clutch currently used in the USG, which in a temperature range of 25 ° C (250-275 ° C) a friction value fluctuation of up to 30%, this can lead to problems.

In Fig. 67 is a USG-circuit in which the coefficient of friction of the power-shift clutch is not correct and a reibwertabhängiger torque buildup of LSK was performed, is shown. In Fig. 67 the case of a load circuit having too high assumed friction coefficient of the power-shift clutch is shown. The previous control strategy of the USG stipulates that in order to be able to design the currently shifted gear of the transmission, the engine torque must be taken over by the powershift clutch. For this purpose, the torque of the LSK is raised to the engine torque according to the strategy in FIG. 67. The force on the shift actuator is then increased in order to select the gear currently shifted.

If the coefficient of friction stored in the control corresponds to the physical coefficient of friction of the powershift clutch, the clutch that is to be opened is torque-free during gear selection. Since the LSK is set at the lowest gear ratio of the transmission, the torque build-up of the LSK corresponds to the engine torque level for tracking a steady reduction in the output torque, and thus the vehicle acceleration follows the following equation:

In FIG. 67, however, a highly discontinuous acceleration curve was noted in the Ausspurvorgang. This is due to the fact that the physical coefficient of friction of the power shift clutch and that assumed in the control do not match.

In this case, the physical coefficient of friction is lower than that stored in the control. This leads to a sudden reduction in the output torque and thus the vehicle acceleration when the track is disengaged, since the transmission control opens the gear when the engine torque is not fully transferred to the LSK. The resulting jump in torque continues to excite vibrations, which can be clearly seen in Fig. 2.

To the influence of fluctuations in the coefficient of friction when building torque to compensate for a power shift, a new strategy was developed that is described below. Here, for. B. a controlled, torque build-up independent of the friction value can be provided during a power shift.

Based on the wheel speeds of the vehicle, the Transmission control continuously calculates the current, filtered vehicle acceleration.

If a power shift is requested, the acceleration level of the vehicle, which should occur during the phase-out phase, is calculated in the course of initialization (power shift clutch takes over the complete engine torque):

Furthermore, the expected increase in the acceleration reduction dA is determined from the current vehicle acceleration, the expected vehicle acceleration after disengagement, and the time for the torque build-up of the power shift clutch.

The slope is required because the calculated, filtered acceleration is one time delay of about 30 ms compared to the physical Has vehicle acceleration. To make statements about the current, physical Being able to accelerate the vehicle is therefore an interpolation carried out with the aid of the calculated acceleration and the gradient.

In the phase of torque build-up of the powershift clutch, that becomes The clutch torque increases until the vehicle acceleration (calculated from the filtered vehicle acceleration and the Acceleration slope) has decreased to the predetermined level. That in the control assumed clutch torque can, if the physical and assumed The friction coefficient of the powershift clutch does not match, so it should also be larger than the current engine torque.

In Fig. 68, a USG-circuit with a controlled torque build-up is the beginning of the load circuit indicated.

This is a USG shift, in which the torque build-up of the power shift clutch can be carried out according to the strategy described. The torque of the LSK is raised significantly above the current engine torque. The vehicle acceleration drops to the predetermined acceleration level during the unraveling process. In Fig. 69 is a flow chart of the controlled torque build-up of the LSK is shown.

The coefficient of friction of a dry clutch is strong in the short and long term Subject to fluctuations. An incorrectly assumed coefficient of friction Power shift clutch of the USG can create a very uncomfortable torque build-up Powershift clutch at the beginning of a shift and thus the Negatively affect the switching comfort of the USG.

It can e.g. B. a control strategy can be proposed in which a controlled torque build-up of the LSK takes place during the load switching.

The control variable is the one calculated from the wheel speeds Acceleration signal of the vehicle used. The moment build-up is therefore regardless of the coefficient of friction of the powershift clutch. Of course, the Control strategy can also be used in a so-called DKG.

A further embodiment of the present invention follows described which implement a cruise control function in particular an automated manual transmission (ASG) concerns.

With the cruise control, an automatic speed control of the Be made in the driver targeted Can set or change the speed setpoint and at any time if necessary manual control (accelerator pedal or service brake) can override.

A prerequisite for the implementation of a cruise control function may be that Vehicle has a system in which either the engine torque is direct or indirectly (throttle valve position using e-gas and / or Ignition angle adjustment) can be specified. The well-known cruise control systems are mainly used in middle class and luxury class vehicles, because with the known systems high costs to implement the cruise control functions occur.

An advantageous development of the present invention provides a cruise control Function in conjunction with an automated manual transmission (ASG), which is also designed for vehicles of the lower middle class or small cars. In this way, a cruise control function can be used extremely advantageously in an advantageous manner realize, since all necessary interfaces are available and the Realization effort is very low.

The input signals required for this such. B. the vehicle Actual speed, the service brake signal is active and / or the Accelerator pedal positions are either recorded directly in the ASG system or can be indirect from suitable substitute values, such as B. the speed from the Wheel speeds can be determined.

Furthermore, the ASG system already has the option of one Torque influence on the part of the ASG control unit, in which the engine torque however only reduced, i. H. can be capped. This Possibility is used for example in circuits. To realize a Cruise control function must also have the possibility of a targeted increase in Engine torque (compared to the pedal value dependent portion) exist. Only the controls necessary for operating the cruise control or switch off the cruise control function or change the The target speed is required in addition to the ASG.

As with the ASG system, all measurands and intervention options (Manipulated variables) are available, the additional effort is limited to an implementation of the speed control loop. Advantageously, the control loop only low dynamic requirements and on the other hand only limited Requirements placed on the control quality. A conventional one is therefore sufficient Control concept with e.g. B. PI or PID controllers.

It can be provided that the implementation of this function is optional in the engine and / or in the ASG control unit. When operating z. B. It should be provided that in order to avoid additional controls, the ASG System, for example, the cruise control function to operating mode "D" - automatic gear change - is restricted. So stand in this Operating mode, the controls for changing gears (up / down) Change of the target speed for the cruise control is available.

The claims filed with the application are Proposed wording without prejudice for obtaining further patent protection. The The applicant reserves the right to add more, so far only in the description and / or To claim drawings disclosed combination of features.

Relationships used in subclaims point to the others Training the subject of the main claim by the features of respective subclaim; they are not considered a waiver of achieving one independent, objective protection for the combination of features to understand related subclaims.

Since the subjects of the subclaims with regard to the prior art reserves the right to make its own and independent inventions on the priority day Applicant before becoming the subject of independent claims or To make divisional declarations. You can continue to work independently Inventions contain one of the objects of the previous Have independent design independent claims.

The exemplary embodiments are not intended to limit the invention understand. Rather, there are numerous within the scope of the present disclosure Changes and modifications possible, especially such variants, elements and combinations and / or materials, for example by combination or modification of individual in connection with that in general Description and embodiments as well as the claims described and in features or elements contained in the drawings or Process steps for the person skilled in the art with a view to solving the problem are and through combinable features to a new item or lead to new process steps or process step sequences, also as far as they Manufacturing, testing and working procedures concern.

Claims (2)

1. Method for operating a motor vehicle with a drive motor, a Gearbox and one in the power flow between the drive motor and gearbox arranged switchable clutch, the transmission according to the principle is a manual transmission with at least one transmission input shaft, the can be operated automatically by means of a first actuating device and the Coupling can be actuated automatically by means of a second actuating device is and a control of the actuators by means of the way that at least one for the operation of the transmission characteristic signal for controlling the clutch recorded and evaluated becomes. 2. Device for a motor vehicle consisting of a drive motor, a Gearbox and one in the power flow between the drive motor and gearbox arranged switchable clutch, the transmission according to the principle is a manual transmission with at least one transmission input shaft, the can be operated automatically by means of a first actuating device and the Coupling can be actuated automatically by means of a second actuating device is and a control of the actuating means by means of a Method is carried out in such a way that at least one for the operation of the Transmission characteristic signal for controlling the clutch is detected and is evaluated,