GB2294301A

GB2294301A - Friction clutch

Info

Publication number: GB2294301A
Application number: GB9524222A
Authority: GB
Inventors: Paul Maucher
Original assignee: LuK Lamellen und Kupplungsbau GmbH
Current assignee: LuK Lamellen und Kupplungsbau GmbH
Priority date: 1991-11-26
Filing date: 1992-11-23
Publication date: 1996-04-24
Anticipated expiration: 2012-11-23
Also published as: GB2294301B; GB9524222D0

Abstract

A friction clutch, particularly for motor vehicles, has a pressure plate 3 which is connected to a housing 2 in a manner rotationally fast but axially displaceable to a limited extent. A pressure spring 4 between the housing 2 and the pressure plate 3 acts against the pressure plate in the direction of a clutch disc 8 clampable between the pressure plate and a counter pressure plate 6, which might be a flywheel. An adjustment device 16 is provided which compensates for wear of the friction linings 7 of the clutch disc 8, and the adjustment device takes account of and at least reduces detensioning of the pressure spring which normally arises as a result of wear of the friction linings over the service life of the clutch. The adjustment device is blockable, that is it cannot be adjusted, under certain conditions which may depend on the operating state of the clutch. <IMAGE>

Description

FRICTION CLUTCH The invention relates to a friction clutch, more particularly for motor vehicles, with a pressure plate which is connected rotationally secured but axially displaceable to a limited extent to a housing wherein between the housing and pressure plate there is at least one active tensioning contact pressure spring which biases the pressure plate in the direction of a clutch disc which can be clamped between the pressure 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 DE-AS 1 267 916.

The object of the present invention is to improve friction clutches of this kind with regard to their function and service life. More particularly as a result of the invention the forces required to operate such friction clutches are to be reduced and a practically constant disengagement force curve is to be ensured throughout their service life. Furthermore the friction clutches according to the invention are to be produced in a particularly simple and economic 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 limited extent to a housing wherein between the housing and pressure plate there is at least one active contact pressure spring which biases the pressure plate in the direction of a clutch disc which can be clamped between the pressure plate and a counter pressure plate, such as a flywheel, wherein an adjustment device is provided which compensates for the wear of the friction linings of the clutch disc and which causes a practically constant 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 over a partial region of the operating path of the operating means causes a gradual reduction of the torque which can be transferred by the friction clutch or clutch disc.

Through a device of this kind it can likewise be achieved that during the engagement process of the friction clutch and with the beginning of the tensioning of the friction linings between the pressure and counter plate a gradual or progressive build-up of the torque transferrable by the friction clutch takes place.

Through the design of a friction clutch according to the invention it is ensured that the contact pressure plate 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 reduction of the torque transferrable by the friction clutch during a disengagement process, it is possible to achieve a reduction or minimization of the disengagement force path or of the maximum disengagement force required. This is due to the fact that the device helps the operation, more particularly the disengagement process, of the friction clutch.To this end the device can have axially resilient pliable means which exert a reaction force on the operating means and/or on the contact pressure spring and/or on the pressure plate and/or on the counter pressure plate wherein this reaction force is directed against the force exerted by the contact pressure spring on the pressure plate and is connected in series.

It can be particularly advantageous if the device of the friction clutch is mounted so that during the disengagement process over a partial section of the axial displacement path of the pressure plate areas biased by the contact pressure spring it causes a gradual reduction of the torque transferrable by the friction clutch or clutch disc.

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

In other cases it can however 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 arrangement of this kind has been proposed for example through DE-OS 37 42 354 and DE-OS 1 450 201.

In 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 through a socalled "lining suspension", eg through lining spring segments provided between the linings. Such precautionary measures are known for example through DE-OS 36 31 863.

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

For the functioning and construction of a friction clutch according to the invention it can be particularly expedient if the device allows an axial resilient flexibility between clutch components wherein the device is arranged and designed so that when the clutch is open 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 rises gradually to the maximum wherein this rise preferably takes place only over a partial area 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 decrease or gradual increase in the torque transferrable 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 compensating any possible deformations which may be present on the clutch components such as more particularly the clutch disc, pressure plate and counter pressure plate.

In order to minimize the forces required for operating the friction clutch according to the invention 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 which thus means that the contact pressure spring has a declining force path, at least over a partial area of its compression or deformation path.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 partial area of the disengagement path the tensioning or deformation of the contact pressure spring is assisted 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 which exists in the disengagement area, the force exerted by the contact pressure spring on the pressure plate or friction linings decreases. The force path which is effectively required to disengage the friction clutch is produced, insofar as no additional overlying spring actions are present, from the difference between the force path applied by the device and the force path of the contact pressure spring.As the pressure plate is lifted from the friction linings and during the release of the clutch disc through the pressure plate the required remaining disengagement force path or the required disengagement force is mainly determined 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 during 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 by bringing the spring characteristic or force characteristic of the device close to the contact pressure spring characteristic up to the release of the clutch disc through the pressure plate only a very low operating force, and in extreme cases practically no operating force at all, need be required for the contact pressure spring.

A particularly advantageous contact pressure spring is a plate spring which on the 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 plate spring can 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 swivel mounted 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 them on the pressure plate is greatest when the friction clutch is in the engaged state but reduces during the disengagement process.This can occur for example through the coil springs being positioned inclined relative to the axis of rotation of the friction clutch.

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

In a particularly advantageous way the friction clutch can have a plate 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 minimum.

It can furthermore be advantageous if the friction clutch is operable by way of a disengagement system engaging on the operating means, such as eg on the tongue tips of the plate spring wherein the disengagement system can have a clutch pedal which is designed similar to the accelerator pedal and is mounted in the inside of the vehicle. Such a design of the clutch pedal can be particularly advantageous since through the design according to the invention the force or force path required for disengaging the friction clutch can be brought to a very low level so that the operating force can be better measured out through a clutch pedal designed like an accelerator pedal.

As a result of the design of a friction clutch according to the invention 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 make the components correspondingly smaller and lighter so that a considerable reduction in the manufacturing costs can be achieved. By reducing the disengagement forces it is furthermore possible to reduce the friction and elasticity losses in the clutch and disengagement system and thus to improve significantly the efficiency of the friction clutch/disengagement system. The entire system can thus have an optimum design whereby the clutch comfort can be considerably improved.

The design according to the invention can be used quite generally with friction clutches and more particularly in those which have been proposed for example by 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 51-126452, JP-GM 3-19131, JP-GM 353628.

The present invention furthermore relates to the earlier applications DE-P 42 07 528.9 and DE-P 42 06 904.1 whose contents belong expressly to the contents disclosed in the present invention.

The use of a friction clutch with an automatic or independent compensation of at least the lining wear whereby an almost constant tensioning force of the clutch disc is guaranteed at least during the service life of the friction clutch - is particularly advantageous in connection with clutch assemblies wherein the friction clutch, clutch disc and the counter pressure plate, such as for example a flywheel, form one assembly unit or module. With an assembly unit of this kind it is advantageous for reasons of cost if the clutch housing is connected to the counter pressure plate by a non-releasable connection such as for example a welded connection or a shape-locking connection, for example through plastic material deformation. Through such a connection it is possible to dispense with the fixing means normally used, such as screws.With assembly units of this kind it is practically impossible to exchange the clutch disc or clutch linings due to exceeding the wear limits without destroying the components, such as for example the clutch housing. By using a wear-adjusting clutch the assembly unit can be designed so that this unit guarantees satisfactory functioning throughout the entire service life of the vehicle. Thus as a result of the design according to the invention the wear reserve of the clutch disc and the adjustment reserve of the friction clutch or clutch module can be made so great that the clutch service life and thus also the service life of the assembly unit can reliably reach at least that of the vehicle.

According to a further embodiment of the invention 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 one flywheel mass connectable with a gearbox and the second flywheel mass is connectable with the driven shaft of an internal combustion engine. Twin-mass flywheels in which the friction clutch according to the invention 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 also belong to the disclosed contents of the present invention so that the features described in these applications can be combined with the present invention.More particularly the clutch housing or clutch cover can be connected to the flywheel mass supporting same by means of a connection which cannot be released without destruction, as shown and described for example for the various embodiments in DE-OS 41 17 579.

By using a friction clutch with a device which compensates at least the wear on the linings it is possible to optimize the design of the friction clutch still further, particularly the energy accumulator which applies the tensioning force for the clutch disc. This energy accumulator can thus be designed so that it practically only applies the tensioning force for the clutch disc which is necessary to transfer the desired torque. The energy accumulator can be formed by at least one plate spring or by several coil springs. Furthermore the use of a selfadjusting friction clutch is advantageous in connection with twin-mass flywheels wherein the rotationally elastic damper mounted between the two flywheel masses is provided radially outside of the clutch disc or outer friction diameter of the friction face 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 accordIng to the ratio of the average friction radii in order to be able to transfer a 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 over the disengagement path of the torque transferrable by the clutch disc according to claim 1 it is however possible to achieve a reduction in the disengagement force whereby an increase in the disengagement force can be avoided or even a reduction in the disengagement force compared to a conventional clutch can be achieved through a corresponding design of the friction clutch.

Through the design of a friction clutch according to the invention it can thereby be ensured that the disengagement force can be kept low despite a reduced outer diameter of the friction lining and the higher contact pressure thereby required. The lower disengagement force also produces a reduction in the load on the rolling bearing through which the two flywheel masses can rotate relative to each other.

Furthermore the wear adjustment increases the service life of the clutch so that a replacement of the parts, more particularly the clutch disc, is no longer necessary during the service life of the vehicle. The clutch cover can thus be fixedly connected to the flywheel mass connectable with the gearbox, for example by rivetting or welding. This is then particularly advantageous if the installation space or contours of the clutch bell are restricted and no longer allow a connection of the clutch cover with the flywheel on the gearbox side in the normal way by screws.

In the case of friction clutches with integrated adjustment means for the wear on the linings, with a conventional fixing of the clutch unit comprising the friction clutch and flywheel on the driven output shaft of an internal combustion engine, axial, rotational and torsional vibrations which are excited by the output shaft of the internal combustion engine,such as in particular the crankshaft, are transferred to the clutch unit. So that the function of the clutch unit or adjustment device is not impaired through such vibrations and more particularly in order that any undesired adjustment of the wear compensating device is suppressed, the inertia forces of those components which act on the device must be taken into consideration when designing the said adjustment device.In order to avoid these undesired side effects caused in particular by the axial and torsion vibrations, and to avoid the higher costs associated therewith when designing an adjustment device for compensating the lining wear, according to a further inventive idea the clutch unit having the adjustment device is substantially neutralized with regard to the axial and bending vibrations excited by the output shaft of the internal combustion engine. This can be achieved in that the clutch unit is connectable to the output shaft of the internal combustion engine by an axially elastic resiliently pliable component.The rigidity of this component is measured so that the axial and torsional or bending vibrations produced by the output shaft of the internal combustion engine on the clutch unit are suppressed or at least dampened to such an extent by the said elastic component that a satisfactory function of the friction clutch is guaranteed, and more particularly the adjustment device thereof. Such elastic components are known for example through EP-OS 0 385 752 and 0 464 997 as well as SAE Technical Paper 9 003 91. The content of these publications should likewise belong to the contents disclosed in the present invention. 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 when the friction clutch is disengaged - through flywheel vibrations and/or vibrations of the plate spring.In the case of clutch units or clutch assemblies without a device at least substantially suppressing these vibrations, such as in particular an axially pliable disc, such vibrations can lead to a modified adjustment independently of the state of wear of the clutch disc wherein the plate spring of the friction clutch could be regulated down in the contact pressure force to a force minimum whereby the transfer of the desired torque would no longer be guaranteed.

According to a further inventive design a friction clutch with an automatic or independent compensation which can be formed in particular according to the present invention can advantageously be used in a drive unit, more particularly for motor vehicles which consists of an automatic or semiautomatic 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 with control or regulation in dependence on the operation of the gearbox.

The friction clutch is preferably operable fully automatically. An automated or fully automated operation of a friction clutch has already been proposed for example by DE-OS 40 11 850.9 so that the method of operation and means required are dispensed with in this specification.

With the hitherto known drive units with automatic or semiautomatic gearbox and conventional friction clutch there have been considerable problems up until now regarding the operation of the clutch and the design of the actuators required therefor,such as eg the piston/cylinder units and/or electromotors. As a result of the comparatively high disengagement forces which are required with conventional clutches, very heavy or large dimensioned actuators are necessary. This means large structural volumes, high weight and high costs. Also large actuators of this kind are comparatively slow in their response time owing to their mass inertia. When using servo cylinders a greater volume flow of pressurised medium is required so that the supply pump must also be made comparatively large in order to guarantee the desired operating time for the corresponding clutch.In order to overcome the aforesaid disadvantages at least in part it has been proposed for example by DE-OS 33 09 427 to reduce the operating force for disengaging the clutch through corresponding compensation springs in order to be able to use smaller dimensioned actuators. Since however in conventional clutches the disengagement force fluctuates quite severely over the service life, that is the disengagement force is relatively low in the new state but rises with increasing wear on the linings during the service life, only a part of the disengagement force normally required can be broken down by a compensation spring.

Taking into account all the tolerances and despite using compensation springs a disengagement capacity of the actuators will be necessary which is greater than that for a new conventional clutch. By using a friction clutch according to the invention with compensations for the wear on the lining in connection with a drive unit, comprising a motor and an automatic or semi-automatic gearbox the disengagement force can be lowered quite considerably compared to the aforesaid state of the art, namely directly in the clutch wherein this disengagement force value or disengagement force curve of the new clutch remains practically unchanged during the entire service life.This results in considerable advantages for the design of the actuators since their drive performance or operating performance can be kept correspondingly lower whereby the forces or pressures occurring in the entire disengagement system are correspondingly lower. The losses which occur 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 37 in which: Figure 1 is a view of a friction clutch according to the invention; Figure 2 is a section according to the line II-II of Figure 1; Figure 3 shows an adjustment ring used with the friction clutch according to Figure 1; Figure 4 is a section along the line IV-IV of Figure 3; Figure 5 shows a support ring used with the friction clutch according to Figures 1 and 2; Figure 6 is a section along the line VI-VI- of Figure 5; Figures 7 and 7a show a spring which exerts a torsional force on the adjustment ring; Figures 8 to 11 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 12 and 13 show a further possibility for creating a friction clutch wherein Figure 13 shows a section along the line XIII of Figure 12; Figure 14 is a view of the adjustment ring used with the friction clutch according to Figures 12 and 13; Figures 15 to 17 show details of a further friction clutch with a compensating device; Figures 18 and 19 are diagrams with various characteristic lines from which can be seen the interaction of the contact pressure plate spring and the lining suspension as well as the ensuing action on the disengagement force curve of the friction clutch; Figure 20 is a partial view of a further friction clutch according to the invention; Figure 20a is a partial view in the direction of arrow A of Figure 20; Figure 21 is a section according to the line XXI of Figure 20;; Figure 22 is a partial view of an adjustment ring usable with a friction clutch according to Figures 20 to 21; Figures 23 and 24 show further embodiments of friction clutches according to the invention; Figure 25 is a view of an adjustment ring which would be usable with a friction clutch according to Figures 12 and 13 or 20 to 21; Figures 26 and 27 show additional variations of friction clutches; Figure 28 is a view of a friction clutch according to the invention; Figure 29 is a section along the line II-II of Figure 28; Figure 30 is a section along the line III-III of Figure 28; Figure 31 is a partial view along the line IV-IV of Figure 28; Figure 32 is a view of an adjustment ring used with the friction clutch according to Figures 28 and 29; Figures 33 and 34 show details of further embodiments of friction clutches according to the invention;; Figure 35 shows a flywheel divided into two masses and provided with a rotary vibration damper and a friction clutch according to the present invention; and Figures 36 and 37 show a torque transfer device with a friction clutch according to the invention.

The friction clutch 1 shown in Figures 1 and 2 has a housing 2 and a pressure disc 3 connected therewith rotationally secured but axially displaceable to a limited extent. A contact pressure plate 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.

In the illustrated embodiment the plate spring 4 has a ringlike foundation body 4a applying the contact pressure force and from which operating tongues 4b emerge radially inwards.

The plate 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 plate spring 4 is axially held and clamped. The swivel support pad 11 provided on the side of the plate spring 4 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 12 is formed by a plate spring or by a plate-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 plate spring 4 and thus also in the direction of the housing 2.The plate spring 13 provided between the pressure disc 3 and the operating plate spring 4 has an outer ring-like 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 plate-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 platespring 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 pressed 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 plate-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 plate-spring like component 13 so that the plate-spring like component 13 is first pretensioned 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 plate-spring like component 13 relative to the housing 2.The support areas of the plate-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 plate spring 4 and where applicable the plate-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 l5a which extends axially through a cut-out section provided between adjoining plate spring tongues 4b and which can be engaged partially by areas 13d formed on the tongue 13c of the plate spring 13 associated therewith.

The plate-spring like component or the plate 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 acting on 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 plate 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 plate spring tongues during the operation of the friction clutch 1.

The swivel pad 12 on the side of the housing is supported on the housing 2 by an adjustment device 16. This adjustment device 16 ensures that in the case of an axial displacement of the swivel bearings 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 plate 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.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 8 to 11.

The adjustment device 16 comprises a spring-biased adjustment element in the form of a ring-like component 17 which is illustrated in Figures 3 and 4. 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 2) 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 sections 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 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. The adjustment element 17 can however also be made as a pressed 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 3, 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. 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 5 and 6, prevents any slipping between the run-up ramps 18 and 24. The angle 23 can be in the range from 5 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 1 and 2 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 27 is installed with resilient tension.

The support ring 25 illustrated in Figures 5 and 6 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 3 and 5 the run-up ramps 18 and the counter run-up ramps 24 are spread out in the circumferential direction in a similar way. 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 2 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 connected 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. 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 2, 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 2, 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 7. In the relaxed state of the leg spring 26 the legs 27,28 are offset 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. The numeral 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 2.

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 1 and 2 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 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 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 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 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 plate spring 4 exerts no axial force on this swivel pad 12 and thus a satisfactory turning of the adjustment ring 17 is guaranteed.

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 8 to 11.

The line 40 in Figure 8 shows the axial force which is produced in dependence on the change in the conicity of the plate spring 4, namely during the deformation of the plate 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 plate spring is shown on the ordinate. The point 41 represents the installation position of the plate spring 4 when the clutch 1 is closed, thus the position wherein the plate 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 of the plate 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 plate spring 4 on the pressure disc 3. It is advantageous if the axial force which can be applied by elastic deformation of the spring segments 10 corresponds at least to the force which is exerted by the plate 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. On exceeding the point 44 the friction linings 7 are released whereby as a result of the degressive characteristic line range of the plate 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 9. This characteristic line 47 corresponds to that which is produced when the plate-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 plate-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 8.The support force to be applied by the sensor spring 13 is reduced compared to the force of the plate spring 4 corresponding to point 44 in accordance with the lever translation of this plate 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 plate 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 plate-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 so that the axial force applied as a whole through the lining suspension when the clutch disc 8 is not tensioned corresponds at least to the disengagement force on the plate spring 4 corresponding to point 44, and is preferably somewhat higher.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.

It is expedient if the pretension of the suspension provided between the linings is in the order of 0.3 mm to 0.8 mm, preferably in the order of 0.5 mm. 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 at least 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 subject of the present invention.

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 according to Figure 10 the accumulating forces exerted first by the lining suspension 10 and the sensor spring 13 on the plate spring 4 as well as the force which is only exerted by the sensor spring 13 on the plate spring 4 after the pressure disc 3 has been lifted from the friction linings 7, is/are greater than or at least the same as the disengagement force engaging in the area 4c of the plate spring tongue tips and changing according to Figure 10 over the disengagement path.

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

In the event of axial wear, more particularly 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 and thus the contact pressure force applied by the plate 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 force equilibrium originally present when disengaging the clutch 1 in the area of the swivel support pad 11 between the operating plate spring 4 and the sensor spring 13. The increase in the plate 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 10 by the chain-dotted line 50. The increase in the disengaging force curve during the disengaging process of the friction clutch 1 overcomes the radial force exerted by the sensor spring 13 on the plate 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 plate spring 4 is supported on the biasing area 3a of the pressure disc 3 so that this plate 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 plate 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 8 this change takes place in the sense of reducing the forces exerted by the plate spring 4. This change takes place until the axial force exerted by the plate 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 8 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 during the disengaging process of the friction clutch 1 the adjustment ring 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 plate spring 4 is guaranteed. After the adjustment process the disengagement force path again corresponds to the line 49 according to Figure 10. The lines 50 and 51 of Figure 10 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 11 shows the forces curve of the force exerted during the disengagement process on the housing 2 or on the plate spring 13 wherein the extreme values were capped. Starting from the engaged position according to Figure 1 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 8) of the plate spring 4.

During the disengagement process the axial force exerted by the plate spring 4 on the housing 2 or swivel pad 12 decreases according to the line 52 of Figure 11, namely to the point 53. If in the case of a conventional clutch wherein the plate 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 plate 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 plate spring 4 in the area of the swivel bearing 5 is taken up by the sensor spring 13.On reaching the point 54 the plate 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. 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 4e and of the axial force exerted in the radial area 3a on the plate spring 4 by the lining suspension 10.On exceeding the point 54 in the disengagement direction the axial force exerted by the plate spring 4 on the swivel pad 11 is taken up by the counter force applied by the sensor plate 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 11 shows that as the disengagement path increases so the disengagement force or the force exerted by the plate 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 plate 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 plate 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 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 plate spring 4 and the sensor spring 13.

In practice the described adjustment takes place continuously or in very small steps so t" t 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 plate spring 4 and a corresponding adaptation of the curve 47 of the sensor spring 13. An axial setting of the lining suspension 10 would only lead to the plate spring 4 occupying a more pressed-through position compared to that shown in Figure 1, whereby the contact pressure force exerted by the plate spring 4 on the pressure disc would be slightly less, as can be seen in connection with the characteristic line 40 according to Figure 8. 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 plate 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 plate 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 9 and designated by reference numeral 47a. A rise in the support force for the operating plate spring 4 with increasing wear can compensate at least partially a drop in the contact pressure force of the operating plate 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 plate 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 settling 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 9.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 plate 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 spring 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 plate spring 4. The axial force applied by the leaf springs 9 is overlapped with the forces applied by the sensor spring 13 and plate spring 4 and with the disengagement force. For better understanding this was not taken into consideration in the description of the diagrams according to Figures 8 to 11.The entire force biasing the operating plate 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 plate 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 plate spring 4 becomes greater. Thus for example over the path 48 according to Figure 9 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 9 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 plate spring 4. Adding the force curve according to the characteristic lines 47b and the plate spring characteristic line produces the resulting force curve which acts axially on the plate spring 4, namely in the sense of pressing the plate spring 4 against the swivel support pad 12 on the cover side. In order to obtain a curve according to the line 47a it is expedient to design to the sensor plate spring so that it has a characteristic line curve corresponding to the line 47c of Figure 9.

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 throughout the service life of the friction clutch the plate 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 plate spring 4 the contact pressure force applied by the plate 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 plate 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 8 represents the resulting force curve of the operating plate 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 8 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 through the plate spring 4 on the sensor spring 13 reduces, namely as a result of the counter moment exerted by the leaf springs 9 on the plate 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 plate spring 4 and the pressure plate 3.When designing the friction clutch 1 it is particularly important that the increase in the tension of the leaf springs 9 which occurs through wear on the linings is less than the increase in the disengagement force which takes place owing to the same wear on the linings and which causes the swivelling through of the sensor spring 13 which is required for the adjustment. Otherwise the contact pressure force of the pressure plate 13 for the friction linings 7 in the engaged state of the friction clutch would drop, and no adjustment could take place.

The friction clutch 101 shown in Figures 12 and 13 mainly differs from the friction clutch 1 shown in Figures 1 and 2 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 2 to 4. 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 14, 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 formed areas which are formed integral with the clutch housing 2.

Thus the stops 128 can be formed for example by indentations or tabs pressed axially out of a sheet metal housing 2. As can be seen in particular from Figures 13 and 14 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 13, the cover 2 can have on its radially inner edge area axial formed areas 130 which overlap the spring 126 in the axial direction. Instead of individual formed 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 plate spring 4.

Guiding the adjustment springs 126 according to Figures 12 to 14 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 is 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 1 and 2. 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. Such blocking means can 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 12 and 13 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 13 the wire ring 11 provided in Figure 2 is omitted and has been replaced by formed 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 plate spring 4.

Figures 15 to 17 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 218 are mounted rotatable on these attachments 215a. The cover 202 is provided with formed 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 15 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 formed 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 218 are turned and the displacement is compensated by the ramps 218 running up on the ramps 224.

The axial support of the sensor plate spring 213 on the housing 202 is carried out by tabs 214 which were formed 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 218 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 14 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 formed 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 plate spring 4 and/or the sensor spring 13.

In the diagram according to Figure 18 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 plate 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 plate spring between two radially spaced supports, namely as described in connection with the characteristic line 40 according to Figure 8 and in conjunction with the plate spring 4.

The plate spring characteristic line 340 can be combined with a lining spring characteristic line 342. As can be seen from Figure 18 the force-path curve of the lining spring segment characteristic line 342 comes close to the contact pressure plate 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 plate spring tongues.This was likewise described in connection with the embodiment according to Figures 1 and 2 as well as the diagrams according to Figures 8 to 11.

In Figure 18 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 plate 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 plate spring characteristic line 440 can be associated with a lining suspension characteristic line corresponding to line 442.

Figure 19 shows the disengagement force curve for the associated characteristic lines 340 and 342 and 440 and 442 which is to be applied to the operating line, such as the plate 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 20, 20a and 21, the sensor plate 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 formed out of the cover material. The cover tabs 502a are formed 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 formed 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 21 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 20a by an axial recess in the cover 502. It is furthermore apparent from Figure 20a 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 502 g for the two turning directions. The stops 502g which prevent 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 1 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 plate 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 plate spring. This means that for an embodiment according to Figures 1 and 2 the basic body 13b which applies the axial tensioning force of the sensor spring 13 is provided radially inside the biasing area 3a between the operating plate spring 4 and the pressure plate 3.

In the embodiment according to Figures 20 and 21 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 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 548 make it much easier to assemble the friction clutch.

The adjustment ring 517 is shown in part in Figure 22. The adjustment ring 517 has radially inwardly aligned formed 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 12 and 13 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 formed out of the cover material or by axial indentations 526 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 23 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 settling 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 plate 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 connect ion 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 formed areas.

With the embodiment according to Figure 24 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 and to this end the cover 702 has tabs 715 which extend through corresponding slits or recesses in the plate spring 704 and which axially support the sensor plate spring 713.

According to another variation the sensor spring 713 could also have on its inner edge area tabs which extend axially through corresponding openings in the plate spring 704 and are supported on the cover side.

The adjustment ring 817 illustrated in Figure 25 can be used with a friction clutch according to Figures 20 to 21. The adjustment ring 817 has radially inwards formed areas 827 which extend radially. The formed 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 528 is provided which is broken or open on the outer circumference. The ring 528 is connected to the radial formed 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 formed 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 26, 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 plate 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 plate spring 905 or slit ends which form the tongues of the plate spring 904, in comparison with the plate 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 plate spring 904 it is possible to prevent damage or overstraining of the plate spring foundation body 904a in the area of the swivel bearing 905.

In Figure 26 an axial formed area 903a is shown in dotted lines which is provided on the pressure plate 903. Through such formed 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 plate spring 904 relative to the clutch 901. The operating plate 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 26 can be dispensed with. Although not shown the external diameter centring can however also be undertaken by tabs or indentations formed 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 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 27 the force directed against the disengagement force of the friction clutch or swivel force of the operating plate 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 plate spring 104 is not supported by a swivel bearing in the disengagement direction in the swivel or tilting area 1005. The bearing of the plate 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 plate spring 1004 is or becomes greater than the disengagement force of the friction clutch 1001. It must thereby be ensured that if there is no wear on the friction linings, the plate spring 1004 always remains adjoining the coverside 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 plate 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.

The friction clutch 1101 illustrated in Figures 28 and 29 has a housing 1102 and a pressure disc 1103 which is connected rotationally secured thereto but axially displaceable to a limited extent. A contact pressure plate spring 1104 is clamped axially between the pressure disc 1103 and the cover 1102 wherein this plate spring 1104 can swivel about a ring-like swivel bearing 1105 supported by the housing 1102 and biases the pressure disc 1103 in the direction of a counter pressure plate 1106, such as for example a flywheel, fixedly connected to the housing 1102 whereby the friction linings 1107 of the clutch disc 1108 are clamped between the friction faces of the pressure disc 1103 and the counter pressure plate 1106.

The pressure disc 1103 is connected rotationally secured to the housing 1102 by circumferentially or tangentially aligned connecting means in the form of leaf springs 1109.

In the illustrated embodiment the clutch disc 1108 has socalled lining spring segments 1110 which ensure a progressive torque build-up during engagement of the friction clutch by allowing a progressive rise in the axial forces acting on the friction linings 1107 through a restricted axial displacement of the two friction linings 1107 towards each other. However a clutch disc could also be used wherein the friction linings 1107 are attached practically rigid axially on at least one support disc and a replacement is provided at another point for the lining spring segments 1110, eg between the plate spring 1104 and pressure disc 1103.

In the illustrated embodiment the plate spring 1104 has an annular foundation body 1104a applying the contact pressure force whereby radially inwardly aligned operating tongues 1104b extend from the foundation body. The plate spring 1104 is thereby installed so that it impinges on the pressure disc 1103 with the radially further outer areas and can tilt about the swivel bearing 1105 with the radially further inner areas. The swivel bearing 1105 comprises two swivel support pads 1111,1112 which are here formed by wire rings and between which the plate spring is axially held and clamped. In order to secure the operating spring 1104 against rotation and to centre and hold the wire rings 1111,1112 relative to the housing 1102, retaining means are fixed on the cover in the form of rivet elements 1115 which each extend with an axially extending shaft 1115a through a cut-out section provided between adjoining plate spring tongues 1104b.

The clutch 1101 has an adjustment device compensating the axial wear on the friction faces of the pressure disc 1103 and the counter pressure plate 1106 as well as the friction linings 1107. This adjustment device consists of a wear compensating unit 1116 provided between the contact pressure plate spring 1104 and pressure disc 1103 as well as of restricting means 1117 designed as path sensors limiting the disengagement path of the pressure disc 1103.

The restricting means 1117 which act as wear sensors each have a socket 1118 which is housed rotationally secured in a bore 1120 of the pressure disc 1103. The socket 1118 forms a slit 1121 through which two leaf springs 1122 extend in the axial direction. The leaf spring elements 1122 are supported on each other wherein at least one leaf spring element is curved, and preferably both leaf spring elements are curved in opposite directions. The leaf spring elements 1122 are housed in the socket with a definite pretension and are thus displaceable in the axial direction of the clutch 1101 against a predetermined friction resistance relative to the socket 1118.The axial length of the leaf spring elements 1122 is selected so that when the friction clutch 1101 is engaged these leaf spring elements have a definite play 1124 relative to an axially fixed clutch component - in the illustrated embodiment relative to the outer edge area 1123 of the housing 1102 - wherein this play corresponds to the predetermined disengagement path of the pressure disc 1103.When the friction clutch is engaged the leaf spring elements 1122 come to rest against the counter pressure plate 1106 with their end 1122a remote from the housing 1102 whereby it is ensured that as the friction linings 1107 wear down so the pressure disc 1103 is moved axially relative to the leaf spring elements 1122 corresponding to this wear on the linings, thus against the action of the friction connection between the leaf spring elements 1122 and the socket 1118 which is preferably made from plastics or from a friction material.

In the illustrated embodiment the bore 1120 in which the socket 1118 is pressed in and fixed both axially and circumferentially is provided in a pressure plate cam 1125 which extends radially outwards and on which a leaf spring element 1109 is fixed by a rivet connection 1109a. Moving the socket 1118 in the direction of the counter pressure plate 1106 can also be avoided in that the socket 1118 has at its end facing the housing 1102 a collar 1118a by means of which it can be supported on the pressure disc 1103. The socket 1118 is thereby prevented from moving out of the bore 1120 in the direction of the housing or clutch cover 1102 in that, as shown in chain-dotted lines in Figure 28, the leaf springs 1109 engage partially radially round the socket 1118 and where applicable additionally clamp same axially fixed in the bore 1120.The socket can furthermore be prevented from turning in that it has a profiled section, more particularly a recess which holds the areas 1119 of the leaf springs 1109 engaging over the socket.

The wear compensation unit 1116 has a compensation component which is biased by the plate spring 1104 and has the shape of a U-section sheet metal ring 1126 which is shown in plan view in Figure 32. The compensation component 1126 has on the side of the base 1127 facing the plate spring 1104 at least one ring-shaped axial projection 1128 or several projections 1128 which are preferably evenly spread out over the circumference and are formed by grooves indented in the sheet metal material. Segment-like projections 1128 ensure that in the area between circumferentially adjoining projections 1128 radial ports are formed between the plate spring foundation body 1104a and the compensation ring 1126 to allow air to pass for cooling. As can be seen in particular from Figure 29 the compensation ring 1126 is centred relative to the pressure disc 1103.To this end the pressure disc 1103 has at least one step 1129 which positions the radially inner axially extending wall 1130 of the compensation ring 1126 centrally relative to the pressure disc 1103. The step 1129 can be formed by a closed surface extending over the circumference or even by segmentlike surfaces provided spaced out from each other over the circumference. The compensation ring 1126 furthermore has a radially outer axially extending wall 1131 which forms together with the inner wall 1130 and the base 1127 a ringlike free space 1126a. Radially outside the compensation ring 1126 has radial extension arms or cams 1132 which form stops which interaction with counter stops 1133 of the axially displaceable components in the form of leaf spring elements 1122 of the wear sensor 1117.The counter stops 1133 are formed by noses formed onto the leaf spring elements 1122 and pointing radially inwards to engage over the extension arms 1132. The axial displacement of the compensation ring 1126 in the direction away from the pressure disc 1103, thus in the direction of the housing is restricted by the counter stops 1133.

Between the compensation ring 1126 and the pressure disc 1103 there is a compensating unit 1134 which during disengagement of the friction clutch 1101 and in the presence of wear on the linings allows an automatic adjustment of the compensation ring 1126 and during engagement of the clutch causes an automatic locking, thus blocking action whereby it is guaranteed that during the engagement phase of the friction clutch 1101 the compensation ring 1126 retains a definite axial position relative to the pressure disc 1103. This definite position can only change during a disengagement process and corresponding to the wear which occurs on the linings.

The adjustment device 1134 comprises several pairs of wedges 1135,1136 preferably spread evenly over the circumference and housed in the annular free space 1126a of the sheet metal ring 1126. The wedges 1136 supported on a ring-like face 1137 of the pressure disc 1103 are connected rotationally secured but axially displaceable to the sheet metal ring 1126. To this end the sheet metal ring 1126 has in the area of its axially extending walls 1130,1131 formed areas in the form of grooves 1138,1139 which in the area of the free space 1126a form projections which engage in corresponding indentations or grooves 1140,1141 of the wedges 1136. The grooves 1140,1141 or formed areas 1138,1139 run in the axial direction of the clutch 1101. As can be seen from Figure 30 the wedges 1135 are housed substantially axially between the base 1127 of the sheet metal ring 1126 and the wedges 1136.The wedges 1135 and 1136 form circumferentially extending and axially rising run-up ramps 1142,1143 by which the wedges 1135,1136 associated with one pair are supported against each other.

The wedges 1135 are supported on the other side on the base 1127 of the ring 1126 and are displaceable circumferentially relative to this ring 1126. The run-up ramps 1142,1143 are tensioned against each other. To this end energy accumulators in the form of coil springs 1144 are housed in the ring chamber 1126a and are supported by one end on a wedge 1136 rotationally secured to the ring 1126 and by the other end on a wedge 1135 which is circumferentially displaceable. In order to hold the energy accumulators 1144 the wedges 1135,1136 have at their ends facing the corresponding energy accumulators projections 1145,1146 which engage in the spring windings and thus secure the spring ends. The springs 1144 are furthermore guided through the wall areas 1130,1131 and the base 1127 of the ring 1126.

In the illustrated embodiment the compensating ring 1126 is secured against rotation relative to the pressure disc 1103.

To this end, as is apparent from Figure 31, the pressure disc 1103 has axial projections in the form of pins 1147 which extend axially through recesses 1148 which are provided in the area of the extension arm 1132. Through this securement against rotation it is ensured that during the operation of the friction clutch the stop areas of the tabs 1132 always remain positioned beneath the restricting noses 1133 of the leaf spring elements 1122.

In the illustrated embodiment, the wedges 1135,1136 are made from a heat-resistant plastics, such as for example duroplastics or thermoplastics which can be additionally reinforced with fibres. The wedges 1135,1136 acting as adjustment elements can be simply made as injection moulded parts. However it is also possible for at least one of the wedges 1135,1136 of one pair to be made from friction material such as for example lining material. The wedges or adjustment elements 1135,1136 can however also be made as pressed sheet metal parts or as sintered parts. The slope angle and extension of the run-up ramps 1142,1143 are designed so that an adjustment of the wear occurring on the friction faces of the pressure disc 1103 and the counter pressure plate 1106 as well as on the friction linings 1107 is guaranteed throughout the entire service life of the friction clutch 1101.The wedge angle 1149 or slope angle 1149 of the run-up ramps 1142,1143 compared to a plane at right angles to the axis of rotation of the friction clutch is selected so that the friction which occurs when the runup ramps 1142,1143 press against each other prevents slipping between these ramps. The angle 1149 can be in the range from 5 to 20 degrees, preferably in the order of 10 degrees depending on the material matching in the area of the run-up ramps 1142,1143. The circumferentially displaceable wedges 1135 are arranged so that they point with their wedge tips in the direction of rotation 1150.

The tensioning by the energy accumulators 1144 of the run-up ramps 1142,1143 as well as the slope angle 1149 are designed so that the resulting axial force acting on the adjustment ring 1126 is less than the required displacement force of the wear sensor 1122 of the restricting means 1117.

Furthermore when designing the plate spring 1104 it must be taken into account that the contact pressure force to be applied by same for the pressure disc 1103 must be increased by the required displacement force for the wear sensor 1122 and by the tensioning force of the leaf springs 1109 tensioned between the cover 1102 and pressure disc 1103.

Furthermore the individual components must be designed so that the supporting wear between the plate spring 1104 and bearing ring 1126 as well as the contact wear between the wear sensors 1122 and the counter pressure plate 1106, and between the wear sensors and the housing 1102 remain slight in comparison with the wear on the linings 1107.

In order to avoid undesired adjustment between the run-up ramps 1142,1143 and the adjustment elements 1135,1136 in Figure 30 small projections can be provided in the area of at least one of the run-up ramps 1142,1143 to hook on to the other ramp. The projections can thereby be designed so that an adjustment is possible to compensate for the wear whilst the ramps are prevented from slipping relative to each other. It can also be particularly expedient if both run-up ramps 1142,1143 have projections which engage in each other.

These projections can be formed for example by a saw-tooth profiled section having a very low height and only allowing a relative displacement of the ramps 1142,1143 in the wear adjustment ring. Such profiled section is shown diagrammatically in Figure 30 over a partial area of the extension of the ramps 1142,1143 and is marked by 1143a. In cases where only one of the run-up ramps 1142,1143 has projections these can be designed so that they have a higher hardness than the material forming the other run-up ramp so that the projections can at least slightly project into or engage on the run-up ramp supporting same.

In order to prevent the curved or wavy leaf springs 1122 from losing their tensioning force owing to the very high temperatures which occur on the pressure disc 1103 during the clutch engagement process, the sockets 1118 are preferably made from a material with a low heat conductivity and high friction value. The wedges 1135,1136 can be made from the same material.

In order to allow improved cooling of the clutch, particularly the pressure disc 1103, radially aligned grooves can be provided spread out over the circumference in the pressure disc 1103, one of these grooves being shown in dotted lines in Figure 29 and marked 1151. These radial grooves 1151 are, seen circumferentially, provide between each two adjoining pairs of wedges and extend between the ring 1126 and pressure disc 1103. However the ring 1126 could have in the area of the springs 1144 axial cut-out sections extending out from the base 1127 whereby radial ports would be formed between the plate spring 1104 and the ring 1126.

In order to increase the wear resistance at the various bearing points the corresponding areas can be provided with a wear-resistant layer such as for example hard chromium plate, molybdenum coating, or special wear-resistant components can be provided in the area of the contact points. Thus for example plastics shoes can be provided on the wear sensors 1122 in the area adjoining the counter pressure plate 1106 and housing 1102.

The leaf springs 1109 transferring the torque to the pressure disc 1103 are pretensioned between the pressure disc 1103 and the housing 1102 so that during disengagement of the friction clutch 1101 they move the pressure disc 1103 in the direction of the housing 1102. It is thereby ensured that the ring 1126 remains adjoining the plate spring 1104 practically throughout the entire disengagement phase and until the restricting means 1117 come into operation.

The disengagement path of the clutch in the area of the tongue tips 1104c is preferably selected so that when the clutch is disengaged, the outer edge of the plate spring 1104 is lifted by a slight amount from the ring 1126. This means that during disengagement of the friction clutch 1101 the plate spring path in the diameter range of the pressure disc biasing through the plate spring 1104 is greater than the removal path 1124 of the pressure disc 1103 fixed by the path restricting means 1122.

The relative position shown in Figure 29 of the individual components corresponds to the new state of the friction clutch. In the event of axial wear, more particularly of the friction linings 1107, the position of the pressure disc 1103 is moved in the direction of the counter pressure plate 1106 whereby at first there is a change in the conicity and thus also in the contact pressure force applied by the plate spring in the engaged state of the friction clutch 1101, namely in the sense of an increase. This change causes the pressure disc 1103 to alter its axial position relative to the wear sensors 1122 axially supported on the counter pressure plate 1106. Owing to the plate spring force acting on the ring 1126, this ring 1126 follows the axial displacement of the counter pressure plate 1103 caused by wear on the linings whereby the stop areas 1132 of the ring 1126 lift axially from the areas serving as a counter stop in the form of noses 1133 of the wear sensor 1122, namely by an amount corresponding substantially to the wear on the linings. The compensating ring 1126 retains its axial position during an engagement process relative to the pressure disc 1103 because the ring is biased by the plate spring 1104 in the direction of the pressure disc 1103 and the wear compensating unit 1134 is self-locking during the engagement process, thus acts as an axial lock.During disengagement of the friction clutch 1101, the pressure disc is biased by the leaf springs 1109 in the direction of the housing 1102 and moves until the wear sensors 1122 come to adjoin the housing 1102 or housing stop areas 1123. Up to this disengagement path which corresponds to the lift path of the pressure disc 1103, the axial position of the ring 1126 relative to the pressure disc 1103 is maintained.

During a continuation of the disengagement process the pressure disc 1103 remains axially still whilst the ring 1126 axially follows the disengagement movement of the plate spring in the area of the biasing diameter, namely until the stop areas 1132 of the ring 1126 come to adjoin the counter stop areas 1133 of the wear sensors 1122. The axial displacement of the compensating ring 1126 is caused by the wedges 1135 which are biased by the springs 1144. These wedges 1135 are moved circumferentially relative to the wedges 1136 until the ring 1126 is tensioned against the counter stops 1133 of the wear sensors 1122.The lifting of the pressure disc 1103 is ensured in the illustrated embodiment by the leaf springs 1109 which are installed between the housing 1102 and pressure disc 1103 so that they have an axial pretensioning which forces the pressure disc 1103 in the direction of the housing 1102. If the plate spring 1104 is swivelled in the disengagement direction then this is lifted by its radially outer area away from the adjustment ring 1126 since the latter, as already described, is held back axially by the wear sensor 1122 relative to the pressure disc 1103. An at least slight lifting of the plate spring 1104 from the adjustment ring 1126 of this kind during a disengagement process is particularly advantageous for the functioning of the adjustment system 1117 + 1134.

The adjustment unit 1117 + 1134 according to the invention ensures that the adjustment on the supporting ring 1126 is always carried out by the adjustment wedges 1135,1136 corresponding to the amount of wear on the linings. This is due to the fact that the adjustment ring 1126 is axially tensioned between the adjustment means in the form of wedges 1135,1136 and the wear sensors 1122, which prevents the compensation component in the form of the ring 1126 from being adjusted by an amount larger than the corresponding lining wear.Furthermore through the design of the adjustment device according to the invention it is guaranteed that even in the event of over movement in the area of the disengagement means, such as the plate spring tongues 1104b, or in the case of axial vibrations of the pressure plate, no adjustment of the adjustment device 1117 + 1134 can take place since the wear sensor 1121 is axially supported by the self-locking wear compensation device 1134, namely by the counter stops 1132, relative to the pressure disc 1103 even in the case of a hard impact on the housing 1102.Thus in the disengaged state of the friction clutch axial forces can act on the wear sensors 1122 in the direction of the counter pressure plate 1106 which are greater than the force-locking connection between the wear sensors 1122 and the pressure disc 1103 without the wear sensors being displaced axially relative to this pressure disc 1103.

With the adjustment device according to the invention it is ensured that the plate spring works practically over the same characteristic line range throughout the entire service life of the clutch and in the engaged state of the friction clutch 1101 has a practically constant tension position and thus also applies a practically constant contact pressure force to the pressure disc 1103.It is thereby possible to use a plate spring with a degressive force characteristic line over the disengagement path, namely preferably in combination with a clutch disc whose linings 1107 are supported resiliently against each other through spring segments 1110 whereby the disengagement force effectively to be applied can be brought to a comparatively low level and can be kept practically constant throughout the service life of the clutch provided the lining spring characteristic line does not change significantly during the service life of the clutch.During disengagement of a clutch of this kind the plate spring 1104 is swivelled about its cover bearing 1105 whereby the spring segments 1110 relax over a predetermined partial area of the axial disengagement path of the pressure disc 1103 and thus the axial force applied by the spring segments 1110 assists in the disengagement process of the friction clutch 1101. This means that a smaller maximum contact pressure force need be applied than that which theoretically results in the engaged position of the clutch 1101 from the installation position of the plate spring 1104 and leaf springs 1109.As soon as the spring or relaxation area of the segments 1110 is exceeded then the friction linings 1107 are released whereby as a result of the degressive characteristic line range in which the plate spring 1104 operates, the disengagement force which is still to be applied is already reduced considerably relative to that which would correspond to the installation point or installation position according to Figure 29. On continuing with the disengagement process the disengagement force further decreases, namely until the minimum or trough point of the preferably sinusoidal characteristic line of the plate spring 1104 is reached.

The adjustment device 1117 + 1134 shown in Figures 28 and 29 can advantageously be designed so that when the friction clutch 1101 rotates the individual spring windings of the adjustment springs 1144 are supported on the outer wall 1131 of the adjustment ring 1126 and the displacement forces applied by the springs 1144 in the circumferential direction are reduced or evenly completely lifted as a result of the friction resistances produced between the spring windings and the adjustment ring 1126. The springs 1144 can thereby remain practically rigid during rotation of the friction clutch 1101 as a result of the friction forces which suppress the spring action.Furthermore the adjustment wedges 1135 can likewise be radially supported on the wall 1131 of the adjustment ring 1126 owing to the centrifugal forces acting on same and can be secured against rotation by the friction forces produced between the wedges 1135 and the adjustment ring 1126. It can thereby be achieved that at least with speed ranges above the idling speed of the internal combustion engine the wear compensation device 1134 cannot be rotated by the springs 1144. The friction clutch 1101 can thus be designed so that a compensation of the friction lining wear only takes place when operating the friction clutch 1101 at idling speed or at least approximately at idling speed. The blocking of the adjustment ring 1126 can however also be provided through a corresponding design of the wear compensation device 1134 so that an adjustment of the lining wear can only take place when the internal combustion engine is stationary, thus when the friction clutch 1101 is not turning or however even at very low speeds.

The material matching between the components 1135,1136 forming the adjustment ramps and the material of the components interacting with these components is preferably selected so that during the operating period of the friction clutch no adhesion can occur between the ramps and the components interacting therewith which would prevent 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 support faces.

In order to prevent adhesive connection between the run-up ramps and counter run-up ramps it is possible to provide at least one device which during disengagement of the friction clutch and during wear adjustment exerts an axial force on the or each adjustment element which causes the ramps to separate or tear loose.

In the new state of the friction clutch 1101, thus in the state which the clutch has before it is fixed on the counter pressure plate 1106 with the interposition of the clutch disc 1108, the wedges 1135 are located in a position drawn back further relative to the wedges 1136 compared to the position shown in Figure 30 so that the adjustment ring 1126 occupies its most drawn-back position in the direction of the pressure disc 1103 and thus the pressure disc 1103/ adjustment ring 1126 unit requires the smallest possible axial structural space. In order to keep the wedges 1135 in their restricted position prior to fitting the friction clutch 1101, the wedges 1135 have engagement areas in the form of recesses 1152 for turning and retaining means.Such retaining means can be provided during manufacture or assembly of the friction clutch 1101 and can be removed after fitting the friction clutch 1101 on the flywheel 1106 whereby the adjustment device 1134 becomes activated. In the illustrated embodiment, as apparent from Figures 30 and 32, circumferentially aligned oblong slits 1153 are provided in the adjustment ring 1126 through which the engagement areas of the retaining means or turning tool can be guided for engaging in the indentations 1152. The circumferentially placed oblong recesses 1153 must thereby have at least one extension which allows a rotation corresponding to the largest possible wear adjustment angle of the wedges 1135 in the circumferential direction.The wedges 1135 which are held in their retracted position in the circumferential direction when the friction clutch is in the new state can be held in this position by the wear sensors 1122 which secure the adjustment ring 1126 in its retracted position. The self-adjusting connections between the wear sensors 1122 and the pressure disc 1103 must be designed so that the displacement force required to move the wear sensors 1122 relative to the pressure disc 1103 is greater than the resulting force acting on the ring 1126 and which is produced by the springs 1144 biasing the wedges 1135.

In the embodiment according to Figures 28 and 29 the ramps 1143 can also be formed directly by the ring 1126, for example by imprinting inclined faces 1143 wherein the ring 1126 must then be rotatable relative to the pressure disc 1103 by the springs 1144. With such a design the wedges 1136 are rotationally secured with the pressure disc 1103 or are formed directly on same. Furthermore with this type of embodiment the stops which are formed as extension arms 1132 must be extended in the circumferential direction corresponding to the required adjustment turning angle of the ring 1126 in order to ensure that an axial restriction remains between the wear sensors 1122 and the ring 1126 throughout the service life of the friction clutch.With the last mentioned embodiment the adjustment ring 1126 can be turned easily from radially outside even when the friction clutch 1101 is fitted, namely in particular through the circumferentially extending stop tabs 1132 which are accessible through radial ports provided on the outer sleeve of the clutch housing 1102. These radial ports can in particular also house the torque transfer cams 1125 of the pressure disc 1103 as well as the leaf springs 1109. The adjustment device according to the invention furthermore has the advantage that its principle can also be used with socalled pull-type friction clutches wherein the plate spring is supported for swivel movement on a cover by a radially outer edge area and biases the pressure disc with radially further inner edge areas. Such a clutch is shown in Figure 33.Between the plate spring 1204 and the pressure disc 1203 there is a wear compensation device 1234 which can be designed like the one described in connection with Figures 28 and 29. The adjustment ring 1226 again interacts by sensor elements 1217 with wear sensors 1222. The adjustment of the wear sensors 1222 relative to the pressure disc 1203 takes place by the stop areas 1222a bearing on the housing or cover 1202. The wear sensors 1222 again support stops 1233 which restrict the axial path of the pressure disc 1203 during a disengagement process. In order to allow a satisfactory functioning of the adjustment device according to Figure 33, the ring 1226 again has at least a possibility of a slight axial movement relative to the wear sensors 1222.This can be achieved in that a corresponding stop connection 1233a is provided with play between the wear sensors 1222 and the ring 1226 or however in that the ring 1226 has radial areas 1226a which are resiliently deformable in the axial direction, thus have an elastic pliability.

With the embodiment illustrated in Figure 34 the wear sensor elements 1317 are housed directly in the foundation body of the pressure disc 1303. The wear sensors 1322 have stop areas 1322a which interact with cover areas 1323 which form counter stops. The cover areas 1323 are in one piece with fastening means 1302a by which the plate spring 1304 is mounted for swivel movement on the cover 1302. In the illustrated embodiment the fastening or retaining means 1302a are formed by tabs made in one piece with the cover material and extending axially through the plate springs 1304. The wear compensation device 1334 is provided radially outside of the wear sensors 1317 provided in the radial area of the plate spring foundation body 1304a.

Through the design of a friction clutch according to the invention it is possible not only to achieve an increase in the clutch service life by using thicker friction linings, thus by increasing the axial lining wear volume, but also to reduce the disengagement forces, namely by using an energy accumulator with a degressive force-path characteristic line over the disengagement path of the friction clutch, in combination with at least one spring means counteracting the energy accumulator acting on the pressure plate wherein the spring means during engagement and disengagement of the friction clutch causes a gradual build up and break down of the torque transferrable by the friction clutch over at least a partial area of the clutch actuation path or pressure plate path. This spring means is preferably connected in series with the contact pressure spring, such as eg the plate spring, of the friction clutch. Through the design of a friction clutch according to the invention it is thus possible to reduce the disengagement force quite considerably and this reduction can remain unchanged throughout the service life of the friction clutch, which means within a comparatively narrow tolerance band.

Furthermore with friction clutches according to the present invention plate springs can be used with comparatively steep force-path curves in the operating range. Such plate springs would lead in conventional clutches to a very sharp rise in the disengagement force in the event of wear on the linIngs.

In the case of clutches without the adjustment according to the invent ion, as the wear on the linings increases so at first the point 41 (Figure 8) corresponding to the engaged state of the friction clutch would move along the line 40 in the direction of the maximum 41a. At this point 41a during a disengagement process a drop in the disengagement force occurs but overall the level of the disengagement force curve increases relative to the disengagement force curve when the friction clutch is in the new state. This means that the area 43 moves left until the point 41 coincides with the maximum 41a. The point 44 moves correspondingly along the characteristic line 40.With further wear on the linings the installation point of the operating plate spring corresponding to the engaged state of the friction clutch gradually moves from the maximum 41a in the direction of point 41b so that the contact pressure force applied by the plate spring decreases gradually. The contact pressure force applied at point 41b by the operating plate spring corresponds to the contact pressure force applied at point 41 in the new state of the friction clutch. As soon as the maximum 41a is exceeded the disengagement force increases during the disengagement process at least over a partial area of the clutch operating path. On reaching the maximum permissible wear path or wear point 41b a rise in the disengagement force takes place over the entire disengagement path 43a.This increase in the disengagement force remains even when, as shown in Figure 8, there is a lining suspension or lining suspension substitute 42a.

When designing the friction clutch and more particularly its adjustment device it must be taken into consideration that the crank shaft of the internal combustion engine generates axial and tumbling vibrations on the flywheel which are also transferred to the friction clutch fixed on the flywheel.

So that the friction clutch and adjustment device can function satisfactorily, that is no undesired adjustment takes place as a result of such vibrations,in the embodiment according to Figures 1 to 27, thus quite generally with embodiments with an adjustment device with force sensor, the adjustment force of this force sensor must be greater than the inertia forces which can act on the force sensor. These forces result in particular through the masses of the main plate spring, of the adjustment ring or of the adjustment elements, through a corresponding mass proportion of the force sensor and where applicable through the masses of other components multiplied with the maximum possible axial acceleration of these components or elements which results from the axial and bending vibrations of the flywheel.Thus for example with an embodiment according to Figure 27 wherein the sensor plate spring 1013 is supported on the clutch pressure plate 1003, the inertia of this clutch pressure plate 1003 must also be taken into consideration.

It must thus always be ensured that the force applied by the sensor spring is larger than the forces acting on same which are formed by the mass of the components acting on the sensor spring as a result of its inertia, multiplied with the maximum possible axial acceleration. These inertia forces can have a disadvantageous effect particularly during the operation of the friction clutch and particularly in the disengaged state of the friction clutch.

In the embodiments according to Figures 29 to 34 the forces which arise as a result of the inertia of the individual components and the axial and rotary vibrations acting on same must be taken into account when designing the wear sensors and the wear compensation devices.

Thus in general when designing a friction clutch with integrated wear compensating means it is necessary to take into account the masses of the elements to which the axial or rotary vibrations can be transferred and which act on the compensating means. In the embodiments according to Figures 28 to 34 the components to be considered are those which affect the function of the ramp mechanism.

Figure 35 illustrates a divided flywheel 1401 which has a first or primary flywheel mass 1402 fixable on a crankshaft (not shown) of an internal combustion engine, as well as a second or secondary flywheel mass 1403. A friction clutch 1404 is fixed on the second flywheel mass 1403 with the interposition of a clutch disc 1405 via which a gearbox (likewise not shown) can be engaged and disengaged. The flywheel masses 1402 and 1403 are mounted rotatable relative to each other by a bearing 1406 which is mounted radially inside the bores 1407 for passing through fixing screws 1408 for fitting the first flywheel mass 1402 on the driven shaft of the internal combustion engine. Between the two flywheel masses 1402 and 1403 there is a damping device 1409 which has coil compression springs 1410 which are set in a ringlike chamber 1411 which forms a toroidal area 1412.The ring-shaped chamber 1411 is at least partially filled with a viscous medium, such as for example oil or grease.

The primary flywheel mass 1402 is mainly formed by a component 1413 which was made of sheet metal. The component 1413 has a substantially radially aligned flange-like area 1414 which supports radially on the inside an integral formed axial attachment 1415 which is enclosed by the bores or holes 1407. The single-row rolling bearing 1406a of the rolling bearing unit 1406 is housed with its inner ring 1416 radially outwards on the end section of the axial attachment 1415. The outer ring 1417 of the rolling bearing 1406a supports the second flywheel mass 1403 which is substantially designed as a flat disc-like body. To this end the flywheel mass 1403 has a central recess in which the bearing 1406a is housed.The substantially radially aligned area 1414 changes radially outwards into a dish-like area 1418 which engages at least partially round and guides and supports the energy accumulators 1410 at least over their outer circumference. The dish-like body 1419 fixed on the area 1418 engages partially round the circumference of the energy accumulator 1410. The body 1419 is welded to the sheet metal body 1413 (at 1420). The toroidal area 1412 is divided, viewed circumferentially, into individual sockets in which the energy accumulators 1410 are housed. The individual sockets, viewed circumferentially, are separated from each other by biasing areas for the energy accumulators 1410 which can be formed by pockets stamped in the sheet metal part 1413 and the dish-like body 1419.The biasing areas 1421 connected to the second flywheel mass 1403 for the energy accumulators 1410 are supported by the clutch cover 1422.

The biasing areas 1421 are formed by radial arms 1421 which in the illustrated embodiment are set on the axial area 1423 of the clutch cover 1422 and which engage radially in the ring chamber 1412, namely between the ends of circumferentially adjoining energy accumulators 1410. The axially aligned cover area 1423 encloses or engages round the second flywheel mass 1403 by a section 1423a, and is fixedly connected to same eg by indentations formed in the section 1423a and engaging in corresponding depressions in the flywheel mass 1403, or by other types of fixing.

The clutch cover 1422 centred on the outer contour of the flywheel mass 1403 has at its end remote from the biasing areas 1421 a substantially radially inwardly aligned ringshaped area 1426 on which a plate spring 1427 is held for swivel movement wherein the plate spring acts as a doublearmed lever. The plate spring 127 biases a pressure plate 1428 with the radially further outer areas whereby the friction linings 1429 of the clutch disc 1405 are clamped axially between the second flywheel mass 1403 and the pressure plate 1428. A lining suspension 1465 is provided between the friction linings 1429.

As can be seen from Figure 35, the ring-shaped space 1422 or its toroidal area 1412 is mounted substantially radially outside of the outermost contours of the second flywheel mass 1403. The component 1413 supporting the toroidal area 1412, serving to connect the first flywheel mass 1402 to the driven shaft of the internal combustion engine and adjoining the latter can lie practically opposite, thus practically adjoin with a slight spacing, the second flywheel mass 1403 radially inside the ring-shaped area 1411 over a comparatively large radial extension, thus by forming an interspace or air gap 1430, which makes it possible to achieve an axially very compact construction for the unit comprising the flywheel 1401, clutch 1404 and clutch disc 1405.The sealing of the ring-like chamber 1411 is ensured by a seal 1431 which operates between the inner areas of the radial wall 1419 and the outer sleeve face of the cover 1422.

Advantageously this interspace 1430 can serve to cool the flywheel 1401, namely by a cooling air stream passing through this interspace 1430. In order to produce such a cooling air circulation the second flywheel mass 1403 has radially inside the friction face 1432 axial recesses 1433 which extend in the direction of the component 1413 on the engine side and open into the interspace 1430. In order to improve the cooling action the second flywheel mass 1403 can have further axial ports 1435 which lie radially further outwards and are connected on the side remote from the friction face 1432 with the interspace 1430 whilst opening radially outside of the friction face 1432 on the side of the flywheel mass 1403 facing the clutch 1404.Radially inside the ports or recesses 1433 the flywheel mass 1403 has further ports 1434 which serve in particular to house and guide the fixing screws 1408.

In order to seal the ring-like chamber 1411 which is filled partially with viscous medium a further seal 1436 is provided which is formed by a membrane-like or plate spring like component which extends radially in the interspace 1430.

The dish-like body 1419 supports a starting gear ring 1439 which is connected to same by a welded connection. The twin-mass flywheel 1402+1403 shown in Figure 35 forms together with the clutch assembly consisting of the clutch 1404 and clutch disc 1405, one structural unit A which is preassembled as such and can thus be despatched and stored and screwed onto the crankshaft of an internal combustion engine in a particularly simple and rational way. To assemble the structural unit A first the clutch 1404 and the second flywheel mass 1403 are connected together with the interposition of the clutch disc 1405.Then the sub-unit comprising the clutch 1404 , flywheel mass 1403 and clutch disc 1405 is brought together axially with the component 1413 whereupon the dish-like body 1419 which is set on the outer edge 1423 of the clutch cover 1422 is brought to adjoin the outer areas of the component 1413 and can be welded to the latter (at 1420). Before axially joining the two components 1413 and 1419 the springs 1410 are placed in the toroidal area 1412. The structural unit A furthermore has already integrated the bearing 1406 which is set on the axial attachment 1415. The fixing screws 1408 are already prefitted or contained in the bores 1407 of the flange area 1414, namely in the form of inbus screws 1408. The screws 1408 are thereby located in a position corresponding to the lower half of Figure 35. The screws are held safe against loss in this position in the assembly or unit A.

The clutch disc 1405 is clamped in a position pre-centred relative to the axis of rotation of the crank shaft between the pressure plate 1428 and the friction face 1432 of the second flywheel mass 1403 and furthermore is located in such a position that the openings 1443 provided in the clutch disc are located in such a position that during the assembly process of the assembly A on the driven shaft of the internal combustion a screwing tool can be passed through.

It can be seen that the openings 1443 are smaller than the heads 1440 of the screws 1408 so that a satisfactory secure hold of the screws 1408 in the assembly A is guaranteed.

Also in the plate spring 1427, namely in the area of their tongues 1427a there are openings or cut-out sections 1444 for passing through the screwing tool. The cut-out sections 1444 can be provided so that they form expansions or enlarged areas in the slits between the tongues 1427a. The openings 1444 in the plate spring 1427, 1443 in the clutch disc 1405 and 1434 in the flywheel mass 1403 thereby conform with one another in the axial direction, namely so that even with a non-symmetrical arrangement of the bores 1407 which is required for a positioned assembly of the unit A on the crankshaft, an assembly tool, such as for example an inbus key can pass satisfactorily through the openings 1444, 1427 and 1443 and engage in the recesses of the heads 1440 of the screws 1408.

A complete assembly A of this kind makes it much easier to fit the flywheel since various work steps can be dispensed with such as the otherwise necessary centring process for the clutch disc, the Insertion of the clutch disc, the fitting of the clutch, the insert ion of the centring pin, the centring of the clutch itself, the insertion of the screws and the screwing on of the clutch and the removal of the centring pin.

The friction clutch 1404 has an adjustment device 1445 which guarantees wear compensation in a similar way to that described in connection with Figures 1 to 27 by means of a sensor spring 1446 and adjustment ring 1447.

The torque transfer device 1501 shown in Figure 36 has a counter pressure plate 1502 connectable rotationally secured to the crankshaft K of an internal combustion engine and on which the friction clutch 1504 is fitted with the interposition of a clutch disc 1505. The clutch disc 1505 can be housed on the input shaft of a gearbox which is not shown in further detail.

The clutch cover 1522 has an axially aligned area 1523 which engages radially outwards axially over the pressure plate 1528 and the friction linings 1529 of the clutch disc 1505.

The end section 1523a of the sleeve-like or tubular cover area 1523 surrounds or engages round the counter pressure plate 1503 and is fixedly connected to same by indentations 1524 formed in the section 1523a and engaging in depressions provided on the outer circumference of the counter pressure plate 1504. The cover 1522 and the counter pressure plate 1503 can however also be connected in another way eg by welded connections or by connections by means of screws or pins which are likewise preferably inserted in the radial direction.

The clutch cover 1522 centred on the outer contour of the counter pressure plate 1503 has a substantially radially inwardly aligned ring-like area 1526 on which a plate spring 1527 is mounted for swivel movement acting as a double-armed lever. The plate spring 1527 biases the pressure plate 1528 with radially outer areas whereby the friction linings 1529 are clamped axially between the counter pressure plate 1503 and the pressure plate 1528. The plate spring 1527 has radial tongues 1527a for operating the clutch 1504 by a disengagement system. In order to transfer the torque between the pressure plate 1528 and the cover 1522 there are torque transfer means, preferably in the form of leaf springs 1521 which are fixedly connected at one end to the cover 1522, preferably by rivet connections 1521a, and at the other end to the pressure plate 1528 by a rivet connection.The connection between the pressure plate 1528 and the leaf spring elements 1521 is preferably formed by a so-called blind rivet connection such as shown in the radial extension area of the friction linings in the upper half of Figure 35.

The friction clutch 1504 or the torque transfer device 1501 has an adjustment device 1545 which ensures the wear is compensated by a sensor spring 1546 and an adjustment ring 1547 in a similar way to that described in connection with Figures 1 to 27.

In Figure 35 and Figure 36 the counter run-up ramps imprinted directly in the cover material are designed so that they each form an air passage opening (1547a in Figure 36) in the rotary direction of the friction clutch. With such a design a better cooling action is obtained for the friction clutch through automatic air circulation as the clutch rotates. More particularly the adjustment ring 1447 or 1547 which is made of plastics is cooled down in this way whereby the thermal load on this ring can be substantially reduced.

The friction clutch 1504 or counter pressure plate 1503 is fixed on the driven shaft K by an elastic or resiliently pliable component 1550 whereby it is rotationally secured but axially displaceable to a limited extent. In the illustrated embodiment, this component 1550 is disc-shaped and its rigidity is measured so that the axial or tumbling and bending vibrations excited by the driven shaft K on the friction clutch 1504 are dampened or suppressed by the elastic component 1550 to an extent which guarantees satisfactory functioning of the friction clutch 1504 and more particularly of the adjustment device 1545. The largest possible uncoupling of the clutch unit 1504 should take place through the axially pliable component 1550 relative to the axial and bending vibrations of the driven shaft of the internal combustion engine, such as the crankshaft.This thereby prevents the function of the clutch unit 1504 or its adjustment device 1545 from becoming impaired. Without the aforementioned uncoupling of the clutch unit 1504 from the crank shaft K an undesired adjustment of the wear compensating device 1545 could take place, namely owing to the mass of the components and the accelerations acting on these as a result of vibrations.

Thus without the component 1550 which filters the vibrations, the inertia forces produced by the adjustment device would have to be taken into account, particularly when designing the said adjustment device 1545 whereby an expensive matching and/or additional means would be required in order to avoid an adjustment of the wear compensating device which is not based on wear of the linings.

With the torque transfer device 1501 according to Figure 36 the wear compensating device 1545 acts between the clutch cover 1522 and the plate spring 1527. The torque transfer device 1501 could however also be fitted with a friction clutch according to Figures 28 to 34 , thus with a friction clutch wherein the wear adjustment device acts between the plate spring and the pressure plate biased by same.

The counter pressure plate 1503 is fixedly connected radially outwards by screw connections 1551 to the axially elastic disc-like component 1550. Instead of screw connections 1551 blind rivet connections could also be used, such as those shown in the upper half of Figure 35 in connection with the fixing of the leaf springs on the pressure plate 1428. Radially inside the connecting areas 1551 between the disc-like component 1550 and the counter pressure plate 1503 there is an axial gap 1552 between these two components 1550 and 1503 wherein this gap determines the maximum amplitude of the axial vibrations between the two components 1550 and 1503 in the one axial direction.The maximum axial displacement of the friction clutch 1504 in the direction of the driven shaft K can be restricted by the radially inner areas of the counter pressure plate 1503 stopping on the disc-like component 1550. Under normal operating conditions, more particularly during satisfactory functioning of the internal combustion engine such a bearing does not take place however. The ring-like counter pressure plate 1503 engages round an axial projection 1553 which is part of a ring or disc-like component 1554. This disc-like component 1554 can be fixedly connected to the radially inner areas of the elastic disc 1550. The elastic disc 1550 and the disc-like component 1554 are centred on an annular projection 1555 of the shaft K and are fixedly connected to this by screw connections 1556.The radially inner areas of the disc-like component 1550 are thereby clamped axially between the end face 1557 of the shaft K and the annular component 1554.

The axial attachment 1553 of the component 1554 has at the end remote from the elastic component 1550 radial areas 1558 which restrict the axial displacement of the friction clutch 1504 or counter pressure plate 1503 in the other axial direction. With the elastic component 1550 which is not stressed there is axial gap 1559 between the areas 1558 and the counter pressure plate 1503. This gap 1559 is dimensioned in the axial direction like the gap 1552. The counter pressure plate 1503 can be housed practically free of play by its inner sleeve face on the axial attachment or projection 1553 so that an axial guide of the counter pressure plate 1503 can be guaranteed.It can however also be expedient if there is at least a small air gap between the inner sleeve face of the counter pressure plate 1503 and the axial attachment 1553 , thus under normal operating conditions there is no contact between these two components.

According to a further development a device for cancelling out the vibration energy is provided in order to dampen the vibrations which are still transferred despite the elastic component 1550. Such a device can be formed by a friction connection, as shown for example in Figure 37. With the embodiment shown in Figure 37 a damping unit 1560 is provided between the inner areas of the counter pressure plate 1503 and the outer sleeve face of the attachment 1553 wherein this damping means can be formed for example by a ring which is wavy in the circumferential direction and whose waves run radially. This ring 1560 is installed clamped in the radial direction whereby friction is provided eg between this ring 1560 and the inner sleeve face of the counter pressure plate 1503 in the presence of axial vibrations. Thus a damped bearing is provided for the counter pressure plate 1503 on the attachment 1553. The wavy ring 1560 can be separated over its circumference, thus can be open.

Radially on the outside the disc-like elastic component 1550 supports a starting gear ring 1561.

The disc-like component 1550, the counter pressure plate 1503, the clutch disc 1505 and the friction clutch 1504 form a structural unit which is preassembled as such and can thus be despatched, mounted and screwed onto the crank shaft K of an internal combustion engine in a particularly simple and rational way. The fixing screws 1556 in the form of inbus screws are likewise already prefitted, thus are contained in the structural unit and are secured against loss.

The clutch disc 1505 is clamped in a position pre-centred relative to the axis of rotation of the crank shaft between the pressure plate 1528 and counter pressure plate 1503 and furthermore in such a position that the openings 1562 provided in same and radially inside the spring damper of the clutch disc 1505 are located in such a position that during the assembly process of the unit on the shaft K a screwing tool 1563 can be passed through. Also the plate spring 1527 has, if necessary, openings or cut-out sections 1564 for passing through the screwing tool 1563. The openings or cut-out sections 1564 of the plate spring 1527 coincide with the openings 1562 of the clutch disc 1505 so that the assembly tool, such as for example an inbus key 1563, can engage satisfactorily in the screw profiled sections of the heads of the screws 1556.

As already described in connection with the other embodiments, by using a friction clutch 1504 with a device 1545 which at least compensates the lining wear, it is possible to optimize the design of the friction clutch, more particularly the energy accumulator 1527 which applies the tensioning force for the clutch disc. This energy accumulator can be designed so that it practically only applies the clamping force which is necessary to transfer the desired torque for the clutch disc. Through the adjustment device 1545 it is ensured that the energy accumulator 1527 retains practically the same installation position in the engaged state of the clutch unit 1501 throughout the entire service life.Furthermore it is possible to reduce the disengagement forces or disengagement force curve through the device likewise provided in the clutch unit 1501 in the form of a lining suspension 1565 which causes a gradual breakdown and build up of the torque transferrable by the unit during the disengagement and engagement of the unit 1501 over at least a partial area of the operating path of the pressure plate 1528. The desired disengagement force curve can thus be determined by suitably matching the forces or force-path characteristics applied by the device, such as a lining suspension, and by the energy accumulator acting on the pressure plate. It is thereby also possible to design the elastic component 1550 in optimum manner relative to the desired damping function for axial, bending or tumbling vibrations, since the reduced disengagement forces acting on this elastic component are of secondary importance. The operating forces required to disengage the clutch can thus be supported by the component without substantial axial displacement of the clutch unit.

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 and elements described in connection with the various embodiments.

Our co-pending patent application 9224460.7 also describes and claims matter disclosed in this specification.

Claims

1. Friction clutch, particularly for motor vehicles, with a pressure plate which is connected with a housing in a manner rotationally fast but axially displaceable to a limited extent, wherein at least one pressure spring between the housing and the pressure plate acts against the pressure plate in the direction of a clutch disc clampable between the pressure plate and a counter pressure plate, such as a flywheel, wherein an adjustment device is provided which compensates at least the wear of the friction linings of the clutch discs, which adjustment device takes account of and at least reduces the detensioning of the pressure spring normally arising as a result of wear of the friction linings, over the service life of the friction clutch, characterised in that the adjustment device is blockable in dependence on the operating state of the friction clutch.

2. Friction clutch according to Claim 1, characterised in that the blocking is dependent on the rotational speed.

3. Friction clutch according to Claim 1 or Claim 2, characterised in that the blocking results from elements dependent on centrifugal force.

4. Friction clutch according to any preceding claim, characterised in that the blocking of the adjustment device results from a frictional connection.

5. Friction clutch according to any preceding claim, characterised in that the blocking of the adjustment device results from a shape locking.

6. Friction clutch according to any preceding claim, characterised in that the blocking of the adjustment device results from a weight dependent on centrifugal force.

7. Friction clutch according to any preceding claim, characterised in that the adjustment device is blocked above a predetermined threshold rotational speed.

8. Friction clutch according to any preceding claim, characterised in that the adjustment device is blocked at the idling speed of the driven engine and at rotational speeds lying above idling speed.

9. Friction clutch according to any preceding claim, characterised in that the adjustment device can be adjusted at speeds below the idle speed.

10. Friction clutch according to any preceding claim, characterised in that the adjustment device can practically be adjusted only at zero rotational speed.

11. Friction clutch according to any preceding claim, characterised in that the adjustment device can only be equalised with lining wear when the friction clutch is not rotating.

12. Friction clutch according to any preceding claim, characterised in that the adjustment device is blocked at least at at least a substantially completely disengaged state of the clutch.

13. Friction clutch according to any preceding claim, characterised in that the adjustment device has an adjustment ring, the adjustment function of which is blocked at at least a substantially completely disengaged state of the friction clutch.

14. Friction clutch according to any preceding claim, characterised in that the adjustment device has an adjustment ring which is tensioned against a clutch component at at least a substantially completely disengaged state of the clutch, so that the adjustment function is blocked.

15. A friction clutch as claimed in any preceding claim, wherein the adjustment device results in a practically constant force biasing 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 over a partial region of the operating path of the operating means causes a gradual reduction of the torque which can be transferred by the friction clutch or clutch disc.

16. A friction clutch as claimed in any preceding claim, wherein the friction clutch has a torque reduction device which during the disengagement process over a partial area of the axial displacement path of the pressure plate areas biased by the contact pressure spring causes a gradual reduction of the torque transferrable by the friction clutch.

17. A friction clutch as claimed in Claim 16, wherein the torque reduction device is provided in the torque path between the operating means or contact pressure spring and the fixing areas of the housing on the counter pressure plate.

18. A friction clutch as claimed in Claim 16 or Claim 17, wherein the torque reduction device is provided in the torque path between the operating means or contact pressure spring and the friction face of the pressure plate.

19. A friction clutch as claimed in any one of Claims 16 to 18, wherein the torque reduction device is provided axially between the friction linings of the clutch disc.

20. A friction clutch as claimed in any one of Claims 16 to 19, wherein the torque reduction device allows an axial resilient pliability between the clutch components wherein the device is arranged so that when the clutch is opened the force acting on the device is at its smallest and when closing the clutch the force acting on the device gradually rises to the maximum wherein this rise takes place at least over a partial area of the closing path.

21. A friction clutch as claimed in any one of Claims 16 to 20, wherein the torque reduction device causes the gradual breakdown or the gradual increase in the torque transferrable by the friction clutch over at least approximately 40 to 70% of the operating path of the operating means.

22. A friction clutch as claimed in any preceding claim, wherein the contact pressure spring has at least over part of the disengagement path of the friction clutch a degressive force-path curve.

23. A friction clutch as claimed in any preceding claim, wherein the contact pressure spring is formed by a plate spring which on one side can be swivelled about a ring-like swivel bearing supporting the housing and on the other side biases the pressure plate.

24. A friction clutch as claimed in Claim 23 wherein the plate spring has an annular body from which radially inwardly aligned tongues emerge to form the operating means.

25. A friction clutch as claimed in Claim 23 or Claim 24, wherein the plate spring is supported for swivel movement on the housing between two supports.

26. A friction clutch as claimed in any one of Claims 23 to 25 wherein the plate spring has a sinusoidal force-path characteristic line and in the engaged state of the friction clutch the operating point of the plate spring is provided on the degressive characteristic line area following the first force maximum and the plate spring has a force ratio of about 1:0.4 to 1:0.7 between the first force maximum and the following force minimum.