DE102011119638A1 - Method for actuating two actuators for actuation of friction clutches of dual clutch transmission of e.g. motorcycle, involves controlling pressure of clutch, and controlling actuator of another clutch in unregulated manner - Google Patents

Abstract

The method involves controlling activation pressure of a friction clutch (20) by a pressure regulating device i.e. pressure regulating valve (40), for implementation of a rotational torque transfer that is introduced by switching a directional control valve arrangement (45). An actuator (34) of another friction clutch (24) is controlled in an unregulated manner and pressurized with unregulated pressure over diaphragms (48, 50). Activation pressure of the latter friction clutch is assigned with a source gear ratio and controlled by the pressure regulating valve. An independent claim is also included for a control arrangement.

Description

The present invention relates to a method for driving a first and a second actuator for actuating a first and a second friction clutch of a dual-clutch transmission.

Furthermore, the present invention relates to a control arrangement with an electric control device and a hydraulic circuit for actuating two friction clutches of a dual-clutch transmission and a drive train with a dual-clutch transmission and a control arrangement.

In the field of automated transmissions for motor vehicles, dual-clutch transmissions have become established for some years. These transmissions are generally advantageous because they have high efficiency. On the other hand, dual-clutch transmissions enable comfortable gear changes without interruption of traction. In this case, dual-clutch transmissions generally have a first friction clutch and a first partial transmission for odd gear ratios and a second friction clutch and a second partial transmission for even gear ratios. When changing gear, the two friction clutches are operated overlapping, so that a traction interruption during the gear change can be avoided.

In general, in such dual-clutch transmissions, the control arrangement of particular importance, which controls the friction clutches. In these control arrangements, special emphasis is placed on precisely controlling the torques transmitted by the respective friction clutches, in order to enable as smooth a torque transfer as possible.

In control arrangements, which have a hydraulic circuit with hydraulic actuators, special emphasis is placed on an exact controllability of the pressure exerted by these actuators contact pressure. For this purpose, the known hydraulic circuits usually contain a plurality of pressure control valves. These pressure control valves are usually designed as solenoid valves and on the one hand expensive to manufacture. On the other hand, such control arrangements generally have a relatively high weight.

At the document EP 1 600 669 B1 a hydraulic control device for the hydraulic actuation of two actuators for the friction clutches of a dual-clutch transmission is known, wherein a cascaded arrangement of a pressure regulating valve and a pressure valve is provided whose control input is connected to an output of the pressure regulating valve, and wherein the inputs of the pressure regulating valve and the pressure valve without Intermediate additional valves are connected to a main pressure.

The second valve is therefore not an electrically controlled solenoid valve but preferably a pressure-controlled valve. Nevertheless, there is a continuing need to reduce the cost and weight of such control arrangements.

It is therefore an object of the invention to provide an improved drive method, control arrangement and powertrain that meet this requirement.

This object is achieved on the one hand by a method for driving a first and a second actuator for actuating a first or a second friction clutch of a dual-clutch transmission, wherein an actuating pressure of one of the two friction clutches is controlled by means of a pressure control device to carry out a torque transfer, whereas the actuator the other friction clutch is driven unregulated.

The above object is further achieved by a control arrangement comprising an electric control device and a hydraulic circuit for actuating two friction clutches of a dual-clutch transmission, wherein the control device is adapted to carry out the above-mentioned method. Finally, the above object is achieved by a drive train with a dual-clutch transmission and such a control arrangement.

The control arrangement according to the invention can be realized in particular with only one pressure control valve, which is preferably designed as an electrically actuated solenoid valve.

The inventive method is based on the assumption that certain comfort restrictions can be accepted for cost-effective dual-clutch transmission. Since the torque transfer of an actuator is driven only unregulated, it is conceivable in dependence on, for example, the temperature that some switching operations can not be performed completely smoothly. Preferably, a switching device is provided, so that the pressure control device can be connected alternatively with the one friction clutch or with the other friction clutch. Due to the low cost and low weight with which such a drive train can be realized, the drive train according to the invention is also suitable, for example, for use in motorcycles. For motorcycles, the above comfort limitations may be less significant than for passenger cars. In general, the actuators can be used to carry out the According to the invention be electric actuators such as electric motors, wherein a pressure control device is then usually a means for controlling an electric current.

Of particular advantage, the inventive method, however, when the actuators are hydraulic actuators, wherein the pressure control device is a pressure control valve, which is preferably actuated electromagnetically, and wherein the actuator of the other friction clutch is acted upon via an aperture with an unregulated pressure.

In this embodiment, the torque transfer takes place in a gear change as follows. The torque of a friction clutch is controlled by means of the pressure control valve. The torque of the other friction clutch, however, is brought to a target level unregulated via a hydraulic or other throttle device. It is ensured by the diaphragm that the torque transmitted by that friction clutch does not abruptly undergo a change but changes with a certain gradient.

In a preferred embodiment, in a traction upshift from a forward gear stage to a target gear stage in the dual clutch transmission, the actuation pressure of that friction clutch associated with the target gear is increased by the pressure control valve, whereas the other friction clutch is emptied via an orifice (e.g., connected to a tank) or any other institution with a lower pressure level).

In this embodiment, upshifts may be performed from a low swell stage to a higher target aisle without traction interruption.

In a further preferred embodiment, in a train downshift from a source gear stage to a target gear in the dual clutch transmission, the operating pressure of that friction clutch, which is associated with the target gear, increased via an orifice unregulated, whereas the operating pressure of the other friction clutch, which is associated with the source gear stage, by means of Controlled pressure control valve is reduced.

In this embodiment, downshifts can also be made under train from a higher output gear stage to a lower target gear stage without a complete interruption of the tractive effort.

Overall, it is advantageous if the torque transfer is initiated by switching a directional control valve arrangement.

In this embodiment, the pressure regulating valve is preferably connectable via the directional control valve arrangement alternatively with the first or the second actuator. The directional control valve arrangement also optionally ensures that the actuator, which is not connected to the pressure control device, is connected via a diaphragm with a different pressure level, which may for example be constant and which is formed in particular by the unpressurized level of the tank of a hydraulic system.

In this embodiment, the basic functionality of the method according to the invention can be realized only by means of an electromagnetically controllable pressure control valve and a preferably likewise electromagnetically controllable directional control valve. Since directional valves are generally much cheaper than electromagnetically controlled pressure control valves, a special cost advantage can be achieved in this embodiment.

Furthermore, in contrast to the aforementioned prior art, it is not necessary to provide a separate pressure-controlled pressure valve for the second friction clutch.

It is also generally preferred if, in the case of a shift from a source gear stage to a target gear stage, the friction gearbox associated with the target gear stage is pre-filled by means of a pre-filling arrangement.

By pre-filling, it is possible to bring the friction clutch (more precisely: the actuator) of the target gear before the actual torque transfer already to a higher filling level. In this way, the filling of the friction clutch can be significantly shortened, so that in particular manually requested circuits can be made much more spontaneous.

It is possible that the Vorbefüllanordnung only in the event of a gear change request is active. Alternatively, it is preferred if the Vorbefüllanordnung provides a permanent inflation pressure for the respective inactive friction clutch (corresponding to the target gear stage associated friction clutch), so that a more spontaneous response to manually requested circuits is possible.

The inventive method is preferably carried out with friction clutches, which are designed as normally open friction clutches, so are generally in an open, not torque transmitting position, for example, by integrated springs. In general, however, it is also possible, instead of such "normally open" friction clutches, to apply the method according to the invention also in the case of "normally closed" friction clutches, which normally (that is, when the actuator does not exert any actuator force) are closed. It is also possible to perform one of the friction clutches as "normally open" and the other friction clutch as "normally closed".

It is understood that the features mentioned above and those yet to be explained below can be used not only in the particular combination given, but also in other combinations or in isolation, without departing from the scope of the present invention.

Embodiments of the invention are illustrated in the drawings and are explained in more detail in the following description. Show it:

1 a schematic representation of a drive train for a motor vehicle with a control arrangement according to the invention;

2 a schematic timing of state variables of the control arrangement in a pull upshift;

3 one of the 2 corresponding representation at a Zugrückschaltung;

4 an alternative embodiment of a control arrangement according to the invention;

5 a further alternative embodiment of a control arrangement according to the invention;

6 a further alternative embodiment of a control arrangement according to the invention;

7 a further alternative embodiment of a control arrangement according to the invention;

8th a further alternative embodiment of a control arrangement according to the invention;

9 a further alternative embodiment of a control arrangement according to the invention; and

10 a further alternative embodiment of a control arrangement according to the invention.

In 1 is a powertrain for a motor vehicle 11 as a motorcycle or a passenger car shown schematically and generally with 10 designated.

The powertrain 10 has a drive motor 12 on, for example, is designed as an internal combustion engine and is controlled by an accelerator pedal or a gas actuator (rotary handle on a motorcycle). The drive motor 12 returns an engine torque M _M. Furthermore, the drive train 10 a dual-clutch transmission 14 on, whose input to the output shaft of the drive motor 12 is connected and its output with a differential 16 connected is. At the entrance of the differential 16 is an output torque M _A. The in the differential 16 introduced drive power is by means of the differential 16 on powered wheels 18L . 18R of the motor vehicle 11 distributed. The dual-clutch transmission 14 has a first friction clutch 20 and a first partial transmission 22 on. Furthermore, the dual-clutch transmission includes 14 a second friction clutch 24 and a second partial transmission 26 , The two friction clutches 20 . 24 are normally open and have a common input member connected to the output of the drive motor 12 connected is. An output member of the first friction clutch 20 is with an input of the first sub-transmission 22 connected to the example, the odd gear ratios are assigned. An output member of the second friction clutch 24 is with an input of the second sub-transmission 26 connected to the example, the even gears are assigned. The entrance of the first subtransmission 22 has a speed n ₁ . The entrance of the second partial transmission 26 has a speed n ₂ .

For automated operation of the dual-clutch transmission 14 (optionally also the drive motor 12 ) is a control arrangement 28 intended. The control arrangement 28 has a first hydraulic actuator 30 for actuating the first friction clutch 20 and a second hydraulic actuator 34 for actuating the second friction clutch 24 on.

Furthermore, the control arrangement includes 28 an electronic control device 32 , The control device 32 is with the partial transmissions 22 . 26 or with actuators provided therein for the automated actuation of clutches. Furthermore, the control device 32 with a hydraulic circuit 36 connected by means of which the actuators 30 . 34 be controlled.

The hydraulic circuit 36 has a pump 38 on, at the output of a main pressure P is pending. Furthermore, the hydraulic circuit includes 36 a pressure control valve 40 which is electromagnetically actuated and with the control device 32 connected is. The line branch in which the main pressure P is present, is also via a check valve 42 with a tank 44 connected. The check valve 42 is designed so that the main pressure P can not exceed a certain maximum level. The pump 38 and the pressure control valve 40 are also connected to the tank.

Furthermore, the hydraulic circuit includes 36 a directional control valve arrangement 45 , In the present case, the Way valve arrangement 45 through a single 4-2-way valve 46 formed, which is driven electromagnetically and with the control device 32 connected is.

A first connection of the directional control valve 46 is with the pump 38 connected. A second connection is with the first actuator 30 connected. A third connection is with the tank 44 connected, and a fourth connection is with the second actuator 34 connected. In the pipe between the pump 38 and the first port is an aperture 48 intended. Further, a second aperture 50 in the line between the third port and the tank 44 intended.

In the illustrated position of the directional control valve 46 the first port is connected to the second port, leaving the first actuator 30 over the aperture 48 is supplied with the main pressure P, so that the friction clutch 20 closed is. The second actuator 34 is about the third and fourth port and the bezel 50 with the tank 44 connected so that the second friction clutch 24 is open. Consequently, in the drive train 10 a power transmission from the drive motor 12 over the first friction clutch 20 and the first partial transmission 22 towards the differential 16 ,

When the directional valve 45 is the second actuator is the second actuator 34 over the aperture 48 connected via the main pressure P, leaving the second clutch 24 closed is. Further, in the second position, the second and third terminals are connected to each other so that the first actuator 30 over the aperture 50 with the tank 44 connected is. In this position of the directional control valve power transmission is generally via the second friction clutch 24 and the second partial transmission 26 ,

A gear change with torque transfer from a friction clutch 20 on the other friction clutch 24 is done by switching the directional control valve 46 , From the in 1 shown position then becomes the first actuator 30 over the aperture 50 with the tank 44 connected so that the pressure in the first actuator 30 gradually breaks down and the clutch 20 does not open abruptly but gradually opens. Further, the pressure regulating valve becomes 40 so driven that the pressure in the second actuator 34 gradually over the aperture 48 is increased to the second friction clutch 24 to close regulated.

Because the pressure reduction in the first actuator 30 over the second aperture 50 unregulated occurs, depending on the preferably adjustable control time of the directional control valve 46 , It may happen that, depending on, for example, the temperature of the hydraulic fluid used, the torque transfer is not completely smoothly.

2 shows an example of an upshift in the in 1 shown powertrain 10 , 2 shows timing diagrams of different states. In a top diagram is the state of the gas actuator 13 shown. This is at the maximum value over the entire duration, corresponding to a "kickdown".

In the diagram below, the rotational speeds n ₁ and n _{2 of} the transmission input shafts are shown in dashed lines. The effective at the transmission output speed (the translation of the transmission is excluded for reasons of clarity) is shown in a solid line.

The timing diagram below shows the clutch torques or the torques at the input of the first subtransmission 22 (TG1) and the second sub-transmission 26 (TG2) are present.

In the area below it is that of the drive motor 12 provided engine torque M _M shown. In the bottom timing diagram, this is at the output of the dual clutch transmission 14 upcoming output torque M _A shown schematically.

In 2 an upshift request is made at time t ₀ (manually or by means of an automated control). In this state, the torque applied to the second partial transmission (TG2) is high, with the second friction clutch 24 is closed (ride, for example, in the gear 2 ). In the first partial transmission 22 (TG1), for example, the next higher gear is preselected (gear ratio 3 ), wherein the first friction clutch 20 is open. Preferably, the clutch torque of the friction clutch 20 on a level KP according to a so-called "kiss point". This may optionally be achieved only by means of a pre-filling device, which will be described below. In the simplest case, the torque applied to the transmission input of the first partial transmission (TG1) can also be zero.

Further, by the drive motor 12 a certain drive torque> 0 is provided, so that at the output of the dual-clutch transmission 14 a drive torque M _A results.

In some embodiments, which will be described below, the directional control valve assembly 45 switched at a time t ₀ . For the basic version of the control arrangement according to the invention, however, initially it is assumed that the directional control valve arrangement 45 to a time t _{1 is} switched. This is from this time the pressure and thus of the first friction clutch 20 transmitted torque regulated by means of the pressure control valve 40 elevated. The pressure in the second friction clutch 24 falls on the other hand, on the aperture 50 unregulated, so that this pressure drop can occur with different gradients, which is in 2 is shown schematically by dashed lines. Due to the gear change in the higher gear, the drive torque M _A falls from the time t ₁ to a value that is reached at time t ₂ . Subsequently, the gear change is basically complete.

Subsequent to t ₂ , however, can still a short torque withdrawal by means of the drive motor 12 respectively. This has the following reason: By reducing the torque, the load can be reduced during the speed adjustment from n ₂ to n ₁ .

At time t ₃ , the gear change is completed. The first friction clutch 20 is then completely closed, and the second friction clutch 24 is open (possibly at level below the kiss point, corresponding to a drag torque level).

In 3 is explained in schematic form by means of the same timing diagrams a train downshift.

The difference to the in 2 shown upshift is that the torque of the source gear (TG1) regulated by means of the pressure control valve 40 is reduced, whereas torque of the target gear stage associated second friction clutch (TG2) unregulated across the aperture 48 is increased. Depending on the preferably adjustable control time of the directional control valve 46 may be a previous or later pressure or torque build up of TG2.

In the 4 to 10 Further alternative embodiments of control arrangements are shown, which are generally in terms of structure and operation of the in 1 shown control arrangement 28 correspond. The same elements are therefore provided with the same reference numerals. The following section essentially explains the differences.

In the control arrangement 28 ' of the 4 includes the directional control valve assembly 45 a first electromagnetically controllable directional control valve 46a and a second electromagnetically controllable directional control valve 46b , The directional valves 46a . 46b are each with the controller 32 connected. The first directional valve 46a connects the pump 38 over a first panel 48a with the first actuator 30 , The second directional valve 46b connects the pump 38 over another first aperture 48b with the second actuator 34 , Because the directional valves 46a . 46b can be controlled independently of one another, the sequence control can be offset in time with a torque transfer. This is advantageous since the process of filling the coupling assigned to the target gear stage generally lasts considerably longer than the opening of the clutch associated with the output gear stage. The directional valves 46a . 46b can each be designed as 3-2-way valves. A connection of the directional valves 46a . 46b is preferably each with its own second aperture 50a . 50b with the tank 44 connected. The first directional valve 46a is in 4 also called SV_1. The second directional valve 46b is in 4 also called SV_2. Furthermore, in 4 the pressure control valve also called LPDV.

Furthermore, in 4 in the hydraulic circuit 36 a pressure measuring arrangement 52 integrated. To the pressure measuring arrangement 52 with only one pressure sensor 56 To realize, is the pressure sensor 56 via a designed as a pressure compensator directional control valve 54 connected. The directional valve 54 is hydraulically controllable on both sides, by means of the actuators 30 . 34 connected lines. The pressure compensator function of the directional control valve 54 is designed so that the pressure sensor 56 always with the line of the actuator 30 connected, in which the lower pressure is present. Advantageously, in this case, with knowledge of the actual filling pressure, the switching off of the active coupling (assigned to the source gear stage) can be coupled directly to the instantaneous pressure of the passive clutch (assigned to the target gear stage).

By switching off the main pressure P, both friction clutches 20 . 24 regardless of the state of the rest of the hydraulic circuit 36 be depressurized and opened. As a result, a functional reliability is achieved in that no stress state can be present in a single error.

Load circuits can by suitable control of the directional control valves 46a . 46b and by controlling the main pressure P by means of the pressure regulating valve 40 be performed. Furthermore, circuits are possible in which a traction interruption takes place (so-called "torque-break" circuits). In this case, the same timing can be realized as in load circuits, but the main pressure P is lowered.

5 shows a further control arrangement 28 '' , in terms of structure and operation generally the control arrangement 28 ' of the 4 equivalent. In the control arrangement 28 '' is additionally a Vorbefüllanordnung 58 intended. The pre-filling arrangement has a pressure limiting valve 60 on, that with the main pressure P as well as with two exchange valves 62 . 64 connected is. The respective other inputs of the shuttle valves 62 . 64 are connected to the outputs of the directional valves SV_1, SV_2.

In this embodiment, it can be achieved that also at the non-active clutch generally a permanent filling pressure is applied, which may correspond, for example, a clutch torque slightly smaller than the kiss point moment. This value is determined by the pressure relief valve 60 set up.

By this measure, the time of pre-filling can be largely eliminated, so that manually requested circuits can be made much more spontaneous. However, in this embodiment, the filling pressure is permanently present and can not be changed.

In 6 is another embodiment of a control arrangement 28 ''' shown in terms of structure and operation generally the control arrangement 28 ' of the 1 equivalent. In the control arrangement 28 ''' of the 6 is the first directional valve 46 ''' designed as a pressure compensator, and the second directional control valve 46b ''' is also designed as a pressure balance. The second directional valve 46b ''' corresponds in its function to the valve 46 of the 1 , An input of the second directional valve 46b ''' is directly connected to the main pressure P. An entrance of the other way valve 46 ''' is via a switching valve 68 connected in the form of a 3-2-way valve with the main pressure P. The switching valve 68 is electromagnetically controllable and with the control device 32 connected.

In this embodiment, the pressure of the passive clutch is not regulated but by the switching valve 68 switched on.

The advantage is the use of a lower cost switching valve 68 instead of a pressure reducing valve. However, pre-filling is not possible in this embodiment.

Load circuits and traction interrupting circuits are made by switching the switching valve 68 as well as by controlling the main pressure P.

In the 7 to 10 Further variants of control arrangements are shown. In the control arrangement 28 ^IV of 7 is a first way valve 46a ^IV in the form of a 3-2-way valve the first actuator 30 assigned. The second actuator 34 is a second directional valve 46b Assigned to ^IV (V_SV_1). The second directional valve 46b ^IV is pressure controlled and is controlled by a switching valve 74 operated, which is designed as electromagnetically actuated 3-2-way valve.

The first directional valve 46a is at the same time designed as a pressure reducing valve, so that the valve can be used to pre-fill the non-active coupling. This is in 7 schematically at 58 ^IV indicated.

In load circuits, the torque transfer by switching the means of the switching valve 74 pilot-operated 4-2-way valve 46b ^IV causes. This process reverses each with an actuator 30 . 34 connected tank and main pressure connections. The pressure build-up and pressure reduction takes place via panels BL1.1 and BL1.2. This hydraulically simple concept allows traction-free handovers. However, since no arbitrary momentum and no torque curve can be represented on a single clutch, the switching comfort (jerk freedom) is limited. In that regard, the operation of the embodiment corresponds to the 1 ,

To the passive clutch regardless of the switching operation of the directional control valve 46b ^IV and independent of the pressure of the active clutch to fill and print (pre-filling) is the first directional control valve in the form of the pressure reducing valve 46a ^IV provided. Its pressure connection is via a shuttle valve 62 . 64 with the two actuators 30 . 34 connected. This determines the working pressure of the pressure reducing valve 46a ^IV the minimum pressure of both clutches. Thus, with the working pressure of the pressure reducing valve 46a ^IV filled the passive coupling and possibly even printed on the kiss point addition.

If the passive coupling is printed beyond the kiss point, it must be after switching the directional control valve 46b the working pressure of the pressure reducing valve 46a ^IV be withdrawn controlled in order to relieve the formerly active clutch completely.

A start-up can be done with the system pressure by controlling the pressure control valve 40 respectively.

Load circuits are controlled by the valves 46b ^IV , 46a ^IV and 40 ,

Traction interrupting circuits can by briefly switching off the main pressure by means of the pressure regulating valve 40 will be realized.

About the panels 78 and 50 in each case the back pressure when switching the OR valves 62 . 64 degraded, and in a slightly delayed form. The aperture 48 serves to generate a defined pressure difference at the two OR valves 62 . 64 : Pressure behind pressure reducing valve 46a ^IV to pressure behind directional valve 46b ^IV .

In 8th is another embodiment of a control arrangement 28 ^V in terms of structure and operation generally the control arrangement 28 ^IV of 7 equivalent. In this embodiment, the pre-filling pressure is achieved by means of a pre-filling arrangement 58 ^V set as a simple pressure reducing valve 76 ^{V is} formed.

Consequently, in this embodiment, the pre-filling pressure is permanently present and can not be changed.

In 9 is another embodiment of a control arrangement 28 ^IV shown, the general structure of the system arrangements 28 ^III and 28 ^IV of 6 and 7 equivalent.

In this embodiment, the switching of the clutch pressure supply is not effected by a freely controllable by the transmission software switching valve but is the pressure of the passive clutch, here in the same way by a pressure reducing valve 76 ^{VI is} regulated, triggered. This release or triggering takes place on the second directional control valve designed as a pressure balance 46b ^VI , as soon as the force of the working pressure of the pressure reducing valve 76 ^VI on the valve spool of the directional control valve 46b ^{VI is} greater than the counterforce of that valve spool, which results from the clutch pressure of the active clutch (= main pressure). Thus, the switching point may be below the main pressure, it is preferred if the surface of the triggering pressure is greater than the area acted upon by the clutch pressure (main pressure). The switching pressure is determined by the main pressure (provided by the pressure regulating valve 40 ) and by the constructive design of the designed as a pressure compensator valve 46b ^VI . By blending in front of the pilot surfaces of the directional control valve 46b ^VI , the switching point can also increasingly from the pressure build-up dynamics of the pressure reducing valve 76 ^VI , which may result in further meaningful control and application possibilities.

After the directional valve 46b ^VI has carried out the hydraulic function change by the triggering and have overlapped the clutch pressures, the likewise designed as a pressure compensator further directional control valve begins 46a ^VI to move. After switching over this valve, the hydraulics are ready for re-triggering, which leads to the initial state.

The release pressure of the pressure reducing valve 76 ^VI is preferably held until the clutch pressures have overlapped. Because only then remains the directional control valve 46b ^VI in the new position. The working pressure of the pressure reducing valve 76 ^VI should be lowered again below the trigger pressure before the designed as a pressure compensator directional control valve 46b ^VI performs the hydraulic function change. Otherwise, triggering would be triggered again at this directional control valve.

So after parking the vehicle in the hydraulic circuit 36 a defined state prevails, it is preferred if the two positions of the designed as a pressure compensator directional valve 46a ^VI be taken restraining. Alternatively, a weak return spring is possible, which forces the slider in a certain position.

The starting takes place in this embodiment by means of the pressure control valve 40 , Load circuits are triggered by triggering the pressure reducing valve 76 triggered. In traction interrupting circuits, the main pressure by means of the pressure control valve 40 reduced.

In 10 is another control arrangement 28 ^VII shown in terms of structure and operation generally the control arrangement 28 ^V the 9 equivalent.

In the control arrangement 28 ^{VII of} the 10 However, an exact timing of the pressure reducing valve 76 ^VII omitted, since the trained as a pressure compensator valve 46a ^VII hydraulically locked itself. This can be done so that this pressure compensator switches only when the working pressure of the pressure reducing valve 76 ^VII drops below the kiss point pressure. This embodiment is functionally more robust than the embodiment of FIG 9 ,

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• EP 1600669 B1 [0006]

Claims (8)

Method for activating a first and a second actuator ( 30 . 34 ) for actuating a first or a second friction clutch ( 20 . 24 ) of a dual-clutch transmission ( 14 ), wherein to carry out a torque transfer, an actuating pressure of one of the two friction clutches ( 20 . 24 ) by means of a pressure control device ( 40 ), whereas the actuator of the other friction clutch ( 20 . 24 ) is driven unregulated. Method according to claim 1, wherein the actuators are hydraulic actuators ( 30 . 34 ), wherein the pressure regulating device ( 40 ) is a pressure control valve and wherein the actuator of the other friction clutch ( 20 . 24 ) via a diaphragm ( 48 . 50 ) is subjected to an uncontrolled pressure. Method according to claim 2, wherein in a pull upshift ( 6 - 7 ) from a source gear stage ( 6 ) to a target gear ( 7 ) in the dual-clutch transmission ( 14 ) the actuation pressure of that friction clutch ( 20 . 24 ), the target gear ( 7 ), by means of the pressure regulating valve ( 40 ) is increased in a controlled manner, whereas the other friction clutch ( 20 . 24 ) via a diaphragm ( 48 . 50 ) is emptied unregulated. A method according to claim 2 or 3, wherein in a train downshift ( 7 - 6 ) from a source gear stage ( 7 ) to a target gear ( 6 ) in the dual-clutch transmission ( 14 ) the actuation pressure of that friction clutch ( 20 . 24 ), the target gear ( 6 ), via an aperture ( 48 . 50 ) is increased unregulated, whereas the actuation pressure of the other friction clutch ( 20 . 24 ), which is assigned to the source gear stage, by means of the pressure regulating valve ( 40 ) is reduced in a controlled manner. Method according to one of claims 1 to 4, wherein the torque transfer by switching a directional control valve arrangement ( 45 ) is initiated. Method according to one of claims 1 to 5, wherein in a shift from a source gear stage to a target gear stage, the target gear stage associated friction clutch ( 20 . 24 ) is pre-filled by means of its Vorbefüllanordnung. Control arrangement ( 28 ) with an electrical control device ( 32 ) and a hydraulic circuit ( 36 ) for actuating two friction clutches ( 20 . 24 ) of a dual-clutch transmission ( 14 ), wherein the control device ( 32 ) is adapted to perform the method according to one of claims 1 to 6. Powertrain ( 10 ) with a dual-clutch transmission ( 14 ) and a control arrangement ( 28 ) according to claim 7.

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