Clutch system

A clutch system includes a clutch arrangement between a drive unit and a transmission, a pump arrangement which supplies pressure medium to the clutch arrangement at a pressure and rate of flow, and an operating medium supply arrangement which supplies an operating medium to the clutch arrangement at a pressure which is less than the pressure of the pressure medium and a rate of flow which is greater than the rate of flow of the pressure medium. The operating medium supply arrangement includes a pressure receiving element which is movable in a housing and has at least one pressure receiving surface which is exposed to the pressure medium, and an interaction element which is coupled to move with the pressure receiving element and has at least one interaction surface which interacts with the operating medium to produce a flow of the operating medium.

BACKGROUND OF THE INVENTION

1. Priority Claim

This is the national phase application of PCT Application No. PCT/EP01/12515 filed on Oct. 30, 2001.

2. Field of the Invention

The invention relates to a clutch system including at least one clutch device, particularly for arrangement in a drivetrain (especially in a motor vehicle) between a drive unit and a transmission, wherein the clutch device has at least one clutch arrangement which is provided for operation under the effect of an operating medium and/or which can be actuated through a pressure medium.

3. Description of the Related Art

Reference is made herein in particular, but not exclusively, to a wet multi-disk clutch arrangement. Further, reference is made herein above all, but not exclusively, to a clutch arrangement which can be actuated by hydraulic means with a hydraulic slave cylinder integrated in the clutch device as is disclosed, for example, in U.S. Pat. No. 5,887,690 or in one of the present Assignee's patent publications concerning multiple clutch devices (particularly dual clutch devices). Reference is made particularly to U.S. Pat. Nos. 6,499,578, 6,523,657, and 6,586,852. The clutch arrangements of the referenced clutch devices are wet multi-disk clutch arrangements. U.S. Pat. No. 6,059,682 can be referred to with respect to wet, hydraulically actuated multi-disk clutch arrangements. A dry hydraulically actuated dual clutch device is known from U.S. Pat. No. 4,714,147.

In principle, there are many possible ways in which an operating medium, possibly a cooling liquid (especially cooling oil), is supplied to the clutch arrangement or to a plurality of clutch arrangements of the clutch device. It is the job of the designer to keep the costs and space requirement as low as possible while providing for a good efficiency and, accordingly, a low energy consumption for supplying the operating medium.

In some cases, other marginal conditions must be accounted for as well. For a clutch device with one or more wet, hydraulically actuated clutch arrangements (particularly multi-disk clutch arrangements), a flow of hydraulic medium must be supplied for controlling the clutch device or clutch devices in order to activate a clutch, i.e. engage a normally open type clutch or disengage a normally closed type clutch. A relatively small volume flow at relatively high pressure is needed for engaging and disengaging. Further, a large volume flow of coolant, particularly cooling liquid (generally oil), is required for wet operation or for cooling of the clutch devices, wherein a volume flow at a lower pressure level is sufficient or even necessary.

It is possible to provide two separate pumps which are driven by motor (e.g., determined by the drive unit and/or transmission) or by electric motor, one of which is designed to supply the pressure medium flow for controlling the clutch device or clutch devices and the other is designed for supplying the flow of operating medium. However, this solution is comparatively costly.

DE 198 00 490 A1 discloses another approach. A suction jet pump is provided in the supply area of a hydraulic fluid serving to cool a multi-disk clutch arrangement in order to achieve high throughflow rates and, therefore, a high cooling output. The suction jet pump receives a flow of hydraulic fluid from a hydraulic pump which provides a hydraulic pressure sufficient for actuating the multi-disk clutch arrangement. The suction jet pump sucks additional hydraulic fluid out of the oil pan.

While the hydraulic pump need not be designed for very high pumping outputs and a separate motor-driven pump for pumping coolant can be dispensed with using the approach disclosed in DE 198 00 490 A1, the approach based on the suction jet pump is disadvantageous because the efficiency of the suction jet pump is comparatively poor due to the high vortex losses of a suction jet pump and the energy consumption is consequently relatively high.

SUMMARY OF THE INVENTION

According to a first embodiment, a clutch system includes at least one clutch device, particularly for arrangement in a drivetrain between a drive unit and a transmission. The clutch device has at least one clutch arrangement which is provided for operation under the action of an operating medium. The operating medium can be supplied to the clutch arrangement by means of a pump arrangement supplying pressure medium and an operating medium supply arrangement. The operating medium supply arrangement is connected to the pump arrangement and provides the operating medium in a flow of the operating medium with reduced pressure relative to a pressure medium whose pressure is given by the pump arrangement and, in relation to the medium volume supplied per time unit averaged over time. The operating medium flow is greater than a pressure medium flow possibly flowing from the pump arrangement to the operating medium supply arrangement. Therein, the operating medium supply arrangement has at least one pressure receiving element and at least one interaction element. The pressure receiving element is arranged so as to be movable in a housing and which is exposed at least on at least one pressure receiving surface to the pressure medium that is provided by the pump arrangement in operation and that can be set in motion through the intermediary of the pressure medium. The interaction element is arranged so as to be movable in the housing or a housing and is coupled with respect to movement with the pressure receiving element and interacts during operation with the operating medium at least on at least one interaction surface. The operating medium flow can be supplied through the intermediary of this interaction element.

Due to the operating medium supply arrangement according of the present invention, the operating medium can be provided economically and in a reliably operating manner with high efficiency and in a comparatively large volume flow. The operating medium supply arrangement can advisably be connected to an operating medium reservoir and can draw the operating medium from the latter to supply the clutch arrangement.

The operating medium supply arrangement may be constructed as a pressure converter arrangement. In one embodiment, the operating medium supply arrangement is constructed in such a way that the effective area of the pressure receiving surface exposed to the pressure medium is smaller per time unit averaged over time than the effective area of the interaction surface interacting with the operating medium.

In one economical embodiment, at least one component is provided which is constructed integrally or in one piece and which forms the pressure receiving element or a pressure receiving element and the interaction element or an interaction element. For example, the component can advantageously be constructed as a piston arranged in a piston housing so as to be displaceable along an axis with a first piston surface exposed to the pressure medium and a second piston surface interacting with the operating medium. The first piston surface serves as a pressure receiving surface and the second piston surface serves as an interaction surface. The piston surfaces can be arranged at opposite axial end faces of the piston. The first piston surface is preferably smaller than the second piston surface.

Another embodiment which is particularly advantageous with respect to the attainable volume flow is characterized in that at least one rotor component forms the pressure receiving element and/or the interaction element.

The particular construction of the operating medium supply arrangement notwithstanding, this can comprise a pump arrangement or can be identifiable as a pump arrangement.

In general, the clutch device can be actuated through the intermediary of the pressure medium supplied by the pump arrangement.

According to a second aspect, the present invention provides, a clutch system includes at least one clutch device particularly for arrangement in a drivetrain, particularly in a motor vehicle between a drive unit and a transmission. The clutch device has at least one clutch arrangement which can be actuated through the intermediary of pressure medium, wherein the pressure medium can be supplied for the actuation based on a pump arrangement supplying an operating medium and based on a pressure medium supply arrangement which is connected to the pump arrangement and which supplies the pressure medium at a pressure that is increased relative to a pressure of the operating medium that is given by the pump arrangement, and possibly in a flow of pressure medium which, with respect to medium volume supplied per time unit averaged over time, is smaller than an operating medium flow possibly flowing from the pump arrangement to the pressure medium supply arrangement, wherein the pressure medium supply arrangement has at least one pressure receiving element which is arranged so as to be movable in a housing and which is exposed in operation at least on at least one pressure receiving surface to the operating medium supplied by the pump arrangement and can be set in motion through the intermediary of the operating medium, and at least one interaction element which is arranged so as to be movable in the housing/a housing and which is coupled with respect to movement with the pressure receiving element and interacts during operation with the pressure medium at least on at least one interaction surface, and the pressure medium can be supplied through the intermediary of the interaction element.

Due to the pressure medium supply arrangement according to the present invention, the operating medium can be provided economically and in a reliably operating manner with high efficiency and with a pressure level that is sufficient for clutch actuation. In this connection, the pressure medium supply arrangement can advisably be connected to a medium reservoir and can draw medium from the latter to supply the clutch arrangement and/or can conduct medium away to the medium reservoir. The pressure medium supply arrangement can also be constructed in such a way that it supplies the operating medium drawn from the pump arrangement as pressure medium.

The pressure medium supply arrangement can advisably be constructed as a pressure converter arrangement. According to an advantageous construction, the pressure medium supply arrangement is constructed in such a way that the effective area of the pressure receiving surface exposed to the operating medium is larger per time unit averaged over time than the effective area of the interaction surface interacting with the pressure medium.

In one economical embodiment, at least one component is provided which is constructed integrally or in one piece and which forms the pressure receiving element and the interaction element. For example, the component can advantageously be constructed as a piston arranged in a piston housing so as to be displaceable along an axis with a first piston surface which serves as a pressure receiving element and which is exposed to the operating medium and a second piston surface interacting with the pressure medium which serves as interaction surface. The piston surfaces can be arranged at opposite axial end areas of the piston. The first piston surface is preferably larger than the second piston surface.

In another advantageous embodiment, at least one rotor component forms the pressure receiving element and/or the interaction element.

Irrespective of the specific construction of the pressure medium supply arrangement, this can comprise a pump arrangement or can be identified as a pump arrangement.

Above all, the clutch device is intended for operation under the influence of an operating medium which can be supplied to the clutch device by the pump arrangement for this operation.

With respect to the clutch system according to the first aspect and with respect to the clutch system according to the second aspect, it is provided that a hydraulic medium is used as operating medium and as pressure medium.

According to a third aspect of the present invention, a clutch system comprises at least one clutch device particularly for arrangement in a drivetrain (particularly in a motor vehicle) between a drive unit and a transmission, wherein the clutch device has at least one clutch arrangement which is provided for operation under the influence of an operating medium and to which the operating medium can be supplied for this operation based on a first pump arrangement (also referred to hereinafter as primary pump arrangement) which supplies the pressure medium and a second pump arrangement (also referred to hereinafter as secondary pump arrangement) which is connected to the first pump arrangement and which can be driven by the pressure medium supplied by the first pump arrangement, wherein the secondary pump arrangement has at least one pump element which is arranged so as to be movable in a housing, can be driven through the intermediary of the pressure medium (also referred to hereinafter as primary medium), interacts with the operating medium (also referred to hereinafter as secondary medium), and conveys the secondary medium (operating medium) during operation by means of displacement and/or by imparting an acceleration and/or by imparting torsion or angular momentum in the direction of the clutch arrangement. The secondary pump arrangement which is provided according to the present invention can be identified as an operating medium supply arrangement according to the first aspect of the present invention.

Further, according to a fourth aspect, the present invention provides a clutch system comprising at least one clutch device particularly for arrangement in a drivetrain (particularly in a motor vehicle) between a drive unit and a transmission, wherein the clutch device has at least one clutch arrangement which can be actuated through the intermediary of pressure medium, wherein the pressure medium can be supplied for this actuation based on a second pump arrangement (also referred to hereinafter as primary pump arrangement) supplying operating medium and a first pump arrangement (also referred to hereinafter as secondary pump arrangement) which is connected to the second pump arrangement and can be driven by the operating medium supplied by the second pump arrangement, wherein the secondary pump arrangement has at least one pump element which is arranged so as to be movable in a housing, can be driven through the intermediary of the operating medium (also referred to hereinafter as primary medium), interacts with the pressure medium (also referred to hereinafter as secondary medium) and, during operation, brings the secondary medium (pressure medium) to a pressure sufficient for clutch actuation by means of displacement and/or imparting an acceleration and/or imparting a torsion or angular momentum. The secondary pump arrangement which is provided according to the invention can be identified as a pressure medium supply arrangement according to the second aspect of the present invention.

As a rule, in any case when the operating medium is a liquid, the second pump arrangement is connected to an operating medium reservoir and can draw operating medium from the latter for supplying to the clutch arrangement. The second pump arrangement can supply the operating medium in a operating medium flow which has a lower pressure than a pressure of the pressure medium that is given by the first pump arrangement and which, with respect to the medium volume supplied per time unit averaged over time, is greater than pressure medium flow possibly flowing from the first pump arrangement to the second pump arrangement or which can be supplied in its entirety in continuous operation by the pump arrangement.

With respect to the drive of the primary pump arrangement, it is intended above all that this primary pump arrangement can be driven electrically and/or through the intermediary of the drive unit and/or the transmission.

At least one of the first and second pump arrangements can comprise at least one hydrostatic pump or a pump constructed as a positive-displacement machine. Above all, it is intended that the first pump arrangement has at least one hydrostatic pump or pump constructed as a positive-displacement pump. Pumps of this kind are particularly suitable for building up high pressures and therefore—to mention one example—can advantageously supply the pressure that may be required for actuating the clutch arrangement. Examples for suitable pumps of the above-mentioned type, depending on requirements, are piston pumps, particularly axial piston pumps and radial piston pumps, toothed gear pumps and toothed ring pumps, vane cell pumps, screw pumps and rotary plunger pumps. Further or alternatively, at least one of the first and second pump arrangements can comprise at least one hydrodynamic pump or a pump constructed as a liquid flow machine. It is intended particularly that the second pump arrangement has at least one hydrodynamic pump, or pump constructed as a flow machine, that can be driven by pressure medium supplied by the first pump arrangement. Pumps of this kind are particularly suited to supply large volume flows, so that, if necessary, a large volume flow of operating medium can easily be supplied for feeding to the at least one clutch arrangement. Examples of pumps of this type are torque or swirl pumps and centrifugal pumps or impeller pumps.

For the application mentioned above with at least one hydraulically actuated wet multi-disk clutch arrangement, because of the comparatively high pressure at a comparatively small volume flow that is required for clutch actuation on the one hand and by reason of the comparatively high volume flow at a comparatively low pressure required for wet operation on the other hand, it seems advantageous when the first pump arrangement comprises at least one hydrostatic pump or pump constructed as a positive-displacement machine and when the second pump arrangement comprises at least one hydrodynamic pump or pump constructed as a flow machine. However, this design of the clutch system which is viewed as especially advantageous is in no way compulsory.

The pump (of the first or second pump arrangement) can have at least one pump element that can be driven in a translational or rotational manner and which can be driven by a motor, particularly an electric motor, in the case of a pump associated with the primary pump arrangement or which can be driven through the intermediary of the primary medium in case of a pump associated with the secondary pump arrangement. For example, the secondary pump arrangement can have at least one piston pump that can be driven by the primary medium supplied by the primary pump arrangement. The piston pump can have at least one piston which is arranged in a pump housing so as to be displaceable along an axis and with a first piston surface exposed to the primary medium and a second piston surface interacting with the secondary medium. The piston surfaces can be arranged at opposite axial end areas of the piston. In order to make use of or achieve the possibly high pressure level of the pressure medium on the one hand and/or to achieve or make use of a high operating medium delivery volume on the other hand, it can be advantageous when the piston surface associated with the pressure medium is smaller than the piston surface associated with the operating medium.

When only one piston is provided, the pump delivers the secondary medium in pulses because of an oscillating piston operation, so that the pump can also be referred to as an impulse pump. In order to provide an at least approximately continuous secondary medium flow and/or to prevent mechanical vibrations deriving from the oscillating movement of the piston, the secondary pump arrangement can have a plurality of piston pumps which can be operated synchronously with offset piston stroke phases. For example, a first piston pump and a second piston pump can be operated synchronously with piston stroke phases that are offset by approximately 180°. As a result of this kind of oppositely running operation of the two piston pumps, mechanical vibrations are sufficiently compensated or prevented.

The expenditure on control of the secondary pump arrangement is particularly low when an intake or suction stroke and a discharge stroke of the piston pump can be triggered alternately by means of a valve arrangement which is connected between the primary pump arrangement and the piston pump of the secondary pump arrangement. Either the suction stroke or the discharge stroke or both can be triggered by applying primary medium pressure to a control input of the piston pump associated with the respective stroke. The piston pump can have a return spring arrangement which is put under increasing tension during a suction stroke and discharge stroke.

It is suggested that at least one pressure reduction valve is provided, by means of which a primary medium pressure applied to a control input of the piston pump can be reduced by allowing the primary medium to flow off in order to trigger the suction stroke or discharge stroke of the piston pump.

The other respective stroke, the suction stroke or discharge stroke, can then be carried out by means of the return spring acting on the piston. When the valve arrangement is arranged between the primary pump arrangement and the piston pump it can form the pressure reduction valve or has this pressure reduction valve. However, the pressure reduction valve can also be a valve that is separate from said valve arrangement.

It is particularly preferable that the pressure reduction valve is integrated in the secondary pump arrangement. This avoids the flow resistance of longer lines so that the primary medium pressure can be reduced especially quickly, and a particularly high pump output can be achieved for the piston pump. The pressure reduction valve can be mechanically actuated through the intermediary of a stroke movement of a/the piston of the piston pump and/or through the intermediary of primary medium acting on the pressure reduction valve. In this connection, it is suggested by way of a further development that a plurality of valve elements of the pressure reduction valve are displaceable between a first relative position identified as opening position and a second relative position identified as the closing position in response to the stroke movement of the piston and/or in response to primary medium acting on at least one of the valve elements. For example, a first valve element and a second valve element can be provided which are pretensioned in the direction of the open position by a pretensioning spring arrangement. The piston can advantageously form a valve element.

In a particularly preferred manner, the pressure reduction valve can be moved into an/the open position in that a valve element executing the stroke movement along with the piston contacts a stop and/or in that the pressure reduction valve is actuated by the piston contacting an actuating stop and can be brought into a/the closing position by primary medium acting on at least one valve element. In this way, a compulsory control of the pressure reduction valve can be achieved to a certain extent, so that it need not be controlled separately.

In general, it is advantageous for a highly efficient secondary pump arrangement when the latter is constructed with at least one valve, particularly a non-return valve, which prevents secondary medium from the secondary pump arrangement (possibly from the piston pump) from flowing back into the secondary medium reservoir and/or with at least one valve, particularly a non-return valve, which prevents secondary medium from the side of the clutch arrangement from flowing back into the secondary pump arrangement (possibly into the piston pump).

According to an advantageous constructional variant, the secondary pump arrangement has at least one swirl pump or centrifugal pump (impeller pump) which can be driven by the primary medium supplied by the primary pump arrangement or at least one rotary positive-displacement pump, possibly a toothed gear pump or toothed ring pump, which can be driven by the primary medium supplied by the primary pump arrangement. As was already mentioned above, swirl pumps or centrifugal pumps can advantageously supply large volume flows, while positive-displacement pumps are particularly well suited for supplying higher pressures. The swirl pump or centrifugal pump or positive-displacement pump can have at least one rotor which interacts with the secondary medium. The rotor can be constructed as at least one blade or at least one rotor having another flow geometry or conveying geometry.

For purposes of driving the rotor, it is suggested that the latter can be driven by means of a hydromotor receiving primary medium from the primary pump arrangement. The hydromotor can be constructed as a flow machine (possibly a turbine) or positive-displacement machine. For example, a toothed gear motor or toothed ring motor, piston motor (particularly radial piston motor or axial piston motor) and a vane cell motor can be used. Generally speaking, the hydromotor can be constructed as a hydrostatic or hydrodynamic motor.

The hydromotor can be constructed as a rotary positive-displacement motor, for example, possibly as a toothed gear motor or toothed ring motor.

With regard to the prefix “hydro” (see hydromotor, hydrostatic pump, and hydrodynamic pump mentioned above) employed herein, this prefix does not exclude the possibility of a gaseous pressure medium and/or operating medium. Thus, in general, “pressure medium” designates a fluid under pressure and “operating medium” designates an operating fluid. Accordingly, the prefix “hydro” may also denote fluids. Therefore, instead of a “hydromotor” and a “hydrodynamic pump”, it would also be possible to speak of a “fluid motor” and a “fluid-dynamic pump”. Generally, however, the pressure medium and operating medium are liquid and can then also be designated as “pressure liquid” and “operating liquid.”

The rotor interacting with the secondary medium and/or the rotor of the hydromotor can comprise a disk and/or a shaft with a flow geometry arrangement, possibly a blade arrangement. Diverse designs of the swirl pump or centrifugal pump are possible. For example, the swirl pump or centrifugal pump can be constructed in such a way that the pumped secondary medium flows against the rotor in a substantially axial or radial-tangential direction and flows off the rotor in a substantially axial or radial-tangential direction. This applies in a corresponding manner to the hydromotor. The latter can be constructed in such a way that the primary medium flows against the rotor in a substantially axial or radial-tangential direction and flows off the rotor in a substantially axial or radial-tangential direction.

It is advantageous for the actual pump as well as for the hydromotor when a plurality of flow-in channels or flow-off channels which are arranged in a substantially rotational-symmetric manner with respect to the rotor are provided in a housing accommodating the rotor for essentially radial-tangential flow against the rotor in question and/or for essentially radial-tangential flow away from the respective rotor. Bearings associated with the rotor then need only absorb comparatively small radial forces and a comparatively low-friction running of the rotor is ensured.

An advantageous construction is characterized in that at least one rotor interacting with the primary medium and at least one rotor interacting with the secondary medium have rotational axes which are essentially coaxial with respect to one another. In this respect, it is suggested by way of further development that at least one operating medium flow is guided substantially in axial direction at least in some areas as an annular flow surrounding at least a pressure medium flow and/or surrounding the rotor interacting with the pressure medium flow. The pressure medium flow and the operating medium flow can pass into a common medium flow which is fed to the clutch device as an operating medium flow, for example. The combination of a pressure medium flow and the operating medium flow and the supply of the combined medium flow to the clutch device, particularly to the clutch arrangement, can also be provided in the other designs described above, and indeed for all of the aspects of the present invention discussed. In this way, it can be achieved, for example, that a certain minimum supply of operating medium (operating liquid, particularly cooling oil) to the clutch device is ensured in case of a viscous operating medium at lower temperatures.

An advantageous embodiment form is characterized in that the secondary pump arrangement has at least two rotors which interact with the secondary medium and which are not arranged coaxial to one another and/or in that the hydromotor has at least two rotors which interact with the primary medium and which are not arranged coaxial to one another. This type of construction is particularly advisable for a secondary pump arrangement operating according to the principle of positive displacement or a hydromotor operating according to the principle of positive displacement.

With respect to the hydromotor, it is suggested by way of a further development that a drive formation, possibly a toothing, of a first rotor of the hydromotor meshes with a driving formation, possibly a toothing, of a second rotor of the hydromotor. With respect to the secondary pump arrangement, it is suggested by way of a further development that a conveying formation, possibly a toothing, of a first rotor of the secondary pump arrangement meshes with a conveying formation, possibly a toothing, of a second rotor of the secondary pump arrangement.

It will often be advisable that at least one rotor of the hydromotor interacting with the primary medium and at least one rotor of the secondary pump arrangement interacting with the secondary medium are coupled with one another by means of at least one common shaft which is shared by at least one rotor of the hydromotor and at least one rotor of the secondary pump arrangement. However, this is in no way mandatory. It can also be provided that a rotor of the hydromotor and a rotor of the secondary pump arrangement are arranged on a common shaft, while another rotor of the hydromotor and another rotor of the secondary pump arrangement are arranged on their own respective shafts which need not necessarily be rotationally coupled with one another as long as there is a coupling by means of the first rotor.

The clutch device (and the clutch arrangement) can be self-priming with respect to the operating medium. In this case, it is advantageous when the second pump arrangement allows operating medium which is sucked in by the clutch device (or clutch arrangement) to pass to the latter in a non-operational state. This makes possible an emergency operation, for example, when the second pump arrangement is defective. Also, it could be conceivable under certain circumstances to put the second pump arrangement out of operation, that is, to put it into the non-operational state, for certain operating states of the clutch system or drivetrain in order to save energy.

It has already been mentioned several times that the clutch arrangement can be actuated through the intermediary of pressure medium. It has already been made clear that it is intended in particular that the clutch arrangement can be actuated by means of pressure medium supplied by the pump arrangement according to the first aspect of the invention or by the first pump arrangement according to the third aspect of the present invention. For this purpose, the pump arrangement or the first pump arrangement is designed to supply pressure medium at a pressure sufficient for actuating the clutch arrangement.

It is noted for the sake of clarity that the clutch device according to the second and fourth aspects of the invention can also be a clutch device which is provided for an operation under the influence of the operating medium supplied by the pump arrangement or by the second pump arrangement.

It has already been made clear that the clutch arrangement is intended above all to be a wet clutch arrangement, that the operation is a wet operation under the influence of the operating medium, and that the operating medium is an operating liquid, possibly a cooling liquid. Of the possible wet clutch arrangements, a multi-disk clutch arrangement is intended above all.

The pressure medium can be a hydraulic pressure medium, particularly a hydraulic oil which may also serve as a cooling liquid.

The clutch device can be a multiple clutch device with a plurality of clutch arrangements, for example, a dual clutch device with two clutch arrangements assigned to a transmission input shaft. In this connection, reference is had particularly to the embodiments disclosed in the present Applicant's patent applications cited above.

It can be advisable to construct the clutch system in such a way that operating medium can be supplied to the clutch arrangement through the intermediary of an operating medium storage which is connected or can be connected to the operating medium supply arrangement or to the pump arrangement or to the second pump arrangement. Further, an advantageous construction can consist in that the clutch arrangement can be actuated through the intermediary of a pressure medium storage which is connected or can be connected to the pump arrangement or to the pressure medium supply arrangement or to the first pump arrangement.

DETAILED DESCRIPTION OF THE PRESENTLY PREFERRED EMBODIMENTS

FIG. 1is a schematic view showing a clutch system200which has a wet dual clutch202with a first, radially inner clutch device204and a second, radially outer clutch device206. The clutch devices204and206are wet clutch devices, for example, wet multi-disk clutch devices which have, as is known in the art, have at least one disk stack, are arranged one upon the other radially in the present embodiment example, and are actuated in each instance by an associated actuating piston of a hydraulic slave cylinder integrated in the dual clutch. Examples of dual clutches of this type are disclosed in the patents cited hereinabove.

A basic problem in clutches of this kind is that a high pressure is generally required at a small volume flow for actuating the clutch devices204,206, especially the disk stacks. In contrast, a low pressure at a high volume flow is required for wet operation of the clutch devices204,206, particularly for cooling the disk stacks. In order to be able to meet both of these conflicting requirements, two different, mutually independent pumps can be provided. However, due to the high space requirement this creates problems with respect to design and also economically because of the comparatively high cost, depending on the type and drive of the pumps.

In contrast, a pump208is provided in the arrangement according toFIG. 1which is driven by an electric motor210and sucks hydraulic medium (particularly hydraulic oil) from a hydraulic reservoir212. This hydraulic medium is conveyed further to valves214,216and218.

The pump208generates a sufficiently high pressure to dependably actuate the clutch devices204,206, so that the associated clutch device is actuated, i.e.—engaged in case of a “normally open” clutch according to the valve state of the valve in question or disengaged in case of a “normally closed” clutch according to the valve state of the respective valve—as soon as one of the valves214and216is switched to open.

The valve218, together with component P which is designated by reference number220, serves to convert the pressure at the input side of valve218at a small volume flow into a low pressure with a large volume flow. Accordingly, the subassembly formed by the valve218and component P220has pressure-converting characteristics. In order to supply a larger volume flow of operating medium, in this case cooling oil, component P220is connected to an oil reservoir222. It is not compulsory that the reservoir222is a separate reservoir from reservoir212. The component P220provides at least one operating medium flow, in this case a flow of cooling oil, which is conducted to the clutch devices204,206, particularly to their disk stacks in case of the present example.

Therefore, component P220or the subassembly formed by component P220and valve218is considered as a pump arrangement. Embodiments of devices that can be used as component P220are presented below. To the extent that these devices are designated as “pump” or “pump arrangement” herein, this in no way confines or limits the subject matter of the invention. However, as is easily recognized, component P220may be at least one pump or one pump arrangement corresponding to one embodiment of the invention. A further embodiment of the invention is directed to the pressure conversion of the comparatively high clutch actuation pressure supplied by the pump208into a lower pressure with a greater volume flow for the clutch cooling, which pressure conversion is achieved according to the embodiment ofFIG. 1.

FIGS. 2aand2bshow a piston device230which can be used as component P220and which has a piston234which is mounted in a housing232so as to be displaceable along an axis A. The piston device230has a connection236at which the piston device230serving as component P220is connected to the valve218. The piston device230is connected to the reservoir222via another connection238. The dual clutch202is supplied with the cooling oil via a third connection240.

FIGS. 2aand2bshow the piston234and a valve element242cooperating with the latter in two different positions, one of which is shown which isFIG. 2awhich is in the half-plane above the axis A and the other isFIG. 2bwhich is shown in the half-plane below the axis A. InFIG. 2a, the position of the piston and of the valve element is occupied by the piston234and the valve element242at the end of a stroke of the piston234which is carried out under the action of a return spring244and which can be identified as the “suction stroke” and of a stroke of the valve element242which moves along with the piston234. A pressure reduction valve246which is formed by the piston234and the valve element242is closed and an interior space250of the piston device is filled with cooling oil that is sucked in via the connection238and a non-return valve252. The piston234and the valve element242are pretensioned or biased with respect to one another in the opening direction of the pressure reduction valve246by a valve spring254. However, the piston234and the valve element242occupy the closed position since the valve element242is pushed by a valve shaft at a housing portion and the spring force of the return spring244is sufficient to displace the piston234against the action of the spring254into a stop position in which the piston234stops at the housing by a piston portion extending in radial direction and closes the pressure reduction valve246in cooperation with the valve element242. This is illustrated inFIG. 2b. The return spring244which preferably acts directly at the above-mentioned piston portion is supported at a closure plate258which closes the interior space250and which is held at the housing232by a snap ring256.

The piston position and valve position shown inFIG. 2aare occupied when a comparatively low pressure prevails in interior space260of the piston device230, so that the return spring244can move the piston234and the valve element242into the stop position shown inFIG. 2b. The valve218is then closed. When pressure oil supplied via connection236by the pump208is conducted into the interior space260at comparatively high pressure by switching the valve218which, in the simplest case, is constructed as a lift valve, this pressure oil acts on a first piston surface262of the piston and an associated pressure receiving surface264of the valve element242. Since forces exerted by the pressure oil in the interior space260on the pressure receiving surface264are conducted away via the valve spring254and a contact of the piston234and valve element242against one another in a positive-locking connection closing the pressure reduction valve246, the actual piston surface or pressure receiving surface262of the piston on the one hand and the pressure receiving surface264of the valve element242on the other hand can be taken into account jointly in the form of an individual effective piston surface associated with the pressure oil.

The pressure oil in the interior space260displaces the piston234and the valve element242together in the direction of the closure plate258against the action of the return spring244. In so doing, cooling oil is displaced from the interior space250via a check valve266and the connection240in the direction of the dual clutch202. Accordingly, the piston234executes a “discharge stroke.” This discharge stroke ends when the piston234strikes against a stop ring268which is held inside the interior space250at the housing232in the manner of a snap ring.

Shortly before the piston234stops against the stop ring268, the valve shaft of the valve element242strikes a stop pin270which is arranged at the closure plate258, so that the valve element242can not take part in the rest of the discharge stroke along with the piston234.

The pressure of the pressure oil in the interior space260is sufficient for the pressure oil acting on the first piston surface262to move the piston234into its stop position while releasing the valve element242until the piston234stops against the stop ring268. The pressure reduction valve246which was previously held in its closed position by the pressure oil in the interior space260against the action of the valve spring254accordingly opens so that pressure oil can flow out of the interior space260through a passage formed between the piston234and the web of the valve element242into the interior space250. This situation is shown inFIG. 2b; a passage can be seen between an annular flange having the pressure receiving surface264and an associated end portion of the piston234.

The piston position and valve position shown inFIG. 2aor a somewhat different position corresponding to a pressure and force equilibrium to be adjusted is occupied until pressure oil is subsequently delivered into the interior space260via the valve218and the connection236. When the valve218is closed, the returning movement of the piston234corresponding to the “suction stroke” and the returning movement of the valve element242into the position shown inFIG. 2ais initiated under the influence of the return spring244. The pressure reduction valve246remains open under the action of the valve spring254until the valve shaft of the valve element242strikes against the housing. Cooling oil is again sucked out of the reservoir222during this “suction stroke.”

The piston device230according toFIGS. 2aand2bhas pressure-converting characteristics. On the side of the interior space250, the piston has a second piston surface280which is appreciably larger than the first piston surface262or the effective piston surface formed by the first piston surface262and the pressure receiving surface264. Due to this, other effective piston surface262,264on the side of the pressure oil drawn from the pump208and the larger piston surface on the side of the cooling oil drawn from the reservoir222, a greater pressure in interior space260can balance out a small pressure in interior space250, and the forces exerted on the piston234by the return spring254are also taken into account in addition. In practical operation of the piston device230, an equilibrium state of this kind need not occur, the pressure in the interior space250can be substantially less than an equilibrium pressure corresponding to a pressure conversion according to the piston surfaces depending on the flow resistance through the check valve266, the connection240, an oil connection to the dual clutch206and an effective flow resistance of the dual clutch202and a return line into the reservoir222during the entire discharge stroke when the pressure reduction valve is closed.

The piston device230makes it possible that through the intermediary of the small piston surface262,264on the pressure oil side and by means of the high pressure of the pressure oil and through the intermediary of the large piston surface280on the cooling oil side at low pressure, a volume quantity conveyed for cooling the clutch is greater than the amount of pressure oil supplied to the piston device via the valve218. Therein, a pressure compensation opening282is provided in the housing238so that pressure compensation opening282is connected to the reservoir222in order to prevent the piston234from sticking in its stop position shown inFIG. 2adue to a vacuum.

The check valve252serves to prevent cooling oil from flowing out of the interior space250back into the reservoir222. Alternatively, this check valve252could also be formed in the piston as is indicated in dashed lines at252′. In this case, a connection of the piston device associated with the pressure compensation opening282will be sufficient and connection238could be dispensed with. The check valve266prevents cooling oil from the side of the dual clutch from flowing back into the interior space250.

The piston device230works to supply fresh oil continuously to the dual clutch202with reciprocating pistons and consequently with a pulsed delivery of fresh oil to the dual clutch202. Therefore, the piston device230may also be called a pulse delivery pump. Vibrations can occur under certain circumstances due to the oscillating motion of the piston. In order to remedy this, two piston devices230can be operated in parallel and in opposite directions, that is, with a phase of the piston stroke that is offset by 180°, approximately in the manner shown inFIGS. 2aand2b. Of course, both piston230devices can be constructed with an individual housing shared by both devices230and with common connections. It is also possible for more than two piston devices230to be connected in parallel and operated with mutually offset stroke phases for a more uniform supply of fresh oil.

FIG. 3shows a rotor device300which can be used as component P 220 in the embodiment ofFIG. 1. The rotor device300has at least one pressure medium connection302which is connected to the pump208either via the valve218or directly to a cooling connection, not shown in the drawing, which is connected to the reservoir222and from which cooling oil can flow via a check valve304in a housing306of the rotor device300which is constructed in multiple parts into a conveying rotor308having at least one conveying geometry, for example, a vane or blade shape.

Pressure oil flowing into the housing306via the connection302flows against at least one drive rotor310which is constructed with at least one pulse geometry or rotational pulse receiving geometry, for example, a vane shape, and is set in rotation by the pressure oil.

FIG. 4shows a possible construction of the drive rotor310in detail. A plurality of drive blades312is shown. The pressure oil which is directed through a supply channel314to the drive blade312sets the drive rotor310in rotation in the direction indicated by arcuate arrow315. After interacting with the rotor310, the pressure oil which has now given off a portion of its energy flows off via a flow-off channel316.

The conveying rotor308is connected with the drive rotor310so as to be fixed with respect to rotation relative to it.FIG. 6andFIG. 7illustrate embodiments of conveying structures, which in this case are blade-shaped, herein conveying blades318. In the embodiment, the drive blades312of the drive rotor310and the conveying blades318of the conveying rotor308are arranged at a rotating disk320which is mounted radially and axially in the housing306. The disk320on the one hand and the drive rotor310formed by the drive blades312and a hub portion322and the conveying rotor308formed by the conveying blades318on the other hand can be constructed in one part or multiple parts. For example, it could possibly be produced as a plastic part in one piece.

According to the preceding, the conveying rotor308is driven in rotation by the pressure oil flowing against the drive rotor310via the channel314through the intermediary of the drive rotor310. The rotation of the conveying rotor308generates a cooling oil flow from the reservoir222via the associated connection of the rotor device, the check valve304to the conveying rotor308and the cooling oil interacting with the conveying rotor308is guided via a flow-off channel330and a connection of the piston device associated with the dual clutch202to the dual clutch302, or more exactly to its clutch devices204and206. The conveying effect of the conveying rotor308relies on torsion imparted to the cooling oil and/or on the action of centrifugal force of the cooling oil accelerated through the intermediary of the conveying rotor308, so that the rotor device300can also be designated as a swirl pump; specifically, as an impeller pump or centrifugal pump.

The flow-off channel316guiding the pressure oil opens at332into the supply channel330, so that the pressure oil is conveyed together with the cooling oil to the dual clutch202. In the embodiments ofFIGS. 6 and 7, the pressure oil supplied by the pump208on the one hand and the cooling oil conveyed from the reservoir222to the dual clutch202on the other hand are the same medium, that is, a hydraulic oil.

The conveying structures, e.g. conveying blades318, are shaped in such a way that large volume flows can be conveyed through the dual clutch202during a rotational movement of the conveying rotor308and cooling oil can flow from the reservoir222through the rotor device300to the dual clutch202when the conveying rotor308is stationary in the case of a self-priming clutch.

In one embodiment, pressure oil flows against the drive blades312of the drive rotor310radially and tangentially and the pressure oil flows off from the drive blades312tangentially or radially. The cooling oil that is sucked in flows against the blade geometries318of the conveying rotor308axially and the cooling oil flows off in radial direction. Other flow conditions are also possible. For example, flow can proceed against the drive rotor310axially.

The comparatively costly sliding bearing or rolling bearing for the bearing support of the rotors308,310can be dispensed with when a lubricating film bearing is provided. The rotor device300can be constructed in such a way that a corresponding lubricating film is formed compulsorily based on the rotational movement, e.g., based on a hydrodynamic effect. However, supply openings can be provided through which small amounts of pressure oil and/or cooling oil are conducted to the lubrication points. An opening of this kind is shown at334.

The rotating disk320serving to support the rotors308,310can be provided at its circumference and/or at its axial end faces with a friction-reducing coating, e.g., PTFE, for a particularly low-friction bearing support. Accordingly, the transition from adhesive friction to sliding friction in particular is accelerated.

The rotational movement imparted to the conveying rotor308is represented inFIGS. 6 and 7by arcuate arrow340. The circular line342represents the outer circumference of the rotating disk320. The dashed circular lines344ofFIGS. 5 and 346ofFIGS. 6 and 7do not represent design details of the rotor device300, but rather correspond to the rotational movement path of the radial outer blade ends of the drive rotor and conveying rotor. The circular line348represents the inner circumference350of a cooling oil channel352in the interior of the housing306which guides cooling oil from the non-return valve304to the conveying rotor308. The radial position of this inner circumference is also shown inFIG. 7by a dashed circular line352. The straight arrows354show the flow direction of the pressure oil to the drive rotor310and away from the latter. The straight arrows356show the flow direction of the cooling oil to the conveying rotor308and away from the latter.

In order to ensure that the drive rotor310starts with the least friction possible, flow can proceed against the drive rotor310or its drive blades312in a rotationally symmetric manner. In the embodiment ofFIG. 5, two supply channels314which are offset by about 180° and two flow-off channels316which are offset relative to one another by 180° are formed in the housing306. For example, three supply channels which are offset by 120°, four supply channels which are offset by 90° relative to one another, etc., could be provided.

The function of the check valve304is referred to in the following. The check valve304generally ensures that no oil flows out of the rotor device300into the reservoir (compensation vessel)222. In particular, pressure oil flowing via the connection302to the drive rotor310and from the latter via the flow-off channel316into the channel360is prevented from flowing off into the reservoir222. The fact is that at low temperatures the cooling oil in the reservoir222can be quite viscous and, therefore, under some circumstances, e.g., after the motor vehicle has stood for a long time, can not be sucked in immediately in sufficient quantities. The pressure oil flowing into the channel360then ensures a minimum supply of cooling medium to the clutch device.

Another rotor device400which can be used as component P (220) in the embodiment ofFIG. 1is shown in a perspective view inFIG. 8. The rotor device400has a rotating shaft404which is rotatably mounted in a housing402. The rotating shaft404has at least one drive rotor410with pulse or rotary pulse receiving structures, in this case especially blades412, and at least one conveying rotor408with conveying structures, such as blades418. The drive rotor410is associated with a high-pressure turbine drive and pressure oil flows against it axially as is indicated by the arrow420. Referring to the embodiment ofFIG. 1, this pressure oil having a relatively high pressure is supplied by the pump208which is connected via the valve218or directly to a connection of the rotor device400associated with the drive rotor410.

The housing402has a radial inner cylinder421and a radial outer cylinder422. The inner cylinder421receives the drive rotor410and guides the pressure oil in an axial flow to the drive rotor410. Between the outer cylinder422and the inner cylinder421, a cooling oil pressure flow represented by arrow424is guided axially to the conveying rotor408which is arranged at the other side of an axial end of the inner cylinder421and whose blades418extend farther radially outward than the outer circumference of the inner cylinder421. The pressure oil which has passed the drive rotor410exits from the inner cylinder421and flows together with the cooling oil in the region of the conveying rotor408.

In addition to the rotors408and410, which are shown inFIG. 8and which can also be referred to as impellers, conducting devices that are stationary with respect to the housing402can be provided in order to make efficient use of the kinetic energy of the pressure oil for the driving of the drive rotor410and therefore (via shaft404) for the driving of the conveying rotor408and in order to convey the cooling oil through the intermediary of the conveying rotor408in an efficient manner. Therein, the conveying effect of the rotating conveying rotor408is based on the fact that angular momentum or swirl is imparted to the pressure oil. Accordingly, one can say that the conveying rotor408forms a swirl conveying pump.

In a corresponding manner, as in the case of the rotor device300, the radial bearing and axial bearing supporting the rotors408,410, that is, the radial bearing and axial bearing supporting the shaft404, can provide a lubricating film bearing which compulsorily forms during the rotational movement, particularly in a hydrodynamic manner. A compulsory lubrication can also be provided by supplying pressure oil and/or cooling oil. Also, for the embodiment ofFIG. 8, a coating, e.g., of the bearing portions of the rotating shaft404, for example, a PTFE coating, is advantageous for reducing friction so that the mixed friction region can be traversed more quickly when accelerating.

When the conveying rotor408is stopped, conveying oil which is sucked in through a self-priming dual clutch can flow through the blade geometries418of the conveying rotor408without excessive flow resistance.

The embodiments ofFIGS. 2aand2band those ofFIGS. 3–8have important differences. With regard to the conveying of the cooling oil, the piston device230works as a hydrostatic pump or, viewed in another way, as a positive-displacement machine. In contrast, the rotor devices300and400which can be designated as swirl conveying pumps with respect to the conveying of cooling oil operate as hydrodynamic pumps or, viewed in another way, as liquid flow machines. Such hydrodynamic pumps or flow machines are particularly well suited for generating high volume flows, whereas hydrostatic pumps or positive-displacement machines, especially piston pumps or piston machines, are not necessarily the first choice for generating higher volume flows. Since, on the other hand, hydrostatic pumps or positive-displacement machines, particularly piston pumps, are generally more suitable for generating high pressures than are hydrodynamic pumps or flow machines, the pump208supplying the pressure oil is constructed as a hydrostatic pump or positive-displacement machine in a particularly preferred construction of the embodiment ofFIG. 1.

In the embodiments described above, a comparatively large flow of cooling medium, particularly cooling oil, which is at comparatively low pressure is generated through the intermediary of a pressure medium, particularly pressure oil, having a comparatively high pressure and through the intermediary of a comparatively small volume flow of this pressure medium. Conversely, a (possibly comparatively smaller) pressure medium flow, particularly a flow of pressure oil, can also be generated with comparatively high pressure through the intermediary of a cooling medium flow, particularly a flow of cooling oil at comparatively low pressure by using a corresponding pressure converter.FIG. 9shows an embodiment for a clutch system according to the present invention. The pump208ais driven by the electric motor210aand is designed for supplying a large-volume flow of cooling oil for cooling the clutch arrangements204aand206aof the dual clutch202a, wherein the pump208agenerates a cooling oil pressure that would not be sufficient in terms of pressure to actuate the clutch devices204aand206a. A pressure converter220a, component P, is connected to the cooling oil pump208a, possibly via a valve218a, and supplies pressure oil at a pressure level sufficient for actuating the clutch devices204aand206aunder the influence of the cooling oil. For this purpose, the clutch devices are connected to component P220avia the valves214aand216a. Component P220acan be connected to the reservoir112aby means of a separate line in order to suck hydraulic oil out of the latter and/or to deliver hydraulic oil to the reservoir depending on the construction and function of component P220a. In one embodiment, component P220acan be constructed as a positive-displacement machine or flow machine, possibly a hydrostatic pump or hydrodynamic pump, in a manner analogous to that discussed with reference to the piston device230and the rotor devices300and400, wherein, however, a small input pressure is converted to a high output pressure. In one embodiment, component P220ais a piston device similar to piston device230, wherein a large piston surface is associated with the cooling oil and a small piston surface is associated with the pressure oil to be put under pressure.

Since a comparatively large volume flow is required for cooling the dual clutch, the pump208acan advantageously be constructed as a hydrodynamic pump.

FIG. 10shows a piston device500which can be used as component P220ain the embodiment ofFIG. 9and which is connected, for example, to the valve218aof the embodiment ofFIG. 9. Analogous to the description of the piston device230ofFIG. 2in the following is a description of piston device500, wherein only the differences between the piston devices will be emphasized.

The piston device500is connected via its connection536and valve218to the cooling oil pump208a. The valve218aaccordingly lets cooling oil into the interior space560corresponding to the pressure generated by the pump208afor bringing about a “discharge stroke” against the action of the return spring544and lets off cooling oil contained in the interior space560in the reservoir212ato reduce the pressure in the interior space560in order to initiate or permit a “suction stroke” under the action of the return spring544. During the suction stroke, hydraulic oil is sucked out of one of the reservoirs212aand222a, for example, into the interior space550via the connection538and the check valve552. During the discharge stroke, corresponding to the axial displacement of the piston534to the right, hydraulic oil is supplied as pressure oil from the interior space550via the check valve566and the connection540to an actuating pressure oil circuit which can be constructed as a pressure oil reservoir in order to supply a uniform pressure level for the clutch actuation. A reservoir221aof this kind is shown in dashed lines inFIG. 9as one embodiment of the clutch system. This reservoir221acan easily be associated with respect to construction and/or function with component P220a, e.g., integrated into component P220a.

Piston534has a first piston surface562which is exposed to the cooling oil in the interior space560and which is appreciably larger than the second piston surface580which is arranged at the other end of the piston and which is exposed to the hydraulic oil that is contained in the interior space550and is to be supplied for clutch actuation. Therefore, also taking into account the axial forces exerted on the piston surface562by the return spring544, the comparatively lower pressure in the interior space560can be converted into a comparatively high pressure in interior space550, so that the pressure oil flowing off via the connection540is at a pressure level which is sufficient for reliable actuation of one or both clutch arrangements204aand206b.

FIG. 11illustrates one embodiment of a clutch system of another approach and which has two mutually independent pumps. A first pump208band a second pump209bare provided and are driven by an electric motor210band211b, respectively. The first pump208bwhich is driven by an electric motor210band which is preferably constructed as a hydrostatic pump or positive-displacement machine supplies pressure medium, particularly pressure oil, at a comparatively high pressure which is sufficient to actuate the clutch devices204band206bof the dual clutch202b. For selective actuation of the clutch devices, these clutch devices are connected to the pump208bvia an associated valve214bor216b.

The second pump209bwhich is driven by an electric motor211band which is preferably constructed as a hydrodynamic pump or turbine pump supplies a comparatively large volume flow of cooling medium, particularly cooling oil, which is sufficient for cooling the clutch devices204band206b. The pressure delivered by the pump209bcan be appreciably less than the pressure delivered by the pump208b.

FIG. 12shows one embodiment with two pumps which are driven independently, preferably by electric motors. The reference numbers inFIG. 11are also used inFIG. 12insofar as the various components of the clutch system ofFIG. 12correspond to the components of the arrangement according toFIG. 11.

Fundamentally, the cooling of the dual clutch202brequires a high volume flow at comparatively low pressure, the electrically driven pump209bis preferably a hydrodynamic pump, for example, an impeller pump. Since the clutch arrangements must generally only be cooled very intensively when stopping or starting, it is sensible to construct and control this pump209bin such a way that the rotational speed is controlled or regulated corresponding to the cooling requirement as is indicated inFIG. 12by a pump symbol with an arrow. Very high conveying flows can be achieved with a hydrodynamic pump, particularly an impeller pump. For example, in a conventional impeller pump the conveying flow increases in proportion to the square of the pumping speed.

InFIG. 12, the cooling oil is supplied to the dual clutch202bvia a heat exchanger600, since a noticeable increase in temperature can come about also in the oil in the oil pan212b, for example, in case of a long slip operation. The oil temperature is maintained at a temperature level sufficient for cooling the dual clutch202bby means of the heat exchanger600. Since the cooling oil can be quite viscous at lower temperatures and in certain cases cooling oil would no longer reach the dual clutch202bin sufficient quantities because of the flow resistance of the heat exchanger600at especially low temperatures, a bypass valve602which is pretensioned by a spring, for example, is provided which opens when the cooling oil pressure downstream of the oil cooler600exceeds a predetermined pressure threshold and lets the cooling oil past the oil cooler600to the dual clutch.

As previously stated, a comparatively small volume flow at relatively high pressure is needed to actuate the actuation piston of the two clutch devices204band206b. Consequently, the pump208bwhich is driven by electric motor210bis preferably a hydrostatic pump, e.g., a toothed gear pump or vane cell pump.

In the embodiment example inFIG. 12, a pressure oil storage604having a gas cushion under pressure is installed in the clutch actuation pressure oil circuit and is charged by the pump208bvia a check valve606and is connected via valve214band valve216bto the actuation slave cylinders of the two clutch devices204band206b. The pressure oil storage604provides for a uniform pressure level which is particularly advisable in case the pump208bis constructed as a piston pump and allows a pump with a particularly small conveying volume to be used as pump208b. Thus, the oil volume delivered by the pump208bper time unit can be less than the pressure oil volume required per time unit during actuation of the dual clutch202b.

The pressure oil circuit between the check valve606and the valves214band216bis protected against excessively high pressure of the pressure oil which could possibly result in damage by means of a pressure limiting valve608. The pressure determined by the fluid level of the storage604is detected by a pressure sensor610in this pressure oil circuit.

Another pressure limiting valve612ensures that the pressure prevailing on the other side of the valves214band216band acting on the hydraulic slave cylinders of the clutch devices does not exceed a maximum value, for example, in order to prevent possible damage. Due to two check valves615and616, one pressure limiting valve is sufficient for monitoring the actuation pressure of two hydraulic slave cylinders.

When the pressure of the cooling oil is not sufficient to cool the clutch at low temperatures, that is, in case of high oil viscosity, for example because the hydrodynamic pump used for pump209bcan not generate sufficient pressure, a valve614is provided in the embodiment according toFIG. 12, through which valve614a small volume flow can be branched off from the cooling oil flow supplied by the pump208bin order to provide a kind of “emergency cooling” of the clutch disks when required. Since the high viscosity of the cooling oil which makes it necessary to open the valve614occurs only at low temperatures at which there is only a low cooling requirement for the dual clutch in any case, a relatively small “emergency cooling oil flow” is sufficient. Moreover, this emergency cooling is only required until the temperature of the oil and, accordingly, the viscosity of the oil is sufficient to ensure adequate pumping output of the cooling oil pump209b. Instead of a valve614, a diaphragm, choke, or the like could also be provided by which a small volume flow is continuously branched off from the pressure oil flow supplied by the pump208bin the cooling circuit. When using the valve614which branches off the cooling oil when needed, the pump208bmay be operated briefly in overload operation in order to supply sufficient cooling oil within the short time period before the oil is sufficiently heated. Since the time periods when this is required are generally only very short, the life of the pump208bis not substantially shortened in this way.

FIG. 13shows a dual clutch12arranged in a drivetrain10between a drive unit and a transmission. The only part of the drive unit, e.g., an internal combustion engine, shown inFIG. 13is a driven shaft14, possibly a crankshaft14, with a coupling end16serving to couple a torsional vibration damper, not shown. The transmission is represented inFIG. 13by a transmission housing portion20defining a transmission housing casing18and two transmission input shafts22and24, both of which are constructed as hollow shafts. Transmission input shaft22extends through transmission input shaft24essentially coaxial thereto. A pump drive shaft which serves to drive an oil pump on the transmission side, not shown inFIG. 13, is arranged in the interior of the transmission input shaft22as will be discussed further. When at least one oil pump driven by electric motor is provided, the pump drive shaft can be dispensed with.

The dual clutch12is received in the transmission housing casing18, wherein the interior of the casing is closed in the direction of the drive unit by a cover28which is pressed into a casing housing opening and/or is secured therein by a snap ring30. When the dual clutch12has wet friction clutches, for example, disk clutches, as in the embodiment shown inFIG. 13, it is generally arranged so as to ensure a tight engagement between the cover28and the clutch housing formed by the transmission housing casing18. This tight engagement can be produced, for example, by means of an O-ring or some other sealing ring.FIG. 13shows a sealing ring32with two sealing lips.

A clutch hub34comprising two annular portions36,38secured to one another for reasons which will be explained in the following serves as input side of the dual clutch12. The clutch hub34extends through a central opening of the cover28in the direction of the drive unit and is coupled via an external toothing42with the torsional vibration damper, not shown, so that there is a torque-transmitting connection between the coupling end16of the crankshaft14and the coupling hub34by means of this torsional vibration damper. If a torsional vibration damper is not desired at this location in the drivetrain, or at all, the clutch hub34can also be coupled directly to the coupling end16. The pump drive shaft26has an external toothing44at the end remote of the transmission which engages in an internal toothing46of the annular portion36of the clutch hub34, so that the pump drive shaft26rotates along with the clutch hub34and accordingly drives the oil pump when a rotational movement is imparted to the clutch hub34, generally by the drive unit and in many operating situations possibly also by the transmission via the dual clutch12(for example, in an operating situation characterized by the term “engine braking”).

The cover28extends radially between an annular circumferential wall portion of the housing casing18defining a radial recess50of the housing casing18and the annular portion38of the hub34. It is advantageous when a seal arrangement and/or pivot bearing arrangement54is provided between a radial inner wall area52of the cover28and the hub34, especially the annular portion38, especially when—as in the embodiment illustrated the cover28is secured to the housing casing18and therefore does not rotate along with the dual clutch12. A seal between the cover and the hub is required particularly when the clutch arrangements of the dual clutch are wet clutches as in the embodiment example. Highly reliable operation is also achieved in case of occurring oscillations and vibrations when the seal arrangement and/or pivot bearing arrangement54are/is secured axially at the cover28and/or at the clutch hub34, for example, by an end portion of the cover edge of radial inner wall area52that is bent radially inward, as is shown inFIG. 13.

A support plate60which serves to transmit torque between the hub34and an outer disk carrier62of a first multi-disk clutch arrangement64is arranged at the annular portion38of the hub34so as to be fixed with respect to rotation relative to it. Referring to the embodiments ofFIGS. 1,9,11and12, the multi-disk clutch arrangement64can be identified, for example, as clutch arrangement204,204aor204b. The outer disk carrier62extends in the direction of the transmission and radially inward to an annular part66at which the outer disk carrier62is arranged so as to be fixed with respect to rotation relative to it and which is rotatably supported at the two transmission input shafts22and24by means of an axial and radial bearing arrangement68in such a way that radial as well as axial forces are supported at the transmission input shafts. The axial and radial bearing arrangement68makes possible a relative rotation between the annular part66on the one hand and transmission input shaft22and transmission input shaft24on the other hand. The construction and operation of the axial and radial bearing arrangement68will be discussed more fully later on.

An outer disk carrier70of a second multi-disk clutch arrangement72is arranged at the annular part66farther axially in the direction of the drive unit so as to be fixed with respect to rotation relative to it, its disk stack74being surrounded annularly by the disk stack76of the first multi-disk clutch arrangement.

Referring to the embodiments ofFIGS. 1,9,11and12, the second multi-disk clutch arrangement72can be identified, for example, as clutch arrangement206,206aor206b. As was already indicated, the two outer disk carriers62and70are connected with one another by the annular part66so as to be fixed with respect to relative rotation and are jointly in a torque-transmitting connection with the clutch hub34and therefore—via the torsional vibration damper, not shown—with the crankshaft14of the drive unit by means of the carrier plate60which is in a positive-locking torque-transmitting engagement with the outer disk carrier62by means of an external toothing.

Referring to the normal flow of torque from the drive unit to the transmission, the outer disk carriers62and70serve as input side of the multi-disk clutch arrangement64and72, respectively.

A hub part80of an inner disk carrier82of the first multi-disk clutch arrangement64is arranged on the transmission input shaft22so as to be fixed with respect to rotation relative to it by means of a spline or the like. In a corresponding manner, a hub part84of an inner disk carrier86of the second multi-disk clutch arrangement72is arranged on the radial outer transmission input shaft24so as to be fixed with respect to rotation relative to it by means of a spline or the like. Referring to the regulating torque flow from the drive unit in the direction of the transmission, the inner disk carriers82and86serve as output side of the first and second multi-disk clutch arrangements64and72.

Referring again to the radial and axial bearing support of the annular part66at the transmission input shafts22and24, two radial bearing subassemblies90and92acting between the radial outer transmission input shaft24and the annular part66serve as radial support of the annular part66. With regard to a support in the direction of the drive unit, the axial bearing support of the annular part66is carried out by means of hub part84, an axial bearing94, hub part80and a snap ring96which secures the hub part80axially to the radial inner transmission input shaft22. The annular part38of the clutch hub34is supported in turn via an axial bearing68and a radial bearing100at the hub part80. The hub part80is supported axially in the direction of the transmission at an end portion of the radial outer transmission input shaft24by the axial bearing94. Hub part84can be supported at the transmission input shaft24directly at an annular stop or the like or at a separate snap ring or the like in the direction of the transmission. Since the hub part84and the annular part66are rotatable relative to one another, an axial bearing can be provided between these components insofar as the bearing92does not function as both axial bearing and radial bearing. The latter is assumed with regard to the embodiment ofFIG. 13.

Great advantages result when, as in the present embodiment, the portions of the outer disk carriers62and70extending in radial direction are arranged on one axial side of a radial plane extending orthogonal to an axis A of the dual clutch12and the portions of the inner disk carriers82and86of the two multi-disk clutch arrangements64,72extending in radial direction are arranged on the other axial side of this radial plane. This allows a particularly compact construction, particularly when—as in the present embodiment—disk carriers of one type (outer disk carrier as in the embodiment example or inner disk carrier) are connected with one another so as to be fixed with respect to relative rotation and serve in each instance as the input side of the respective multi-disk clutch arrangement64,72with respect to the flow of force from the drive unit to the transmission.

Actuation pistons for actuating the multi-disk clutch arrangements64,72are integrated in the dual clutch12in order to actuate the multi-disk clutch arrangements64,72for engagement in the case of the present embodiment. An actuation piston110associated with the first multi-disk clutch arrangement64is arranged axially between the radially extending portion of the outer disk carrier62of the first multi-disk clutch arrangement64and the radially extending portion of the outer disk carrier70of the second multi-disk clutch arrangement72and is guided so as to be axially displaceable at both outer disk carriers and at the annular portion66by means of seals112,114,116and so as to seal a pressure chamber118formed between the outer disk carrier62and the actuation piston110and a centrifugal force pressure compensation chamber120formed between the actuation piston110and the outer disk carrier70. The pressure chamber118communicates, via a pressure medium channel122formed in the annular portion66, with a pressure control device, such as a control valve (possibly valve214,214aor214b) which is connected to a pressure medium supply, possibly the oil pump already mentioned above or pump208or208bor component P (220a). The pressure medium channel122is connected to the pressure control device via a connection sleeve which receives the annular part66and which is possibly fixed with respect to the transmission. With regard to the annular part66, for a simpler manufacture particularly with respect to the pressure medium channel122and another pressure medium channel, this annular part66is produced in two parts with two sleeve-like ring part segments which are inserted one into the other as shown inFIG. 13.

An actuation piston130associated with the second multi-disk clutch arrangement72is arranged axially between the outer disk carrier70of the second multi-disk clutch arrangement72and a substantially radially extending wall part132arranged at an axial end region of the annular part66remote of the transmission so as to be fixed with respect to rotation relative to it and so as to be tight against fluid and is guided so as to be axially displaceable by means of seals134,136and138at the outer disk carrier70, the wall part132and the annular part66and so as to seal a pressure chamber140formed between the outer disk carrier70and the actuation piston130and a centrifugal force pressure compensation chamber142formed between the actuation piston130and the wall part132. The pressure chamber140is connected via another pressure medium channel144(already mentioned) in a manner corresponding to pressure chamber118at a/the pressure control device (possibly valve216,216aor216bin connection with the pump208or pump208bor component P220a). Pressure applied by the (respective) pressure medium source (possibly oil pump) can be applied to the two pressure chambers118and140selectively (possibly also simultaneously) by means of the pressure control device(s) in order to actuate the first multi-disk clutch arrangement64and/or the second multi-disk clutch arrangement72for purpose of engagement. Diaphragm springs146,148serve to restore, that is, to release, the dual clutches12; the diaphragm spring148associated with the actuation piston130is received in the centrifugal force pressure compensation chamber142.

The pressure chambers118and140are completely filled with pressure medium (in this case hydraulic oil) when during normal operating states of the dual clutch12and the actuating state of the multi-disk clutch arrangements64,72, per se, depends on the pressure of the pressure medium applied to the pressure chambers. However, since the outer disk carriers62and70, including the annular part66and the actuation piston110and130and the wall part132, rotate together with the crankshaft14in driving operation, pressure increases in the pressure chambers118and140which are caused by centrifugal force are brought about even when no pressure is applied to the pressure chambers on the part of the pressure control device, which increases could lead to an unwanted engagement or at least slippage of the multi-disk clutch arrangements at least at higher rotational speeds. The centrifugal force pressure compensation chambers120,142, which receive a pressure compensation medium and in which pressure increases occur correspondingly as a result of centrifugal force, which pressure increases compensate for the pressure increases caused in the pressure chambers by centrifugal force, are provided for this reason.

It is also possible to fill the centrifugal force pressure compensation chambers120and142permanently with pressure compensation medium, for example, oil, wherein, as the case may be, a volume compensation could be provided to absorb pressure compensation medium displaced in the course of actuating the actuation pistons. In the embodiment ofFIG. 13, the centrifugal force pressure compensation chambers120,142are first filled with pressure compensation medium during operation of the drivetrain, specifically in connection with the supply of cooling fluid, particularly cooling oil in the present embodiment, to the multi-disk clutch arrangements64and72via an annular channel150formed between the annular part66and the outer transmission input shaft24, with which the bearings90,92which pass the cooling oil are associated. The cooling oil which is supplied, by component P220or pump208aor209bflows from a transmission-side connection between the annular part and the transmission input shaft24in the direction of the drive unit through bearing90and bearing92and then flows in a partial flow between the end portion of the annular part66remote of the transmission and the hub part84radially outward in the direction of the disk stack74of the second multi-disk clutch arrangement72, enters the area of the disks of the disk stack79through the passages in the inner disk carrier86, flows radially outward between the disks of the disk stack74and through friction lining grooves, or the like, of these disks, enters the area of the disk stack76of the first multi-disk clutch arrangement64through passages in the outer disk carrier70and through passages in the inner disk carrier82, flows radially outward between the disks of disk stack76and through lining grooves, or the like, in these disks and then, finally, flows off radially outward through passages in the outer disk carrier62. The centrifugal force pressure compensation chambers120,142are also connected to the cooling oil supply between the annular part66and the transmission input shaft24, specifically by means of radial bore holes152,154in the annular part66. Due to the fact that the cooling oil serving as pressure compensation medium in the pressure compensation chambers120,142flows off out of the pressure compensation chambers due to the absence of centrifugal force when the drive unit is stationary, the pressure compensation chambers are refilled again during the operation of the drivetrain (of the motor vehicle).

Since a pressure application surface of the actuation piston130associated with the pressure chamber140is smaller and also does not extend as far radially outward as a pressure application surface of the piston130associated with the pressure compensation chamber142, at least one fluid level limiting opening156is formed in the wall part132and adjusts a maximum radial fluid level of the pressure compensation chamber142providing the required centrifugal force compensation. When the maximum fluid level is reached, the cooling oil supplied via the bore hole154flows off through the fluid level limiting opening156and combines with the cooling oil flowing radially outward between the annular part66and the hub part84. For the piston110, the pressure application surfaces of the piston which are associated with the pressure chamber118and the pressure application surfaces of the piston which are associated with the pressure compensation chamber120are of the same size and extend in the same radial area so that corresponding fluid level limiting means are not required for the pressure compensation chamber120.

Additional cooling oil flows preferably occur in operation. Accordingly, at least one radial bore hole160is provided in the transmission input shaft24, a further partial flow of cooling oil flows via this radial bore hole160and via an annular channel between the two transmission input shafts22,24and splits into two partial flows, one of which flows radially outward between the two hub parts80and84(through the axial bearing94) and the other partial flow flows radially outward between the end area of the transmission input shaft22remote of the transmission and the hub part80and between this hub part80and the annular portion38of the clutch hub34(through the bearings98and100).

Since the cooling oil flowing radially outward could collect in the neighborhood of a radial outward portion of the actuation piston110associated with the first multi-disk clutch arrangement64and could hinder the engaging movement of the piston110due to centrifugal force at least at higher rotational speeds, this piston110has at least one pressure compensation opening162which enables the cooling oil to flow from one side of the piston to the other. Consequently, collection of the cooling oil can come about on both sides of the piston with corresponding compensation of pressure forces exerted on the piston as a result of centrifugal force. Further, other forces based on an interaction between the cooling oil and the piston are prevented from hindering the required axial piston movements, for example, hydrodynamic forces or the like as well as the piston attaching itself by suction to the outer disk carrier62.

It is also possible to provide at least one cooling oil flow-off opening in the radially extending, radial outer area of the outer disk carrier62of the first multi-disk clutch arrangement64. A cooling oil flow-off opening of this kind is indicated by dashed lines at164. In order nevertheless to ensure a sufficient flow of cooling oil through the disk stack76of the first multi-disk clutch arrangement64, a cooling oil conducting element (generally a cooling fluid conducting element) may be provided. It is indicated in dashed lines inFIG. 13that an adjoining end disk166of the disk stack76could have a cooling oil conducting portion168, so that the end disk166itself serves as cooling oil conducting element.

For the purpose of a simple construction of the pressure control device for actuation of the two multi-disk clutch arrangements64,72, provision was made in the embodiment ofFIG. 13to at least partially compensate for a torque transmission capability which is given per se for the radial inner multi-disk clutch arrangement72with respect to an actuation pressure and which is smaller (because of a smaller effective friction radius than the radial outer clutch arrangement64) compared to the other clutch arrangement64. Thus, the pressure application surface of the piston130associated with the pressure chamber140is greater than the pressure application surface of the piston110associated with the pressure chamber118, so that the axially directed forces acting on piston130are greater than those exerted on piston110at the same hydraulic oil pressure in the pressure chambers.

Advantageously, good use of the available installation space may be made by means of a radial offsetting of the seals associated with the piston, especially also an axial overlapping of at least some of the seals.

Steps for preventing the risk of overheating can be taken with the disk stacks74,76in addition to the supply of cooling oil already described and the forming of cooling oil passages (only shown schematically inFIG. 13) in the disk carriers. Accordingly, it is advantageous to use at least some of the disks as “heat buffers” which temporarily store heat which occurs, for example, during slip operation and which momentarily overburdens the possibilities for heat flow-off by means of the cooling fluid (in this case cooling oil) or by means of heat conduction via the disk carriers, so that the heat can be guided off subsequently, for example, in a disengaged state of the respective multi-disk clutch arrangement. For this purpose, disks without friction linings in the radial inner (second) multi-disk clutch arrangement72are constructed so as to be thicker axially than friction lining carrying elements of disks carrying friction linings in order to provide a comparatively large material volume with corresponding heat capacity for the disks without friction linings. These disks should be produced from a material having a considerable heat storage capability (heat capacity), for example, steel. When using conventional friction linings made of paper, for example, the friction lining carrying disks can only store a little heat intermediately, since paper has a poor thermal conductivity.

The heat capacity of the friction lining carrying elements carrying the friction linings can likewise be made available as a heat accumulator when lining materials with high conductivity are used instead of lining materials with low conductivity. The use of friction linings of sintered material having a comparatively high heat conductivity is possible. However, the use of sintered linings is problematic in that sintered linings have a degressive curve of the friction coefficient over a slippage speed (relative rotational speed ΔN between the rubbing surfaces), that is, in that dμ/dΔN<O. A degressive curve of the friction coefficient is disadvantageous insofar as this can promote a self-excitation of vibrations in the drivetrain or at least can not dampen vibrations. Therefore, it is advantageous when disks with friction linings of sintered material as well as disks with friction linings of another material with a progressive friction coefficient curve over the slippage speed dμ/dΔN>O are provided in a disk stack74,76, so that a progressive friction coefficient curve over the slippage speed or at least approximately a neutral friction coefficient curve over the slippage speed dμ/dΔN=O results on the whole for the disk stack and consequently a self-excitation of vibrations in the drivetrain is at least not promoted or—preferably—rotational vibrations in the drivetrain are even damped (because of a considerable progressive friction coefficient curve over the slippage speed).

In the embodiment ofFIG. 13, the disk stack74of the radial inner multi-disk clutch arrangement72is constructed without sintered linings since the radial outer multi-disk clutch arrangement64is preferably used as starting clutch with corresponding slippage operation. The latter, that is the use of the radial outer multi-disk clutch arrangement64as a starting clutch, is advantageous insofar as this multi-disk clutch arrangement can be operated with smaller actuation forces (for the same torque transmission capability) because of the greater effective friction radius, so that surface pressing can be reduced compared to the second multi-disk clutch arrangement72. It is also helpful in this respect when the disks of the first multi-disk clutch arrangement64are formed with a somewhat greater radial height than the disks of the second multi-disk clutch arrangement72. If desired, however, friction linings of sintered material can also be used for the disk stack74of the radial inner (second) multi-disk clutch arrangement72, preferably—as previously stated—in combination with friction linings from another material such as paper.

While all inner disks carry a friction lining and all outer disks are disks without linings in the disk stack74of the radial inner multi-disk clutch arrangement72, wherein the end disks defining the disk stack axially are outer disks and are accordingly disks without lining, the inner disks in the disk stack76of the first multi-disk clutch arrangement64are disks without linings and the outer disks, including the end disks166,170, are disks carrying friction linings.

Preferably, at least the end disks166and170have lining carrying elements which are substantially thicker axially than the lining carrying elements of the other outer disks and are constructed with linings of sintered material in order to make the lining-carrying elements of the two end disks having a comparatively large volume usable as heat buffers. As in the disk stack74, the disks without linings are axially thicker than the friction lining carrying elements of the disks carrying friction linings (with the exception of the end disks) in order to provide a comparatively large heat capacity for temporary storage of heat. The outer disks located on the axial inner side should have, at least in part, friction linings of another material exhibiting a progressive friction coefficient curve in order to achieve at least an approximately neutral friction coefficient curve over the slippage rotation speed for the disk stack in its entirety.

The person skilled in the art can easily gather further details of the dual clutch12according to the described embodiment example fromFIG. 13. Accordingly, the axial bore hole in the annular portion36of the clutch hub34in which the internal toothing46for the pump drive shaft is formed is closed in an oil-tight manner by a stopper180which is held therein. The carrier plate60is fixed axially at the outer disk carrier62by two retaining rings172,174, the retaining ring172also supports the end disk170axially. A corresponding retaining ring is also provided for supporting the disk stack74at the outer disk carrier70.

FIG. 14shows a rotor device700which can be used as component P220in the embodiment ofFIG. 1or as component P220aof the embodiment ofFIG. 9. In the embodiment ofFIG. 1, the rotor device has at least one pressure medium connection702connected to the pump208via the valve218or directly to a cooling oil connection703which is connected to the reservoir222and from which cooling oil can flow, possibly via a check valve, not shown, into a housing706of the rotor device700constructed in multiple parts to two conveying rotors708aand708bwhich have conveying structures in the form of toothings meshing with one another. The two conveying rotors708aand708bform an external toothed gear pump.

Pressure oil flowing into the housing706via the connection702strikes two drive rotors710aand710bwhich have meshing drive geometries in the form of toothings and are set in rotation by the pressure oil. The two drive rotors710aand710bform a hydromotor710in a rotary positive-displacement type construction, especially an external toothed gear motor. The drive rotor710ais connected with the conveying rotor708aso as to be fixed with respect to rotation relative to it and the drive rotor710bis connected with the conveying rotor708bso as to be fixed with respect to rotation relative to it. InFIG. 16, cooling oil conveyed by the conveying rotors708aand708bflows over a flow-off channel716in the direction of the dual clutch202in order to supply the disk clutches204and206with cooling oil. The pressure oil which has passed the drive rotors710aand710bflows off via a channel717which preferably passes into the channel716.

As is shown inFIG. 14, the drive rotor710aand the conveying rotor708ahave a common rotating shaft, e.g., such that two toothed gears are cut into a rotating shaft. The same can apply to the drive rotor710band the conveying rotor708bas is shown inFIG. 14. As was mentioned, the pair of rotors (possibly a pair of toothed gears)710a,710bare driven by the pump208, preferably a hydrostatic pump. Since the rotor pair (possibly a toothed gear pair)708a,708bis located on the same shaft with a respective drive rotor and, as can be seen inFIG. 14, is longer axially (has teeth which are wider axially) than the drive rotors, a larger volume flow can be conveyed at a lower pressure. According toFIG. 14, the two pairs of rotors and therefore the volume flows (pressure oil volume flow and cooling oil volume flow) are separated from one another by a divided separating or sealing disk721. The sealing disk is shown inFIG. 17and comprises two half disks721aand721bwhich define two openings723aand723bfor the rotor shafts725aand725bin the assembled state.

Another embodiment form of a rotor device800functioning according to the same principle is shown schematically inFIG. 18. Therein, the drive rotors810aand810bof the hydromotor810have a smaller diameter than the conveying rotors808aand808b, so that a volume flow which is greater than a pressure oil flow driving the drive rotors810aand810bcan be conveyed again by means of the conveying rotors but, on the other hand, has a lower pressure than this flow of pressure oil. The drive rotor810aand the conveying rotor808ahave a common rotating shaft825a. The drive rotor810band the conveying rotor808bhave their own rotating shaft825band825c, respectively. These two rotors810band808bare in a drive connection, via their toothings meshing with the toothings of the rotors810aand808a, with these rotors, that is, rotors810aand808a.

In summary, the invention is directed to a clutch system comprising at least one clutch device particularly for arrangement in a drivetrain between a drive unit and a transmission. The clutch device can be actuated through the intermediary of pressure medium and/or provided for operation under the influence of an operating medium. According to one aspect of the invention, it is suggested that a secondary medium serving as pressure medium for actuation or serving as operating medium for the operation of the clutch arrangement can be supplied by means of a pump arrangement supplying a primary medium and by means of a secondary medium supply arrangement or secondary pump arrangement connected to the pump arrangement, wherein the secondary medium supply arrangement or secondary pump arrangement has at least one pressure receiving element which is arranged so as to be movable in a housing and which is exposed at least on at least one pressure receiving surface in operation with the primary medium supplied by the first pump arrangement, which pressure receiving element is set in motion through the intermediary of the operating medium, and at least one interacting element which is arranged in the/a housing so as to be movable, is coupled with respect to movement with the pressure receiving element and interacts in operation with the secondary medium at least on at least one interaction surface, the secondary medium can be supplied by means of this interaction element possibly through displacement and/or imparting an acceleration and/or imparting angular momentum or swirl.