GB2305474A

GB2305474A - A friction clutch

Info

Publication number: GB2305474A
Application number: GB9626436A
Authority: GB
Inventors: Karl-Ludwig Kimmig; Christoph Wittmann; Paul Maucher; Wolfgang Reik; Rolf Meinhard
Original assignee: LuK Lamellen und Kupplungsbau GmbH
Current assignee: LuK Lamellen und Kupplungsbau GmbH
Priority date: 1992-07-11
Filing date: 1993-07-07
Publication date: 1997-04-09
Anticipated expiration: 2013-07-07
Also published as: GB2305475A; GB2305475B; GB9626436D0; GB9626437D0; GB2305474B

Abstract

A friction clutch with a pressure plate 403 which is connected to a housing 402 so that it cannot rotate relative to the housing but can move axially relative to the housing by a limited amount. A contact pressure plate spring 404 biases the pressure plate 3 in the direction of a clutch disc 408 which can be clamped between this pressure plate and a counter pressure plate (not shown). An adjustment device 416 is provided which automatically compensates for wear of the friction linings of the clutch disc and causes a practically constant force biasing the pressure plate 403 through the contact pressure plate spring. The adjustment device causes axial movement of a swivel support pad 412 on the housing side of the spring 404. The contact pressure plate spring 402 is supported by a supporting force whose size can change in dependence on the speed and which is directed towards the swivel pad 412 on the housing side. A plate spring 413 may provide the supporting force, the force increasing with speed due to the effort of centrifugal force acting on axially extending tongues 450 of the spring 413.

Description

Friction Clutch The invention relates to a friction clutch which has a pressure plate which can be connected rotationally secured but axially displaceable to a restricted amount to a counter pressure plate wherein at least a contact pressure spring biases the pressure plate in the direction of a clutch disc which can be clamped between this and the counter pressure plate and wherein an adjustment device compensating at least the wear of the friction linings of the clutch disc is provided which causes a practically uniform force biasing of the pressure plate through the contact pressure spring, furthermore the friction clutch has operating means for engaging and disengaging which can be operated by means of a release member such as for example a disengagement fork which is swivel mounted on a gearbox housing.

A clutch assembly which is constructed and operated in this way has been proposed by FR-OS 2 582 363. The operating means of a clutch assembly of this kind can be biased by disengagement systems or by disengagement means and release members, such as has been proposed by US-PS 4 368 810, US-PS 4 326 617, DE-OS 27 52 904 and DE-OS 27 01 999.

With clutch assemblies and friction clutches with an integrated adjustment device which compensates at least the wear of the friction linings of the clutch disc there is a problem, particularly in connection with so-called mechanical disengagement systems wherein the movements of the clutch pedal are transferred to the operating means of the friction clutch through a rod linkage and/or Bowden cable with the interposition of at least one release bearing, that as a result of the tolerances present in the entire kinematic chain it is not guaranteed that the areas of the release bearing biasing the operating means always have the same axial position relative to the areas of the operating means to be biased so that a comparatively large variation of the disengagement path of the friction clutch and of the operating path transferred to the operating means can exist.The function of the adjustment device can at least be impaired through this variation and in extreme cases the adjustment function of this device can no longer be provided. Furthermore cases can occur where the operating means cover an unacceptably large path whereby undesired adjustment can occur which causes the friction clutch either to no longer open satisfactorily or the pretensioning and installation position of the contact pressure spring to change so that the force applied by the latter is no longer sufficient to guarantee a satisfactory torque transfer.

The object of the present invention is to avoid these disadvantages and to provide a clutch assembly of the kind mentioned at the beginning wherein a satisfactory functioning of the adjustment device is guaranteed compensating the wear of the friction linings. Furthermore the assembly is to be manufactured in a particularly simple and cost-effective manner.

According to the invention there is provided a friction clutch, more particularly for motor vehicles, with a pressure plate which is connected rotationally secured but axially displaceable to a restricted amount to a housing, wherein between the housing and pressure plate is a contact pressure plate spring which on the one hand can swivel about a swivel bearing supported by the housing and formed by swivel support pads mounted on each side of the contact pressure plate spring and on the other hand - in the assembled state of the friction clutch on a counter pressure plate - biases the pressure plate in the direction of a clutch disc which can be clamped between this and the counter pressure plate, such as a flywheel, wherein the swivel support pad on the side of the housing can be axially moved by an automatic adjustment device which compensates at least the wear of the friction linings of the clutch disc and the contact pressure plate spring stands under the action of a supporting force whose size can change in dependence on the speed and which is directed towards the swivel pad on the housing side.

This supporting force is preferably permanent so that the diaphragm spring is supported against the disengagement force only with force locking engagement, namely through a spring force and not by keyed engagement means. The diaphragm spring is thereby installed over its working area with a degressive characteristic line, namely so that the supporting force and the diaphragm spring force are matched with each other so that with the proposed installation position of the diaphragm spring and without any wearconditioned change in the conicity and throughout the disengagement path of the diaphragm spring the supporting force is greater than the force applied by the diaphragm spring and counteracting the supporting force, and with a wear-conditioned change in the conicity of the diaphragm spring the supporting force over partial areas of the disengagement path of the diaphragm spring is less than the form of the force applied by the diaphragm spring against the supporting force. The supporting force can thereby be applied by a single spring element or at least substantially by a single spring element or spring element system. By "supporting force" is meant the sum of all the spring forces acting against the diaphragm spring - insofar as they are noticeable-, thus eg also or only the forces which are active through the (torque-transfer or lift) leaf springs, the (residual) suspension of a lining suspension or their "substitute".

This arrangement is particularly advantageous in the case of clutch assemblies where the operating means move in the axial direction of the disengagement movement in dependence on at least the wear of the friction linings since a practically play-free force transfer can be ensured between the release bearing or disengagement means and the operating means. It is also thereby guaranteed that the operating means can be constantly moved by the same amount. Thus practically no play can exist in the force flux between the release bearing and/or the disengagement means and operating means.

It can be particularly advantageous if the compensating device is provided and acts axially between the release bearing and the operating means. The compensating device can however also be provided at another point, eg in active connection between the release bearing and disengagement means. It is advantageous if the release bearing is positioned on an axial guide preferably provided on the gearbox side, such as for example a guide tube enclosing the gearbox input shaft.

Particularly with clutch assemblies having a friction clutch which has a housing, such as for example a sheet metal lid, fixable on the counter pressure plate, with a base facing the release bearing, it can be expedient if the compensating device is mounted and is active axially between the operating means and the base. Furthermore it can be advantageous if the contact pressure spring is formed by a diaphragm spring which can be axially tensioned between a clutch housing and the pressure plate and which has a resilient annular foundation body and tongues running radially inwards from same to form the operating means.

In order to ensure a satisfactory adjustment through the compensating device it can be particularly advantageous if this compensating device in the engaged state of the clutch assembly or friction clutch automatically and independently guarantees the desired adjustment but automatically and independently blocks same during the operating of the friction clutch.

The compensating device can have a ring-like component which axially adjoins the operating means also in the engaged state of the friction clutch. Through this ring-like component it is possible to compensate any change which may occur in the distance between the biasing areas of the operating means and the release bearing. For the functioning of the compensating device it can be advantageous if this has axially rising adjustment ramps or run-up ramps wherein these can be provided on the ring-like component.

The run-up ramps can interact with cylinder shaped or ball like rolling bodies for adjustment. It can however be particularly advantageous if the run-up ramps interact with counter run-up ramps since then through suitable selection of the run-up angle of these ramps a self-locking can occur in the event of axial tensioning of the ramps. The counter run-up ramps can likewise be supported by a ring-like component.

In order to ensure an economic production of the friction clutch it can furthermore be advantageous if at least a part of the compensating device is made from plastics. Such plastics parts can be made by injection. Particularly suitable plastics are thermoplastics such as polyamide.

In a particularly advantageous way the components having the adjustment ramps can be axially displaced when operating the clutch assembly or friction clutch. Furthermore it can be expedient if the component supporting the run-up ramps and counter run-up ramps are rotatable relative to each other wherein one of these components can be rotationally fixed relative to the friction clutch, more particularly relative to the clutch housing.

The compensating device can be designed so that - seen in the disengagement direction of the clutch assembly - it adjusts or acts like a freewheel, but is self-locking in the direction opposite the disengagement direction. To this end the run-up ramps and/or counter run-up ramps can be designed so that they have in the axial direction a pitch angle which lies between 50 and 20 , preferably in the order of 70 to 110. The adjustment ramps are advantageously designed so that a self-locking takes place through friction engagement.

It is thus to be guaranteed in each case that the adjustment ramps have a self-locking engagement so that no additional means are required in order to avoid undesired resetting.

When necessary such means can however be provided.

In order to ensure a satisfactory functioning of the automatic compensating device it can be expedient if at least a component supporting the run-up ramps and/or a component supporting the counter run-up ramps is springbiased in the adjustment direction. The spring biasing can thereby be carried out advantageously so that the functioning of the remaining springs, such as in particular the contact pressure or diaphragm spring and the spring biasing the axially pliable support bearing is not or practically not affected. A particularly advantageous design can be guaranteed in that the components having the run-up ramps and counter run-up ramps are biased or tensioned in the adjustment direction by at least one energy accumulator such as a coil spring provided between the same.

Through such tensioning these two components, viewed axially, are pressed in opposite directions, thus moved axially away from each other through the energy accumulator and adjustment ramps. When the clutch is engaged the compensating device can thereby be tensioned play-free axially between the biasing areas of the operating means and the clutch lid and/or release bearing.

The clutch assembly can have a device for restricting the disengagement path of at least the operating means. To this end a restricting stop can be provided which limits the path of the release bearing and/or of the disengagement means in the disengagement direction. The restricting stop can advantageously be formed by a component forming the compensating device striking against the clutch lid after a certain disengagement path. The restriction can however also be produced if the release bearing has areas which come to rest against an axially fixed component after a certain disengagement path. Furthermore it can be advantageous if the release bearing also has in the engagement direction a restriction which can likewise be formed by a stop.

Advantageously the compensating device is formed so that the release bearing is axially supported by same in the engaged state of the clutch assembly. The constant operating path for the clutch assembly can also be guaranteed in that a component forming the compensating device has path restricting areas which interact with stop areas and act in the disengagement and engagement directions. Advantageously this component can be formed by the component of the disengagement device biased by the release bearing wherein the restricting stops can be provided on the clutch housing or can be formed by this housing. The restriction of the operating path of the clutch assembly can however also be carried out by providing corresponding stops on the component which guides the release bearing in the axial direction.Advantageously these stops interact with a component which is connected to the non-rotating bearing ring of the release bearing. Restricting the disengagement path in at least an axial direction can however also be carried out between the rotating bearing ring of the release bearing and a component turning same, such as eg the clutch housing.

It can be advantageous, particularly in order to minimize the disengagement force path and the maximum disengagement force necessary, if means are provided which during the disengagement process over at least a part of the operating path of the operating means causes a gradual breakdown of the moment which can be transferred by the friction clutch or clutch disc. These means can be formed for example by a so-called lining suspension which are provided between the friction linings of the clutch disc which can be clamped between the pressure plate and counter pressure plate.

When the contact pressure spring, which can be formed preferably by a diaphragm spring, is supported for swivel movement on the housing between two support bearings - of which the one facing the pressure plate is spring-loaded in the direction of the contact pressure diaphragm spring, the maximum disengagement force exerted on the spring-loaded support bearing by the contact pressure spring during disengagement of the friction clutch increases as the lining wears down and becomes greater than the counter force or supporting force acting on the spring-loaded support bearing.When using leaf spring elements provided for torque transfer between the pressure plate and clutch housing, and/or a so-called lining suspension such as is known for example through DE-OS 36 31 863, the forces exerted by these springs on the contact pressure spring have to be taken into account when fixing the force which acts on the spring-loaded support bearing, namely because these forces are superimposed on each other. This means that the increased disengagement force which is temporarily set in the presence of sufficient lining wear must be greater than the resulting force arising from the aforesaid forces and in relation to the swivel diameter of the diaphragm spring in order to enable adjustment. It can be particularly expedient if the spring-loaded support bearing is axially displaceable.Advantageously the contact pressure diaphragm spring can have such a force characteristic line that, starting from its structurally defined installation position' in the friction clutch, in the event of a relaxation caused by wear on the friction lining the force which can be applied by same and thus also the level of the disengagement force path increases and in the event of a more deformed or more tensioned position in relation to the defined installation position the maximum force which can be applied by it decreases in the event of a disengagement process.

Through such an arrangement and design of the contact pressure diaphragm spring it is possible to guarantee that in the event of wear on the linings a counterbalance can always be set again, at least between the maximum disengagement force of the friction clutch and the counter force acting on the spring-loaded support bearing, or the resulting counter force acting on the contact pressure diaphragm spring in the area of the rolling diameter.

The clutch assembly or friction clutch can be installed advantageously so that the axially displaceable springloaded support bearing moves together with the pressure plate over the wear reserve of the friction clutch. During the adjustment of the adjustment device which takes place gradually or in small stages over the service life of the friction clutch, the spring-loaded support bearing can be slightly moved in the direction of the pressure plate.

Through this displacement it can be ensured that the diaphragm spring which is then supported on the pressure plate undergoes an additional deformation so that the force exerted by it, as already mentioned decreases until the counter force acting on the spring-loaded support bearing or the resulting counter force already mentioned is balanced with the disengagement force. Thus during displacement of the spring-loaded support bearing the maximum disengagement force of the clutch or contact pressure diaphragm spring decreases again.

It can be particularly advantageous if the contact pressure diaphragm spring is installed in the friction clutch so that it has at least over a part of the disengagement area, and preferably over the entire disengagement area of the clutch, a declining force-path characteristic line. The installation position of the contact pressure spring can thereby be designed so that in the disengaged state of the friction clutch the contact pressure spring practically reaches the minimum or trough point of its sinusoidal forcepath line.

The counter force exerted on the spring-loaded support bearing can advantageously be produced by an energy accumulator which applies basically a constant force, at least over the intended adjustment area. A correspondingly designed diaphragm spring installed in the pretensioned state in the friction clutch is particularly suitable for this.

The invention is not restricted to the above described friction clutches but can be used generally in the case of friction clutches or clutch assemblies having an adjustment device which compensates at least the wear of the friction linings of the clutch disc.

Automatic adjustment devices which are to cause a practically constant force biasing of the pressure plate through the contact pressure diaphragm spring are known for example from DE-OS 29 16 755 and 35 18 781. The adjustment devices which are displaceable in dependence on at least one sensor are thereby mounted and act between the pressure disc and the contact pressure diaphragm spring.Owing to the articulated connection between the pressure disc and the housing by means of tangentially mounted leaf springs whose force must be relatively light because it is directed against the contact pressure force of the diaphragm spring the pressure disc which has a comparatively large mass can vibrate axially when the friction clutch is disengaged, thus can lift off from the diaphragm spring whereby the function of the clutch is not only impaired but also the clutch becomes a safety risk since the adjustment device makes adjustments in the opened state until the pressure plate adjoins the clutch disc, thus can no longer separate the clutch. For this reason such adjustment devices have not proved successful in practice.

It is desirable to remove these disadvantages and to provide adjustment measures of the kind already mentioned which can be used in practice on a wide basis and even during rough operation, which have a simple construction and a reliably safe function, which furthermore require a small installation space and which are economical to produce.

Furthermore the disengagement forces required should be slight, remain slight throughout the service life and the service life of the friction clutches should be further increased.

The energy accumulator which applies at least most of the supporting force can preferably be a spring which changes its shape during the adjustment, eg a diaphragm spring. The energy accumulators which apply the supporting force can however also be formed by the leaf springs.

A diaphragm spring applying the supporting force can be mounted directly on the diaphragm spring, eg at the radial level of the axially displaceable support on the cover side.

It is particularly advantageous if the adjustment device is mounted axially between the diaphragm spring and cover. The adjustment assembly can have in particular run-up faces such as ramps.

Through the invention it is ensured that the diaphragm spring, seen over the service life of the friction clutch, practically always has the same conicity and tension when the friction clutch is engaged and that there is a practically constant force biasing of the pressure plate and thus clutch disc - independently of the wear on the friction linings, the pressure plate itself or other elements such as the supports on the cover or pressure plate side, the diaphragm spring or friction face of the flywheel disc -.

Through the measures described here, it is furthermore ensured that the mass of the pressure plate is not increased by that of the adjustment device. It is furthermore housed in an area in which it is protected against the effects of wear on the disc and in which it is further removed from the source of the friction heat.

A particularly advantageous design of a friction clutch can be achieved if the contact pressure diaphragm spring is supported for swivel movement on the housing between two pressure pads - of which the one facing the pressure plate is spring-loaded in the direction of the contact pressure diaphragm spring-, wherein the force acting on the springloaded support pad by the contact pressure diaphragm spring during disengagement of the clutch increases with wear on the linings and then becomes greater than the counter force or supporting force acting on the spring-loaded support pad.

The contact pressure diaphragm spring thereby has such a characteristic line curve that, starting from its structurally defined installation position in the friction clutch, with a relaxation direction conditioned by the friction lining wear, the force then applied by same and thus also the required disengagement force first increase and with a further deformed or tensioned position relative to the defined installation position, the force which can be applied by same decreases during the disengagement process.

Through such a design and arrangement of the contact pressure diaphragm spring it is ensured that when wear appears on the linings, an equilibrium can always be set between the force exerted by the contact pressure diaphragm spring on the support pad during disengagement, and the counter force acting on the spring-loaded support pad, because, when the supporting force is exceeded by the force exerted by the diaphragm spring on the support pad, the diaphragm spring moves the sensor spring away from the support pad on the cover side and the adjustment device can be turned further by the force of the feed mechanism. Thus the support pad is axially displaced until the force exerted by the sensor prevents further rotation and further axial displacement of the support pad.

It can be particularly advantageous, as already mentioned, if the contact pressure diaphragm spring is installed in the friction clutch so that it has a declining force characteristic line at least over a part of the disengagement area, preferably practically over the entire disengagement area of the friction clutch. The installation position of the contact pressure diaphragm spring can thereby be such that in the disengaged state of the friction clutch the contact pressure diaphragm spring reaches or exceeds practically the minimum or trough point of its sinusoidal force-path curve.

The counter force exerted on the spring-loaded support pad can advantageously be produced by an energy accumulator which applies a substantially constant force at least over the proposed adjustment area. A suitably designed diaphragm spring installed in the pretensioned state in the friction clutch is particularly suitable and advantageous for this purpose.

The adjustment device can advantageously be used with friction clutches having a contact pressure diaphragm spring which with the radially outer areas biases the pressure plate and with the radially further inner areas is mounted between two swivel support pads on the housing. With this type of construction the diaphragm spring can act as a double-armed lever.

The invention is however not restricted to friction clutches with diaphragm springs which have at the same time the disengagement lever moulded on in the form of diaphragm spring tongues, but also extends to other clutch constructions wherein eg the diaphragm spring is operated by additional levers.

In order to ensure satisfactory adjustment of the wear or an optimum contact pressure force for the friction clutch it can be particularly advantageous if the counter support pad provided on the side of the contact pressure diaphragm spring remote from the spring-loaded support pad is designed so that it can be automatically or independently displaceable axially in the direction of the pressure plate but can be automatically or independently locked in the counter direction by a device. The adjustment of the counter support pad, thus the pad on the cover side can be carried out by an energy accumulator which biases this counter support pad in the direction of the pressure plate or against the contact pressure diaphragm spring.The counter support pad can thus be adjusted automatically corresponding to the displacement of the spring-biased pad conditioned through the wear on the linings whereby a playfree swivel bearing of the contact pressure diaphragm spring can be guaranteed.

The counter support pad can be axially displaceable by means of an adjustment device provided between the contact pressure diaphragm spring and cover. The cover device can thereby have a ring-shaped thus inherently joined component which is axially biased by the contact pressure diaphragm spring at least in the engaged state of the friction clutch.

By rotating the ring-shaped component when wear starts to appear and during the disengagement process it is possible to adjust the swivel bearing according to the wear on the linings. To this end the adjustment device or ring-shaped component of this adjustment device can advantageously have axially rising adjustment ramps. Furthermore it can be advantageous if the ring-shaped component supports the counter support pad wherein the latter can be formed by a wire ring. This wire ring can be housed in a circumferential ring groove of the component and can be connected to same by keyed engagement. The keyed engagement can be produced by a snap fit connection.

The run-up ramps can interact with cylindrical or ball-like rolling bodies for adjustment. It can be particularly advantageous however if the run-up ramps interact with corresponding counter run-up ramps since then by suitably selecting the run-up angle of these ramps a self-locking action can occur when the ramps are axially stressed. The counter run-up ramps can be supported by a ring-like component which can be mounted between the component supporting the run-up ramps and the cover. A particularly simple construction can however by ensured by fitting the counter run-up ramps in the housing. The latter can occur particularly easily in the case of sheet metal housings since the counter run-up ramps can be imprinted. Imprinting can thereby be carried out in the radially aligned areas of the housing.

In order to ensure a favourable economic production of the friction clutch it can furthermore be advantageous if at least part of the adjustment device is made of plastics.

Such plastics parts can be made by injection moulding.

Thermoplastics such as eg polyamide are particularly suitable as the plastics. The use of plastics is therefore possible because the adjustment device is located in an area which is only exposed by a minimum amount to the effects of heat. Furthermore as a result of the lighter weight a lower mass inertia moment is produced.

The adjustment device can be designed so that - seen in the disengagement direction of the friction clutch - it acts like a freewheel, but is self-locking in the direction opposite the disengagement direction. To this end the runup ramps and/or the counter run-up ramps are designed so that they have in the axial direction a slope angle which is between 4 and 20 degrees, preferably in the order of 5 to 12 degrees. The run-up ramps and/or counter run-up ramps are advantageously designed so that self-locking takes place through friction engagement. The self-locking can however also be achieved or assisted by positive keyed locking wherein for example one of the ramps is soft and the other has a profiled section or wherein both ramps have profiled sections. Through these measures it is ensured that no additional means are required to avoid undesired resetting.

The adjustment device can be particularly advantageous and simple if the circumferentially acting feed mechanism is designed as a spring which is installed in the pretensioned state and which resiliently biases in the adjustment direction at least one component supporting the run-up ramps and/or one component supporting the counter run-up ramps or run-up areas. The spring biasing can advantageously be carried out so that the function of the remaining springs, such as in particular the operating diaphragm and the spring biasing the axially pliable support pad, is not or hardly affected.

In many cases it can be advantageous if the adjustment device has several displaceable adjustment elements such as eg radially and/or circumferentially displaceable adjustment wedges or rolling bodies. It can furthermore be advantageous if the adjustment device is dependent on speed.

Thus for example the centrifugal force acting on individual elements of the adjustment device can be used to operate and/or lock the adjustment device at certain operating stages of the internal combustion engine. More particularly the adjustment device can be blocked by means dependent on centrifugal force from a certain speed, which can take place for example at at least approximately idling speed or speeds below idling speed so that the wear adjustment takes place only at low speeds. This has the advantage that no undesired adjustments are made which could result from vibrations at high speeds.

A particularly simple and functionally reliable structure of the adjustment device can be ensured in that the parts which have the run-up ramps and/or counter run-up ramps and counter run-up areas and which are displaceable relative to the housing are spring tensioned. If there is only one corresponding component with the corresponding ramps or areas which is displaceable relative to the housing then this is biased. It can thereby be particularly advantageous if the spring tension produces a force in the circumferential direction.

It can furthermore be particularly advantageous for the construction and functioning of the friction clutch if the sensor spring designed as a disc spring, such as a diaphragm spring is supported by its radially outer area on an axially fixed component such as the housing, and with the radially further inner areas biases the rolling support pad remote from the cover. This rolling support pad can also be made in one piece with the sensor spring so that the sensor diaphragm spring also forms the support pad. In order to hold the sensor spring in the tensioned position the housing can support supporting areas. These supporting areas can be formed by individual support elements attached to the housing.It can however also be advantageous if the supporting areas are in one piece with the housing, eg indentations or cut-out and moulded areas can be provided on the housing which engage axially underneath the sensor spring for support.

It can be particularly advantageous for the function of the friction clutch, more particularly to minimize the disengagement force curve or the maximum disengagement force required, if the clutch disc which can be clamped between the pressure plate and counter pressure plate has friction linings between which is provided a so-called lining suspension, such as is known for example from DE-OS 36 31 863. The use of such a clutch disc assists in the operation, more particularly the disengagement process of the friction clutch. This is due to the fact that in the engaged state of the friction clutch the tensioned lining suspension exerts on the pressure plate a reaction force which counteracts the force exerted by the contact pressure diaphragm spring or operating diaphragm spring on this pressure plate.During the disengagement process during the axial displacement of the pressure plate the latter is first forced back by the resiliently tensioned lining suspension wherein at the same time as a result of the comparatively sharply falling characteristic line section of the contact pressure diaphragm spring which exists in the disengagement area, the force exerted by the contact pressure diaphragm spring on the pressure plate is reduced. With the reduction in the force exerted on the pressure plate by the contact pressure diaphragm spring, the resetting force exerted by the lining suspension on this pressure plate also decreases.

The force effectively required to disengage the friction clutch is produced from the difference between the resetting force of the lining suspension and the contact pressure force of the contact pressure diaphragm spring. After the relaxation of the lining suspension, thus as the pressure plate lifts off from the friction linings and during release of the clutch disc through the pressure plate the required disengagement force is mainly determined by the contact pressure diaphragm spring. The force-path characteristic of the lining suspension and the force-path characteristic of the contact pressure diaphragm spring can advantageously conform with each other so that during release of the clutch disc through the pressure plate the force required to operate the contact pressure diaphragm spring is located at a lower level.Thus by adapting the lining suspension characteristic to or matching it with the contact pressure diaphragm spring characteristic until the clutch disc is released by the pressure plate it is possible that only a very slight operating force, and in extreme cases practically none at all, is required for the contact pressure diaphragm spring to overcome the remaining output.

Furthermore the characteristic of the contact pressure diaphragm spring can be designed so that when the clutch disc is released the force then still set against swivelling by the contact pressure diaphragm spring, or the force required to swivel the contact pressure diaphragm spring is located at a very low level compared to the contact pressure force applied by this contact pressure diaphragm spring in the engaged state of the friction clutch. Designs are also possible wherein on releasing the clutch disc through the pressure plate only a very slight or practically no force is required to operate the contact pressure diaphragm spring to disengage the clutch. Such friction clutches can be designed so that the operating forces lie in the order of between 0 and 200 N.

The friction clutch can be designed so that at least approximately when releasing the clutch disc through the pressure plate the axial force applied by the contact pressure diaphragm spring is located at zero wherein on continuing with the disengagement process the force applied by the contact pressure diaphragm spring can be negative, thus a change in the force action of the contact pressure diaphragm spring takes place. This means that when the friction clutch is fully disengaged it remains open practically by itself and the engagement process can only be initiated again through outside force action.

There is also disclosed a friction clutch, more particularly for motor vehicles with a pressure plate which is connected rotationally secured but axially displaceable to a restricted amount to a housing wherein between the housing and pressure plate there is at least one active contact pressure spring which can be tensioned and which biases the pressure plate in the direction of a clutch disc which can be clamped between this plate and a counter pressure plate such as a flywheel.

Clutches of this kind are known for example from DE-OS 24 60 963, DE-PS 24 41 141 and 898 531 as well as from DE-AS 1 267 916.

The object of the present invention is also to improve such friction clutches with regard to their functioning and service life. More particularly the forces required to operate such friction clutches are to be reduced and a practically constant disengagement force path is to be sought throughout the service life of the clutches.

Furthermore the clutches are to be manufactured in a particularly simple economic manner.

This is achieved in that an adjustment device which automatically compensates the wear on the friction linings of the clutch disc is provided which causes a practically uniform force biasing of the pressure plate through the contact pressure spring, and the friction clutch has operating means for engagement and disengagement as well as a device which during the disengagement process, at least over part of the operating path of the operating means and/or disengagement path of the pressure plate causes a gradual breakdown of the moment which can be transferred from the friction clutch or clutch disc.Through a device of this kind it can be achieved that during the engagement process of the friction clutch and at the beginning of the tensioning of the friction linings between the pressure and counter pressure plate a gradual or progressive build-up of the moment which can be transferred by the friction clutch takes place.

Through the design of a friction clutch as set out here, the contact pressure diaphragm spring, seen over the service life of the friction clutch, practically always has the same pretension when the friction clutch is engaged and thus a practically constant force biasing of the pressure plate is provided. Furthermore through the additional device which causes a gradual breakdown of the moment which can be transferred by the friction clutch during a disengagement process it is possible to reduce or minimize the disengagement force path or maximum disengagement force required. This is due to the fact that the device assists the operation, more particularly the disengagement process, of the friction clutch.To this end the device can have axially resiliently pliable means which exert on the operating means and/or on the contact pressure spring and/or on the pressure plate and/or on the counter pressure plate a reaction force which is directed against and connected in series with the force exerted by the contact pressure spring on the pressure plate.

It can be particularly advantageous if the device of the friction clutch is arranged so that during the disengagement process over a part of the axial displacement path of the pressure plate areas biased by the contact pressure spring it causes a gradual breakdown of the moment which can be transferred by the friction clutch or clutch disc.

In many cases the device can advantageously be provided in the force flow between the swivel bearing of the operating means or between the contact pressure spring and the fastening points, such as screws of the housing on the counter pressure plate.

In many other cases however it can also be advantageous if the device is provided in the force flow between the swivel bearing of the operating means or between the contact pressure spring and the friction face of the pressure plate.

An assembly of this kind has been proposed for example through DE-OS 37 42 354 and DE-OS 1 450 201.

In many other cases it can be particularly advantageous if the device is provided axially between two friction linings of the clutch disc mounted back to back, thus is formed by a so-called "lining suspension", eg through lining spring segments provided between the linings. Such devices are known for example from DE-OS 36 31 863.

A further possibility of obtaining a progressive moment build-up and breakdown has been proposed by DE-OS 21 64 297 wherein the flywheel is designed in two parts and the component forming the counter pressure plate is supported axially resiliently relative to the component which is connected to the output shaft of the internal combustion engine.

For the functioning and construction of a friction clutch it can be particularly expedient if the device allows an axial resilient pliability between the clutch components wherein the device is mounted and designed so that when the clutch is opened the force acting on the device is at its lowest and during the closing process of the clutch, thus over the engagement path of the clutch the force acting on the device gradually rises to the maximum wherein this rise preferably takes place only over a part of the closing path or engagement path of the operating means or pressure plate.

It can be particularly advantageous if the device is designed so that the gradual decline or gradual increase in the moment transferable by the friction clutch takes place over at least approximately 40 to 70 % of the operating path of the operating means and/or of the maximum axial path of the pressure plate. The remaining area of the corresponding path is required for the satisfactory separation of the force flow and for the compensation of any possible deformations which may occur on the clutch components, such as in particular the clutch disc, pressure plate and counter pressure plate.

In order to minimize the forces required to operate the friction clutch it can be particularly advantageous if the contact pressure spring has, at least over a part of the disengagement path of the friction clutch, a degressive force-path curve, this means that the contact pressure spring has at least over a part of its compression or deformation path a declining force curve.It can thereby be achieved that during the disengagement process of the friction clutch the spring force of the device counteracts the force of the contact pressure spring so that over a part of the disengagement path the pretensioning or deformation of the contact pressure spring is supported by the spring force of the device wherein at the same time, as a result of the degressive or declining force-path curve of the contact pressure spring present in the disengagement area, the force exerted by the contact pressure spring on the pressure plate or friction linings decreases. The force curve effectively required to disengage the friction clutch arises, if no additional overlying spring actions are present, from the difference between the force curve applied by the device and the force curve of the contact pressure spring. When the pressure plate lifts off from the friction linings and the clutch disc is released through the pressure plate, the remaining disengagement force curve required or the disengagement force required is determined mainly by the contact pressure spring. The force-path characteristic of the device and the force-path characteristic of the contact pressure spring can be matched with each other so that with the release of the clutch disc through the pressure plate the force required to operate the contact pressure spring is located at a comparatively low level. Thus through the spring characteristic or force characteristic of the device approaching or equalling the contact pressure spring characteristic up to release of the clutch disc through the pressure plate only a very small operating force and in extreme cases none at all need be necessary for the contact pressure spring.

Particularly suitable as a contact pressure spring is a diaphragm spring which on one side can be swivelled about a ring-like swivel bearing supported by the housing and on the other side biases the pressure plate. The diaphragm spring can thereby have a ring body from which tongues extend radially inwards to form the operating means. The operating means can however also be formed by levers which are mounted for swivel movement for example on the housing. The contact pressure force for the pressure plate can however also be applied by other types of springs, such as eg coil springs which are mounted in the friction clutch so that the axial force exerted by these on the pressure plate is greatest in the engaged state of the friction clutch and this force reduces during the disengagement process.This can be produced for example by positioning the coil springs inclined relative to the axis of rotation of the friction clutch.

It can be particularly advantageous if the diaphragm spring is supported for swivel movement on the housing between two support pads in order to form a so-called clutch of the depressed type. With clutches of this type the operating means for disengaging the friction clutch are normally biased in the direction of the pressure plate. The invention is however not restricted to clutches of the depressed type but also includes clutches of the pull-type construction wherein the operating means for disengaging the friction clutch are normally biased in the direction away from the pressure plate.

In a particularly advantageous way the friction clutch can have a diaphragm spring which is designed so that it has a sinusoidal force-path curve and which is installed so that in the engaged state of the friction clutch its operating point is provided on the degressive characteristic line area following the first force maximum. It can thereby be particularly advantageous if the plate spring has a force ratio of 1 : 0.4 to 1: 0.7 between the first force maximum and the following force minimum.

It can furthermore be particularly advantageous if the friction clutch can be operated by a disengagement system engaging on the operating means such as eg on the tongue tips of the diaphragm spring wherein the disengagement system can have a clutch pedal which is designed like an accelerator pedal and is mounted in the inside of the motor vehicle. Designing the clutch pedal in this way can be particularly advantageous since the force required for disengaging the friction clutch or force curve can be brought to a very low level so that a better ability for measuring the operating force is possible through a clutch pedal designed like an accelerator pedal.

Through this design of a friction clutch and the possibility connected therewith of reducing the maximum contact pressure spring forces which occur during the service life of the friction clutch it is possible to reduce the components accordingly and to lower their strength so that considerable costs can be saved in production. By reducing the disengagement forces the friction and elasticity losses in the clutch and in the disengagement system are reduced and thus the degree of efficiency of the friction clutch/disengagement system is significantly improved. Thus the entire system can be designed in optimum manner and the clutch comfort substantially improved.

The design according to the invention can be used generally in friction clutches and more particularly in those as known for example from DE-PS 29 16 755, DE-PS 29 20 932, DE-OS 35 18 781, DE-OS 40 92 382, FR-OS 2 605 692, FR-OS 2 606 477, FR-OS 2 599 444, FR-OS 2 599 446, GB-PS 1 567 019, US-PS 4 924 991, US-PS 4 191 285, US-PS 4 057 131, JP-GM 3 25026, JP-GM 3-123, JP-GM 2-124326, JP-GM-1-163218, JP-OS 51126452, JP-GM 3-19131, JP-GM 3-53628.

The use of a friction clutch with an independent or automatic compensation of at least the lining wear - whereby an approximately constant tensioning force of the clutch disc is guaranteed at least throughout the service life of the friction clutch - is particularly advantageous in connection with clutch assemblies wherein the friction clutch, clutch disc and counter pressure plate, such as for example a flywheel, form an assembly unit or module. With an assembly unit of this kind it is advantageous for reasons of costs if the clutch housing is connected with the counter pressure plate through a non-detachable connection such as for example a welded connection or keyed engagement, for example through plastic material deformation. Through such a deformation it is possible to dispense with the fixing means normally used, such as screws.With assembly units of this kind it is in practice not possible to replace the clutch disc or clutch linings when the wear limit is exceeded without destroying components such as for example the clutch housing. By using a wear-adjusting clutch it is possible to design the assembly unit so that this guarantees a satisfactory functioning throughout the entire service life of the vehicle. Thus as a result of the design set out here it is possible to make the wear reserve of the clutch disc and the adjustment reserve of the friction clutch or clutch module with such large dimensions that the service life of the clutch and thus also the service life of the assembly unit can reliably match that of the vehicle.

It can be particularly advantageous if a friction clutch having a wear adjustment device is combined with a so-called twin mass flywheel wherein the friction clutch can be mounted with the interposition of a clutch disc on the flywheel mass connectable with a gearbox and the second flywheel mass can be connected with the output shaft of an internal combustion engine. Twin mass flywheels where the friction clutch can be used are known for example from DE-OS 37 21 712, 37 21 711, 41 17 571, 41 17 582 and 41 17 579.

The entire contents of these applications are incorporated herein by reference so that the features described in these applications can be combined in any way with the features described here. More particularly the clutch housing or clutch cover can be connected to the flywheel mass supporting same through a connection which cannot be released without destruction, as shown and described for example for various embodiments in DE-OS 41 17 579.

By using a friction clutch with a device which compensates at least the lining wear it is possible to optimize the design of the friction clutch, more particularly the energy accumulator which applies the tensioning force for the clutch disc. This energy accumulator can thus be designed so that it applies practically only the clamping force for the clutch disc necessary to transfer the desired torque.

The energy accumulator can be formed by at least a diaphragm spring or by several coil springs. Furthermore the use of a self-adjusting friction clutch in connection with twinmass flywheels is advantageous wherein the rotationally elastic dampener mounted between the two flywheel masses is provided radially outside the clutch disc or the outer friction diameter of the friction surface of the flywheel mass connectable with the gearbox. With twin mass flywheels of this kind the friction diameter of the clutch disc must be smaller than in the case of conventional clutches so that the contact pressure force must be increased corresponding to the ratio of the average friction radii in order to be able to transfer definite engine torque. When using a conventional clutch this would lead to an increase in the disengagement force.By using a wear-adjusting clutch with a progressive breakdown of the torque transferred by the clutch disc over the disengagement path according to claim 1 it is however possible to achieve a drop in the disengagement force whereby an increase in the disengagement force can be avoided or even by designing the friction clutch accordingly it is possible to achieve a reduction in the disengagement force compared to a conventional clutch.

It can thus be ensured that despite reduced outer diameters of the friction lining and the higher contact pressure force thereby required the disengagement force can be kept low.

Through the lower disengagement force the strain on the rolling bearing through which the two flywheel masses rotate relative to each other is also reduced. Furthermore through the wear adjustment the service life of the clutch is increased so that a replacement of the parts, more particularly the clutch disc during the service life of the motor vehicle is no longer necessary. Thus the clutch cover can be fixedly connected to the flywheel mass which can be connected to the gearbox, for example by rivetting or welding. This is particularly advantageous if there is restricted space for installation or restricted contours for the clutch bell which no longer allow the clutch cover to be connected to the flywheel on the gearbox side in the convectional way through screws.

With a friction clutch having an integrated adjustment device for the wear on the linings, in the case of conventional fixings for the clutch unit comprising the friction clutch and flywheel on the output shaft of an internal combustion engine, axial, torsional and twisting vibrations are transferred to the clutch unit which are excited by the output shaft of the internal combustion engine such as in particular the crankshaft. So that the clutch unit and adjustment device are not impaired in their functioning by such vibrations and in particular an undesired adjustment of the wear compensating device is suppressed, the inertia forces of those components which act on this device must be taken into account when designing the adjustment device.In order to avoid these undesired side effects which are caused in particular through axial and twisting vibrations, and to avoid the expense connected therewith for designing an adjustment device for compensating the lining wear, the clutch unit having the adjustment device can be substantially uncoupled from the axial and bending vibrations excited by the output shaft of the internal combustion engine. This can happen in that the clutch unit is connectable by an axially elastic or resiliently pliable component with the output shaft of the internal combustion engine.The rigidity of this component is measured so that the axial and twisting or bending vibrations produced by the output shaft of the internal combustion engine on the clutch unit are at least damped to a certain extent or suppressed through this elastic component so that a satisfactory functioning of the friction clutch, more particularly its adjustment device is guaranteed. Elastic components of this kind are known for example through EP-OS 0 385 752 and 0 464 997 as well as SAE Technical Paper 9 003 91. The contents of these publications are likewise incorporated herein by reference.

By using an elastic component it is possible to avoid undesired wear adjustment caused by axial vibrations of the pressure plate relative to the clutch cover - particularly in the case of a disengaged friction clutch - through flywheel vibrations and/or vibrations of the plate spring.

Vibrations of this kind can lead in the case of clutch assemblies or clutch units without a device at least substantially suppressing these vibrations, such as in particular an axially pliable disc, to a modified adjustment independently of the wear state of the clutch disc wherein the diaphragm spring of the friction clutch could be regulated down in the contact pressure force towards a force minimum whereby the transfer of the desired moment would no longer be guaranteed.

A friction clutch with an independent or automatic compensation can advantageously be used in a drive unit, particularly for motor vehicles, which consists of an automatic or semi-automatic gearbox and a friction clutch which is mounted between a drive motor, such as an internal combustion engine and a gearbox and which can be operated controlled at least in dependence on the operation of the gearbox. The friction clutch is preferably operated full automatically. An automated or fully automated operation of a friction clutch is known for example from DE 03 40 11 850.9 so that reference is made to this specification regarding the method of operation and means required.

With the drive units known up until now with automatic or semi-automatic gearboxes and a conventional friction clutch there have been considerable problems with the operation of the clutch and the design of the actuators required for this, such as eg the piston/cylinder units and/or electric motors. As a result of the comparatively high disengagement forces required with conventional clutches very strong or large dimensioned actuators are required. This means a large structural volume, high weight and high costs. Also actuators made this big are comparatively slow in their response time as a result of the mass inertia. When using servo cylinders a larger volume flow of pressurised medium is moreover required so that even the supply pump has to be made comparatively large in order to ensure the desired operating time for the corresponding friction clutch.In order partly to overcome the aforesaid drawbacks it has been proposed for example through DE-OS 33 09 427 to reduce the operating force for disengaging the clutch through corresponding compensation springs in order then to be able to use smaller dimensioned actuators. Since however the disengagement force in the case of conventional clutches fluctuates a great deal during its service life, that is the disengagement force is relatively low in the new state and rises during the service life with increasing wear on the linings, only a part of the disengagement force normally required can be broken down through a compensation spring.

Taking into account all the tolerances, despite using compensation springs the disengagement capacity of the actuators required will be greater than that for a new conventional clutch. By using a friction clutch with a lining wear compensation in conjunction with a drive unit consisting of a motor and an automatic or semi-automatic gearbox, the disengagement force can be lowered quite considerably compared to the aforesaid prior art, namely directly in the clutch wherein this disengagement force value or disengagement course path of the new clutch remains practically unaltered throughout the entire service life.

Substantial advantages arise here for the design of the actuators since their drive capacity or operating capacity can be kept correspondingly lower wherein even the forces and pressures occurring in the entire disengagement system are correspondingly lower. The losses occurring in the disengagement system as a result of friction or elasticity of the components are avoided or reduced to a minimum.

The invention will now be explained in detail with reference to Figures 1 to 48 in which: Figure 1 is a sectional view through a clutch assembly designed according to the invention; Figure 2 shows the compensation device on an enlarged scale and in cross-section; Figure 3 shows a section in the direction of arrow III of Figure 2; Figure 4 shows the adjustment ring adjoining the disengagement means of the friction clutch in plan view according to arrow IV of Figure 2; Figure 5 is a sectional view along the line V-V of Figure 4; Figure 6 shows the counter adjustment ring used with the clutch assembly of Figure 1 in plan view according to arrow III of Figure 2; Figure 7 is a sectional view according to line VII VII of Figure 6; Figure 8 shows a detail of a varied embodiment of the compensation device shown in Figure 2;; Figure 9 is a sectional view of a further detail of a clutch assembly according to the invention; Figures 10 and 11 show wear adjustment rings which can be used in the clutch assemblies according to the invention, eg according to Figure 9; Figure 12 is a sectional view through a clutch assembly according to the invention; Figure 13 is a partial view in the direction of arrow XIII of Figure 12; Figure 14 shows a further embodiment of a friction clutch according to the invention; Figure 15 is a diagrammatic illustration of a disengagement system for a clutch assembly according to the invention and Figure 16 shows a further embodiment according to the invention of a friction clutch having a brake for the adjustment ring.

Figure 17 is a view of a friction clutch according to the invention; Figure 18 is a sectional view according to line II-II of Figure 17; Figure 19 shows an adjustment ring used with the friction clutch according to Figures 17 and 18; Figure 20 is a sectional view according to line IV-IV of Figure 19; Figure 21 shows a support ring used with the friction clutch according to Figures 17 and 18; Figure 22 shows a cross-section according to the line VI-VI of Figure 21; Figure 23 and 23a show a spring which exerts a turning force on the adjustment ring; Figures 24 to 27 are diagrams with different characteristic lines from which can be seen the interaction of the individual spring and adjustment elements of the friction clutch according to the invention;; Figures 28 and 29 show a further embodiment of a friction clutch according to the invention wherein Figure 28 is a cross-section along the line XIII of Figure 28; Figure 30 is a view of the adjustment ring used with the friction clutch according to Figures 28 and 29; Figures 31 to 33 show details of a further friction clutch with a compensation device; Figures 34 and 35 are diagrams with different characteristic lines from which can be seen the interaction of the contact pressure diaphragm spring and the lining suspension as well as the effect arising therefrom on the disengagement force curve of the friction clutch; Figure 36 is a partial sectional view of a further friction clutch according to the invention; Figure 36a is a partial sectional view in the direction of arrow A of Figure 36; Figure 37 is a sectional view along the line XX1 of Figure 36;; Figure 38 is a partial sectional view of an adjustment ring which can be used with a friction clutch according to Figures 36 to 37; Figures 39 and 40 show further variations of the friction clutches according to the invention; Figure 41 is a view of an adjustment ring which can be used with a friction clutch according to Figures 28 and 29 or 36 to 37; Figures 42 to 45 show additional variations of the friction clutches; Figures 46 to 48 show details of another embodiment of a friction clutch wherein Figure 47 is a partial sectional view according to arrow A of Figure 46 and Figure 48 is a sectional view along arrows B-B of Figure 47.

The clutch assembly shown in Figure 1 has a friction clutch 1 with a housing 2 and a pressure disc 3 which is connected to the housing rotationally secured but axially displaceable to a restricted amount. A contact pressure diaphragm spring 4 tensioned axially between the pressure disc 3 and the cover 2 can swivel about a ring-like swivel bearing 5 supported by the housing 2 and biases the pressure disc 3 in the direction of a counter pressure plate 6, such as for example a flywheel, which is fixedly connected to the housing 2 whereby the friction linings 7 of the clutch disc 8 are clamped between the friction faces of the pressure disc 3 and the counter pressure plate 6.

The pressure disc 3 is connected rotationally secured to the housing 2 by circumferentially or tangentially directed leaf springs 9. In the illustrated embodiment the clutch disc 8 has so-called lining spring segments 10 which guarantee a progressive torque build-up during engagement of the friction clutch 1 by enabling a progressive rise in the axial forces acting on the friction linings 7 through the restricted axial displacement of the two friction linings 7 towards each other. However a clutch disc could be used wherein the friction linings 7 were fitted axially practically rigid on a support disc.

In the illustrated embodiment the diaphragm spring 4 has a ring-like foundation body 4a applying the contact pressure force and from which operating tongues 4b protrude radially inwards. The diaphragm spring 4 is thereby installed so that it biases the pressure disc 3 with radially further outer areas and can be tilted about the swivel bearing 3 with radially further inner areas.

The swivel bearing 5 comprises two swivel support pads 11, 12 between which the diaphragm spring 4 is axially held or clamped. The swivel support pad 11 provided on the side of the diaphragm spring 4 facing the pressure disc 3 is biased with force axially in the direction of the housing 2. To this end the swivel support pad 11 is part of a diaphragm spring or a diaphragm-spring-like component 13 which is supported resiliently with its outer edge area 13a on the housing 2 whereby the swivel support pad 11 moulded on radially inside is axially biased against the operating diaphragm spring 4 and thus also in the direction of the housing 2. The diaphragm spring 13 provided axially between the pressure disc 3 and the operating plate spring 4 has a ring-like area 13b from whose inner edge extend radially inwardly aligned tongues 13c which form the swivel support pad 11.

In order to support the diaphragm spring like component 13 a bayonet-type connection or lock is provided in the illustrated embodiment between the housing 2 and the tonguelike extension arms 13a of the diaphragm spring like component 13.

The diaphragm-spring-like component or diaphragm spring 13 is designed as a sensor spring which produces an at least approximately constant force over a predetermined operating path. The clutch disengagement force acting on the tongue tips 4c is taken up at least substantially by this sensor spring 13 wherein an at least approximately counter balance is always produced between the force produced by the disengagement force on the swivel support pad 11 and the counter force exerted by the sensor diaphragm spring 13 on this swivel support pad 11. By disengagement force is meant the maximum force which is exerted on the tongue tips 4c or on the release areas of the diaphragm spring tongues during operation of the friction clutch 1.

The swivel support pad 12 on the housing side is supported on the housing 2 by an adjustment device 16. This adjustment device 16 ensures that with an axial displacement of the swivel support pads 11 and 12 in the direction of the pressure disc 3 and counter pressure plate 6 no undesired play can take place between the swivel support pad 12 and the housing 2, and between the swivel support pad 12 and diaphragm spring 4 respectively. It is thereby ensured that no undesired dead or idle paths arise during operation of the friction clutch 1 whereby an optimum degree of efficiency and thus satisfactory operation of the friction clutch 1 are provided. The axial displacement of the swivel support pads 11 and 12 occurs in the event of axial wear on the friction faces of the pressure disc 3 and of the counter pressure plate 6 and friction linings 7.

The adjustment device 16 comprises a spring-biased adjustment element in the form of a ring-like component 17 which has circumferentially extending and axially rising run-up ramps 18 which are spread out round the circumference of the component 17. The adjustment element 17 is installed in the clutch 1 so that the run-up ramps 18 face the housing floor 2a.

The adjustment ring 17 is spring-loaded in the circumferential direction, namely in the adjustment turning direction, thus in the direction which as the ramps 18 run up the counter ramps 19 imprinted in the cover base 2a causes an axial displacement of the adjustment ring 17 in the direction of the pressure disc 3, this means in the axial direction away from the radial housing section 2a.

The functioning of the automatic adjustment of the swivel bearing 5 and adjustment device 16 as well as further embodiments of the adjustment device 16 will be described in further detail in connection with Figures 17 to 48.

The clutch assembly comprises a compensation device 20 which ensures that the operating means of the friction clutch 1 formed by the diaphragm spring tongues 4b can be operated play-free in the axial direction and can be displaced over a constant path 21. The compensation device 20 is provided between the release member 22 comprising a release bearing, and the tongue tips 4c. The release member 22 is axially displaceable on a guide tube 23, shown diagrammatically, for operating the friction clutch 1. The guide tube 23 is supported by a gearbox housing (not shown in further detail) and encloses the gearbox input shaft on which the clutch disc 8 can also be fitted secured against rotation.The force required for axially moving the release member 22 is applied by a member 24 which in the illustrated embodiment is formed by a disengagement fork, shown diagrammatically, which can likewise be mounted on the gearbox side. However release members 22 can also be used which are operated hydraulically or pneumatically, thus release members which have a piston/cylinder unit which can be biased by pressurised medium.

The compensation device 20 is shown on an enlarged scale in Figures 2 and 3 and comprises an adjustment element in the form of a ring-like component 25 which is shown in Figures 4 and 5. The ring-like adjustment element 25 has in the illustrated embodiment two radially off-set circumferentially extending and axially rising sets of runup ramps 26, 27 which are each spread out over the circumference of the component 25. As can be seen in particular from Figure 5, the radially inner run-up ramps 26 are off-set circumferentially opposite the radially outer run-up ramps 27, namely by about half a ramp length or one ramp division. The adjustment element 25, as can be seen from Figures 1 and 2, is supported by its end face 25a directly on the tongue tips 4c. The run-up ramps 26, 27 are axially remote from the operating means 4b.The adjustment element 25 is spring-loaded in the circumferential direction, namely in the adjustment turning direction, thus in the direction which as the ramps 26, 27 run up on the counter ramps 28, 29 of the support ring 30 shown in more detail in Figures 6 and 7, causes an axial displacement of the adjustment ring 25 in the direction of the pressure disc 3, this means in the axial direction away from the release member 22.

As can be seen from Figures 6 and 7 the counter run-up ramps 28, 29 likewise form two sets of run-up ramps off-set from each other both radially and circumferentially. The ramps 26, 27 of the adjustment element 25 and 28, 29 of the support ring 30 are matched with each other and engage axially in each other. Through the circumferentially offset ramps it is ensured that a satisfactory central guide is provided between the adjustment element 25 and support ring 30. As can be seen in particular from Figure 2, the two components 25 and 30 of the compensation device 20 are set axially inside each other. The installation angle 31 (Figure 7) of the counter run-up ramps 28, 29 of the support ring 30 correspond to the angle 32 (Figure 5) of the run-up ramps 26, 27 of the adjustment element 25.The support ring 30 can be connected rotationally secured to the housing 2, but can be displaced relative to same over the clutch actuation path 21 axially to a restricted amount. The axial restriction is provided by radial areas 33 of the support ring 30 which in the engaged state of the friction clutch 1 adjoin the radially inner areas of the cover base 2a. This contact is caused by the resiliently biased operating means 4b. During disengagement of the friction clutch 1 the path restriction is ensured through a shaped sheet metal member 34 which is provided on the side of the support ring 40 remote from the operating means 4b and can be biased by the release member 22 in the diameter area 35. This shaped sheet metal part 34 likewise has radial areas 36, which during disengagement of the friction clutch 1 can come to adjoin the radially inner areas of the cover base 2a.

In the illustrated embodiment the adjustment ring 25 as well as the support ring 30 is made from heat-resistant plastics such as eg from thermoplastics which can be additionally strengthened through fibres. These components can thereby be made simply as injection moulded parts.

The run-up ramps 26, 27 and counter run-up ramps 28, 29 are designed circumferentially so that they allow at least a turning angle between the two components 25 and 30 which guarantees throughout the entire service life of the friction clutch 1 an adjustment of the wear which occurs on the friction faces of the pressure disc 3 and counter pressure plate 6 as well as on the friction linings 7. This adjustment angle can be in the order of between 30 degs. and 90 degs. depending on the design of the run-up ramps. In the illustrated example this turning angle which is marked by reference numeral 36 in Figure 3 is in the order of 75 degs. The installation angle 31 and 32 of the ramps and counter ramps respectively can be in the order of between 6 degs and 14 degs, preferably in the order of 8 degs, wherein the actual angle 31 and 32 of the ramps and counter ramps respectively changes over the radial extension of these ramps since the same height difference must be bridged for any given turning angle. This means that the ramp angle 31 and 32 becomes smaller with an increasing diameter.

The force biasing in the circumferential direction required for the adjustment of the element 25 is guaranteed by means of energy accumulators which in the illustrated embodiment are formed by two coil springs 38, 39 arranged in a curve and tensioned between the support ring 30 and the adjustment element 25. These coil springs 38, 39 are supported on the support ring 30 which is rotationally secured with the cover 2 and turn the adjustment ring 25 as soon as the operating means or diaphragm spring tongues 4b move away axially from the cover base 2a and release member 22 as a result of wear on the linings. As can be seen in particular from Figures 3 and 6 the coil springs 38, 39 are each contained in a socket 40, 41 of the ring 30 extending channel or toroidal fashion in the circumferential direction.As can be seen from Figure 2 a socket 40 of this kind which matches in cross-section the windings of the energy accumulators 38, 39 extends over more than half the circumference of the crosssection of a spring 38 or 39 wherein as can be seen from Figures 3 and 6 a slit-like opening 42, 43 remains on each side facing the operating means 4b and a slit-like opening 44, 45 remains on each side of the support ring 30 remote from the operating means 4b. The springs 38, 39 are secured axially relative to the support ring 30 by the faces defining the sockets 40, 41. In order to thread in the coil springs 38, 39 the sector-shaped sockets 40, 41 each have a thread-in area 46, 47 which has a radial insert width which corresponds at least to the outer diameter of the windings of the coil springs 38, 39.The energy accumulators 38, 39 can be pushed through these thread-in areas 46, 47 inclined into the sector-shaped sockets 40, 41. After the coil springs 38, 39 which are still relaxed have entered the sector-shaped sockets 40, 41 the adjustment element 25 is assembled with the support ring 30. To this end the axial noses 48, 49 provided on the adjustment ring 25 and which simultaneously form the biasing areas and support areas for the coil springs 38, 39 are each inserted into an axial slit area 50, 51 circumferentially adjoining the thread-in areas 46, 47 whereby the biasing areas 48, 49 come to lie against an end area of the relaxed coil springs 38, 39. The relaxed state of an energy accumulator 38 or 39 can be seen in Figure 3 and is marked by numeral 39a.The other end area of the coil springs 38, 39 is supported on the bottom 53,53a of the sector-shaped sockets 40, 41 which is present in the circumferential direction. The springs 38, 39 can be pretensioned by turning between the adjustment ring 25 and support ring 30. After a certain relative turning angle which is larger than the angular extension of the thread-in areas 46, 47 the biasing areas 48, 49 of the adjustment ring 25 each come to lie axially above an end area of a slit 44, 45 so that the adjustment ring 25 and supporting ring 30 can be moved towards each other until the run-up ramps 26, 27 and counter run-up ramps 28, 29 contact one another. The slits 44, 45 and the axial noses 48, 49 are matched with each other so that an axially active snap connection is provided between the two components 25, 30.To this end the axial noses 48, 49 have at their end areas a hook-like section 48a which can adjoin radially aligned areas of the support ring 30. The springs 38, 39 are brought to their tensioned angular length 54 corresponding to the new state of the friction clutch 1 through an additional relative rotation between the two parts 25 and 30 corresponding to angle 37 (Figure 3). The two parts 25, 20 can then be secured in this position through means not shown. These means can comprise for example a keyed engagement lock which acts between the two components 25 and 30 and which can be removed after fitting the friction clutch 1 on the counter pressure plate 6 whereby the compensation device 20 is activated. The possible adjustment angle for compensating in particular the wear on the lining corresponds to the turning angle marked in Figure 3 by numeral 37.After this turning angle 37 the axial noses 48, 49 come to adjoin the end areas of the slits 44, 45 present in the adjustment direction of the ring 25. A tensioned position of a coil spring 38, 39 corresponding to this position is marked in Figure 3 by numeral 38a.

In the new state of the friction clutch 1 the axial cams 26, 27 and 28, 29 forming the run-up ramps and counter run-up ramps engage furthest in each other axially. This means that the rings 25 and 30 which lie one on top of the other require the smallest axial space.

In the illustrated embodiment the restriction of the operating path in the disengagement direction of the friction clutch 1 is ensured through the shaped sheet metal part 34. According to an embodiment not shown the stop areas required for this and which interact for example with the cover 2 can also be provided on the release member 22, namely on the bearing ring which turns with the friction clutch or on a component which is connected to the ring.

The axial restriction of the operating path of the friction clutch 1 in at least one of the axial directions could also be formed by at least one axial stop provided on the guide tube 23 for the release member 22.

Furthermore the release member 22 could act directly on the operating means 4b and a corresponding compensation device could be provided between the release member 22 and the disengagement means 24.

It is expedient if the release member 22 is biased in the direction of the operating means 4b with a pretension which does not impair the function of the friction clutch 1 and the compensation device 20.

As can be seen from Figures 2 to 4, the adjustment ring 25 has radially inside cams 55 which form engagement areas for a turning and restraint means which can when necessary adjoin the housing 2 or the support ring 30 on the other side to secure against rotation. Such restraint means can be provided during the manufacture and assembly of the friction clutch 1 or compensation device 20 and can be removed after the friction clutch 1 is fitted on the flywheel 6.

The detail shown in Figure 8 represents a varied embodiment of the lower half of the compensation device 20 shown in Figures 1 and 2. With the variation according to Figure 8 the axial restriction between the compensation device 120 and the housing 102 is carried out in the engaged state of the friction clutch by hook-like axial extension arms 133 which are formed integral with the shaped sheet metal part 134. The extension arms 133 are moulded on the outer edge of the sheet metal part 134 which serves as the pressure member, and engage axially through the cover 102. At their free end facing the diaphragm spring 104 the extension arms 133 have radially outwardly aligned areas 133a which engage radially behind the cover 102 on its side facing the diaphragm spring 104. Through a design of this kind it is ensured that the axial forces exerted by the diaphragm spring 104 on the compensation device 120 can be supported by the pressure member 134 which is made from sheet metal so that larger axial forces can be taken up by the compensation device 120 than with the compensation device 20 according to Figure 2 wherein the stops are formed by the areas 33 of the support ring 30 which is made from plastics. Such axial forces on the compensation device 20 or 120 can occur inter alia during transport, thus when the friction clutch is not fitted, since in this state the main diaphragm spring 4 or 104 is axially supported by spring tongues on the support ring or plastics compensation element 30, 130, respectively.

The pressure member 134 which is made from sheet metal can have at least two, preferably three or more hook-like extension arms 133 which are preferably spread out symmetrically or uniformly round the circumference. The sheet metal thickness of the pressure member 134 can be made according to the axial forces which are to be supported.

The ring 130 which is made from plastics is connected rotationally secured to the pressure member 134. As in Figure 2 the pressure member or shaped sheet metal part 134 likewise has radially on the outside, areas 136 which, viewed circumferentially, extend between the hook-like extension arms 133 and serve to restrict the disengagement path or to avoid unacceptably high over-movement by stopping against the housing 102.

The detail of a friction clutch 201 shown in Figure 9 has a substantially similar construction to the right lower area of the friction clutch 1 according to Figure 1. Figure 9 shows in part the clutch housing 202, the swivel bearing 205 for the diaphragm spring 204, the adjustment device 216 and the compensation device 220. With regard to the functioning of the adjustment device 216 and the compensation device 220 reference is made to the description of Figures 1 to 8 and to the German Patent Applications P 43 06 505.8 and P 42 39 289.6 whose contents are expressly to be regarded as integrated into the present application.

With the varied embodiment shown in Figure 9 an antirotation device 260 is provided for the adjustment element in the form of an adjustment ring 217.

The anti-displacement or -rotation device 206 ensures when the friction clutch 201 is not fitted a definite position of the adjustment element 217 relative to the remaining components such as in particular the housing 202. More particularly it can be guaranteed through the anti displacement device 260 that in the new state of the friction clutch 201 the adjustment element 217 can be held in its retracted position, thus practically in the zero position in which no adjustment takes place, and this is achieved even though the diaphragm spring 204 does not bias the adjustment ring 217 in the area of the swivel or support pad 212.The latter is due to the fact that when the friction clutch 201 is not fitted or when the friction clutch 201 is provided for despatch, the main diaphragm spring 204 is supported axially by its spring tongues on the compensation device 220, as can be seen in connection with Figures 1 and 2. As a result of this support the main diaphragm spring 204 presses the force sensor in the form of a diaphragm spring 213 axially away from the housing 202 or adjustment ring 217 whereby an axial tensioning of the adjustment ring 216 is no longer ensured in the direction of the housing 202. Without the anti-rotation device 260 the ring 217 could be displaced.When fitting the friction clutch 201 on the output shaft of an internal combustion engine the ring 217 would thus not have the desired retracted position which ensures adjustment of the wear which occurs in particular on the friction linings of the clutch disc. In Figure 9 the position of the components which correspond to the friction clutch fitted on a flywheel is shown in solid lines. The position of the diaphragm spring 204 and the sensor spring 213 which corresponds to a new friction clutch not yet fitted is shown by dotted lines.

As can be seen, in the non-fitted state of the friction clutch 201 an axial space or air gap is present between the adjustment element 217 or the ring-like support 212 and the diaphragm spring 204.

The anti-adjustment device 260 for the adjustment device 216 which (260) is provided inter alia for transporting the friction clutch 201 has at least one security element 261 which at least in the non-fitted state of the friction clutch 201 and where applicable in the engaged state of the fitted friction clutch 201 is held rotationally secured relative to the housing 202 and interacts with same in order to secure the adjustment element 217 against rotation. The security element 261 can have for example, as shown in Figure 10, individually radially aligned arms 262 which are connected radially on the outside fixed to the adjustment ring 217 and can be clamped radially on the inside between the stop 233 provided on the side of the housing 202 facing the diaphragm spring 204, and the housing 202.A forcelocking connection is thereby provided between the adjustment element 217 and the housing 202. The arms 262 can be formed like leaf springs and are connected together radially on the inside by a ring-like area 263. In the illustrated embodiment the arms 262 are screwed to the adjustment element 217, but these arms 262 could also be rivetted to the adjustment element 217 or even have areas which are embedded in the plastics forming this element in order to provide the rotationally secured connection with the adjustment element 217.

The compensation device 220 or its restriction stop 233 forms in conjunction with the housing 202 and the axially clamping areas of the security element 261 a brake or clutch for the adjustment element 217 which is active when the friction clutch 201 is not operated.

During disengagement of the clutch the brake action or clamping of the security element 261 or of the tabs 262 is lifted so that the adjustment element or adjustment ring 217 can adjust where necessary.

It is advantageous if the security element 261 or the leafspring like tabs forming same have in the axial direction a low spring rate or spring stiffness but are comparatively rigid or spring-stiff in the circumferential direction.

With the embodiment of an adjustment ring 317 shown in Figure 11 the axially elastic tabs 362 are injection moulded directly on the plastics ring 317. The tabs 362 can be connected radially on the inside by a circular ring shaped area in a similar way to that shown in connection with Figure 10.

The clutch unit or friction clutch 401 shown in Figures 12 and 13 has a housing formed by a sheet metal cover 402, a pressure disc 403 connected rotationally secured to the housing but axially displaceable to a restricted amount, as well as a contact pressure diaphragm spring 404 clamped between this disc and the cover 402. The contact pressure diaphragm spring 404 is mounted as a double-armed lever relative to the housing 402 whereby it is held for tilting or swivel movement in a swivel bearing 405. With the areas which lie radially further out relative to the ring-like swivel bearing 405, the diaphragm spring 404 biases the pressure plate 403 in the direction of the friction linings 407 of a clutch disc 408 which can be clamped between the pressure plate 403 and a flywheel. The torque transfer between the pressure disc 403 and the cover 402 is carried out by leaf springs 409 which can be pretensioned in the direction of lift of the pressure disc 403 from the friction linings 407.

The diaphragm spring 404 has a ring-like foundation body 404a as well as radially inwardly directed tongues 404b extending therefrom.

The swivel bearing 405 comprises two swivel support pads 411, 412 between which the diaphragm spring 404 is axially held and clamped. The swivel support pads 411 and 412 are arranged and designed similar and have the same function as the swivel support pads 11 and 12 described in connection with Figure 1. With regard to the components acting on the swivel support pads 411, 412, as well as the automatic adjustment function of the swivel bearing 405, reference is made to the description of Figure 1 and of Figures 17 to 48.

The disengagement means of the friction clutch 401 formed by the diaphragm spring tongues 404b can be axially operated by a disengagement device 422 whereby the conicity of the diaphragm spring 404 can be changed. The disengagement device 422 can have a compensation device 420, similar to that described in connection with Figures 1 to 7. With clutch release systems having a self-adjusting release bearing, such a compensation device 420 is not however required. With a disengagement system of this kind the disengagement device 422 can be connected to the release bearing ring which rotates with the clutch 401 at least during the disengagement process.

In order to prevent an unacceptably large release path of the clutch release means formed by the diaphragm spring tongues 404b, path restrictor means 436 are provided on the clutch 401 or on the housing 402 for the diaphragm spring tongues 404b. The path restrictor means 436 restrict the swivel path or swivel angle of the diaphragm spring 404 by axially supporting the diaphragm spring tongues 404b and thus by axially absorbing the disengagement force acting on the disengagement device 422.

In the illustrated embodiment the path restrictor means 436 are formed by a ring-like stop area 436 formed by the radially inner sections of the cover 402 and adjoined by the tongue tips 404c after a predetermined axial path 421. The ring-like stop area 436 is designed so that it lies at least approximately on the disengagement area of the diaphragm spring tongues, thus that diameter on which the disengagement device 422 comes to adjoin the diaphragm spring tongues 404b. The stop area 436 is mounted axially between the spring tongues 404b or the spring tongue tips 404c and the clutch disc 408.

The ring-like stop area 436 is connected by radially aligned ribs or webs 437 to the cover body 402a. As can be seen from Figure 13, with the illustrated embodiment six such webs are provided. In many cases however only three such webs need only be provided. With clutch designs where particularly large disengagement forces are required more webs can be provided, for example nine.

Starting from the cover base 402b or from the cover body 402a the webs 437 run radially inwards axially inclined in the direction of the pressure disc 403 or clutch disc 408 respectively. The stop area 436 is off-set axially in the cover chamber relative to the cover base 402b. The spring tongues 404b engage through the openings 438 formed between the ring-like stop area 436, the radially further outer lying cover body 402a and the connecting ribs 437. With the illustrated embodiment the diaphragm spring tongues 404b are for this purpose angled or installed radially inwards over a part of their length in the axial direction opposite the path of the webs 437. As can be seen from Figure 13, the diaphragm spring tongues 404b form groups of three which are each associated with an opening or recess 438. Between the individual groups of three there are slits 439 for holding the webs 437.The slits 439 and webs 437 are thereby matched with each other so that a satisfactory swivelling of the diaphragm spring 404 is possible.

The diaphragm spring tongues 404b are threaded into the openings 438 during the fitting of the friction clutch 401.

To this end the diaphragm spring 404 has in the relaxed state which is shown in dotted lines in Figure 12, an inner diameter 440 in the area of the tongue tips 40c which is greater than the outer diameter 441 of the ring-like stop area 436. The diaphragm spring 404 can thereby be pushed, substantially fully relaxed, by its spring tongues 404b axially into the openings 438 of the cover 402. During the assembly of the friction clutch 401 or at the latest during the fitting of the friction clutch 401 eg on a flywheel, the diaphragm spring 404 is swivelled whereby the inner diameter 440 restricted by the diaphragm spring tongue 404b is reduced. With the friction clutch mounted on a flywheel the diaphragm spring 404 has its operating position, and the tongue tips 404c define an inner diameter 442 which is smaller than the outer diameter 441 of the stop area 436.

The diaphragm spring 404a is held for swivel movement on the housing and the tongues 404 are designed so that even after covering the swivel path 421 the inner diameter defined by the tongues is smaller than the outer diameter of the stop area 436.

The maximum possible axially restricted operating path 421 is dimensioned so that the clutch 401 after reaching the maximum permissible wear on the linings 407 at least still has the full ideal release path which is required for satisfactory functioning, thus satisfactory separation of the clutch assembly 401. The clutch 401 or the sensor spring 413 and adjustment device 416 which guarantee an automatic lining wear compensation in the clutch are designed so that in the new state of the friction clutch 401 a faulty axial displacement of the swivel bearing 405 does not occur even when completely travelling the path 421.

The action or interplay between the stop area 436 and the diaphragm spring tongues 404b will now be explained and demonstrated in the following with reference to a numerical example: The prescribed disengagement path of the friction clutch 401 taking into account the existing tolerances amounts to 8.4 to 10 mm. The clutch 401 is designed so that in the new state a faulty axial adjustment of the swivel bearing 405 would only be possible with disengagement paths above 14 mm.

The stop 436 is designed and positioned so that in the new state of the friction clutch the areas coming to adjoin the stop 436, namely the tongue tips 404c can cover an axial path 421 of 12.5 mm. When the diaphragm spring tongues adjoin the stop 436 and apply the maximum disengagement force the cover can spring in axially by about 0.5 mm so that overall a maximum axial path 421 of 13 mm is possible.

Assuming that a maximum lining wear of 3 mm is possible on the linings 307 throughout the service life of the friction clutch 401 the diaphragm spring is moved through axial displacement of its swivel bearing 405 by this 3 mm in the direction of the clutch disc. The maximum possible disengagement path 421 is thus reduced from about 13 mm to about 10 mm so that the clutch at the end of the service life still lies within the required disengagement tolerance of 8.4 to 10 mm.

With the illustrated embodiment the stop 436 is formed integral with the cover 402. This stop could however also be formed by a separate component which is connected to the cover 402. The webs 437 can also be formed by separate components or can be integral with the stop 436 which is designed as an independent component.

The clutch unit 401 shown in Figures 12 and 13 furthermore has a device or means which cause an increase in the support force on the diaphragm spring 404 during operation of the clutch unit 401 at least in parts of the speed range in which the clutch unit 401 rotates during use. Through this rise in support force it is possible to prevent as a result of breakdown factors appearing at least in a certain speed section during operation of the clutch unit 401 an inadmissible adjustment appearing as a result of an undesired axial expansion or yielding of the sensor device in the form of the senspr spring 413 interacting with the swivel support pad where the axial expansion is not due to wear on the friction linings 407.

In Figure 12 means 450 which are dependent on speed or centrifugal force are provided in order to increase the axial force acting on the rolling support pad 411. The means 450 dependent on centrifugal force are formed by tongues 450 moulded onto the outer periphery of the sensor diaphragm spring 413 and standing up axially in the direction of the cover 402. As can be seen from Figure 12a the diaphragm-spring-like sensor spring 413 has radially outwardly aligned tongue-like extension arms 413a which as apparent from Figures 12 and 13 are supported axially on the cover 402. Between the extension arms 413a and the areas 451 of the cover 402 axially supporting same is a bayonettype connection or lock 451.The bayonet-type connection 452 is designed so that by axially bringing together the sensor spring 413 and the housing 402 and through the following relative rotation between these two components the extension arms 413a come to lie axially over the support areas 451 of the housing 402. When assembling the sensor spring 413 and cover 402 and before rotating these two components the sensor spring 413 is first tensioned axially elastically and after rotation is relaxed whereby the extension arms 413a are supported with pretension on the cover 402. As can be seen from Figure 12a a tongue 450 is provided either side of a radial extension arm 413a.With a rotating clutch unit 401 and as a result of the centrifugal force acting on the tongues 450 a force is produced which is superimposed on the force applied by the sensor spring 413 as a result of its pretension, thus is added to same whereby the support force for the operating diaphragm spring 404 is increased in the area of the swivel support pad 411. This force which is produced additionally on the swivel support pad 411 by the tongues 450 becomes greater with increasing speed.This force increase can however be restricted in that from a specific speed level the tongues 450 are deformed or swivelled as a result of the centrifugal force acting on same so that they are supported radially outwards on the housing 402 so that then no or practically no further increase of the extra supporting force produced by the means 450 dependent on centrifugal force is present in the area of the swivel support pad 11.

When considering the axial force ratios or force equilibrium between the swivel support pad 411 and the diaphragm spring 404 it is furthermore necessary to take into account the leaf-spring-like torque transfer means. These leaf-spring like torque transfer means 409 can be pretensioned between the housing 402 and the pressure disc 403 so that throughout the entire service life of the clutch unit 401 the pressure disc 403 is tensioned with force by the torque transfer means 409 axially against the diaphragm spring 404. Thus the axial force applied by the torque transfer means 409 counteracts the force exerted by the diaphragm spring 404 on the pressure disc 403 and thus is added with the axial force applied by the sensor spring 413 on the diaphragm spring 404 wherein these two forces then axially counteract the disengagement force acting on the tongue tips 404c.The actual sensor force which counteracts an axial displacement of the diaphragm spring 404 when the clutch unit 401 is not rotating is thus formed by the resulting force produced from the torque transfer means 409 and sensor spring 413 and acting on the diaphragm spring 404. With rotation of the clutch unit 401 this resulting force is still superimposed on a speed- or centrifugal force- dependent force which is produced by the tongues 450.

With a clutch disc 408 which has a device eg in the form of a lining suspension 453 which during operation of the friction clutch unit 401 guarantees over a part of the lift path of the pressure disc 403 a gradual breakdown or gradual build-up of the torque which can be transferred by the clutch disc 408, this device 450 assists up to release of the friction linings 407 or clutch disc 408 through the pressure disc 403, the axial support of the diaphragm spring 404 relative to the adjustment element in the form of an adjustment ring 417. It is thereby guaranteed that at least approximately until the friction linings 407 are released the adjustment ring 417 remains axially tensioned between the diaphragm spring 404 and the housing or cover 402 and thus no adjustment can take place.If during disengagement of the clutch unit 401 the pressure disc 403 is lifted from the linings 407 then with a clutch unit without the tongues 405 only the resulting force produced through the leaf spring like torque transfer means 409 and the sensor diaphragm spring 413 still acts as the axial tensioning force on the main diaphragm spring 404. This resulting sensor force counteracts the disengagement force introduced onto the tongue tips 404c. In certain speed areas, more particularly at high engine speeds, vibrations excited by the engine can occur for example which cause an axial vibration of the pressure disc 403.If the pressure disc 403 vibrates axially then this pressure disc 403 can temporarily be lifted from the main or diaphragm spring 404 whereby the resulting sensor force temporarily drops since the axial force then produced by the leaf spring like torque transfer means 409 no longer acts on the diaphragm spring 404. This has the result that the force ratio required for an intended adjustment of the device 416 between the diaphragm spring 404 or the disengagement force acting on same and the resulting supporting force acting on this diaphragm spring 404 is destroyed, namely with such operating states of the clutch unit 401 the axial supporting force acting on the diaphragm spring 404 is too low whereby the clutch is prematurely and undesirably adjusted and thus the operating point of the diaphragm spring 404 is moved in the direction of the diaphragm spring minimum.Furthermore with specific operating states of the engine, more particularly with higher engine speeds, particularly high crankshaft circumferential accelerations occur which as a result of the inertia of the adjustment ring 417 produce circumferential forces which as a result of the adjustment ramps 418, 419 acting between the adjustment ring 417 and the housing 402 can produce an axial component on the diaphragm spring 404 which is directed against the resulting sensor force whereby an undesired adjustment can likewise result. As a result of the vibrations occurring the friction engagement between the run-up ramps 418, 419 can also be reduced so that the axial force which is produced by the adjustment spring 417a acting on the adjustment ring 417 in the circumferential direction and which acts on the diaphragm spring 404 is increased whereby an undesired adjustment is likewise assisted.

In order to overcome the said disadvantages of a clutch unit 401 without aids 450 which are dependent on centrifugal force, in the illustrated embodiment according to Figures 12 to 13 tongues 450 are provided which are dependent on centrifugal force. These centrifugal-force-dependent means 450 compensate the harmful effects which are dependent on speed in that they produce an assisting force which rises in dependence on speed or centrifugal speed and which is connected in parallel with the force produced by the sensor spring 411.

The centrifugal force dependent means can thereby be designed so that an adjustment in the clutch unit 401 which is required for lining wear is only possible at stationary or low speeds of the clutch unit 401. When the clutch unit 401 is rotating or above a speed at which critical vibrations can occur the adjustment device 416 can be practically blocked.

With the embodiment of a friction clutch 501 illustrated in Figure 14 the sensor spring 513 is mounted radially inside the diaphragm spring swivel bearing 505. The sensor spring 513 has a ring-like foundation body 513a from which extend radially inwardly directed tongues 513b. Through these tongues 513b the sensor spring 513 is supported on the ringlike stop area 536 which is arranged and designed like the ring-like stop area 436 according to Figures 12 and 13. The sensor tongues 513b are supported on the side of the stop area 536 facing the diaphragm spring tongue tips 504c.

Radially outwards the foundation body 513a likewise has tongues 513c which adjoin the diaphragm spring 504 in order to support it axially.

Fitting the sensor spring 513 on the cover 502 can be carried out by deforming the spring conically in the tension direction until the inner diameter 540 defined by the inner tongues 513b is greater than the outer diameter 541 of the stop area 536. The support tongues 513b can thereby be fed into the openings 538 of the cover 502 in a similar way to that described in connection with the tongues 404b and the openings 438 of Figures 12 and 13. After the tongues 513b have been inserted into the openings 538 the sensor spring 513 can be relaxed whereby the inner end areas of the tongues 513b are moved to a smaller diameter and come to rest against the stop area 536.

A further possibility for fitting the sensor spring 513 on the cover 502 consists in lifting at least parts of the inner tongues 513b axially in the direction of the cover 502 so that they define a larger inner diameter 504 than the outer diameter of the stop area 536. After the sensor spring or tongues 513b have been inserted into the openings 538 of the cover, the tongues 513b can be bent back so that they come to rest with their radially inner areas pretensioned against the stop area 536. By bending back the tongues 513b they are swivelled from the position shown in dotted lines in Figure 14 into the position shown in solid lines through plastics deformations in the diaphragm spring material. For plastically deforming the sensor tongues 513b they can be axially supported on the tongues 504b or the tongue tips 504c of the diaphragm spring 504.For bending the tongues 513b it is possible to use a tool which supports the tongues 504b of the operating diaphragm spring 504 from above and biases the sensor spring tongues 513b from below, namely approximately on the diameter area on which the tongues 513b are bent.

The stops 436 and 536 for defining the disengagement path or swivel angle of the diaphragm springs 404 and 504 have the advantage that they are integrated in the corresponding clutch 401 or 501 and are active in the area of the diaphragm spring tongues 404b or 504b whereby it can be ensured that when the diaphragm spring tongues 404b, 504b contact the stops 436, 536 the diaphragm spring tongues cannot or only insignificantly be deformed in the axial direction. It can thereby also be ensured that the diaphragm spring tongues 404b, 504b themselves do not come to adjoin a component of the clutch disc 408 in their position corresponding to the disengaged state of the friction clutch 401, 501. In Figure 12 the position of the diaphragm spring 404 corresponding to the disengaged clutch is shown in dotted lines and marked by numeral 450.It can thus be avoided that in the disengaged state of the friction clutch 401 the diaphragm spring tongues 404b can come to adjoin or slip on the clutch disc 408 rotating relative to this clutch 401.

With the illustrated embodiments according to Figures 12 to 14 the tops 436, 536 are provided in the area of the diaphragm spring tongue tips 04c, 504c. These stops can however also be formed differently and can be off-set radially outwards relative to the inner tongue tips 404c, 504c. With a design of this kind it is however expedient if the radial lever arm provided between the tongue tips 404c, 504c and the stops which are then radially further outwards is selected so that an inadmissible bending of the diaphragm spring tongues 404b, 504b does not take place as a result of the disengagement force acting on same and the support through the stops.

The said oversized or unacceptably large disengagement path can be caused by the disengagement system or operating system which acts on the clutch operating means which are formed by the diaphragm spring tongues in the case of the embodiments described and illustrated. This operating system normally contains a release bearing which acts on the operating means of the friction clutch, an operating member such as a clutch pedal, and a force transmission train provided between the release bearing and the operating member. This force transmission train can have a receiver and a transmitter cylinder.With disengagement systems having a transmitter and receiver cylinder an unacceptable disengagement path lying over the normal disengagement path can be caused in that as a result of a rapid engagement and disengagement of the friction clutch the receiver cylinder cannot return fast enough, that is it does not come to the end position so that with a second disengagement which rapidly follows the receiver cylinder itself does indeed cover a path corresponding to the normal disengagement path but an overall disengagement path for the clutch is produced which corresponds to the sum of the normal disengagement path and the remaining resetting path which has not taken place. An overall operating path for the friction clutch can thereby appear which considerably exceeds the maximum permissible disengagement path proposed for the clutch.

This means that the over-movement reserve provided for operation in the friction clutch can thus be exceeded.

Through the measures and stops 36, 436, 536 according to the invention it is thus possible to prevent an unacceptably large disengagement path or over movement when operating the friction clutches wherein however the normal disengagement path provided which is required for the service life of the clutch is guaranteed.

Thus according to the invention, in the case of clutches in general and clutches having an adjustment device compensating at least the wear of the friction linings of the clutch disc, in particular, at least one stop can be provided in the clutch operating train which avoids an overmovement of the clutch operating means when operating the latter. A stop of this kind can restrict for example the disengagement path of the release bearing of the swivel path of the plate spring. Such a stop can however also be provided at another place. Furthermore the operating path of the friction clutch can be restricted to a defined constant value by providing a defined restriction such as a stop in both the disengagement direction and in the engagement direction.

A restriction of this kind can advantageously take place in the area of the disengagement bearing since in this area the tolerance chain between the operating means, such as the diaphragm spring tongues of the friction clutch and the component restricted to a specific path is small.

With the presence of a restriction of this kind or a stop of this kind disengagement is carried out against a practically rigid restriction whereby an overstrain can occur on the components, more particularly those of the disengagement system or in the case of foot-operated systems it can also be undesirable for the operator. According to a further development of the invention a resiliently or elastically pliable device and/or a device restricting the pressure in the disengagement system is therefore provided in the operating train of the friction clutch wherein this device has a pretensioning force or requires a minimum deformation force or opening force which is at least somewhat greater than the maximum force necessary or the maximum pressure necessary to operate the clutch.It is thereby ensured that when the stop becomes active the clutch pedal can be pressed down further or the operating motor can carry out a movement up to a defined position. The pliable means provided in the operating train of the friction clutch can be provided between the clutch operating means and the release bearing or between the latter and the disengagement operating means such as eg the clutch pedal or disengagement motor.

In Figure 15 a disengagement system 601 is shown wherein various possibilities are given for arranging a device for restricting the maximum force which can be exerted by the release bearing 622 on the clutch operating means 604 and/or the clutch housing 602. In Figure 15 an axial stop 636 is provided which after a determined movement of the release bearing 622 comes to adjoin the housing 602, in a similar way to that described in connection with the stop 36 according to Figures 1 and 2. The disengagement path restriction could however also be carried out in a different way, such as eg described in connection with Figures 12 to 14. The disengagement system 601 has a transmitter cylinder 650 and a receiver cylinder 651 which are connected by a pipe 652.The piston 653 of the receiver cylinder 651 supports the release bearing 622 and is housed axially displaceable in a housing 654. The pressure chamber 655 is supplied with hydraulic medium such as eg oil through the pipe 652. The cylinder unit 650 has a housing 656 which forms in conjunction with the piston 657 provided therein a pressure chamber 658 of variable volume. The pressure chamber 658 is connected with the pressure chamber 655 by the pipe 652. A resetting spring 659 is provided in the pressure chamber 658 for the piston 657. The piston 657 is axially movable by a clutch pedal or an operating motor such as eg an electromotor or pump. The pressurised medium circuit of the disengagement system 601 is connected with a pressurised medium reservoir 660. The transmitter cylinder 650 is preferably connected by a pipe 661 directly to the pressurised medium reservoir 660.

In order to restrict the clutch operating force acting on the disengagement means 604 and/or the housing 602, in the embodiment according to Figure 15, at least one device is provided in the pressurised medium circuit of the disengagement system 601 which restricts the pressure arising in the pressurised medium circuit during operation of the friction clutch, to a definite value. In the embodiment according to Figure 15 this device is formed by at least a pressure restrictor valve. In Figure 15 there are various arrangements shown for a pressure restrictor valve of this kind. A pressure restrictor valve 662 of this kind can be provided for example in the pipe system 652 and can have a return line 663 into the pressurised medium reservoir 660.However instead of the pressure restrictor valve 662 it is also possible to provide a pressure restrictor valve 664 which can be supported by the housing 654 or can even be integrated in same, which is connected with the pressurised chamber 655 and is connected by a return line 665 to the pressurised medium reservoir 660.

Figure 15 shows a further alternative embodiment for arranging a pressure restrictor valve 666. The pressure restrictor valve 666 is connected with the pressure chamber 658 of the transmitter cylinder and can be supported by the housing 656 or can be integrated in same. Furthermore the pressure restrictor valve 666 has a return into the pressurised medium container 660. To this end the pressure restrictor valve 666 can have its own pipe or can however have a connection with the pipe 661.

A further possibility for arranging a pressure restrictor valve 667 consists in integrating this valve in the piston 657 of the transmitter cylinder 650. This valve 667 must likewise have on the relaxation side a connection with the pressurised medium reservoir 660 or at least with an intermediate reservoir.

Instead of an excess pressure valve a hydro-reservoir could also be provided in the pressurised medium circuit to restrict the maximum pressure occurring in the disengagement system wherein after the stops become active to restrict the disengagement path the hydro-reservoir relaxes the system by supplying pressurised medium and thus acts in practice as a buffer or spring accumulator.

The clutch unit 701 shown in Figure 16 has, as described in connection with the preceding figures, an adjustment device 716 for automatically compensating the wear which occurs on the friction linings 707 of the clutch disc 708. In the illustrated embodiment the basic construction and method of operation of the adjustment device 716 corresponds to those of Figures 12 and 13. The adjustment element or adjustment ring 717 has bearing and stop areas 770 which can interact with the diaphragm spring 704 during a disengagement process of the clutch unit 701.The axial relative arrangement of the stop areas 770 in relation to the areas 771 of the diaphragm spring 704 interacting with same is designed so that during the disengagement process the diaphragm spring areas 771 are supported at least indirectly and preferably directly axially on the contact bearing areas 770 supported by the adjustment ring 717. This mutual support is preferably carried out at least approximately when reaching or slightly exceeding the ideal disengagement path 772 or corresponding swivel angle of the diaphragm spring 404 in the area of the tongue tips 704c. Exceeding the ideal disengagement path 772 in this way can be carried out as a result of a faulty or incorrectly set disengagement system.

Through the axial support of the diaphragm spring 704 on the contact bearing areas 770, the adjustment ring 717 is secured against undesired rotation. The diaphragm spring 704 thus acts in practice as a brake for the adjustment ring 717 on exceeding a predetermined disengagement path 772.

With the illustrated embodiment the contact bearing or stop areas 770 are formed by a ring-like projection 773 moulded on the ring 717 radially outside of the swivel bearing 705.

Instead of a ring-like radial projection it is possible to use several radial extension arms 773 spread out over the circumference. The projection or extension arms 773 extend in the illustrated embodiment up to the outer edge of the diaphragm spring 704. As soon as the specific disengagement path 772 is reached the diaphragm spring 704 is supported with its outer area 771 on the stop areas 770 of the adjustment ring 717. On exceeding the specific disengagement path 772 the swivel diameter for the diaphragm spring 704 is enlarged since this is moved from the diameter of the swivel bearing 705 to the contact diameter between the areas 771 of the diaphragm spring 704 and the stop areas 770.Through this displacement a reduction in the disengagement force required in the area of the ends 704c of the tongues also occurs since the lever ratio of the diaphragm spring changes from i to i+l, namely because the diaphragm spring mounted as a double-armed lever at first up to the disengagement path 772 is swivelled practically as a single-armed lever on exceeding the path 772. Through this reduction in the disengagement force it is also guaranteed that the diaphragm spring 704 is forced in the direction of the housing 702 or adjustment ring 717 through the axial resulting support force or biasing force applied inter alia through the sensor spring 713 and leaf springs 709. The diaphragm spring 704 can thus not be moved in its entirety axially away from the adjustment ring 717 or from the cover 702.On exceeding the determined disengagement path 772 the sensor spring 713 becomes axially resiliently deformed, namely because the diaphragm spring then lifts from the adjustment ring 717 in the area of the swivel bearing 705.

The projection or extension arm 773 could advantageously be injection moulded on the adjustment ring 770 which is made of plastics. The maximum force which acts axially on the release member 773 is produced from the difference of the minimum disengagement force in the area of the diaphragm spring tongues 704c and the axial sensor or support force for the diaphragm spring 704 which is applied through the sensor spring 713 and the leaf spring elements 709. The extension arms 773 are designed so that they withstand this maximum force without substantial deformation.

A further significant advantage lies in the fact that the axial lift of the pressure plate 703 remains practically constant and thus on exceeding the path 772 the axial force applied by the leaf springs 709 on the main diaphragm spring 704 drops no further. Since the force applied by the leaf springs 709 represents a part of the resulting sensor force, the over-movement security of the friction clutch 701 is increased through the remaining residual tension of these leaf springs. Thus for example in the case of motor car clutches an over-movement of about 0.5 to 2 mm can be achieved in the area of the tongue tips without impairing the function of the adjustment device 716.

The lift restriction of the pressure plate 703 can also be carried out in that the pressure plate 703 is axially supported on the sensor spring 713 on exceeding a specific disengagement path. To this end corresponding moulded areas such as cams, projections or the like can be provided on the sensor spring 713 and/or on the pressure plate 703.

The friction clutch 1 shown in Figures 17 and 18 has a housing 2 and a pressure disc 3 connected therewith rotationally secured but axially displaceable to a limited extent. A contact pressure diaphragm spring 4 is tensioned axially between the pressure disc 3 and the cover 2 wherein this spring 4 can swivel about a ring-like swivel bearing 5 supported by the housing 2 and biases the pressure disc 3 in the direction of a counter pressure plate 6, such as for example a flywheel, fixedly connected to the housing 2 by screws 6a whereby the friction linings 7 of the clutch disc 8 are clamped between the friction faces of the pressure disc 3 and the counter pressure plate 6.

The pressure disc 3 is connected rotationally secured to the housing 2 by circumferentially or tangentially aligned leaf springs 9. In the illustrated embodiment the clutch disc 8 has so-called lining spring segments 10 which, as known, ensure a progressive torque build-up during engagement of the friction clutch 1 by allowing a progressive rise in the axial forces acting on the friction linings 7 through a restricted axial displacement of the two friction linings 7 towards each other. However a clutch disc could also be used wherein the friction linings 7 are placed axially practically rigid on a support disc.In such a case a "lining spring set" could be used, thus a suspension in series with the diaphragm spring, eg a suspension between the cover and flywheel, between the cover and support pad on the cover side as well as between the plate spring and pressure plate or through the cover elasticity.

In the illustrated embodiment the diaphragm spring 4 has a ring-like foundation body 4a applying the contact pressure force and from which operating tongues 4b emerge radially inwards. The diaphragm spring 4 is thereby installed so that it biases the pressure disc 3 with radially further outer areas and can tilt about the swivel bearing 5 with radially further inner areas.

The swivel bearing 5 comprises two swivel support pads 11,12 which are here formed by wire rings and between which the diaphragm spring 4 is axially held and clamped. The swivel support pad 11 provided on the side of the diaphragm spring o facing the pressure disc 3 is biased with force axially in the direction of the housing 2 by means of an energy accumulator 13. The energy accumulator 13 is formed by a diaphragm spring or by a diaphragm-spring-like component 13 which is supported by its outer edge area 13a on the housing 2 and with radially further inner sections axially biases the swivel support pad 11 towards the operating diaphragm spring 4 and thus also in the direction of the housing 2.

The diaphragm spring 13 provided between the pressure disc 3 and the operating diaphragm spring 4 has an outer ringlike edge area 13b from whose inner edge tongues extend radially inwards, these tongues being supported on the swivel pad 11.

In order to support the diaphragm -spring-like component 13 in the illustrated embodiment additional means 14 are fixed on the housing 2 and form a swivel support pad for the diaphragm -spring-like component 13. These additional means can be formed by stuck-on or rivetted-on segment-like individual parts 14 which can be spread out evenly round the circumference. The means 14 can however also be formed by a circular ring-shaped inherently closed component.

Furthermore the support means 14 can be moulded out directly from the housing 2, eg by indentations formed in the axial area of the housing 2 or by tongue-like cut-out sections which after insertion and tensioning of the diaphragm spring-like component 13 are pressed under the outer edge area of this component 13 by deforming the material.

Furthermore a bayonet-like connection or lock can be provided between the support means 14 and the diaphragm-like component 13 so that the diaphragm -spring-like component 13 is first pre-tensioned and its radially outer areas can be brought axially over the support means 14. Then the support areas of the component 13 can be brought to adjoin the support means 14 by suitably turning the diaphragm -springlike component 13 relative to the housing 2. The support areas of the diaphragm -spring-like component 13 can thereby be formed by extension arms projecting radially outwards on the ring-like foundation body 13b.

In order to rotationally secure the operating diaphragm spring 4 and where applicable the diaphragm -spring-like component 13 and in order to centre the wire rings 11, 12, axially extending centring means in the form of rivet elements 15 are fixed on the housing 2. The rivet elements 15 each have an axially extending shaft 15a which extends axially through a cut-out section provided between adjoining diaphragm spring tongues 4b and which can be engaged partially by areas 13d moulded on the tongue 13c of the diaphragm spring 13 associated therewith.

The diaphragm -spring-like component or the diaphragm spring 13 is designed as a sensor spring which over a predetermined operating path produces an at least substantially constant force. The clutch disengagement force applied to the tongue tips 4c is taken up by this sensor spring 13 whereby an at least approximately counter balance always prevails between the force produced by the disengagement force on the swivel pad 11 and the counter force exerted by the sensor diaphragm spring 13 on the said swivel pad 11. By disengagement force is meant the maximum force which is exerted on the tongue tips 4c or on the disengagement lever of the diaphragm spring tongues during the operation of the friction clutch 1 and thus counteracts the sensor spring 13.

The swivel pad 12 on the side of the housing is supported on the housing 2 by an adjustment device 16 which is provided in the axial space between the diaphragm spring 4 and housing 2. This adjustment device 16 ensures that in the case of an axial displacement of the swivel support pads 11 and 12 in the direction of the pressure disc 3 or in the direction of the counter pressure plate 6 no undesired play can occur between the swivel pad 12 and the housing or between the swivel pad 12 and the diaphragm spring 4. It is thereby ensured that no undesired dead or idle paths arise during operation of the friction clutch 1 whereby an optimum degree of efficiency and thus satisfactory operation of the friction clutch 1 are provided.The axial displacement of the swivel pads 11 and 12 arises in the case of axial wear on the friction faces of the pressure disc 3 and counter pressure plate 6 as well as of the friction linings 7. In the devices according to the invention the adjustment is however also carried out in the event of wear on the swivel support pads 11,12, the axially opposing areas of the diaphragm spring and in the event of wear on the diaphragm spring in the area of the pressure plate support pad cams (at 3a) or the areas of the diaphragm spring opposite same.

The way in which the automatic adjustment of the swivel bearing 5 takes place will be explained in detail in connection with the diagrams according to Figures 24 to 27.

The adjustment device 16 comprises a spring-biased adjustment element in the form of a ring-like component 17 which is illustrated in Figures 19 and 20. The ring-like component 17 has circumferentially extending and axially rising run-up ramps 18 which are spread out over the circumference of the component 17. The adjustment element 17 is installed in the clutch 1 in such a way that the runup ramps 18 face the housing base 2a. The swivel pad 12 formed by a wire ring is positioned centrally in a groovelike socket 19 (Figure 18) on the side of the adjustment element 17 remote from the run-up ramps 18. The socket 19 can thereby be designed so that the swivel pad 12 is also secured axially on the adjustment element 17.This can happen for example in that the areas of the adjustment element 17 adjoining the socket 19 securely clamp the swivel pad 12 at least in sect ions or form a snap connection for the swivel pad 12. When using different materials for the swivel pad 12 and the adjustment element 17 it can be expedient in order to compensate for the expansion differences which occur with wide temperature changes if the swivel pad 12 which is designed as a wire ring is open, thus is separated at at least one point over the circumference whereby the wire ring 12 can move circumferentially relative to the socket 19 and thus the wire ring 12 can adapt to the diameter changes of the socket 19.

In the illustrated embodiment the adjustment element 17 is made from plastics, such as eg from heat-resistant thermoplastics which can be additionally reinforced with fibres. The adjustment element 17 can thereby be simply made as an injection moulded part. As already mentioned an adjustment element of plastics with a lower specific weight produces a lower mass inertia weight whereby the sensitivity to pressure vibrations is also reduced. Furthermore the swivel support pad could be formed directly by the plastics ring. The adjustment element 17 can however also be made as a shaped sheet metal part or by sintering. Furthermore with a suitable choice of material the swivel pad 12 can be made integral with the adjustment element 17. The swivel pad 11 can be formed directly by the sensor spring 13.To this end the tips of the tongues 13c can have corresponding indentations or moulded areas such as eg grooves.

The adjustment ring 17 is centred by the axially aligned areas 15a of the rivets 15 which are evenly distributed over the circumference. To this end the adjustment ring 17 has centring contours 20 which are formed by circumferentially extending recesses 21 which lie radially inside the swivel pad 11. In order to form the recesses 21 the adjustment ring 11 has on the inner edge area radially inwardly extending cams 22 which define radially inner contours of the recesses 21.

As can be seen from Figure 19, 5 run-up ramps 18 are each provided, seen circumferentially, between the evenly distributed recesses 21. The recesses 21 are designed in the circumferential direction so that they allow at least a turning angle of the adjustment ring 17 relative to the housing 2 which ensures throughout the entire service life of the friction clutch 1 an adjustment of the wear which occurs at the friction faces of the pressure disc 3 and the counter pressure plate 6 as well as on the friction linings 7 as well as where applicable of the wear on the clutch itself, thus eg the pressure pads 11,12, the interposed diaphragm spring areas, the pressure plate cams (at 3a) or the areas of the diaphragm springs 4 opposite same.This adjustment angle can be in the order of between 8 and 60 degrees, preferably in the order of 10 to 30 degrees depending on the design of the run-up ramps. In the illustrated embodiment this turning angle lies in the area of 12 degrees wherein the starting angle 23 of the run-up ramps is likewise in the area of 12 degrees. This angle 23 is selected so that the friction which occurs when the runup ramps 18 of the adjustment ring 17 press on the counter run-up ramps 24 of the supporting ring 25 shown in Figures 21 and 22, prevents any slipping between the run-up ramps 18 and 24. The angle 23 can be in the range from 4 to 20 degrees depending on the material pairing in the area of the run-up ramps 18 and counter run-up ramps 24.

The adjustment ring 17 is spring-loaded in the circumferential direction, namely in the adjustment turning direction, thus in the direction which as the ramps 18 run up on the counter ramps 24 of the support ring 25 causes an axial displacement of the adjustment ring in the direction of the pressure disc 3, this means therefore in the axial direction away from the radial housing section 2a. In the embodiment illustrated in Figures 17 and 18 the spring tension on the adjustment ring 17 is ensured through at least one ring-like leg spring 26 which can have for example two windings and which has at one of its ends a radially aligned leg 27 which is rotationally secured with the adjustment ring 17 whilst at the other end it has an axially aligned leg 28 which is hung from the housing 2 secured against rotation. The spring 26 is installed with resilient tension.

The support ring 25 illustrated in Figures 21 and 22 is likewise formed by a ring-like component which has counter run-up ramps 24 which form complementary faces with the faces defined by the run-up ramps 18 wherein the faces defined by the run-up ramps 18 and counter run-up ramps 24 can also be congruent. The starting angle 29 of the counter run-up ramp 24 corresponds to the angle 23 of the run-up ramps 18. As can be seen through a comparison of Figures 19 and 21 the run-up ramps 18 and the counter run-up ramps 24 are spread out in the circumferential direction in a similar way. The support ring 25 is connected rotationally secured to the housing 2. To this end the support ring 25 has circumferentially spread out recesses 30 through which extend the rivetting attachments of the rivets 15.

In Figure 18 a further ring-like leg spring 26a is shown in chain-dotted lines which like the leg spring 26 can be bent down at the end areas in order to ensure a rotationally secured connection with the housing 2 on one side and with the adjustment element 17 on the other side. This spring 26a is likewise installed with resilient tension so that it exerts a torsional force on the adjustment element 17. The use of two leg springs 26,26a can be advantageous in many cases since during rotation of the friction clutch 1 the spring force is intensified owing to the centrifugal forces acting on the spring 26 or 26a. By using two leg springs the increase in force occurring for example on the spring 26 can be compensated by the force applied by the leg spring 26a.To this end the leg springs 26 and 26a are coiled so that they produce, at least under the effect of centrifugal force on the adjustment element 17, forces which act circumferentially in the opposite direction. The two leg springs 26,26a can have one or more windings, moreover these leg springs 26,26a can have different winding diameters, as shown in Figure 18, wherein the centrifugal forces normally connected therewith and acting on the springs 26,26a and which would produce different size circumferential forces on the adjustment element 17, can be compensated at least approximately by a corresponding design of the wire thickness and/or the winding number of the individual springs 26,26a. In Figure 18, the spring 26 is mounted radially inside the adjustment element 17 and the spring 26a radially outside of this adjustment element 17.Both springs could however be mounted through a corresponding design radially inside or radially outside the adjustment element 17.

The leg spring 26 is shown in plan view in Figure 23. In the relaxed state of the leg spring 26 the legs 27,28 are off-set by an angle 31 which can lie in the order of between 40 and 120 degrees. In the illustrated embodiment this angle 31 is in the order of 85 degrees. The numeral 32 represents the relative position of the leg 27 relative to the leg 28 as positioned in new clutch linings 7 in the friction clutch 1. 33 represents the position of the leg 27 which corresponds to the maximum permissible wear on the friction linings 7. The adjustment angle 34 lies in the illustrated embodIment in the order of 12 degrees. The spring 26 is designed so that in the relaxed state of this spring 26 only one wire winding 35 runs between the two legs 27,28. In the remaining circumferential area two wire windings lie axially one above the other.The spring 26a is designed in a similar way to the spring 26 but has a larger winding diameter and another tensioning direction in relation to the adjustment element 17 according to Figure 18. The force exerted by the spring 26 on the adjustment ring 17 is however greater than that of the spring 26a.

When the friction clutch 1 is new the axial cams 18a,24a forming the run-up ramps 18 and counter run-up ramps 24 engage axially in each other the furthest, this means that the rings 17 and 25 lying one on top of the other require the smallest axial structural space.

In the embodiment according to Figures 17 and 18 the counter run-up ramps 24 or the cam-like attachments 24a forming same are formed by a single component. The counter run-up ramps 24 can however also be formed directly by the housing 2 for example by imprinting cam-like attachments which can extend into the housing space. The imprinting method is particularly advantageous in the case of sheet metal housing and covers which are formed in one piece.

In order to hold the adjustment ring 17 in its retracted position prior to fitting the friction clutch 1, the ring has in the area of the cams 22 engagement areas 36 for a turning and retaining means which can be supported on the other side on the housing 2. Such retaining means can be provided when manufacturing or assembling the friction clutch 1 and can be removed from the clutch after fitting the friction clutch 1 on the flywheel 6 whereby the adjustment device 16 is activated. To this end in the illustrated embodiment oblong recesses 37 are placed circumferentially in the cover or housing 2 and an indentation or recess 38 is provided in the adjustment ring 17. The circumferentially placed oblong recesses 37 must thereby have at least one such extension so that the adjustment ring 17 can be turned back corresponding to the largest possible wear adjustment angle.After assembling the friction clutch 1 it is also possible to pass a turning tool radially through the slits 37 in the cover and into/up to the recesses 38 of the adjustment ring 17. The ring 17 can then be turned back by means of the tool so that it moves in the direction of the radial area 2a of the housing 2 and occupies its smallest axial spacing relative to this area 2a. The adjustment ring 17 is then secured in this position for example by a clip or pin which engages in a flush recess of the cover and adjustment ring 17 to prevent these two components from turning. After fitting the friction clutch 1 on the flywheel 6 the pin is removed from the recess so that as already mentioned the adjustment device 16 is released. The slits 37 in the housing 2 are designed so that during or after dismantling the friction clutch 1 from the flywheel 6 the adjustment ring 17 can be brought into its retracted position. To this end the clutch 1 is first disengaged so that the operating diaphragm spring 4 exerts no axial force on this swivel pad 12 and thus a satisfactory turning of the adjustment ring 17 is guaranteed.

A further possibility of bringing the components of the friction clutch already fixed on an internal combustion engine into a functionally correct position, consists in turning back or resetting the adjustment element or adjustment ring 17 only after it has been fitted on the internal combustion engine or flywheel thereof. To this end the friction clutch 1 can be operated for example by a secondary tool and the then practically relaxed ring 17 can be moved into its position set back from the pressure plate.

The friction clutch 1 is then again engaged so that the ring 17 at first retains this retracted position.

The ring-like adjustment element or the supporting ring 25 can also each have two radially off-set circumferentially extending and axially adjoining sets of run-up ramps which are each spread over the circumference of these components.

The radially inner run-up ramps can thereby be off-set circumferentially relative to the radially outer run-up ramps, namely by about half a ramp length or a ramp division. Through the circumferentially off-set ramps it is ensured that a satisfactory central guide is achieved between the adjustment element 17 and the support ring 25.

The method of functioning of the friction clutch 1 described above will now be explained in detail in connection with the characteristic lines entered in the diagrams according to Figures 24 to 27.

The line 40 in Figure 24 shows the axial force which is produced in dependence on the change in the conicity of the diaphragm spring 4, namely during the deformation of the diaphragm spring 4 between two supports whose radial distance corresponds to the radial distance between the swivel bearing 5 and the radially outer support diameter 3a on the pressure disc 3. The relative axial path between the two support pads is shown on the abscissa and the force produced by the diaphragm spring is shown on the ordinate.

The point 41 represents the flat position of the diaphragm spring which is preferably selected as the installation position of the diaphragm spring 4 when the clutch 1 is closed, thus the position wherein the diaphragm spring 4 exerts the maximum contact pressure force on the pressure disc 3 for the corresponding installation position. The point 41 can be moved up or down by altering the conical installation position, thus the setting, of the diaphragm spring 4 along the line 40.

The line 42 represents the axial expanding force applied by the lining spring segments 10 and acting between the two friction linings 7. This axial expanding force counteracts the axial force exerted by the diaphragm spring 4 on the pressure disc 3. It is advantageous if the axial force which is required for the possible elastic deformation of the spring segments 10 corresponds at least to the force which is exerted by the diaphragm spring 4 on the pressure disc 3 wherein this latter force can advantageously be even greater.During disengagement of the friction clutch 1 the spring segments 10 relax, namely over the path 43. The disengagement process of the clutch 1 is assisted through this path 43 which corresponds to a corresponding axial displacement of the pressure disc 3, this means that a lower maximum disengagement force need be applied than that which would correspond to the installation point 41 where the lining spring segments 10 are not present (in the absence of a lining suspension). On exceeding the point 44 the friction linings 7 are released whereby as a result of the degressive characteristic line range of the diaphragm spring 4 the disengagement force which is then still to be applied is reduced considerably compared to that which would correspond to the point 41.The disengagement force for the clutch 1 decreases until the minimum or trough point 45 of the sinusoidal characteristic line 40 is reached. On exceeding the minimum 45 the required disengagement force rises again wherein the disengagement path is selected in the area of the tongue tips 4c so that even when exceeding the minimum 45 the disengagement force does not exceed the maximum disengagement force arising at the point 44, and advantageously remains below same. The point 46 should thus not be exceeded.

The spring 13 serving as the force sensor has a path-force curve corresponding to the line 47 of Figure 25. This characteristic line 47 corresponds to that which is produced when the diaphragm -spring-like component 13 is changed in its conicity from the relaxed position, namely between two swivel support pads which have a radial distance which corresponds to the radial distance between the swivel support pads 11 and 14. As the characteristic line 47 shows, the diaphragm-spring-like component 13 has a spring path 48 over which the axial force produced by same remains practically constant. The force produced in this area 48 is thereby selected so that this corresponds at least approximately to the disengagement force of the clutch arising in the point 44 of Figure 24.The support force to be applied by the sensor spring 13 is reduced compared to the force of the diaphragm spring 4 corresponding to point 44 in accordance with the lever translation of this diaphragm spring 4. This translation ratio lies in most cases in the order of between 1 : 3 to 1 : 5 but in many cases can however also be greater or smaller.

The said diaphragm spring translation corresponds to the ratio between the radial distance of the swivel bearing 5 to the support 3a and the radial distance of the swivel bearing 5 to the contact bearing diameter 4c eg for a disengagement bearing.

The installation of the diaphragm -spring-like element 13 in the friction clutch is selected so that in the region of the swivel bearing 5 the element can undertake an axial spring path in the direction of the friction linings 7 which not only corresponds at least to the axial adjustment path of the pressure disc 3 in the direction of the counter pressure plate 6 which arises as a result of the friction surface and friction lining wear, but also guarantees an at least approximately constant axial support force for the swivel bearing 5. This means that the linear area 48 of the characteristic line 47 should have a length which at least corresponds to the said wear path and is preferably greater than this wear path since installation tolerances can then thereby be compensated at least in part.

In order to obtain a practically constant or defined release point 44 of the friction linings 7 when disengaging the friction clutch 1 it is possible to use a so-called double segment lining suspension between the friction linings 7, thus a lining suspension wherein individual spring segments are provided back to back in pairs wherein the individual pairs of segments can have a certain axial pretension relative to each other. By pretensioning the spring means provided between the linings it can be achieved that the embedding losses occurring throughout the operating period of the segments into the reverse side of the linings can be at least substantially compensated or balanced. By embedding losses is meant the losses which arise as the segments are worked into the reverse side of the linings.

By correspondingly defining the axial spring path between the two friction linings 7 as well as by providing a definite pretension of the suspension acting between the friction linings it can furthermore be achieved that when disengaging the friction clutch 1 the pressure plate 3 is forced back over a definite path 43 by the suspension provided between the linings. In order to obtain a definite path 43 the axial path between the friction linings can be defined by corresponding stops both in the relaxing direction and in the tensioning direction of the lining suspension 10. The lining suspensions used in connection with the present invention are advantageously those already known for example from Patent Application P 42 06 880.0 which can be expressly added to the content and subject of the present invention.

In Figure 26 the line 49 shows the force required to disengage the clutch through a disengagement element engaging on the area 4c of the diaphragm spring in order to move the pressure plate from point 41 to point 44 (Figure 24). The line 49 furthermore shows the route of the tongue tips of the diaphragm spring in the area 4c.

In order to ensure an optimum functioning of the friction clutch 1 or of the adjustment device which guarantees an automatic compensation of the lining wear, it is advisable that - seen over the disengagement force curve 49 which actually occurs according to Figure 10 - the accumulating forces exerted first by the lining suspension 10 and the sensor spring 13 on the diaphragm spring 4 are greater than the force exerted by the diaphragm spring 4 on the support pad 11. Also after the pressure disc 3 has been lifted from the friction linings 7 the force still exerted by the sensor spring 13 on the diaphragm spring 4 is to be greater or at least the same as the required disengagement force engaging in the area 4c of the diaphragm spring tongue tips and changing according to Figure 26 over the disengagement path (according to line 49).The force exerted thereby by the sensor diaphragm spring 13 on the support pad 11 should furthermore be dimensioned so that rotation of the ring 17 standing under the force of the spring 26, and thus axial displacement of the diaphragm spring, is prevented, at least approximately until the point 41 of the rising branch of the characteristic line 40 corresponding to the installation position of the diaphragm spring is not exceeded.

The considerations up until now correspond to a quite specific installation position of the diaphragm spring 4, and no wear on the friction linings 7 has been taken into account.

In the event of axial wear, eg of the friction linings 7, the position of the pressure disc 3 moves in the direction of the counter pressure plate 6 resulting in a change in the conicity of the diaphragm spring (the tongue tips 4c move to the right, seen by an observer) and thus also a change in the contact pressure force applied by the diaphragm spring in the engaged state of the friction clutch 1, namely in the sense of an increase. This change causes the point 41 to move in the direction of point 41' and point 44 in the direction of point 44'. This change destroys the equilibrium force originally present when disengaging the clutch 1 in the area of the swivel support pad 11 between the operating diaphragm spring 4 and the sensor spring 13.

The increase in the diaphragm spring contact pressure force for the pressure disc 3 caused by the wear on the linings also causes a displacement in the path of the disengagement force in the sense of an increase. The resulting disengaging force path is shown in Figure 26 by the chaindotted line 50. The increase in the disengaging force curve during the disengaging process of the friction clutch 1 overcomes the axial force exerted by the sensor spring 13 on the diaphragm spring 4 so that the sensor spring 13 yields substantially in the area of the swivel bearing 5 by an axial path corresponding substantially to the wear of the friction linings 7.During this sagging phase of the sensor spring 13 the diaphragm spring 4 is supported on the biasing area 3a of the pressure disc 3 so that this diaphragm spring 4 changes its conicity and thus also the energy stored therein or the torque stored therein and consequently also the force exerted by the diaphragm spring 4 on the swivel pad 11 and by the sensor spring 13 on the pressure disc 3.

As can be seen in connection with Figure 24 this change takes place in the sense of reducing the forces exerted by the diaphragm spring 4 on the pressure plate. This change takes place until the axial force exerted by the diaphragm spring 4 in the area of the swivel pad 11 on the sensor spring 13 is balanced with the counter force produced by the sensor spring 13. This means that in the diagram according to Figure 24 the points 41' and 44' again move in the direction of points 41 and 44. Once this counterbalance is restored then the pressure disc 3 can again be removed from the friction linings 7.During this adjustment phase of the wear, thus during the disengaging process of the friction clutch 1 the sensor spring 13 yields, the adjustment element 17 of the adjustment device 16 is turned by the pretensioned spring 26 whereby the swivel support pad 11 moves along corresponding to the wear on the linings, and thus a satisfactory swivel bearing 5 of the diaphragm spring 4 is again guaranteed. After the adjustment process the disengagement force path again corresponds to the line 49 according to Figure 26. The lines 50 and 51 of Figure 26 represent the axial path of the pressure disc 3 with a disengagement force- path curve corresponding to lines 49,50.

The diagram according to Figure 27 shows the forces curve over the disengagement path of the force exerted during the disengagement process on the housing 2 or on the diaphragm spring 13 wherein the extreme values were capped. Starting from the engaged position according to Figure 17 at first a force acts on the housing 2 and thus also on the pressure disc 3 which corresponds to the installation point 41 (Figure 24) of the diaphragm spring 4. During the disengagement process the axial force exerted by the diaphragm spring 4 on the housing 2 or swivel pad 12 decreases according to the line 52 of Figure 27, namely to the point 53.If in the case of a conventional clutch wherein the diaphragm spring is swivel mounted axially fixed on the housing, thus the swivel pad 11 were connected axially inflexible with the housing 2, on exceeding the point 53 in the disengagement direction, an axial direction reversal would take place in the force action through the diaphragm spring 4 on the housing 2 radially level with the swivel bearing 5. With the clutch according to the invention in the area of the swivel bearing 5 the force produced by the axial reversal through the diaphragm spring 4 in the area of the swivel bearing 5 is taken up by the sensor spring 13. On reaching the point 54 the diaphragm spring 4 is lifted off from the biasing area 3a of the pressure disc 3.Up to at least this point 54 the disengagement process of the friction clutch 1 is assisted by the axial force applied by the lining suspension 10, because it acts against the diaphragm spring force. The force applied by the lining suspension 10 thereby decreases with an increasing disengagement path in the area 4c of the tongue tips and with an increasing axial disengagement stroke of the pressure disc 3. The line 52 thus represents the resulting line seen over the disengagement process, of the disengagement force acting in the tongue tip area 4c and of the axial force exerted in the radial area 3a on the diaphragm spring 4 by the lining suspension 10. On exceeding the point 54 in the disengagement direction the axial force exerted by the diaphragm spring 4 on the swivel pad 11 is taken up by the counter force applied by the sensor diaphragm spring 13 wherein these two forces are balanced at least after the relaxing of the friction linings 7 through the pressure disc 3 and with a continuation in the disengagement process the axial force applied by the sensor spring 13 in the area of the swivel bearing 5 is preferably somewhat greater than the ensuing disengagement force. The part 55 of the characteristic line 52 of the diagram according to Figure 27 shows that as the disengagement path increases so the disengagement force or the force exerted by the diaphragm spring 4 on the swivel pad 11 becomes smaller compared to the disengagement force arising at point 54.

The chain-dotted line 56 corresponds to a state of the friction clutch 1 wherein wear has occurred in the area of the friction linings 7 but no adjustment has been made in the area of the swivel bearing 5. It can also be seen here that the change in the installation position of the diaphragm spring 4 caused by wear causes an increase in the forces exerted on the housing 2 and on the swivel pad 11 and on the sensor spring 13.This has the particular result that the point 54 moves in the direction of point 54 which has the result that during the renewed disengagement process of the friction clutch 1 the axial force exerted by the diaphragm spring 4 on the sensor spring 13 in the area of the swivel pad 11 is greater than the counter force of the sensor spring 13, whereby the adjustment process already described is carried out through the sensor spring 13 springing out axially. Through this adjustment process caused by the spring 26, thus through the rotation of the ring 17 and the axial displacement of the pad 12, the point 54' is again moved in the direction of point 54 whereby the desired counter balanced state is again produced in the area of the swivel suspension 5 between the diaphragm spring 4 and the sensor spring 13.

In practice the described adjustment takes place continuously or in very small steps so that the great displacements in the points and characteristic lines shown in the drawings for better comprehension of the invention normally do not occur.

Some function parameters or operating points can change over the operating time of the friction clutch 1. Thus for example improper operation of the friction clutch 1 can result in an overheating of the lining suspension 10 which can lead to setting and thus a reduction in the axial suspension of the lining suspension or lining segments 10.

However an operationally reliable functioning of the friction clutch can be guaranteed by a suitable configuration of the characteristic line 40 of the diaphragm spring 4 and a corresponding adaptation of the curve 47 of the sensor spring 13. An axial settling of the lining suspension 10 would only lead to the diaphragm spring 4 occupying a more pressed-through position compared to that shown in Figure 17, whereby the contact pressure force exerted by the diaphragm spring 4 on the pressure disc would be slightly less, as can be seen in connection with the characteristic line 40 according to Figure 24. Furthermore a corresponding axial deformation of the sensor spring 13 and thus a corresponding axial displacement of the swivel support pad 11 would result.

According to a further inventive idea the resulting support force acting on the operating diaphragm spring 4 can rise as the wear on the friction linings 7 increases. The rise can thereby be restricted to a partial area of the maximum wear path permitted overall of the friction linings 7. The rise in the support force for the operating diaphragm spring 4 can thereby be produced through a corresponding design of the sensor spring 13. A corresponding characteristic line curve over the area 48 is shown chain-dotted in Figure 25 and designated by reference numeral 47a. A rise in the support force for the operating diaphragm spring 4 with increasing wear can compensate at least partially a drop in the contact pressure force of the operating diaphragm spring 4 for the pressure plate 3, conditioned by a decrease in the lining suspension, eg by the segments embedding in the linings.It can thereby be particularly advantageous if the support force for the operating diaphragm spring 4 rises proportional to the setting of the lining suspension or proportional to the embedding of the segment in the linings.

This means that the said support force is to rise as the disc thickness reduces in the area of the linings, thus with a reduction in the distance between the friction faces of the linings as a result of the segment embedding and/or a setting of the lining suspension and/or the lining wear. It is thereby particularly advantageous if the force rise takes place so that it is greater over a first partial area than in a following second partial area whereby the two partial areas are located within the area 48 according to Figure 25.

The latter design is advantageous because the largest part of the said embedding between the spring segments and the linings mainly takes place within a time span which is slight compared to the total service life of the friction clutch and then the ratios between the spring segments and friction linings practically stabilize. This means that after a certain embedding there is no more significant change regarding the embedding. A rise in the support force for the operating diaphragm spring can however also take place over at least a part of the friction wear of the friction linings.

With the above description of the adjustment process for compensating the friction lining wear any axial forces possibly applied by the leaf springs 9 were not taken into consideration. The disengagement process is assisted by a pretensioning of the leaf springs 9 in the sense of a removal of the pressure disc 3 from the corresponding friction lining 7, thus in the sense of the pressure disc 3 pressing against the diaphragm spring 4. The axial force applied by the leaf springs 9 is overlapped with the forces applied by the sensor spring 13 and diaphragm spring 4 and with the disengagement force. For better understanding this was not taken into consideration up until now in the description of the diagrams according to Figures 24 to 27.

The entire force biasing the operating diaphragm spring 4 in the disengaged state of the friction clutch 1 against the rolling support pad 12 on the cover side is achieved by adding the forces which are mainly exerted by the leaf spring elements 9, by the sensor spring 13 and by the existing disengagement force on the operating diaphragm spring 4. The leaf spring elements 9 can thereby be installed between the cover 2 and pressure plate 3 so that as the wear on the friction linings 7 increases so the axial force exerted by the leaf springs 9 on the operating diaphragm spring 4 becomes greater. Thus for example over the path 48 according to Figure 25 and thus also over the wear compensation path of the adjustment device 16 the axial force applied by the leaf springs 9 can have a curve according to line 47b.It can also be seen from Figure 25 that with the increasing sagging of the sensor spring 13 there is an increase in the resetting force which is exerted by the leaf springs 9 on the pressure plate 3 and which also acts on the operating diaphragm spring 4. Adding the force curve according to the characteristic lines 47b and the diaphragm spring characteristic line produces the resulting force curve which acts axially on the diaphragm spring 4, namely in the sense of pressing the diaphragm spring 4 against the swivel support pad 12 on the cover side. In order to obtain a curve according to the line 47a, wherein at the beginning of the adjustment area 47d there is at first an initial force rise which changes into a somewhat constant force area, it is expedient to design the sensor diaphragm spring so that it has a characteristic line curve corresponding to the line 47c of Figure 25. Adding the force curve according to line 47c and the force curve according to line 47b then produces the force curve according to line 47a. By suitably pretensioning the leaf springs 9 it is thus possible to reduce the support force to be applied by the sensor spring or support force curve. By suitably designing and arranging the leaf spring elements 9 it is likewise possible to compensate at least partially a reduction in the lining suspension and/or embedding of the lining spring segments in the linings. It can thereby be ensured that the diaphragm spring 4 retains substantially the same operating point and same operating area so that throughout the service life of the friction clutch the diaphragm spring 4 exerts an at least approximately constant contact pressure force on the pressure plate 3.Furthermore the resulting axial force which is produced by the adjustment springs 26 and/or 26a acting on the adjustment element 17 and which counteracts the sensor spring 13 and/or the leaf springs 9 must be taken into account when designing the friction clutch, more particularly the sensor spring 13 and/or the leaf springs 9.

When designing the friction clutch 1 with pretensioned leaf springs 9 it must still be taken into account that the axial force exerted by the pressure plate 3 on the friction linings 7 is influenced by the pretension of the leaf springs 9. This means that when the leaf springs 9 are pretensioned in the direction of the operating diaphragm spring 4 the contact pressure force applied by the diaphragm spring 4 is reduced by the pretensioning force of the leaf springs 9. With a friction clutch 1 of this kind this results in a contact pressure force curve for the pressure plate 3 or for the friction linings 7 which is produced by overlapping the contact pressure force curve of the diaphragm spring 4 with the tension curve of the leaf springs 9.Assuming that - considered over the operating range of the friction clutch 1 - the characteristic line 40 according to Figure 24 represents the resulting force curve of the operating diaphragm spring 4 and pretensioned leaf springs 9 in the new state of the friction clutch 1, a reduction in the distance between the pressure plate 3 and counter pressure plate 6 following wear on the linings would move the resulting curve in the sense of a reduction. In Figure 24 a line 40a is shown in dotted lines which corresponds for example to an overall lining wear of 1. 5 mm.Through this displacement of the line 40 in the direction of line 40a which occurs over the service life of the friction clutch, the axial force exerted during disengagement of the friction clutch 1 by the diaphragm spring 4 on the sensor spring 13 reduces, namely as a result of the counter moment exerted by the leaf springs 9 on the diaphragm spring 4 as the wear increases. This counter moment exists owing to the radial distance between the swivel bearing 5 and the biasing diameter 3a between the operating diaphragm spring 4 and the pressure plate 3.

The friction clutch 101 shown in Figures 28 and 29 mainly differs from the friction clutch 1 shown in Figures 17 and 18 in that the adjustment ring 117 is loaded in the circumferential direction by coil springs 126. The adjustment ring 117 corresponds in its function and method of operation regarding compensating the wear on the friction linings, to the adjustment ring 17 according to Figures 18 to 20. In the illustrated embodiment three coil springs 126 are provided evenly spread out round the circumference and pretensioned between the clutch housing 2 and adjustment ring 117.

As can be seen in particular from Figure 30, the adjustment ring 117 has on the inner circumference radial projections or steps 127 on which the coil springs 126 arranged in an arc can be supported in the circumferential direction with one of their ends in order to bias the adjustment ring 117.

The other end areas of the springs 126 are supported on stops 128 supported by the clutch housing 2. In the illustrated embodiment these stops 128 are formed by screwlike connecting elements which are connected to the cover 2.

These stops 128 can however also be formed by axial moulded areas which are formed integral with the clutch housing 2.

Thus the stops 128 can be formed for example by indentations or tabs shaped axially out of a sheet metal housing 2. As can be seen in particular from Figures 29 and 30 the ring 117 can be formed on the inner circumference so that a guide 129 is provided at least substantially in the area of the extension of the springs 126 and preferably also over the turning angle of the ring 117 required for adjusting the wear or over the relaxation path of the springs 126 wherein this guide ensures that the springs 126 are held axially and supported radially. In the illustrated embodiment, the spring guides 129 are formed by indentations which, seen in cross-section, are substantially semi-circular, whereby their boundary faces substantially conform with the crosssection of the coil springs 126.

A design of this kind has the advantage that when the friction clutch rotates there is a satisfactory guide for the springs 126 so that these cannot escape axially. As extra security for the coil springs 126, as shown in Figure 29, the cover 2 can have on its radially inner edge area axial moulded areas 130 which overlap the spring 126 in the axial direction. Instead of individual moulded areas 130 the cover 2 can also have an axial inner edge 130 extending over the circumference. The inner edge 130 can serve to restrict the relaxation of the diaphragm spring 4.

Guiding the adjustment springs 126 according to Figures 28 to 30 has the advantage that as the clutch unit 1 rotates so the individual windings of the springs 126 can be radially supported through the centrifugal force action on the adjustment ring 117 whereby the adjustment forces applied by the springs 126 in the circumferential direction are reduced or even completely eliminated owing to the friction resistances produced between the spring windings and the adjustment ring 117. Thus during rotation of the friction clutch 101 the springs 126 can stay practically rigid (owing to the friction forces suppressing the spring action). It can thereby be achieved that at least at speeds above the idling speed of the internal combustion engine the adjustment ring 117 cannot be turned by the springs 126.It can thereby be achieved that compensation of the friction lining wear only takes place when operating the friction clutch 101 at idling speed or at least approximately at idling speed. The blocking of the adjustment ring 117 can however also be carried out so that an adjustment as a result of wear on the lining can only take place when the internal combustion engine is stationary, thus when the friction clutch 101 is not rotating.

Blocking the adjustment process during rotation of the friction clutch 1 or when exceeding a specific speed can also be advantageous in the case of an embodiment according to Figures 17 and 18. To this end means can be provided for example on the housing 2 which through centrifugal force acting on the adjustment element 17 provide security against rotation, namely against the adjustment force produced by the leg spring 26 and/or 26a. The blocking means can thereby be formed by at least one weight which can be forced radially outwards through centrifugal force action and which is supported for example on the inner edge of the ring 17 where it can produce a friction which produces on the ring 17 a stopping moment which is greater than the turning moment exerted by the adjustment springs on the ring 17.

Support means supported by the housing 2 can also be provided for radially supporting at least one partial area of the extension of the springs 126. In the embodiment according to Figures 28 and 29 the supporting means can be formed in one piece with the stops 128. To this end the stops 128 can be made angular so that they each have a circumferentially extending area which projects into the spring 126 at least over a partial section of the extension thereof. At least a part of the spring windings can thereby be guided and supported at least in the radial direction.

As can be seen from Figure 29, the wire ring 11 provided in Figure 18 is omitted and has been replaced by moulded areas 111 formed in the area of the tongue tips of the sensor spring 113. To this end the tongues 113c are ball-shaped in the area of their tips on their side facing the operating diaphragm spring 4.

Figures 31 to 33 show a further embodiment of a wear adjustment system according to the invention wherein individual adjustment elements 217 are used in place of one ring-shaped adjustment ring. These adjustment elements are spread out evenly over the circumference of the cover 202.

The adjustment elements 217 are formed by knob or disc like components which have a circumferentially extending and axially rising run-up ramp 218. The annular adjustment elements 217 have a central recess or bore 219 through which extend the axial pin-like attachments 215a which are supported by the cover so that the ring-like adjustment elements 217 are mounted rotatable on these attachments 215a. The cover 202 is provided with moulded areas 225 which form counter run-up ramps 224 for the ramps 218. A spring element 226 is tensioned between an adjustment element 217 and the cover 202 and biases the adjustment element 217 in the turning direction which causes an adjustment. As is apparent from Figure 31 the spring element 226 can extend round an axial attachment 215a, thus can be formed like a coil spring.On the end areas of a spring 226 there are moulded areas such as eg bent areas or arms for supporting the one spring end on the housing 202 and the other spring end on the corresponding adjustment element 217. In the event of an axial displacement of the plate spring 204 or sensor spring 213 in the area of the swivel support pad 205 the adjustment elements 217 are turned and the displacement is compensated by the ramps 218 running up on the ramps 224.

The axial support of the sensor diaphragm spring 213 on the housing 202 is carried out by tabs 214 which were moulded out of the axially aligned area of the housing 202 and forced inwards underneath the outer areas of the sensor spring 213.

The ring-like adjustment elements 217 have the advantage that they can be designed substantially independent of centrifugal force regarding their adjustment action.

Instead of the rotating or turning adjustment elements 217 shown in Figure 31 it is also possible to use individual wedge-shaped adjustment elements which can be displaced in the radial and/or circumferential direction for adjusting the wear. These wedged adjustment elements can have an oblong recess through which an axial attachment 215a can extend for guiding the corresponding adjustment element.

The wedged adjustment elements can have an adjusting effect owing to the centrifugal force acting on same. However energy accumulators can also be provided which bias the wedged adjustment elements in the adjusting direction. For a satisfactory guide of the wedged adjustment elements the housing 202 can have moulded areas. The wedge faces of the adjustment elements running at a certain run-up angle relative to a plane running at right angles to the axis of rotation of the friction clutch can be provided on the side of the housing and/or on the side of the operating plate spring. When using individual wedge elements of this kind it is expedient to make them from a light material in order to reduce the centrifugal forces acting on them to a minimum.

The material match between the components forming the adjustment ramps is preferably selected so that throughout the operating life of the friction clutch there is no adhesion between the run-up ramps and counter run-up ramps which would prevent the adjustment. In order to avoid such adhesion at least one of these components can be provided with a coating at least in the area of the ramps or counter ramps. Through such coatings it is possible to avoid corrosion when using two metal components.Adhesion or sticking between the components forming the adjustment ramps can furthermore be avoided if the components which are supported on one another and form the ramps and counter ramps are made from a material with different coefficients of expansion so that as a result of the temperature fluctuations which occur during operation of the friction clutch the surfaces which form the adjustment ramps and which are in contact complete a movement relative to each other. The components forming the run-up ramps and counter run-up ramps are thereby always held movable relative to each other. Sticking or jamming between these parts cannot then occur since these parts are always released from each other through the different expansions of these parts.A release of the run-up ramps can also be achieved in that as a result of the different strength and/or design of the parts the centrifugal forces acting on these parts cause different expansions or movements which again prevent the parts from sticking or jamming.

In order to prevent an adhesive connection between the runup ramps and counter run-up ramps it is also possible to provide means which exert axial force on the or each adjustment element when the friction clutch is disengaged or when the wear is adjusted. To this end the adjustment element 17, 117 can be coupled axially to a component which has areas which move axially in the event of wear. This coupling can be carried out in particular in the area of the swivel bearing 5, namely with the operating diaphragm spring 4 and/or the sensor spring 13.

In the diagram according to Figure 34 a contact pressure plate spring characteristic line 340 is shown which has a trough point or minimum 345 in which the force applied by the contact pressure plate spring is comparatively light (about 450 Nm). The maximum of the diaphragm spring with the force-path-characteristic line 340 lies in the order of 7,600 Nm. The characteristic line 340 is produced by deforming a diaphragm spring between two radially spaced supports, namely as described in connection with the characteristic line 40 according to Figure 24 and in conjunction with the diaphragm spring 4.

The diaphragm spring characteristic line 340 can be combined with a lining spring characteristic line 342. As can be seen from Figure 34 the force-path curve of the lining spring segment characteristic line 342 comes close to the contact pressure diaphragm spring characteristic line 340 or the two characteristic lines only run at a slight distance from each other so that the corresponding friction clutch can be operated with a very slight force. In the active area of the lining suspension the theoretical disengagement force is produced from the difference between two vertically superposed points of the lines 340 and 342. Such difference is characterised by 360. The disengagement force actually required is reduced by the corresponding lever translation of the operating elements, such as eg the diaphragm spring tongues.This was likewise described in connection with the embodiment according to Figures 17 and 18 as well as the diagrams according to Figures 24 to 27.

In Figure 34 a further operating plate spring characteristic line 440 is shown in dotted lines which has a minimum or a trough point 445 in which the force applied by the diaphragm spring is negative, thus does not act in the engagement direction of the corresponding friction clutch but in the disengagement direction. This means that when exceeding the point 461 during the disengagement phase the friction clutch automatically remains open. The diaphragm spring characteristic line 440 can be associated with a lining suspension characteristic line corresponding to line 442.

In order to obtain minimal disengagement forces, endeavours should be made to keep the path of the lining spring characteristic line 442 as parallel as possible with the diaphragm spring characteristic line 440.

Figure 35 shows the disengagement force curve over the disengagement path for the associated characteristic lines 340 and 342 and 440 and 442 which is to be applied to the operating lever, such as the diaphragm spring tongues, in order to disengage the corresponding friction clutch. As can be seen, the disengagement force curve 349 which is associated with the characteristic lines 340,342, is always in the positive force area, that means that a force is always required in the disengagement direction in order to hold the clutch in the disengaged state. The disengagement force curve 449 which is associated with the characteristic lines 440 and 442 has a partial area 449a in which the disengagement force first decreases and then changes from the positive to the negative force area so that the corresponding friction clutch requires no holding force in the disengaged state.

With the embodiment of a friction clutch 501 shown in Figures 36, 36a and 37, the sensor diaphragm spring 513 is supported axially on the clutch cover 502 by a bayonet-like connection 514. To this end the sensor spring 513 has tabs 513d extending radially from the outer circumference of the ring-like foundation body 513b and supported axially on radial areas 502a, in the form of tabs moulded out of the cover material. The cover tabs 502a are moulded out of the substantially axially aligned edge area 502b of the cover wherein it is expedient if to this end the tabs 502a are at first moulded at least partially by a free cut-out section 502c or 502d from the cover material. By cutting round the tabs 502a at least partially they can be shaped more easily into their ideal position.As can be seen in particular from Figure 37 the tabs 502a and the extension arms or tongues 513d are adapted to each other so that it is possible to centre the sensor spring 513 relative to the cover 502. With the illustrated embodiment the tabs 502a have a small axial step 502e for this purpose.

In order to guarantee a satisfactory positioning of the sensor spring 513 relative to the housing 502 whilst producing the bayonet-like locking connection 514, at least three tabs 502a preferably evenly distributed over the circumference of the cover 502 are adapted in relation to the other cover areas so that after a defined relative rotation between the sensor spring 513 and the cover 502 the corresponding extension arms 513d come to adjoin a circumferential stop 502f and thus further relative rotation between the sensor spring 513 and the cover 502 is avoided.

The stop 502f is formed in the illustrated embodiment as shown in particular in Figure 36a by an axial recess in the cover 502. It is furthermore apparent from Figure 36a that at least some, preferably three tabs 502a form a further turning restriction 502g between the cover 502 and the tongues 513d of the sensor spring 513. In the illustrated embodiment the same tabs 502a form the turning restrictions 502f and 502g for the two turning directions. The stops 502g which hinder unlocking between the sensor spring 513 and the cover 502 are formed by axial radially aligned angled edges of the tongues 502a. The circumferential stops 502f and 502g provide a definite positioning in the circumferential direction of the sensor spring 513 relative to the cover 502.In order to produce the locking connection 514 the sensor spring 513 is tensioned axially in the direction of the cover 502 so that the tongues 513d project axially into the free sections 502c and 502d and come to lie axially above the cover supports 502a. The cover 502 and the sensor spring 513 can then be turned relative to each other until some of the tongues 513d come to adjoin the turning restrictions 502f. A partial relaxation of the sensor spring 513 then takes place so that some of the tongues 513d, viewed circumferentially, come to lie between the corresponding stops 502f and 502g and all the tongues 513d rest on the supports 502a on the cover side. Through the design of the bayonet-like lock 514 according to the invention it is ensured that when fitting the friction clutch 501 the tongues 513d do not come to lie next to the support pads 502a on the cover side.

With the embodiments illustrated up until now the circular ring-shaped body, eg 513b which applies the actual spring force of the sensor spring 513 is provided radially outside of the biasing area or supporting area between the pressure plate and operating diaphragm spring. However in many cases it can also be expedient if the circular ring-shaped body of the sensor plate spring is provided radially inside of the biasing diameter between the pressure plate and operating diaphragm spring. This means that for an embodiment according to Figures 17 and 18 the foundation body 13b which applies the axial tensioning force of the sensor spring 13 is provided radially inside the biasing area 3a between the operating diaphragm spring 4 and the pressure plate 3.

In the embodiment according to Figures 36 and 37 the counter run-up ramps 524 on the cover side are formed by cam-like indentations which are formed in the sheet metal housing 502. Furthermore with this embodiment the coil springs 526 clamped between the housing 502 and the adjustment ring 517 are guided by guide pins 528 which are formed in one piece with the adjustment ring 517 and which extend in the circumferential direction. As is particularly apparent from Figure 21, these guide pins 528 can have in the axial direction an oblong cross-section which matches the inner diameter of the springs 526. The guides 528 extend at least over a partial area of the longitudinal extension of the springs 526 into same. At least a part of the spring windings can thereby be guided and supported at least in the radial direction. Furthermore the springs 526 are prevented from bending or springing out axially. The pins 528 make it much easier to assemble the friction clutch.

The adjustment ring 517 is shown in part in Figure 38. The adjustment ring 517 has radially inwardly aligned moulded areas 527 which support the pin-like circumferentially extending guide areas 528 for the coil springs 526. In the illustrated embodiment the spring socket areas 528 are formed in one piece with the plastics ring 517 which is made as an injection moulded part. The spring guide areas or spring socket areas 528 can however also be formed by individual components or can be all formed together by a single component part which is or are connected to the adjustment ring 517 eg by a snap-fitting locking. Thus all the guide areas 528 can be formed by a ring which can be open if required over the circumference and which is coupled to the adjustment ring 517 by at least three locking parts, preferably designed as a snap-fitting lock.

As with the description of Figures 28 and 29 the coil springs 526 can also be supported radially eg owing to the centrifugal force action, on suitably designed areas of the cover 502 and/or of the adjustment ring 517.

The supports on the cover side for the coil springs 526 are formed by axially extending wings moulded out of the cover material or by axial indentations 526a forming walls. These supporting areas 526a for the springs 526 are thereby preferably designed so that the corresponding ends of the springs are guided and thus secured axially and/or radially against inadmissible movement.

With the embodiment of a clutch 601 illustrated in Figure 39 the sensor spring 613 is provided on the side of the housing 602 remote from the pressure plate 603. By arranging the sensor spring 613 outside of the interior of the housing which holds the pressure plate 603 the thermal strain of the sensor spring 613 can be reduced whereby the danger of this spring 613 setting through thermal overloading is avoided.

Also improved cooling of the spring 613 occurs on the outer side of the housing 602.

The support of the swivel support pad 611 provided on the side of the operating diaphragm spring 604 remote from the cover is achieved by spacer rivets 615 which extend axially through corresponding recesses of the plate spring 604 and housing 602 and are axially connected to the sensor spring 613. With the illustrated embodiment the spacer rivets 615 are rivetted to the sensor spring 613. Instead of spacer rivets 615 other means can also be used which provide a connection between the rolling support pad 611 and the sensor spring 613. Thus for example the sensor spring 613 could have in the radially inner area axially extending tabs which support the rolling pad 611 with corresponding radial areas or which could even form the rolling pad 611 directly through corresponding moulded areas.Instead of the elements 615 rivetted fixedly to the sensor spring it is also possible to use differently designed elements attached eg for articulated movement to the sensor.

With the embodiment according to Figure 40 the sensor spring 713 extends radially inside the swivel bearing 715 for the operating plate spring 704. The sensor spring 713 is supported on the cover 702 by its radially inner areas. To this end the cover 702 has tabs 715 which extend through corresponding slits or recesses in the diaphragm spring 704 and which axially support the sensor diaphragm spring 713.

The adjustment ring 817 illustrated in Figure 41 can be used with a friction clutch according to Figures 20 to 21. The adjustment ring 817 has radially inwards moulded areas 827 which extend radially. The moulded areas 827 have radial attachments 827a which form support areas for the coil springs 826 which are tensioned circumferentially between the clutch cover and the adjustment ring 817. In order to guide coil springs 826 and make it easier to assemble same, a ring 828 is provided which is broken or open on the outer circumference. The ring 828 is connected to the radial moulded areas 827a. To this end the moulded areas 827a can have circumferentially extending indentations or grooves which are designed so that they form a snap fitting connection in conjunction with the ring 828.The supports for the adjustment springs 826 on the cover side are formed by axial tabs 826a of the clutch cover. The axial tabs 826a each have an axial incision 826b to hold the ring 828. The incisions 826b are thereby designed so that the ring 828 has relative to the tabs 826a an axial displacement possibility, at least corresponding to the wear path of the friction clutch. It is particularly expedient for this if the indentations formed in the radial moulded areas 827a for holding the ring 828, and the cut-out sections 826b are formed in opposite directions, viewed in the axial direction, or in other words that the indentations in the moulded areas 827a are open in one axial direction and the cut-out sections 826b are open in the other axial direction.

With the embodiment of a friction clutch 901 illustrated in Figure 42, the support of the operating plate spring 904 in the disengagement direction takes place in a central area of the foundation body 904a of the diaphragm spring 904.

Radially outwards the foundation body 904a is supported on the pressure plate 903 and extends radially inwards over the swivel bearing 905. This means that the swivel bearing 905 is comparatively far away from the inner edge of the foundation body 904a of the diaphragm spring 904 or slit ends which form the tongues of the diaphragm spring 904, in comparison with the diaphragm spring clutches known up until now. In the illustrated embodiment the radial width ratio of the foundation body areas provided radially inside the swivel bearing 905 in relation to the foundation body areas provided radially outside of the swivel bearing 905 is in the order of 1 : 2. It is expedient if this ratio lies between 1 6 and 1 : 2.Through such a support of the operating diaphragm spring 904 it is possible to prevent damage or overstraining of the diaphragm spring foundation body 904a in the area of the swivel bearing 905.

In Figure 42 an axial moulded area 903a is shown in dotted lines which is provided on the pressure plate 903. Through such moulded areas 903a provided on the pressure plate 903, more particularly in the area of the bearing cams 903b it is possible to centre the operating diaphragm spring 904 relative to the clutch 901. The operating diaphragm spring 904 can thus be held in a radial direction relative to the cover 902 through an external diameter centring so that the centring rivets or bolts 915 likewise illustrated in Figure 42 can be dispensed with. Although not shown the external diameter centring can however also be undertaken by tabs or indentations moulded out of the material of the cover 902.

With the friction clutch 901 the sensor spring 913 is designed so that the foundation body 913a applying the force is provided radially inside the cams 903b. In order to support the operating plate spring 904 on one side and for its own support on the cover 902 on the other side the sensor spring 913 has radial extension arms or tongues which extend on one side from the foundation body 913a radially inwards and on the other side from the foundation body 913a radially outwards.

With the modified embodiment of a friction clutch 1001 shown in Figure 43 the force directed against the disengagement force of the friction clutch or swivel force of the operating diaphragm spring 1004 is applied by a sensor spring 1013 which is axially tensioned between the housing 1002 and the pressure plate 1003. With this type of embodiment the operating diaphragm spring 1004 is not supported by a swivel bearing in the disengagement direction in the swivel or tilting area 1005. The bearing of the diaphragm spring 1004 on the swivel pad or support pad 1012 on the cover side is guaranteed through the pretensioning force of the sensor spring 1013.This sensor spring is designed so that during the disengagement process of the friction clutch 1001 the axial force applied by this sensor spring 1013 on the diaphragm spring 1004 is or becomes greater than the necessary disengagement force of the friction clutch 1001. It must thereby be ensured that if there is no wear on the friction linings, the diaphragm spring 1004 always remains adjoining the cover-side support or the swivel support pads 1012. To this end there must be conformity between the individual axially operating and overlapping forces, in the same way as described in connection with the previous embodiments.These forces which are produced by the sensor spring 1013, by the lining suspension, through the leaf spring elements possibly provided between the pressure plate 1003 and housing 1002, through the operating diaphragm spring 1004, through the disengagement force for the friction 1001 and through the adjustment spring elements acting on the adjustment ring 1017, must be suitably adapted to each other.

With the friction clutch 1101 according to Figure 44 the sensor spring 1113 is supported radially outside the ringshaped supporting area 1112 on the cover side. In the illustrated embodiment the mutual support between the operating diaphragm spring 1104 and the sensor spring 1113 is also provided radially outside of the supporting diameter 1103a of the operating diaphragm spring 1104 on the pressure plate 1103. For support on the cover 1102 the sensor spring 1113 has radially outside moulded areas in the form of radially outwardly pointing arms 1113b which are axially supported on the cover 1102 by a bayonet lock 1514 and are secured against rotation in a similar way to that described in connection with Figures 36 to 37.To fit the sensor spring 1113, the cover 1102 has corresponding axial recesses 1502b in which the radially outer support arms of the sensor spring 1113 can be inserted axially to produce the bayonet lock 1514. The abutment of the diaphragm spring 1104 against the swivel pad or support pad 1112 on the cover side is guaranteed by the pretensioning force of the sensor spring 1113.

The function of the clutch is explained in detail in connection with Figure 43. The force of the sensor spring is thereby designed so that it corresponds to the disengagement force in the adjustment point. If after wear on the linings (or wear on other areas) and thus a change in the diaphragm spring angle and a higher diaphragm spring force have occurred, then the diaphragm spring is first moved about the support pad 1012 and up to near the adjustment point. Since at this point the disengagement force is then equal to the sensor force together with the lining spring - residual force - the diaphragm spring swivels on further disengagement about the support pad on the pressure plate until a force equilibrium is again produced between the disengagement force and the sensor force.The diaphragm spring is thereby removed from the support pad on the cover side and is released for adjustment. During further disengagement the disengagement force drops down again, the sensor force prevails and via the pressure plate presses the diaphragm spring against the support pad 1012 on the cover side whereby the diaphragm spring swivels further round this pad. As the diaphragm spring moves from the support on the cover side to the support on the pressure plate side the diaphragm spring alters its function as a double-armed lever. It is temporarily supported on the pressure plate with the disengagement force now existing on the pressure plate and is thereby temporarily lifted from the support pad on the cover side.After further disengagement movement the force of the sensor spring prevails as a result of the drop in force connected therewith and this sensor spring again presses the diaphragm spring against the support pad on the cover side whereby the adjustment device is blocked and the adjustment process is terminated. The diaphragm spring is then active again as a double-armed lever for the further disengagement path. The diaphragm spring is to be designed taking into account all the spring forces which act directly or indirectly against the diaphragm spring. To this belong in particular the forces which are produced by the operating diaphragm spring and the components of the corresponding compensating or adjustment process which can be displaced axially relative to the cover.

The embodiment according to Figure 44 furthermore has the advantage that in the engaged state of the friction clutch the diaphragm spring 1104 is in practice tensioned or active as a double-armed lever and the diaphragm spring 1104 is thus tensioned between the support 1112 on the cover side and the support 1103a on the pressure plate side, but during disengagement of the friction clutch 1101 the diaphragm spring is practically only supported on the sensor spring 1113 and is swivelled about the support area 1113a, with simultaneous axial displacement of the support arm 1113a so that it is then in practice acting as a single-armed lever.

With a suitable design and adaptation, the sensor spring 1113 according to Figure 44 can - as with the sensor diaphragm springs of the other Figures - be supported on any diameter of the operating diaphragm spring 1104. Thus the sensor spring 1113 can be supported on the diaphragm spring 1104 even on a diameter which is located between the swivel area 1105 on the cover side and the support diameter 1103a on the pressure plate side. Furthermore the support for the sensor spring 1113 on the diaphragm spring 1104 could also be provided radially inside the support diameter 1105 on the cover side. The axial supporting force to be applied by the sensor spring 1113 becomes greater the smaller its supporting diameter 1113a becomes on the diaphragm spring 1104.Furthermore the spring area with practically constant force of the sensor spring 1113 must become greater the further away the support diameter 1113a between the springs 1104 and 1113 is removed from the support diameter 1105 on the cover side of the diaphragm spring 1104.

The embodiment according to Figure 45 has an adjustment device 1216 which acts in a similar way to that described with the previous Figures, particularly in connection with Figures 17 to 30. The operating diaphragm spring 1204 is swivel mounted between two ring-shaped rolling supports 1211 and 1212. The support pad 1211 adjoining the pressure plate 1203 is biased by the sensor spring 1213. The friction clutch 1201 has a device 1261 which ensures that throughout the service life of the friction clutch, the ramps of the adjustment ring do not remain sticking to the counter ramps provided on the cover side. In the illustrated embodiment the counter ramps are provided on a support ring 1225 rotationally secured on the cover, in the same way as described in connection with Figure 18. Sticking between the ramps and counter ramps would result in the desired wear adjustment no longer taking place.

The device 1261 forms a tear-loose mechanism which during disengagement of the friction clutch 1201 and in the presence of wear on the friction linings 1207 can exert an axial force on the adjustment ring 1217 whereby the adhesive connection which may be present between the ramps and counter ramps is released. The mechanism 1261 comprises an axially resilient element 1262 which in the illustrated embodiment is axially connected to the diaphragm spring 1204. The element 1262 has a ring-shaped membrane-like or diaphragm -spring-like resilient foundation body 1262a which is connected radially outwards to the diaphragm spring 120^.

From the radially inner edge area of the ring-shaped foundation body 1262a there extends axial tabs 1263 which are spread out round the circumference and which project through axial recesses in the diaphragm spring 1204. At their free end area the tabs 1263 have stop contours in the form of bent areas 1264 which interact with counter stop contours 1265 of the adjustment ring 1217. The counter stop contours 1265 are formed by recesses radially formed in the ring 1217 or by a circumferential groove . The distance between the stop contours 1264 and the counter stop contours 1265 in the engaged state of the friction clutch is measured so that over at least a large part of the clutch disengagement phase there is no contact between the contours 1264 and counter contours 1265.Preferably the stop contours 1264 only come to adjoin the counter stop contours 1265 when the friction clutch is fully disengaged, whereby the element 1262 can be tensioned elastically between the adjustment ring 1217 and the diaphragm spring 1204. It is thereby ensured that as soon as an axial displacement of the swivel support pad 1211 takes place following wear on the linings, the adjustment ring 1217 is automatically lifted off from the run-up ramps on the cover side. Furthermore the mechanism 1261 should prevent an adjustment of the ring 1217 when the disengagement path is too much, for example owing to a faulty basic adjustment of the disengagement system.This is achieved in that when the swivel angle of the diaphragm spring 1204 is too great in the disengagement direction, the resilient element 1262 tensions the adjustment ring 1217 against the diaphragm spring 1204 whereby the adjustment ring 1217 is secured against turning relative to the diaphragm spring 1204. It must thus be ensured that when exceeding the point 46 according to Figure 24 in the disengagement direction the adjustment ring 1217 is held rotationally secured relative to the diaphragm spring 1204 since on exceeding the point 46 the retaining force of the sensor spring 1213 is overcome whereby even in the absence of wear on the clutch disc an adjustment would take place. This would result in a change in the operating point, thus in a change in the installation position of the diaphragm spring 1204, namely in the direction of a smaller contact pressure force.This means that in Figure 24 the operating point 41 would move along the characteristic line 40 in the direction of the minimum marked by 45.

With an embodiment of a friction clutch which is designed according to the details according to Figures 46 to 48, the individual coil springs 1326 are set on tabs 1328 which are formed integral with the clutch cover 1302. The tabs 1328 are moulded out of the sheet metal material of the cover 1302 by forming an eg punched out U-shaped cut section 1302a. The tabs 1328, seen circumferentially, extend in an arc-shape or tangentially and are preferably at least approximately at the same axial level as the directly adjoining cover areas. From Figure 48 it can be seen that with the illustrated embodiment the tab 1328 is off-set by half the material thickness relative to the base areas 1302b of the cover. The width of a tab 1328 is measured so that the coil spring 1326 provided thereon is guided both in the radial and axial direction.

The adjustment ring 1317 biased in the adjustment direction by the springs 1326 has on its inner circumference radially inwardly pointing moulded areas or extension arms 1327 which extend between the cover 1302 and the plate spring 1304.

The extension arms 1327 have radially inwards an axially directed fork or U-shaped moulded area 1327a whose two axially directed tines 1327b engage round a spring guide tab 1328 on both sides. To this end the two tines 1327b extend axially through the cut-out section 1302a of the cover 1302.

The adjustment springs 1326 are supported on the moulded areas 1327a or their tines 1327b.

The adjustment ring 1317 is supported in a similar way by its run-up ramps on the counter run-up ramps 1324 which are imprinted in the cover 1302, as already described in connection with the previous Figures. The cover indentations forming the counter run-up ramps 1324 are however designed so that these form an air passage opening 1324 in the turning direction of the clutch. Through such a design during rotation of the corresponding friction clutch an improved cooling action is achieved through an automatic air circulation. In particular the adjustment ring 1317 which is made of plastics is thereby cooled down whereby the thermal load on even this component can be substantially reduced.

According to a further embodiment the sensor force which acts on the operating diaphragm spring of the friction clutch can be applied for example by leaf spring elements provided between the clutch housing and pressure plate wherein these leaf spring elements can couple the pressure plate and housing together so that they are rotationally secured but axially displaceable relative to each other to a limited extent. With such a design no special sensor spring would be necessary but the leaf spring elements 9 of the friction clutch 1 according to Figures 1 and 2 could be designed so that they additionally undertake the function of the sensor diaphragm spring 13. Both the sensor spring 13 and the rolling ring 11 can thereby be dispensed with.The leaf spring elements 9 must thereby be designed so that during operation of the friction clutch 1 and without any wear on the linings the operating diaphragm spring 4 remains adjoining the rolling support pad 12 on the cover side. As soon as corresponding wear appears on the friction linings 7 whereby the disengagement force of the diaphragm spring 4 increases, the leaf spring elements 9 must allow an adjustment of the diaphragm spring 4 corresponding to the wear. The leaf spring elements installed in the friction clutch advantageously have a practically linear force-path characteristic line at least over the maximum required adjustment path of the friction clutch or pressure plate.

This means that the leaf spring elements 9, as described in connection with Figure 25, should have a characteristic line range 48 according to the characteristic line 47 or 47a.

The invention is not restricted to the embodiments described and illustrated but also embraces variations which can be formed by a combination of individual features or elements described in connection with the present invention.

This application is divided from Patent Application 9616447.0 which itself is divided from Patent Application 9404648.9. Patent Application 96 is also divided from Patent Application 961447.0.

Application 9404648.9 describes and claims a clutch assembly including a friction clutch with a pressure plate which, in use, can be non-rotatably connected with a counter pressure plate with limited freedom of axial movement wherein at least a contact pressure spring biases the pressure plate in the direction of a clutch disc which can be clamped between this pressure plate and the counter pressure plate, furthermore an adjustment device is provided which compensates at least for the wear on the friction linings of the clutch disc and which causes a practically uniform force biasing of the pressure plate through the contact pressure spring, and wherein the friction clutch includes operating means for engagement and disengagement which are operated by means of a release member which can be axially displaced by a disengagement device wherein in dependence on at least the wear on the friction linings the operating means are axially moved in the direction of the disengagement movement and a compensation device is provided in the force flow between the disengagement device and the operating means to at least approximately compensate for the changes in axial displacement of the operating means arising from wear.

Application 9616447.0 describes and claims a friction clutch with a pressure plate which is non-rotatably, but axially limitedly displaceably, connected to a housing, wherein at least a plate spring - in the assembled state of the friction clutch on a counter pressure plate - biases the pressure plate in the direction of a clutch disc which can be clamped between this pressure plate and the counter pressure plate, furthermore an adjustment device is provided which automatically compensates at least the wear of the friction linings of the clutch disc and causes a practically constant force biasing the pressure plate through the plate spring, and the friction clutch has operating means for engagement and disengagement which are operable by axially displaceable disengagement means, wherein means are provided for restricting the disengagement movement of the operating means to hinder the application of an unallowably high biasing and/or deformation to the operating means.

Application 96 (Agents ref P1371P3D) describes and claims a friction clutch with a pressure plate which is connected rotationally secured to, but axially displaceable to a restricted amount in a housing, wherein at least one contact pressure spring is active between the housing and the pressure plate, and an adjustment device is provided which automatically compensates at least for the wear of the friction clutch, the adjustment device being located in a starting position when the friction clutch is not installed on a counter pressure plate, and throughout the service life of the clutch disc causing a practically constant force biasing the pressure plate through the contact pressure spring, the friction clutch having operating means for engagement and disengagement, and the adjustment device, in the non-installed state of the friction clutch, being secured against excessive (unintended) displacement from its starting position.

Claims

1. A friction clutch, more particularly for motor vehicles, with a pressure plate which is connected rotationally secured but axially displaceable to a restricted amount to a housing, wherein between the housing and pressure plate is a contact pressure plate spring which on the one hand can swivel about a swivel bearing supported by the housing and formed by swivel support pads mounted on each side of the contact pressure plate spring and on the other hand - in the assembled state of the friction clutch on a counter pressure plate - biases the pressure plate in the direction of a clutch disc which can be clamped between this and the counter pressure plate, such as a flywheel,

wherein the swivel support pad on the side of the housing can be axially moved by an automatic adjustment device which compensates at least the wear of the friction linings of the clutch disc and the contact pressure plate spring stands under the action of a supporting force whose size can change in dependence on the speed and which is directed towards the swivel pad on the housing side.

2. A friction clutch as claimed in Claim 1 wherein the adjustment device acts between the cover and the plate spring.

3. A friction clutch as claimed in any preceding claim, wherein the supporting force biases the swivel support pad provided on the side of the contact pressure plate spring remote from the housing.

4. A friction clutch as claimed in any preceding claim, wherein the supporting force is dependent on centrifugal force.

5. A friction clutch as claimed in any preceding claim, wherein the contact pressure plate spring is supported only with force-locking engagement on the housing against the disengagement force.

6. A friction clutch as claimed in any preceding claim, wherein the supporting force is produced by at least one plate-spring-like energy accumulator which changes its shape in the event of wear-conditioned adjustment of the contact pressure plate spring or support pad on the side of the cover.

7. A friction clutch as claimed in any preceding claim, wherein the supporting force increases with rising speed.

8. A friction clutch as claimed in any preceding claim, wherein means dependent on centrifugal force are provided which produce a supporting force which acts parallel to the basic supporting force produced by a plate-spring-like element.

9. A friction clutch as claimed in Claim 8 wherein the means dependent on centrifugal force allow wear adjustment only during the stationary or low speed state of the friction clutch.

10. A friction clutch as claimed in any preceding claim, wherein the adjustment function of the adjustment device is practically blocked above a certain speed.

11. A friction clutch as claimed in Claim 8 or Claim 9, wherein the centrifugal-dependent means are formed by tongues which are supported by a plate-spring-like element producing the basic supporting force.

12. A friction clutch as claimed in Claim 11 wherein the tongues are integral with the plate-spring-like element.

13. A friction clutch as claimed in any preceding claim, and which has operating means for engagement and disengagement which are operable by means of a disengagement member, wherein means are provided for restricting the disengagement movement of the operating means.

14. A friction clutch as claimed in Claim 13, wherein the restricting means are provided at least on the housing.

15. A friction clutch as claimed in Claim 13 or Claim 14, wherein the restricting means have at least one restricting stop which is supported by the housing.

16. A friction clutch as claimed in any one of Claims 13 to 15, wherein the restricting means are active between at least one component part of the disengagement member and the clutch housing.

17. A friction clutch as claimed in any one of Claims 13 to 16, wherein the restriction of the disengagement movement is produced by the operating means stopping against a restricting stop supported by the housing.

18. A friction clutch as claimed in any one of Claims 13 to 17, wherein the restricting means limit the angular swivel movement of the contact pressure plate spring.

19. A friction clutch as claimed in any one of Claims 13 to 18, wherein the restricting means are integrated in the friction clutch.

20. A friction clutch as claimed in any one of Claims 13 to 19, wherein the restriction of the disengagement path is produced by the radially inner plate spring tongue tips stopping against a stop supported by the housing.

21. A friction clutch as claimed in any one of Claims 13 to 20, wherein the restricting stop supported by the housing is provided axially between the plate spring tongues and the pressure plate.

22. A friction clutch as claimed in any one of Claims 13 to 21, wherein the restricting stop supported by the housing is formed by a ring-shaped stop area which is connected to the housing by radially aligned webs.

23. A friction clutch as claimed in Claim 22, wherein the webs, starting from the housing, run radially inwards and axially inclined in the direction of the pressure plate and thereby extend through slits between the plate spring tongues.

24. A friction clutch as claimed in any preceding claim, wherein the contact pressure plate spring has a ring-shaped foundation body from which extend radially inwardly directed tongues which form operating means.

25. A friction clutch as claimed in any preceding claim, and which can be operated by means of a disengagement system acting on the operating means, wherein means are provided for restricting the disengagement force acting on the operating means.

26. A friction clutch as claimed in any preceding claim, wherein a lining suspension is provided between the friction linings of the clutch disc.

27. A friction clutch as claimed in any preceding claim, which is a depressed clutch.

28. A friction clutch any preceding claim, wherein means are provided in a disengagement system of the clutch to restrict the operating force which can be exerted on the operating means.

29. A friction clutch as claimed in Claim 25 or 28 wherein the restriction of the operating force is produced by means which are provided in the force transfer path of the disengagement system and which yield on exceeding a certain operating force.

30. A friction clutch as claimed in Claim 29, wherein the yielding means are resilient or elastic.

31. A friction clutch as claimed in any one of Claims 25 to 30 wherein the disengagement system has a pressurised medium circuit in which a valve is provided to restrict the maximum disengagement force.

32. A friction clutch as claimed in any one of Claims 25 to 31, wherein the means restricting the maximum operating force are provided in the force transfer path between a clutch pedal or an operating motor and the operating means of the friction clutch, such as plate spring tongues.

33. A friction clutch as claimed in any one of Claims 25 to 32 wherein the disengagement system has a pressurised medium circuit in which means are provided to restrict the maximum pressure.

34. A friction clutch as claimed in any one of Claims 25 to 33, wherein the means provided to restrict the force acting on the operating means of the friction clutch set a maximum force which is at least slightly greater than the maximum force required to operate the clutch.

35. A friction clutch as claimed in one of Claims 28 to 34, wherein the force restricting means are formed by a buffer, such as eg a hydro or spring accumulator.

36. A friction clutch as claimed in any preceding claim, wherein an operating train is provided for engaging and disengaging the friction clutch between the train and a drive motor, wherein the drive motor is allocated a securing device through which on exceeding a predetermined operating force or a predetermined torque of the drive motor the torque or operating force supplied by the drive motor is restricted to a predeterminable amount.

37. A friction clutch as claimed in any preceding claim, wherein the adjustment device is located in a starting position in the state where the friction clutch is not mounted on a counter pressure plate and which furthermore throughout the service life of the clutch disc exerts a practically constant biasing force on the pressure plate through the contact pressure spring, and wherein the friction clutch has operating means for engagement and disengagement, wherein the adjustment device in the nonassembled state of the friction clutch is secured against excessive (unintended) displacement from its starting position.

38. A friction clutch as claimed in Claim 37, wherein the securing of the adjustment device is lifted at least after the friction clutch is mounted on the counter pressure plate.

39. A friction clutch as claimed in one of Claims 37 or 38 wherein the adjustment device has a rotatable ring which is secured against turning through security means at least in the non-assembled state of the friction clutch.

40. A friction clutch as claimed in one of Claims 37 to 39 wherein the adjustment ring is provided axially between the contact pressure plate and the housing, furthermore has radial extension arms which are adjoined by the contact pressure plate spring when a certain conicity is exceeded.

41. A friction clutch as claimed in one of Claims 37 to 40 wherein the means which when the friction clutch is not fitted cause undue displacement of the adjustment device also block the adjustment device when the friction clutch is fitted and a certain disengagement path of the operating means is exceeded.

42. A friction clutch as claimed in Claim 41 wherein the blocking of the adjustment device is produced by areas provided on the contact pressure plate spring adjoining areas of an adjustment ring of the adjustment device.

43. A friction clutch as claimed in Claim 42 wherein the outer edge area of the contact pressure plate spring is supported on radial extension arms of the adjustment ring when a predetermined conicity of the contact pressure plate spring is exceeded.

44. A friction clutch as claimed in any preceding claim, wherein the swivel bearing supported by the housing can be axially moved by an automatic adjustment device which at least compensates the wear on the friction linings of the clutch disc, which is moved along by a feed mechanism and which operates between the cover and a diaphragm spring, wherein the diaphragm spring stands under the action of a supporting force in the direction of the swivel bearing, and wherein means are provided to limit the swivel movement of the diaphragm spring on disengagement of the clutch.

45. A friction clutch as claimed in Claim 44 wherein the diaphragm spring is installed with a degressive characteristic line over its working range.

46. A friction clutch as claimed in one of Claims 44 and 45 wherein the diaphragm spring is only supported with force engagement against the disengagement force.

47. A friction clutch as claimed in one of Claims 44 to 46 wherein the supporting force and diaphragm spring force are matched with each other so that with the proposed installation position of the diaphragm spring and without wear-conditioned changes in the conicity and over the disengagement path of the diaphragm spring the supporting force is greater than the force applied by the diaphragm spring and counteracting the supporting force, but with a wear-conditioned change in the conicity of the diaphragm spring the supporting force over the partial areas of the disengagement path of the diaphragm spring is less than the force applied by the diaphragm spring against the supporting force.

48. A friction clutch as claimed in one of Claims 44 to 47 wherein the supporting force is applied by at least one energy accumulator such as a spring which changes its shape over a wear-conditioned adjustment of the diaphragm spring or support pad on the cover side.

49. A friction clutch as claimed in one of Claims 44 to 48 wherein the adjustment device is mounted axially between the diaphragm spring and cover.

50. A friction clutch as claimed in one of Claims 44 to 49 wherein the adjustment device contains run-up surfaces such as ramps.

51. A friction clutch as claimed in one of Claims 44 to 50 wherein the supporting force is applied by a diaphragmspring-like element.

52. A friction clutch as claimed in Claim 51, wherein the diaphragm spring applying the supporting force is mounted at the radial level of the axially displaceable support on the diaphragm spring.

53. A friction clutch as claimed in one of Claims 44 to 52 wherein the contact pressure diaphragm spring is supported for swivel movement on the housing between two support pads - of which the one facing the pressure plate is springloaded in the direction of the contact pressure diaphragm spring.

54. A friction clutch as claimed in Claim 53, wherein the support pad spring-loaded by the supporting force is axially displaceable.

55. A friction clutch as claimed in one of Claims 44 to 54, wherein during the displacement of the spring-loaded support pad the disengaging force of the contact pressure diaphragm spring decreases.

56. A friction clutch as claimed in one of Claims 44 to 55, wherein the spring-loaded support pad is moved so far until a force equilibrium is set between the disengagement force of the contact pressure diaphragm spring acting on the support pad and the counter force exerted on this support pad.

57. A friction clutch as claimed in one of Claims 44 to 56, wherein the counter force exerted on the spring-loaded support pad is produced by an energy accumulator which has a substantially constant force over the proposed adjustment area.

58. A friction clutch as claimed in one of Claims 44 to 57, wherein the energy accumulator 13 producing the supporting force acts as a sensor.

59. A friction clutch as claimed in one of Claims 44 to 58, wherein the counter support pad provided on the side of the contact pressure diaphragm spring remote from the springloaded support pad is displaceable axially in the direction of the pressure plate but is lockable in the counter direction.

60. A friction clutch as claimed in one of Claims 44 to 59, wherein the feed device which moves the adjustment device further on is a spring.

61. A friction clutch as claimed in one of Claims 44 to 60, wherein the adjustment device has a cohesive ring-like component which is axially biased by the contact pressure diaphragm spring in the engaged state of the friction clutch.

62. A friction clutch as claimed in one of Claims 44 to 61, wherein the adjustment device has axially rising adjustment ramps.

63. A friction clutch as claimed in Claim 62, wherein the adjustment ramps are provided on the ring-like component.

64. A friction clutch as claimed in one of Claims 61 to 63, wherein the ring-like component supports the counter support pad.

65. A friction clutch as claimed in one of Claims 62 to 64, wherein the run-up ramps interact with corresponding counter run-up ramps.

66. A friction clutch as claimed in Claim 65, wherein the counter run-up ramps are supported by a ring-like component which is mounted between the component supporting the run-up ramps and the cover.

67. A friction clutch as claimed in Claim 66, wherein the counter run-up ramps are installed directly in radially aligned areas of the housing.

68. A friction clutch as claimed in any preceding claim wherein the adjustment device - viewed in the disengagement direction of the friction clutch - acts like a freewheel but is self-locking in the direction opposite the disengagement direction.

69. A friction clutch as claimed in one of Claims 62 to 68, wherein at least the run-up ramps have a rising angle which lies between 40 and 200, preferably in the order of 50 to 120.

70. A friction clutch as claimed in one of Claims 62 to 69, wherein the run-up ramps have a rising angle which causes a self-locking action through the friction engagement of the run-up ramps with counter run-up ramps of another component.

71. A friction clutch as claimed in one of Claims 62 to 70, wherein at least one component supporting the run-up ramps and/or a component supporting the counter run-up ramps or counter run-up areas is spring-biased in the adjustment direction.

72. A friction clutch as claimed in any preceding claim wherein the adjustment device has several displaceable adjustment elements.

73. A friction clutch -as claimed in any preceding claim wherein the adjustment device is dependent on speed.

74. A friction clutch as claimed in any preceding claim wherein the adjustment device is blocked in dependence on speed.

75. A friction clutch as claimed in any preceding claim wherein the adjustment device is blocked at speeds above a certain limit.

76. A friction clutch as claimed in any preceding claim wherein the adjustment device acts at idling speed or speeds below the idling speed.

77. A friction clutch as claimed in any preceding claim wherein the adjustment device is activated practically at zero speed.

78. A friction clutch as claimed in any preceding claim wherein after releasing the clutch disc through the counter pressure plate the contact pressure diaphragm spring changes from a positive force-path curve to a negative force-path curve.

78. A friction clutch as claimed in any preceding claim wherein the parts of the adjustment device having the run-up ramps and/or counter run-up ramps or areas and displaceable relative to the housing are resiliently tensioned.

79. A friction clutch as claimed in Claim 78, wherein the spring tension produces a force in the circumferential direction.

80. A friction clutch as claimed in one of Claims 44 to 79, wherein a sensor spring applying the counter force is supported by its radially outer area on the housing.

81. A friction clutch as claimed in one of Claims 44 to 80, wherein support areas are provided on the housing for a sensor spring producing the counter force.

82. A friction clutch as claimed in any preceding claim wherein a lining suspension or a lining suspension substitute is provided between the friction linings of the clutch disc.

83. A friction clutch as claimed in Claim 82, wherein the lining suspension provided between the friction linings of the clutch disc has a path-force characteristic which over the spring path of the lining suspension is close to the path-force characteristic of the force exerted on the pressure plate by the contact pressure diaphragm spring.

84. A friction clutch as claimed in any preceding claim wherein in the disengaged state of the friction clutch the force required to operate the contact pressure diaphragm spring or friction clutch lies in the order of between -150 to +150 Nm.

85. A friction clutch as claimed in any preceding claim wherein after releasing the clutch disc through the counter pressure plate the contact pressure diaphragm spring changes from a positive force-path curve to a negative force-path curve.

Amendments to the claims have been filed as follows Claims 1. A friction clutch, more particularly for motor vehicles, with a pressure plate which is connected rotationally secured but axially displaceable to a restricted amount to a housing, wherein between the housing and pressure plate is a contact pressure plate spring which on the one hand can swivel about a swivel bearing supported by the housing and formed by swivel support pads mounted on each side of the contact pressure plate spring and on the other hand - in the assembled state of the friction clutch on a counter pressure plate - biases the pressure plate in the direction of a clutch disc which can be clamped between this and the counter pressure plate, such as a flywheel, wherein the swivel support pad on the side of the housing can be axially moved by an automatic adjustment device which compensates at least the wear of the friction linings of the clutch disc and the contact pressure plate spring stands under the action of a supporting force whose size can change in dependence on the speed and which is directed towards the swivel pad on the housing side.

5. A friction clutch as claimed in any preceding claim, where in the contact pressure plate spring is supported only with force-locking engagement on the housing against the disengagement force.

34. A friction clutch as claimed in any one of Claims 25 to 33, wherein the means provided to restrict the force acting on the operating means of the friction clutch 6eC a maximum force which is at least slightly greater than the maximum force requirud tv operate he clutch.

35. A friction clutch as claimed in one of Claims 26 to 34, wherein the force restricting means are formed by a buffer such as eg a hydro or spring accumulator.

36. A friction clutch as claimed in any preceding claim in combination with an operating train provided between the Clutch ano a clutch actuator tor engaging and disengaging the clutch, wherein the act our ie a=socist - d with a security device through which on exceeding a predetermined actuation force or a predetermined torque of the actuator, the torque or actuation force supplied to the clutch is restricted to a prodeterminable amount.

37. A friction clutch as claimed in any preceding claim, wherein the swivel bearing supported by the housing can be axially moved by an automatic adjustment device which at least compensates the wear on the friction linings of the clutch disc, which is moved along by a feed mechanism and which operates between the cover and a diaphragm spring, wherein the diaphragm spring stands under the action or n supporting force in the direction of the swivel bearing. and wherein means are provided to limit the swivel movement of the diaphragm spring on disengagement of the clutch.

38. A friction clutch as claimed in Claim 37 wherein the diaphragm spring is installed with a degressive characteristic line over its working range.

35. A friction clutch as claimed in ons of Claims 37 and 38 wherein the diaphragm spring is only supported with force engagement against the disengagement force.

4C. A friction clutch as claimed in one of Claims 37 to 39 wherein the supporting force and diaphragm spring force are matched with each other so that with the proposed installation position of the diaphragm spring and without wear-conditioned changes in the conicity and over the disengagement path of the diaphragm spring the supporting force is greater than the force applied by the diaphragm spring and counteracting the supporting force, but with a wear-conditioned change in the conicity of the diaphragm spring the supporting force over the partial areas of the disengagement path of the diaphragm spring is less than the force applied by the diaphragm spring against the supporting force.

41. A friction clutch as claimed in one of Claims 37 to 40 wherein the supporting force is applied by at least one energy accumulator such as a spring which changes its shape over a wear-conditioned adjustment of the diaphragm spring or support pad on the cover side.

42 A friction clutch as claimed in one of Claims 37 to 4@ wherein the adjustment device is mounted axially between the diaphragm spring and cover 43. A friction clutch as claimed in one of Claims 37 to 42 wherein the adjustment device contains run-up surfaces such as ramps.

44. A friction clutch as claimed in one of Claim 37 to 43 wherein the supporting force is applied by a dnaphzagm- spring-like element.

45. A friction clutch ac claimed in Claim 44, wh.crcin the diaphragm spring applying the supporting force is mounted at the radial level of the axially displaceable support on the diaphragm spring.

46. A friction clutch as claimed in one of Claims 37 to 45 wherein the contact pressure diaphragm spring is supported for swivel movement on the housing between two support pads - of which the one facing the pressure plate is springloaded in the direction of the contact pressure diaphragm spring.

47. A friction clutch as claimed in Claim 46, wherein the support pad spring-loaded by the supporting force is axially displaceable.

48. A friction clutch as claimed in one of Claims 37 to 47, wherein during the displacement of the spring-loaded support pad the disengaging force of the contact pressure diaphragm spring dccroaoco.

49. A friction clutch as claimed in one of Claims 37 to 48, wherein the spring-loaded support pad is moved so far until a force equilibrium is set between the disengagement force of the contact pressure diaphragm sprint acting on the support pad and the counter force exerted on this support pad.

SO. A friction clutch as claimed in one of Claims 37 to 49.

wherein the counter force exerted on the spring-loaded support pad is produced by an energy accumulator which has a substantially constant force over the proposes adjustment area.

51. A friction clutch as claimed in one of Claims 37 to SO, wherein the energy accumulator 13 producing tne supporting force acts as neor 52. A friction clutch as claimed in one of Claims 37 to 51, wherein the counter support pad provided on the side of the contact pressure diaphragm spring remote from the springloaded support pad is displaceable axially in the direction of the pressure plate but is lockable in the counter direction.

53. A friction clutch ae claimed in one of Claims 37 to 52, wherein the feed device which moves the adjustment device further on is a spring.

54. A friction clutch as claimed in one of Ciairnc 37 to 53, wherein the adjustment device has a cohesive ring-like component which is axially biased by the contact pressure diaphragm spring in the engaged state of the friction clutch.

55. A friction clutch as claimed in one of Claims 37 to 54, wherein the adjustment device has axially rising adjustment ramps.

56. A friction clutch as claimed in Claim 55, wherein the adjustment ramps are provided on the ring-like component.

57. A friction clutch as claimed in one of Claims 54 to 56, wherein the ring-like component supports the counter support pad.

58. A friction clutch as claimed in one of Claims SS to 57, wherein the run-up ramps interact with corresponding counter run-up ramps.

59. A friction clutch aa claimed in Claim SG, wherein the counter run-up ramps are supported by a ring-like ccttonent which is mounted between the component supporting the run-up ramps and the cover.

60. A friction clutch as claimed in Claim 59, wherein the counter run-up ramps are installed directly in radially aligned areas of the housing.

61. A friction clutch as claimed in any preceding claim wherein the adjustment devices vicwcd in the diaongagement direction of the friction clutch - acts like a freewheel but is self-locking in the direction opposite the disengagement direction.

62. A friction clutch as claimed in one of Claims 55 to 61, wherein at least the run-up ramps have a rising angle which lies between 40 and 20 , preferably in the order of 50 to 12 .

63. A friction clutch as claimed in one of Claims 55 to 62, wherein the run-up ramps have a rising angle which causes a sel oclcing action through the friction engagement of the run-up ramps with counter run-up ramps of another e.omponent..

64. A friction clutch as claimed in one of Claims 55 to 63, wherein at least one component supporting the run-up ramps a or a component supportingo the counter run-up ramps or counter run-up areas is spring-biased in the adjustment direction.

65. A friction clutch as claimed in any preceding claim wherein the adjustment device has several displaceable adjustment elements.

6s. A friction clutch ss claimed in any preceding claim wherein the adjustment device is dependent on speed.

67. A friction clutch as claimed in any preceding claim wherein the adjustment device is blocked in dependence on speed.

68. A friction clutch as claimed in any preceding claim wherein the adjustment device is blocked at speeds above a certain limit.

69. A friction clutch as claimed in any preceding claim wherein the adjustment device acts at idling speed or speeds below the idling speed.

70. A friction clutch as claimed in any preceding claim wherein the adjustment device is activated practically at zero speed.

71. A friction clutch as claimed in any preceding claim wherein the parts of the adjustment device having the run-up ramps and/or counter run-up ramps or areas and displaceable relative to the housing are resiliently tension.

72. A friction clutch as claimed in Claim 71, wherein the spring tension produces a force in the circutrferetial direcLiwa.

73. A friction clutch as claimed in one of Clams 37 to 72, wherein a sensor spring applying the counter force is supported by its radially outer area on tne noosing.

74. A friction clutch as claimed in one of Claims 37 to 73, wherein support areas are provided on the housing fcr a sensor spring producing tne counter force.

75. A friction clutch as claimed in any preceding claim wherein a lining suspension or a lining suspension substitute is provided between the friction linings of the clutch disc.

76. A friction clutch as claimed in Claim 7s, wherein the lining suspension provided between the friction linings of the clutch disc has a path-force characteristic which over the spring path of the lining suspension is close tu the path-force characteristic of the force exerted on the pressure plate by the contact pressure diaphragm spring.

II. A friction clutch as claimed in any preceding claim wherein in the di:enga.d state of the friction clutch the force required to operate the contact pressure diaphragm spring or friction clutch lies in the order of between -150 to +150 Nm.