GB2278894A

GB2278894A - Friction clutch with wear adjustment

Info

Publication number: GB2278894A
Application number: GB9410392A
Authority: GB
Inventors: Karl-Ludwig Kimmig; Rolf Meinhard; Christoph Wittmann
Original assignee: LuK Lamellen und Kupplungsbau GmbH
Current assignee: LuK Lamellen und Kupplungsbau GmbH
Priority date: 1993-05-26
Filing date: 1994-05-24
Publication date: 1994-12-14
Anticipated expiration: 2014-05-24
Also published as: JPH0754863A; FR2709337A1; ES2113783A1; RU2166679C2; CN1062058C; BR9402077A; CN1120133A; GB2278894B; JP3715660B2; DE4418026B4; GB9410392D0; FR2709337B1; DE4418026A1; ES2113783B1

Abstract

Friction clutch 1 for use with a clutch disc, more particularly for a motor vehicle, comprises a pivot support for a plate spring 4 which biases an axially displaceable pressure plate 3 towards the clutch disc and a counter pressure plate 6, and has a device 16 compensating for the wear at least of the friction linings 7. The force of a further energy accumulator 26 is superimposed on the plate spring 4 at least approximately from the path in the disengagement direction from which the clutch disc no longer or only insignificantly biases the pressure plate. The addition of the characteristic lines of the springs 4 and 26 produces a desired disengagement force curve over the necessary air path of the pressure plate 3. Leaf springs 9 and a sensor spring 13 provide a resulting force to take up the clutch disengagement force acting on the fingers of the spring 4. In a modification (FIG 7), the further energy accumulator is provided radially inside the plate spring pivot, and the clutch may be of the pull-type (FIGS 8 and 10). <IMAGE>

Description

FRICTION CLUTCH The invention relates quite generally to friction clutches, more particularly those where there is an adjustment device compensating the wear of at least the friction linings of the clutch, more particularly those such as described or mentioned for example in German Patent Applications P 42 39 291.8, P 43 06 505.8, P 42 39 289.6, P 42 31 131.4, P 42 43 567.6 and P 43 17 587.2. The subject of the aforementioned patent applications belongs completely to the contents of the present application, for which these applications are integrated in the first copy of the application documents of the present application.

With self-adjusting clutches of this kind attempts are made to obtain a very low disengagement force despite the high contact pressure force whilst this disengagement force is to be kept as constant as possible throughout the service life of the clutch, that is in particular throughout the wear of the clutch linings.

In order to achieve a low disengagement force with a simultaneously high contact pressure force it is necessary to use plate springs having a very steep force drop. Since the disengagement force path is to have the lowest force fluctuations possible, for such clutches the remaining ie the available force-path curve of the plate spring is not sufficient for a safe complete disengagement with additional path reserve for tolerances, namely in particular because the plate spring with the steeply declining characteristic line already has again after a relatively short path a very steeply rising curve. The chain-dotted characteristic line in Figure A shows a typical force characteristic line of this kind with reference to a diagram wherein the force when the clutch is closed lies at approximately 1 mm spring path and the force with the release of the linings at below 2 mm and at 3 mm a traceable rise in the characteristic line can be detected wherein this point at 3 mm corresponds to the minimum disengagement path required. In addition both path tolerances, installation tolerances and tolerances in the components themselves and elasticity losses of the clutch are to be included. Furthermore additional path fluctuations through the tolerances of the disengagement system are to be added so that the required spring path extends to at least 3.5 mm in the diagram. This then means however a generally severe rise in the disengagement force or however for example in the case of the clutch with sensor plate spring according to Patent Application P 42 39 291.8, possibly an undesired displacement of the adjustment ring in the adjustment device.

The object of the present invention is to avoid the abovementioned disadvantages, that is to provide a clutch which has over the full disengagement path, inclusive of the possible tolerances, the lowest possible disengagement force which is as constant as possible and wherein an inadmissible or undesired force rise is avoided throughout the maximum possible disengagement path. Furthermore a clutch should be provided wherein the manufacture itself or its components can be simple and economic and wherein these components can be dimensioned as light as possible In order to achieve this at least in part in the case of conventional clutches with high disengagement force expensive hydraulic or pneumatic servo support systems were used or even so-called super dead-centre systems in order to reach tolerable pedal operating forces in the disengagement system between the clutch and pedal.

Such solutions have however the quite considerable drawback that the large disengagement force of the clutch is introduced into the operating system through the disengagement bearing and therefore very great elasticity and friction losses occur both in the clutch and in the operating system - conditioned by the high disengagement forces -, with the additional disadvantage that both the clutch components and the entire operating system including for example the axial bearing of the engine, have to be made with sufficient dimensions for these high forces, which is only possible with correspondingly expensive components.

The object of the present invention is also to remove the drawbacks of these systems.

According to the invention this is achieved in that a plate spring is used whose force decline between the points 1 and 2 of the spring path in Figure A has approximately the same force decline as that previously described. The plate spring has a substantially lower force minimum wherein the minimum force can even be smaller than 0, namely can be negative. A plate spring of this kind, as can be seen from the solid line in Figure A, has at its two intersecting points with the 10 000 N line a spacing of more than 2 mm, as opposed to the spring shown in chain-dotted line which has only about 1 mm in its intersecting points with the 10 000 N line. This means that the overall path length for this force area is practically twice as long with the plate spring with the solid line. A clutch with a characteristic line of this kind, namely according to the solid line, has however quite considerable disadvantages for the clutch operation since in the first path area a positive force would occur, followed by a drop to negative force and then a further rise to positive force. This means that the corresponding force change in the disengagement system would occur throughout the disengagement path and this would not be controlled satisfactorily through the pedal.

Even if a plate spring were used where the minimum has not dropped so low, as illustrated with the solid line wherein the force at minimum would lie just slightly above 0, thus a positive force were constantly maintained, the clutch function could only be controlled with difficulty through the very strong force change during disengagement. Through a further inventive step, namely the use of at least one additional so-called compensation spring and as will be described later in further detail and which mainly acts in the area of the force minimum of the clutch plate spring, this strong force drop of the characteristic line shown solid in Figure A is avoided and thus the required disengagement comfort is reached namely by using the extended curve branch of the characteristic line drawn solid and the non-permissible force drop is avoided at the same time.

Further features and expedient developments as well as advantages of the invention are apparent from the following description of Figures 1 to 13 in which: Figure A shows a typical force characteristic line for a clutch in accordance with the prior art; Figure 1 is a sectional view of a clutch assembly designed according to the invention; Figure 2 is a partial view according to arrow II of Figure 1; Figure 3 is a prefitted partial unit for use with the friction clutch shown in Figure 1; Figures 4 to 6 are diagrams with different characteristic lines from which can be seen the interaction of the individual spring and adjustment elements of the clutch assembly according to the invention; Figure 7 is a sectional view of an embodiment of a friction clutch according to the invention; Figures 8 to 8b show further design embodiments for a clutch according to the invention; Figure 9 is a diagram with various characteristic lines from which can be seen the interaction of the individual spring and adjustment elements of the friction clutch according to Figure 8; Figure 10 shows an additional inventive design of a friction clutch and Figures 11 to 13 show various embodiments of a friction clutch according to the invention wherein Figure 12 shows a development in the circumferential direction of the adjustment ring used in Figure 11 and Figure 13 shows a section along the line XIII-XIII of Figure 12.

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

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

In the illustrated embodiment the plate spring 4 has a ringlike base member 4a applying the contact pressure force and from which operating tongues 4b extend radially inwards.

The plate spring 4 is thereby installed so that it biases the pressure disc 3 with radially further outer areas and tilts round the swivel bearing 5 with radially further inner areas.

The swivel bearing 5 comprises two swivel support pads 11, 12 which here are formed by wire rings and between which the plate spring 4 is axially held and tensioned. The swivel support pad 11 provided on the side of the plate spring 4 facing the pressure disc 3 is force biased axially in the direction of the housing 2 by an energy accumulator 13. The energy accumulator 13 is formed by a plate spring or by a plate spring-like component 13 which is supported with radially outer edge areas 13a on the housing 2 and with radially further inner sections 13c axially biases the swivel support pad 11 against 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 a ring-like base member 13b from whose inner edge tongues 13c extend radially inwards and are supported on the swivel support pad 11. Radially outside extension arms 13a are formed on the base member 13b and interact with support areas 14 moulded directly out of the housing 2. Between the support areas 14 and the extension arms 13a of the plate spring-like component 13 is a bayonet-type connection or lock so that after the plate spring-like component 13 is initially pretensioned and its radially outer areas or extension arms 13a were brought axially over the support areas 14 the extension arms 13a of the component can be brought to adjoin the support areas 14 by suitably turning the plate spring-like component 13 relative to the housing 2.

In order to secure the operating plate spring 4 axially extending centring means in the form of rivets are fixed on the housing 2. These rivets each have an axially extending shaft 15a which extends axially through a cut out recess provided between adjoining plate spring tongues 4b.

The plate spring-like component or plate spring 13 is designed as a sensor spring which can produce over a predetermined work path an at least approximately constant force. If the leaf spring elements 9 between the housing 2 and the pressure disc apply an axial force then this is superimposed with the axial force applied by the sensor spring 13. In the case of leaf spring elements 9 which are fitted in the friction clutch 1 so that they bias the pressure disc 3 axially in the direction of the housing 2 or plate spring 4 then the axial forces applied by the leaf spring elements 9 and by the sensor spring 13 are added which then form a so-called resulting sensor force acting on the plate spring 4. When designing the sensor spring 13 it is therefore necessary always to take into consideration the superimposing forces. The axial force applied by the leaf spring elements 9 is likewise superimposed on the force applied by the plate spring 4 on the pressure disc 3 so that when pretensioning the leaf spring elements 9 in the sense of lifting the pressure disc 3 from the clutch disc 8 the axial tension force applied by the pressure disc 3 onto the friction linings 7 is smaller by the force applied by the leaf spring elements than the axial force applied by the plate spring 4 on the pressure disc 3. The resulting sensor force applied by the leaf spring elements 9 and the sensor spring 13 takes up the clutch disengagement force acting on the tongue tips 4c whereby at least when releasing the friction linings 7 an at least approximate counterbalance prevails between the force produced by the disengagement force on the swivel support pad 11, and the resulting sensor force exerted on this swivel support pad. By disengagement force is meant the force which during operation of the friction clutch 1 is exerted on the tongue tips 4c or on the disengagement lever of the plate spring clutch. This disengagement force can then be changed, viewed over the disengagement path in the area of the tongue tips 4c.

The swivel support pad 12 on the housing side is supported on the housing 2 by an adjustment device 16. This adjustment device 16 ensures that with axial displacement of the swivel support pads 11 and 12 towards the pressure disc 3 and towards the counter pressure plate 6 respectively no undesired play can occur between the swivel support pad 12 and the housing 2 or between the swivel support pad 12 and the plate spring 4. It is thereby ensured that no undesired dead or idling paths occur when operating the friction clutch 1 whereby an optimum degree of efficiency and thus satisfactory functioning of the friction clutch 1 is provided. The axial displacement of the swivel support pads 11 and 12 takes place with axial wear on the friction faces of the pressure disc 3 and the counter pressure plate 6 as well as on the friction linings 7. The method of operating the automatic adjustment of the swivel bearing 5 will be explained in further detail in connection with the diagrams according to Figures 4 to 6.

The adjustment device 16 comprises a spring-biased adjustment element in the form of a ring-like component 17 which has circumferentially extending and axially rising run-up ramps 18 which are spread out round the circumference of the component 17. The adjustment element 17 is installed in the clutch 1 so that the run-up ramps 18 face the housing floor 2a. On the side of the adjustment element 17 remote from the run-up ramps 18 the swivel support pad 12 formed by a wire ring is positioned centrally in a groove-like socket.

In the illustrated embodiment the adjustment element 17 is made of plastics, such as eg a heat-resistant thermoplastics which can be additionally strengthened with fibres. The adjustment element 17 can thereby be simply made as an injection moulded part. The adjustment ring 17 is centred through the axially aligned areas 15a of the rivets 15 spread evenly over the circumference.

The adjustment ring 17 is supported by its run-up ramps 18 on the counter run-up ramps 19 imprinted in the cover base 2a. The cover indentations forming the counter run-up ramps 19 are designed so that these each form in the rotary direction of the clutch 1 an air passage 20a. This design ensures as the clutch 1 rotates a better cooling of the components forming the clutch 1, more particularly of the adjustment ring 1 which is made of plastics. The cover indentations are designed so that they cause an automatic air circulation within the clutch installation area defined by the cover 2.

The ramps 18, 19 are designed circumferentially with regard to their length and starting angle 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 occurring on the friction faces of the pressure disc 3 and counter pressure plate 6 as well as on the friction linings 7. This adjustment angle can be in the order of between 8 and 60 degrees, preferably in the order of between lo and 30 degrees, depending on the design of the run-up ramps. The starting angle of the ramps 18, 19 can be in the area of 4 to 12 degrees. This angle is selected so that the friction which occurs as the run-up ramps 18 and counter run-up ramps 19 press against each other prevents slipping between these ramps 18, 19.

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

As can be seen in connection with Figure 2, the spring biasing of the adjustment ring 17 is ensured through individual coil springs 20 which extend in the circumferential direction of the cover 2 and are tensioned between the adjustment ring 17 and the housing 2.

Preferably three such coil springs 20 are provided which are spread out evenly over the circumference. The individual coil springs 20 are mounted or threaded on tabs 21 which are designed integral with the clutch cover 2. The tabs 21 are shaped out of the sheet metal material of the cover 2 by forming an eg punched-out U-shaped cut 22. The tabs 21 extend in an arc or tangentially, viewed circumferentially, and are preferably at least approximately at the same axial height as the directly adjoining-cover areas. The width of the tabs 21 is measured so that the coil springs 20 mounted thereon are guided both in the radial and axial direction.

The adjustment ring 17 biased in the adjustment direction by the springs 20 has on its inner circumference radially inwardly pointing moulded areas or extension arms 23 which have radially inwards an axially aligned fork or U-shaped moulded area 24. The U-shaped moulded areas 24 each form two axially directed tines 25 which engage on both sides round a spring guide plate 21. To this end the two tines 25 extend axially in or through a cut-out section 22 of the cover 2. The tines 25 are biased by the adjustment rings 20.

When the friction clutch 1 is new the axial protrusions forming the run-up ramps 18 and counter run-up ramps 19 engage the furthest axially in each other, this means that the ring 17 and thus also the swivel bearing 5 are displaced the furthest possible amount in the direction of the cover floor 2a.

The friction clutch 1 has an additional energy accumulator 26 which is formed by a plate spring-like component. The plate spring-like component 26 has a ring-like base member 27 from which the extension arms protrude radially inwards in the form of tongues 28. The plate spring-like component 26 is provided axially between the cover base 2a and the contact pressure or operating plate spring 4 and is held positioned relative to the latter. To this end the tongues 28 are bent down axially relative to the ring-like base member 27 so that they project axially through openings 29 provided circumferentially between the tongues 4b of the plate spring 4, and engage with radially inwardly bent areas 30 underneath the swivel pad 11 formed by a wire ring. The axial positioning of the plate spring-like component 26 relative to the plate spring 4 takes place through the leaf spring-like suspension means 31 which are connected axially fixed to the plate spring 4 and axially bias areas 30 of the tongues 28 whereby the plate spring-like component 26, the plate spring 4 and the ring-like rolling support provided axially between the plate spring 4 and the end areas 30 of the tongues 28 are axially tensioned. The tongues 28 are mounted circumferentially off-set relative to the tongues 13c of the sensor spring 13.

As can be seen from Figure 3, the plate spring-like component 26, the plate spring 4 and the ring-like swivel pad 11 are combined by the leaf spring-like spring elements 31 into one preassembled unit which can be fitted as such in the housing or cover 2 when mounting the friction clutch 1.

Figure 3 shows the position of the relaxed plate spring relative to the plate spring-like component 26.

It can be seen from Figure 1 that even in the assembled state of the friction clutch 1 fitted on the counter pressure plate 6 there is an axial space or play 32 between the plate spring 4 - which is located in a position corresponding to the engaged state of the friction clutch 4 - and the plate spring-like component 26 - which is located in the relaxed position. The distance 32 between the radially outer areas of the two ring-like spring base members 4a and 27 must be measured so that during disengagement of the friction clutch 1 the plate spring 4 which is swivelled around the swivel bearing 5 can only bias the plate spring-like component 26 after a swivel angle or disengagement path which corresponds at least approximately to a release of the clutch disc 8. By releasing the clutch disc 8 or friction linings 7 is meant the operating state of the friction clutch 1 where the friction linings 7 are practically no longer tensioned between the friction faces of the pressure disc 3 and counter pressure disc 6, thus the state of the friction clutch 1 where practically no torque can be transferred from the counter pressure plate 6 to the clutch disc 6. In this operating state of the friction clutch 1 the lining spring segments 10 are relaxed. The distance 32 can preferably be dimensioned so that the plate spring 4 comes to adjoin the plate spring-like component 26 shortly after release of the friction linings 7. The plate spring-like component 26 serves as a compensation spring which adapts the disengagement force curve of the friction clutch 1 after release of the clutch disc 8 to the desired force-path characteristic. Through a corresponding design of the compensation spring 26 it is possible in the disengagement path remaining after release of the friction linings 7 to "linearize" the disengagement force path which means that over this remaining disengagement path the disengagement force to be applied can be kept practically constant or however at least the change in the disengagement force can be reduced significantly over this path.

The 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 4 to 6.

The line 33 in Figure 4 represents the resulting axial force path arising in dependence on the change in conicity of the plate spring 4 and taking into account the force applied by the leaf spring elements 9, namely with 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 resulting force produced by the plate spring 4 and the leaf spring elements 9 is shown on the ordinate. Point 34 represents the installation position of the plate spring 4 when the clutch 1 is closed, thus the position where the plate spring 4 for the corresponding installation position exerts the maximum contact pressure force on the pressure disc 3. The point 34 can be moved up or down along the line 33 by changing the conical installation position of the plate spring 4.

Line 35 represents the axial expanding force which is applied by the lining spring segments 10 and acts between the two friction linings 7. Furthermore all the spring actions which act in the same way as the lining suspension are contained in this characteristic line, such as the cover elasticity, elasticity of the swivel bearing or, where applicable, elastic means between plate spring and pressure plate support or the like. 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 required for the maximum possible elastic deformation of the spring segments 10 corresponds at least to the force exerted by the plate spring 4 on the pressure disc 3 in the engaged state of the friction clutch 1. When disengaging the friction clutch 1 the spring segments 10 relax , namely over the path 36. The disengagement process of the clutch 1 is assisted through this path 36 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 34 if the lining spring segments 10 were not present.

On exceeding point 37 the friction linings 7 are released wherein as a result of the degressive characteristic line area of the plate spring 4 the disengagement force which is then still to be applied is considerably reduced compared to that which would correspond to point 34. The disengagement force of the clutch 1 would decrease without the compensation spring 26 until the point 38 located on the abscissa axis is reached. On exceeding point 38 in the disengagement direction a change in the direction of the axial force produced by the plate spring 4 takes place so that when exceeding the point 38 the plate spring 4 automatically snaps round in the disengagement direction and automatically moves in the direction of the minimum or trough point 38a or 39a of the sinusoidal characteristic line 33. When exceeding the point 38 in the disengagement direction the force applied by the plate spring becomes negative so that without the compensation spring 26 the friction clutch 1 would automatically remain open. On exceeding the minimum 38a during disengagement of the friction clutch 1 the negative force applied by the plate spring 4 again decreases to the point 39 lying on the abscissa axis. On exceeding the point 39 in the disengagement direction the force produced by the plate spring 4 becomes positive again whereby on reaching the point 39a the force required for pushing through the plate spring 4 agrees with the force associated with the point 37.

The flat position 33a of the plate spring 4 has been entered in the diagram according to Figure 4. As flat position of the plate spring 4 is designated the position of the deformed plate spring 4 at which the ring-like resilient base member 4a runs parallel to a plane standing at right angles to the axis of rotation of the plate spring 4.

As can be seen from the line section running through points 37, 38, 38a, 39, 39a a significant change in force in the disengagement force curve takes place without the additional spring 26 after release of the friction lines 7 through axial lifting of the pressure disc 3. This change is disadvantageous since an exact measuring of the engagement and disengagement path in this area is difficult, particularly as a result of the different force changes either side of the minimum 38 (here negative). This is the case both with foot-operated friction clutches and those operated by a servo motor. In order to avoid this disadvantage and to obtain a desired disengagement force curve over the necessary air path 40 of the pressure disc 3, the plate spring-like component 26 is provided which in the illustrated embodiment has a force-path characteristic line according to the broken line 41 (taking into account the spacing of its supports on the plate spring 4 compared to the spacing of the supports of the plate spring 104 on the pressure plate cams 3a and on the swivel support pads 11,12). As can be seen from Figure 4, the plate spring 4 and the plate spring-like component 26 have at least inside the air path 40 opposite force-path curves. In the illustrated embodiment the plate spring-like component 26 only acts over a partial area 42 of the air path 40. By air path is meant the path which the pressure disc 3 can still cover axially after release of the friction linings 7 when operating the friction clutch. As can be seen from Figure 4, with a disengagement process of the friction clutch 1 the plate spring-like component 26 is only used after the point 37 where the friction linings 7 are released. The force path which is produced through the superposing or addition of the spring characteristic lines 33 and 41 is marked 43.

This path 43 starts at point 43.

The starting point 43 for the plate spring-like component 26 is determined through the axial spacing 32 between the outer contour of the plate spring 4 and the radially outer area of the plate spring-like component 26. The air path 40 is designed so that even when reaching the full disengagement path of the friction clutch the disengagement force corresponding to the end point 45 of the air path 40 is smaller than the disengagement force corresponding to the point 37. This is, as will be described later, necessary in order to avoid undesired adjustment in the adjustment device 16.

The disengagement path for the friction clutch 1 required in the area of the tongue tips 4c or contact diameter 4d, eg for a disengagement bearing, is suitably enlarged relative to the possible axial displacement path 46 of the pressure disc 3 by the leverage of the plate spring 4 which can be drawn from Figure 4. This plate spring - or leverage corresponds to the ratio of the radial spacing between the swivel bearing 5 and operating or contact bearing diameter 4d to the radial distance between the swivel bearing 5 and support diameter 3a between plate spring 4 and pressure disc 3. This translation ratio lies in m path 47 of Figure 4 corresponding to the leverage of the plate spring 4 wherein however the axial path required in the area of the operating diameter 4d compared to the relaxation path 36 of the spring segments 10 is correspondingly larger by this translation ratio. With the design of the friction clutch 1 according to Figures 1 to 4 only the operating or main plate spring 4 is swivelled over a part of the disengagement path, namely until the outer edge of this plate spring comes to rest against the additional or compensation spring 26. The main plate spring 4 is then swivelled together with the compensation spring 26 wherein the force path curves of these two springs are superimposed on each other and produce a resulting curve 43 which extends at least over a partial area of the air path 40. As can be seen from Figure 4, the force minimum 38a of the plate spring 4 can be selected very much lower and can also assume negative values so that the minimum 38a comes to lie below the abscissa. In the latter case the plate spring 4 forms a so-called snap spring which has a tensioned position in which it can remain without outside force action. The compensation spring 26 acts at least in the areas adjoining the minimum 38a of the main plate spring 4.

The compensation spring 26 acts so that in the air path 40 the path of the force required to disengage the friction clutch 1 is lifted so that the variation occurring in the disengagement force curve can be considerably reduced compared to that which would occur in the characteristic line area of the force-path curve 33 of the plate spring 4 extending over the air path 40.

The spring 13 serving as a force sensor has a path-force curve corresponding to line 48 of Figure 5. The characteristic line 48 corresponds to that which is produced when the plate spring-like component 13 is changed in its conicity from the relaxed positioned, namely between two swivel pads which have a radial distance which corresponds to the radial distance between the swivel pads and supports 11 and 14.

The total force biasing the operating plate spring 4 against the rolling support 12 on the cover side during disengagement of the friction clutch 1 and after release of the clutch disc 8 is produced by adding the forces which are mainly exerted by the leaf spring elements 9, the sensor spring 13 and by the existing disengagement force on the operating plate spring 4 in the operating area 4d. The leaf spring elements 9 can be installed between the cover 2 and pressure plate 3 so that as the wear on the linings 7 increases and throughout the closing path 46 of the clutch 1 the axial force exerted by the leaf springs 9 on the operating plate spring 4 becomes greater. Thus the axial force applied by the leaf springs 9 has a rising path according to line 50 over the path 49 according to Figure 5 and thus also over the wear compensation path of the adjustment device 16. It can also be seen from Figure 5 that as the springing of the sensor spring 15 increases so the resetting force exerted by the leaf springs 9 on the pressure plate 3 and also acting on the operating plate spring 4 also increases. Adding the force path according to the characteristic line 50 and the plate spring characteristic line 48 produces the resulting force path 51 which acts axially on the plate spring 4, namely in the sense of pressing the plate spring 4 against the swivel pad 12 on the cover side. Thus by suitably pretensioning the leaf springs 9 it is possible to reduce the support force or support force curve which is to be applied by the sensor spring 13 at least over the path 49. In the illustrated embodiment the sensor spring 13 has a decreasing or negative force-path curve over the path 49. By suitably designing and arranging the leaf spring elements 9 it is also possible to compensate at least partially an increase in the disengagement force caused by a reduction in the lining suspension and/or the lining spring segments embedding in the linings in that the sensor force 51 rises slightly during displacement of the plate spring 4 in the direction of the counter pressure plate. It can thereby be ensured that the plate spring 4 retains substantially the same operating point 34 when the clutch 1 is closed or the same operating area 46 so that the plate spring 4 exerts an at least approximately constant contact pressure force on the pressure plate 3 throughout the service life of the friction clutch. Furthermore when designing the friction clutch, more particularly the sensor spring 13 and/or the leaf springs 9, it is necessary to take into account the resulting axial force which is produced by the adjustment springs 20 acting on the adjustment element 17 and which counteracts 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 affected by the pretension of the leaf springs 9. This means that when pretensioning the leaf springs 9 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. Thus with a friction clutch 1 of this kind a resulting contact pressure force curve for the pressure plate 3 and friction linings 7 is formed which is produced by superimposing the contact pressure force curve of the plate spring 4 on the tension curve of the leaf springs 9. Assuming that - seen over the operating area 46 of the friction clutch 1 - the characteristic line 33 according to Figure 4 represents the resulting force curve from the operating plate spring 4 and pretensioned leaf springs 9 when the friction clutch 1 is new, as the distance between the pressure plate 3 and the counter pressure plate 6 is reduced, for example as a result of the lining wear, a displacement of the resulting curve would arise in the sense of a reduction, namely as a result of the counter moment exerted with increasing wear through the leaf springs 9 on the plate spring 4. This counter moment exists between the operating plate spring 4 and pressure plate 3 as a result of the radial distance between the swivel bearing 5 and the biasing diameter 3a. When designing the friction clutch 1 it is particularly important that the increase in the tensioning fore of the leaf springs 9 which arises through wear on the linings is preferably smaller than (at the most the same as) the increase in the disengagement force in the operating area 4d which arises as a result of the same lining wear and which causes the swivelling of the sensor spring 13 required for adjustment.

Otherwise the contact pressure force of the pressure plate 3 on the friction linings 7 in the engaged state of the friction clutch as well as the force exerted by the plate spring on the swivel support 11 during release of the friction linings would drop. Thus practically no adjustment could then take place because points 34 and 37 would move in the direction of minimum.

The resulting curve 51 in Figure 5 of the leaf springs 9 and/or of the spring 13 has a spring path 49 over which the resulting axial force remains substantially constant or preferably rises slightly. The force produced in this area 49 is thereby selected so that this disengagement force of the clutch corresponding to point 37 in Figure 4 is at least approximately the same. The resulting supporting force to be applied by the sensor spring 13 and the leaf springs 9 is reduced relative to the force of the plate spring 4 corresponding to point 37 according to the leverage of this plate spring 4.

The installation position of the plate spring-like element 13 in the friction clutch 1 is selected so that this can carry through in the area of the swivel bearing 5 an axial spring path in the direction of the friction linings 7 which corresponds at least to the axial adjustment path of the pressure disc 3 in the direction of the counter pressure plate 6 which (adjustment path) arises in particular as a result of the friction faces and friction lining wear. The at least approximately linear area 49 of the characteristic line 51 can preferably have a greater length than the said wear path since installation tolerances can also be compensated at least in part thereby.

In order to obtain a practically constant or definite release point 37 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 in pairs back to back wherein the individual pairs of segments can also have a certain axial pretension relative to each other. Through the pretension of the spring means provided between the linings it is possible to achieve an at least substantial compensating of the embedding losses of the segments in the back of the linings which occurs during the operating period. By embedding losses are meant the losses which occur as the segments work into the backs of the linings.

It is expedient if the pretension of the suspension provided between the linings is in the order of 0.2 mm to 0.6 mm.

Through a corresponding restriction of the axial relaxation path between the two friction linings 7 as well as through a definite at least slight 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 36 in Figure 4 through the suspension provided between the linings. In order to obtain a definite path 36 the axial path between the friction linings can be restricted through corresponding stops both in the relaxation direction and also in the tensioning direction of the lining suspension 10. As lining suspensions it is possible to use advantageously in connection with the present invention those which are already known for example through Patent Application P 42 06 880.0 which should be added expressly to the subject of the present application.

In order to ensure optimum functioning of the friction clutch 1 or of the adjustment device which ensures an automatic compensation of the lining wear it is advisable if, viewed over the disengagement force curve 52 according to Figure 6, the resulting force initially exerted on the plate spring 4 through the lining suspension 10, sensor spring 13 and leaf springs 9, as well as the resulting force then still exerted on the plate spring 4 by the sensor spring 13 and leaf springs 9 after lifting the pressure disc 3 from the friction linings 7, is at least slightly greater than but at least of equal amount as the disengagement force which engages in the operating area 4d of the plate spring tongue tips 4c and change according to Figure 6 over the disengagement path.

The consideration up until now corresponds to a quite definite installation position of the plate spring 4, and no wear on the friction linings was 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 whereby 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 occurs, namely in the sense of an increase. This alteration causes the point 34 to wander in the direction of point 34' and point 37 in the direction of point 37'. This change destroys the force equilibrium which originally exists 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 of the path of the disengagement force in the sense of an increase. The disengagement force curve thereby resulting is shown in Figure 6 through the broken line 53. By increasing the disengagement force curve, during the disengagement process of the friction clutch 1 the resulting axial force exerted by the sensor spring 13 and leaf springs on the plate spring 4 is overcome so that the sensor spring 13 yields in the area of the swivel bearing 5 by an axial path which corresponds substantially to the wear of the friction linings 7. During this sagging phase of the sensor spring 13 the plate spring 4 swivels round the biasing area 3a of the pressure disc 3 so that the 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 support pad 11 or the sensor spring 13 and on the pressure disc 3. As can be seen in connection with Figure 4, this change takes place in the sense of reducing the force applied 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 support pad 11 on the sensor spring 13 has reached a balance with the counter force produced by the sensor spring 13 and the leaf springs 9. This means that in the diagram according to Figure 4 the points 34' and 37' wander again towards points 34 and 37. After this balanced state has again been produced then the pressure disc 3 can again lift from the friction linings 7. During this adjustment phase of the wear during a disengagement process of the friction clutch 1 the adjustment element 17 of the adjustment device 16 is turned by the pretensioned spring 20 whereby the swivel pad 12 also shifts corresponding to the lining wear, and thus a play-free swivel bearing 5 for the plate spring 4 is guaranteed. After the adjustment process the disengagement force curve again corresponds to the line 52 according to Figure 6. Lines 54 and 55 of Figure 6 represent the axial path of the pressure disc 3 in the event of a disengagement force path curve corresponding to lines 52 and 53.

In practice the described adjustment takes place continuously and in very small stages so that the great point displacements and characteristic line displacements shown in the diagrams for clearer understanding of the invention do not occur.

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

Through a corresponding design of the characteristic line 33 of the plate spring 4 and/or of the characteristic line 50 of the leaf springs 9 and/or corresponding adaption of the curve 48 of the sensor spring 13 it is however possible to ensure an operationally reliable function of the friction clutch.

The torque transfer means 9 between the housing 2 and pressure disc 3 can also be designed so that they apply the full force necessary for axially supporting the plate spring 4 during disengagement of the friction clutch 1. With a design of this kind of the torque transfer means 9 it would be possible to dispense with the plate spring 13. The torque transfer means 9 must be designed so that they have a pathforce curve which guarantees a satisfactory adjustment function of the wear compensation device 16 at least throughout the service life of the friction clutch 1.

The clutch unit shown in Figure 7 with a friction clutch 101 has a similar construction to the clutch unit according to Figure 1. The friction linings 107 of the clutch disc 108 can be axially tensioned between the friction face of the counter pressure plate 106 formed by a flywheel and the friction face of the pressure disc 103. Between the clutch housing 102 and the pressure disc 103 is clamped a main plate spring 104 which is mounted to tilt in the swivel bearing 105 relative to the housing 102. The swivel pad 111 axially supporting the plate spring 104 during disengagement of the clutch 101 is formed directly by the plate springlike component 113 serving as the sensor spring 113. The plate spring-like component 113 is tensioned axially between the housing 102 and the plate spring 104 so that an axial force is exerted on the plate spring 104 which is directed opposite the disengagement force acting in the biasing area 104d of the plate spring tongue tips 104c. Between the plate spring 104 and the housing 102 there is an adjustment device 116 with an adjustment ring 117 which is rotatable relative to the housing 102 and which is biased in the adjustment direction by at least one spring 120. The adjustment device 116 acts in a similar way to the adjustment device 16 according to Figure 1 so that the description of Figure 1 should be referred to with regard to the adjustment function for compensating the wear on the friction linings 107.

When designing the friction clutch 101 according to Figure 7 the additional plate spring 126 which acts after relaxing the friction linings 107 through the pressure disc 103 in order to obtain the desired disengagement force curve over the air path of the friction clutch 101, is provided radially inside the swivel bearing 105. The material expense for the additional spring 126 serving as the compensation spring is thereby significantly reduced compared to an embodiment according to Figure 1 where the additional spring 26 is provided radially outside of the swivel bearing 5.

The additional spring 126 is mounted axially between the adjustment ring 117 and the plate spring 104. The additional spring 126 has a ring-like base member 127 which has on the outer circumference radial extension arms 128 which are spread out circumferentially round the base member and are tensioned or clamped axially in the area of the swivel bearing 105 between the adjustment ring 117 and the plate spring 104.

As can be seen from the top half of Figure 7 the additional plate spring 126 can have tongues 122 which extend radially inwards from the base member 127 and which are bent in the axial direction so that they project axially through openings 129 provided between the plate spring tongues 104b and with a radially angled end area 130 in the engaged state of the friction clutch 1 engage behind the plate spring 104 in the area of the tongues 104b with a definite axial spacing or play 132.

The lower half of Figure 7 shows a further variation for setting a definite play. With this variation individual rivet elements 131 are provided spread out over the circumference of the plate spring 104. These rivet elements extend with their shaft towards the base 102a of the housing 102 and have a top formed by a head 132a for the additional plate spring 126. The additional plate spring 126 has radially inside forked extension arms 126a which engage round the shaft of the rivet elements 131.

The embodiment according to Figure 7 has the advantage that here no additional elements are necessary for securing the additional plate spring 126 in the friction clutch. With the embodiment according to Figure 1 the elements 31 and the fixing means for same were-necessary for this purpose. The additional plate spring 126 is always held tensioned here in the friction clutch 101, namely in the engaged state of the clutch 101 between the main plate spring 104 and the adjustment ring 117. During disengagement of the friction clutch 101 the additional plate spring 126 is tensioned through the main plate spring 104 after exceeding the play 132. To this end the additional plate spring 126 is supported radially outwards on the main plate spring 104 in the area of the swivel bearing 105 and radially inwards on the stop areas provided in the area of the plate spring tongues 104b and interacting with the counter stop areas 130 or 126a of the additional plate spring 126.

This arrangement of the additional plate spring 126 ensures a simple mounting with simultaneously secure tensioning of the additional plate spring 127. Figure 7 shows the additional plate spring 126 in its relaxed position. The axial assembly of the additional plate spring 126 shown in the upper half of Figure 7 with the main plate spring 104 can be produced through a bayonet-type connection, thus an axial push and turn connection. To this end the axially aligned tongues 122 are pushed through openings suitably formed in the area of the plate spring tongues and then moved by turning between the two plate springs 104 and 126 into a slit of smaller radial extension whereby the stop areas 130 come to lie axially relative to counter stop areas of the plate spring 104. The two plate springs 104 and 126 are positioned and guided relative to both the housing 102 and each other in their position fitted in the friction clutch 101 through holders in the form of bolts 115.

Regarding the design of the disengagement force curve of the friction clutch 101 the additional plate spring 126 fulfils the same function as the additional spring 26 according to Figure 1, and reference is made here to the description of Figures 1 to 6.

The friction clutch 101 has means which at least in the partial areas of the speed range in which the clutch 101 rotates during use, increase the supporting force which counteracts the disengagement force. This can thereby prevent inadmissible adjustment from taking place during operation of the friction clutch 101 as a result of breakdown factors which occur for example at higher speeds.

In the friction clutch 101 these means are formed by means dependent on centrifugal force, namely by weights mounted on the outer periphery of the sensor plate spring 113 in the form of moulded-on tongues 156 which stand up axially in the direction of the cover 102. As the clutch 101 rotates and as a result of the centrifugal force acting on the tongues 156, a force is produced which is superimposed on, thus added to, the force applied by the sensor spring 113 as a result of the pretension. The supporting force for the operating plate spring 104 is thereby increased in the area of the swivel pad 111. This additional force produced on the swivel pad 111 by the tongues 156 becomes greater as the speed increases.

At certain speed ranges, particularly at higher engine speeds, vibrations can occur excited eg through the engine and causing an axial vibration of the pressure disc 103 in the disengaged state of the friction clutch 101. If the pressure disc 103 vibrates axially then this pressure disc 103 is lifted at least temporarily from the main plate spring 104 whereby the resulting sensor force temporarily drops since the axial force then produced by the leaf spring-like torque transfer means 109 no longer acts on the plate spring 104. Without means 156 dependent on centrifugal force this would have the result that the structurally set force ratio required for intended adjustment of the device 116 between the plate spring 104 or the disengagement force acting on same and the resulting supporting or sensor force acting on the plate spring 104 would be destroyed, namely this supporting force would drop to an inadmissibly low level whereby the clutch 101 would adjust prematurely or undesirably. The operating point of the plate spring 104 would move in the direction of force minimum. Speed-dependent destroying effects can be compensated by the speed or centrifugal force dependent means in the form of tongues 156. This takes place, as already mentioned by a speed or centrifugal force dependent design of an additional supporting force which is connected in parallel with the force produced by the sensor spring 113 and/or by the leaf spring elements 109.

The friction clutch 201 shown in Figure 8 forms a so-called pull-type friction clutch. The plate spring 204 is supported radially outside on a wear compensation ring 218 provided between the radial areas 202a of the housing 202 and the plate spring 204. With its radially further inner areas the plate spring 204 biases the cams 213 of a pressure disc 203. On the side of the plate spring 204 remote from the pressure disc 203 is a wear sensor 237 which is supported by the plate spring 204 and is locked with this through a bayonet-type connection. To this end the wear sensor 237 designed as a plate spring has radially inwards axial hook-like extension arms 241 which form an axially locking push-turn connection in conjunction with the axial recesses 204 provided in the plate spring. The wear sensor 237 prevents the wear sensor ring 220 from being able to adjust in the absence of any wear on the friction linings 207. The wear sensor ring 220 is provided concentrically and radially inside the wear compensation ring 218.

The wear compensation ring 218 and the wear sensor ring 220 each have circumferentially extending and axially rising run-up ramps 219,223 which are spread over the circumference of the rings 218, 220. The rings 218, 220 are installed in the clutch 201 so that the run-up ramps 219, 223 formed by the wedged or cam-like moulded areas face the housing floor 202a.

The run-up ramps 219, 223 are supported axially on counter run-up ramps 221, 222 which in the illustrated embodiment are fitted directly in the housing 202, namely in the cover base 202a, eg by imprinting.

The run-up ramps 219 and 223 as well as the associated counter run-up ramps 221, 222 are designed circumferentially so that these allow at least a turning angle of the rings 218, 220 relative to the housing 202 which ensures throughout the entire service life of the friction clutch at least a compensating of the wear which occurs on the friction faces of the pressure disc 203, counter pressure plate 206 and friction linings 207. It is particularly expedient if the pitch angle of the axially interacting runup ramps 219, 223 and the outer run-up ramps 221, 222 is selected so that the friction which occurs when the associated ramps 219 and 221 and 223, 222 press against each other prevents slipping between the superposed ramps, thus in practice a self-locking action occurs through friction.

This pitch angle can lie in the order of 3 to 12 The wear compensating ring 218 is spring loaded in the circumferential direction, namely in the adjustment direction, thus in the direction which by running up the ramps 219 and 221 causes an axial displacement of the wear compensation ring 218 in the direction of the pressure disc 203, thus in the axial direction away from the radial housing section 2a. This spring loading of the wear compensation ring 218 is guaranteed in the illustrated embodiment by at least one coil spring 229 which, viewed in the radial direction, is mounted between the two rings 218, 220. The wear sensor ring 220 is likewise spring loaded circumferentially in the adjustment direction, namely as apparent from Figure 8a, through at least one coil spring 229 which is mounted where it can be tensioned between the rings 218, 220 which are arranged actively in series. The coil spring 228 extends circumferentially and is tensioned between an axially angled plate 241a of the plate spring 237 and a radial cam 234 moulded onto the inner circumference of the ring 218. The wear sensor ring 220 has on its outer contour at least one radial cam 235 which overlaps with the radial cam 234. On or in the cams 234, 235 there are sockets for holding and guiding the at least slightly pretensioned coil spring 229. The relative rotation of the wear compensation ring 218 opposite the wear sensor ring 220 can be defined by the cam 234 stopping on the cam 235. The two springs 228 and 229 are connected together actively in series. As can be seen from Figure 8b at least one spring 228 can be associated with each ring 218, 220. The rings 218, 220 have support areas 234a, 235a which are circumferentially off-set. The plate spring 237 likewise has individual circumferentially off-set support or biasing areas 241a, 241b for the individual springs 228.

In the absence of wear the wear sensor 237 prevents inadmissible adjustment of the wear sensor ring 220 which in turn again prevents an inadmissible adjustment of the wear compensation ring 218. If there is no wear present on the friction linings 207 then the cams 234, 235 adjoin one another. In order to guarantee this, throughout the entire service life of the friction clutch 201 the turning moment exerted by the coil spring 228 on the wear compensation ring 218 is greater than the torque produced by the coil spring 229 between the two rings 218, 220.

The plate spring 237 serving as the wear sensor is mounted on the operating plate spring with a specific axial pretension acting in the direction of the wear sensor ring 220. The pretension with which the plate spring 237 adjoins the operating plate spring 204 and counteracts an adjustment of the wear sensor ring 220 is selected so that the wear sensor ring 220 cannot turn in the closed wear-free state of the clutch and after wear adjustment has been made in the clutch. The axial force exerted on the wear sensor ring 220 by t tensioned whereby an uncontrolled adjustment of the friction clutch 201 could occur.

The additional spring 226 which acts at least over a partial area of the air path of the friction clutch 201 is mounted axially between the pressure disc 203 and the operating plate spring 204. The relaxed additional spring 226 designed as a plate spring is set up conically in the direction of the pressure disc 203 wherein the outer edge of the additional spring 226 is fixed axially relative to the operating plate spring 204 so that the plate spring 226 is held for swivel movement relative to the operating plate spring 204. The additional spring 226 ensures the same function as the additional spring 126 according to Figure 7 or the additional spring 26 according to Figure 1 and reference is made in this respect to the previous description.

The plate spring-like wear sensor 237, the operating plate spring 204 and the additional plate spring 226 have axially aligned recesses through which passes at least one axially extending pin 203a which is fixedly anchored in the pressure disc 203. The components 237, 204 and 226 are secured by this pin 203a from turning both relative to the pressure disc 203 and also relative to each other.

The tongues 204b of the operating plate spring 204 support radially inwards an operating element in the form of a traction plate 260 which is axially displaceable through a disengagement mechanism such as eg a disengagement bearing.

The traction plate 260 is held axially on the radially inner areas of the tongues 204b and has areas 260a through which the radially inner areas of the tongues 226a of the additional plate spring 226 can be biased.

The ring-like biasing area 260a is formed by the radially outer sections of the plate 260. Between the tongues 226a of the additional spring 226 and the biasing areas 260a interacting with same is an axial space 232 which ensures that the additional plate spring 226 only acts over the air path, thus after release of the friction linings 207 through the pressure disc 203.

Starting from the new state shown in Figure 8 of the friction clutch 201 mounted on the counter pressure plate 206 with the interposition of the clutch disc 208, when disengaging the clutch the plate spring 204 is swivelled inwards radially to the right so that the plate spring 204 is supported radially outwards on the rolling support pad 212 supported by the wear compensation ring 218. During the disengagement phase the sensor plate spring 237 is tensioned axially between the plate spring 204 and the wear sensor ring 220, namely until the angular play L between the sensor spring 237 and plate spring 204 defining the lift stroke of the pressure disc 203 is used up so that then the plate spring 204 can be supported axially on the wear sensor ring 220. When continuing the disengagement movement the plate spring 204 is swivelled round the ring-like support area 220a provided on the wear sensor ring 220 whereby the plate spring 204 relaxes the radially outer rolling support pad 212 so that if there is any wear this can be compensated by a corresponding axial adjustment of the ring 218. The plate spring 204 is thus swivelled at first during the disengagement phase like a one-armed lever round the outer rolling support pad 218. After exceeding the gap or angle L the ring-like swivel area of the plate spring 204 is moved radially inwards into the area 220a of the wear sensor ring 220 so that by continuing the disengagement movement the plate spring 204 is then swivelled or acts in the manner of a double-armed lever. This radial shift in the ring-like rolling support pad of the plate spring 204 during operation of the friction clutch 201 changes the translation ratio or lever arm ratio which determines the force required for operating the plate spring 204, from I to I-i so that as soon as the plate spring 204 is supported on the wear sensor ring 220 an increase in the disengagement force takes place.

By translation ratio I is meant the ratio of the distance between the engagement area of the operating plate 260 on the plate spring tongues 204b and the contact area plate spring 204 with rolling support pad 212, to the distance between this contact area and the biasing area of the plate spring 204 for the cams 213 of the pressure disc 203. The aforesaid change in translation is based on the assumption that the support between the plate spring 204 and the pressure disc 203 takes place at least approximately on the same diameter as the support of the plate spring 204 on the wear sensor ring 220. The further the support area between the plate spring 204 and wear sensor ring 220 is moved radially outwards towards the rolling support pad 212 the smaller becomes the rise in the disengagement force when the plate spring 204 adjoins the wear sensor ring 220. In an embodiment where the support diameter between the plate spring 204 and wear sensor ring 220 is greater than the support diameter between the plate spring 204 and pressure disc 203, a greater translation ratio is set during swivelling of the plate spring 204 about the wear sensor ring 220, than the aforesaid translation ratio I-i. The translation ratio which is set as the friction clutch 201 is disengaged must not however be greater than the translation ratio I of the plate spring 204.

As soon as wear appears on the friction linings 207 during a coupling phase the plate spring 204 changes its conicity, namely the tongue tips 204c then move left together with the plate 260. Through this change in conicity the wear sensor ring 220 becomes relaxed so that this can adjust according to the wear which has occurred on the linings. With the onset of wear the wear sensor ring 220 first leads the wear compensation ring 218, as shown in Figure 8a. As the wear sensor ring 220 turns so a spacing 245 proportional with the wear arises between the stop cams 234, 235 of the two rings 218, 220. With a disengagement process which now follows the wear compensation ring 218, as already mentioned, is relaxed by the plate spring 204 so that this ring can adjust according to the play 245. The plate spring 204 thereby once more occupies a conicity or operating position which corresponds with the new state. With increasing wear the plate spring 204 is shifted axially away from the cover base 202a wherein a corresponding angular correction of the installation position of the plate spring 204 takes place over the entire adjustment range. The corresponding correction is each time dependent on the wear measured and detected by the wear sensor ring 220.

The interaction of the individual springs 204, 226, 237 will now be explained in detail in connection with the characteristic lines entered in the diagram according to Figure 9.

Line 261 in Figure 9 represents part of the resulting axial force curve produced in dependence on the change in conicity of the plate spring 204 and taking into account the force applied by the torque transfer means, such as for example the leaf spring elements between the housing and pressure disc 203. This curve arises in the event of deformation of the plate spring 204 between the swivel bearing 212 and the cams 213. The complete characteristic line 261 is sinusoidal and drops down again in an extension to the left.

The complete curve of the line 261 is thus similar to the curve of the line 33 according to Figure 4. On the abscissa is the relative axial path between the two support pads or supports 212, 213, and on the ordinate is the resulting force produced by the plate spring 204 and the torque transfer means. The point 262 represents the installation position of the plate spring 204 when the clutch 201 is closed. The line 263 represents the axial expanding force between the friction linings 207 applied by the lining spring segments 210. During disengagement of the friction clutch 201 the spring segments 210 relax over the path 264.

Over this path 264 which also corresponds to a corresponding axial shift of the pressure disc 203 the disengagement process of the clutch is assisted through the lining suspension 210 provided between the linings 207. The curve of the force to be applied in the area of the cams 213 to disengage the clutch 201 over the path 264 is shown in Figure 9 by broken line 265. The forces 266 arising over the path 264 correspond to the relevant difference 267 between the force curve 261 of the plate spring 204 and the force curve 263 of the lining suspension 210. The curve of the disengagement force for the clutch 201 actually to be applied in the area of the tongue tips 204c is reduced relative to the force curve 265 by the translation ratio I of the plate spring 204. On exceeding the point 268 at the end of the disengagement path part 264 the friction linings 207 are released whereby as a result of the degressive characteristic line area of the plate spring 204 the disengagement force which is then to be applied is considerably reduced compared to that which would correspond to point 262. The disengagement force for the clutch 201 would decrease without the additional spring 226 until the minimum or trough point 269 of the sinusoidal characteristic line 261 is reached.

In the illustrated embodiment the plate spring 204 is designed so that the minimum 269 of the characteristic line 261 comes to lie underneath the abscissa axis. On exceeding the point 270 lying on the abscissa axis the plate spring 204 thus automatically tries to occupy a position corresponding to point 261 and thus forms a so-called snap spring which has two different locked states, namely a relaxed and tensioned state.

As can be seen from the part of the line 261 shown broken through points 268, 269, 270, a significant change in force in the disengagement force curve is present without the additional spring 226, after release of the friction linings 207 through axial lifting of the pressure disc 3. By using the compensation spring 226 it is possible to keep the disengagement force curve 272 to a low relatively constant level throughout the remaining disengagement path 271 of the disengagement force curve 272 adjoining the release path 264. The curve 272 in Figure 9 again corresponds to the force curve which is to be applied in the area of the cams 213 in order to swivel the plate spring 204 together with the plate spring 226. The disengagement force curve in the area of the tongue tips 204c is however reduced up to the support of the plate spring 204 on the wear sensor ring 220 relative to the curve 272 according to Figure 9 by the translation ratio I of the plate spring 204. The path corresponding to the air path of the pressure disc 203 is marked by 273 in Figure 9. The path corresponding to the air path of the pressure disc 203 is marked by 273 in Figure 9. The additional spring 226 has a force-path characteristic line according to line 274. The plate spring 204 and the additional spring 226 have an oppositely directed force-path curve at least over the remaining disengagement path 271. The force curve 272 over the air path 273 is produced by adding the curves of the lines 261 and 274 which exist over the air path 273. The resulting curve 272 starts here at point 268, thus at the start of the air path 273. The contact of the areas 260a of the operating plate 260 against the tongues 226a of the additional spring 226 thus corresponds to the point 268 in Figure 9.

After passing through the air path 273 an axial support of the plate spring 204 is produced through the wear sensor ring 220, as already described. The translation ratio which is decisive for operating the plate spring 204 thereby changes from I to I-i. The resulting slight increase in the disengagement force is characterised by 275 in Figure 9.

This increase in disengagement force remains throughout the adjustment path 276 which adjoins the air path.

The adjustment path 276 which is required for compensating the wear on the friction linings 207 is very slight in relation to the overall disengagement path and can be in the area of a few tenths of a millimetre or even less.

In order to obtain the required operating forces arising in the area of the tongue tips 204c the forces or force curves which can be drawn from Figure 9 are to be divided by the existing translation ratio I up to point 275 and 1-1 after point 275. In order to obtain the corresponding operating paths in the area of the tongue tips 204c the paths which can be derived from Figure 9 are to be multiplied with the translation ratio I up to point 275 and 1-1 after point 275.

The translation ratio 1-1 is only decisive for the path section 276 extending beyond point 275.

Through the design of a friction clutch according to the invention it is possible to reduce the losses caused by sagging of the tongues when the clutch is disengaged to a very small amount, namely because in the disengaged state of the friction clutch the forces acting on the tongue tips 4c, 104c and 204c are very low and also, as can be seen from Figure 9, can be negative. The latter means that the plate spring in the negative area of the characteristic line 261 automatically tends towards the disengaged position. This is however compensated by the additional spring 226. This strain on the tongues of the operating plate spring which is low compared to conventional clutches makes it possible to reduce the disengagement path since as already mentioned at least the path losses through the elastic sagging of the tongues and of the cover practically do not exist.

Furthermore the design of a friction clutch according to the invention and shown in Figure 8 makes it possible to minimize the path losses during operation of the clutch 201 which are due to the suspension of the housing or cover 202.

This can be achieved in that the axial suspension produced by the plate spring 204 on the clutch cover 202 when the friction clutch 201 is closed is equalized through a corresponding design of the friction clutch 201 with the cover suspension when the clutch is disengaged, thus when the plate spring 204 is supported on the wear sensor ring 220. When the friction clutch is closed the axial force exerted by the plate spring 204 on the cover 202 is at its greatest, but the free bending length between the biasing diameter of the rolling support pad 212 and the cover screws 202b is at its smallest. When the friction clutch is disengaged the axial force produced by the plate spring 204 and the additional spring 226 and absorbed by the cover 202 is considerably less than the contact pressure force of the plate spring 204 when the friction clutch 201 is closed.

The free lever or spring length of the cover 202 between the support diameter 220a of the wear sensor ring 220 and the screws 202b is however significantly greater than the radial distance between the rolling support pad 212 and these screws 202b.

The friction clutch 301 shown in Figure 10 has, apart from the design of the plate spring 337 serving as the wear sensor, the same construction and the same method of operation as the clutch 201 according to Figure 8. The operating plate spring 304, the wear sensor spring 337 and the additional spring 326 undertake the same functions as those component parts 204, 226 and 237 described in connection with Figure 8 and the diagram according to Figure 9. The wear compensation ring 318 and the wear sensor ring 320 interact with the associated housing 302 in the same way as the rings 218 and 320 according to Figure 8 through runup ramps and counter run-up ramps.

The plate spring 326 is axially pretensioned by spring means 344 between the disengagement ring 360 and the plate spring 304 for axially fixing same when the clutch is engaged.

The sensor plate spring 337 has in its pretensioned position which is shown in Figure 10, a different axial set-up than the sensor plate spring 237 of Figure 8. The plate spring 337 is supported radially inwards on the operating plate spring 304 and with a radially further outer ring-like area biases the wear sensor ring 320 axially in the direction of the housing 302. The spring 337 has radial extension arms 342 with which it adjoins the support pad 343 of the adjustment ring 318 and is held pretensioned to an air play L when the clutch is closed. Radially outwards the sensor plate spring 337 has tongues 341a directed axially towards the housing 302 and serving to support the circumferentially acting adjustment springs 328.

The clutch unit shown in Figure 11 with a friction clutch 401 has the same construction as the clutch unit according to Figure 7 except for the installation position and the slightly different method of operation of the additional spring 426.

In many cases as a result of the clutch characteristic required and the installation space available for the clutch 401, the main plate spring 404 cannot be designed in optimum manner as far as its path-force characteristic is concerned, particularly over the required disengagement path (46 in Figure 4). Thus during the disengagement process the point (39a in Figure 4) of the plate spring characteristic line (33 in Figure 4) after which the disengagement force becomes greater than the axial supporting force provided in the adjustment point (37 in Figure 4) for the plate spring 404 cannot be exceeded. This means that then the point 39a lies practically on or just behind the point 45 of Figure 4, thus at the end of the required disengagement path or when exceeding this disengagement path shortly after same.

Exceeding this path in this way would have the result that the adjustment ring 417 would be axially relaxed through an inadmissible large path of the plate spring 404 and would adjust accordingly. Thus an adjustment could take place even when there is no sign of wear on the clutch linings 407. This would result in a change in the operating point, thus a change in the installation position of the plate spring 404 in the engaged state of the friction clutch 401 namely in the direction of a smaller pretensioning and contact pressure force. This means that with a clutch unit of this kind the operating point marked by 37 in the diagram according to Figure 4 would shift or be displaced along the characteristic line 33 corresponding to the over movement in the direction of the minimum 38a. The torque transferable by the friction clutch would thereby reduce accordingly which can lead to failure of the clutch.

In order to avoid this type of undesired adjustment of the adjustment device 416 the additional spring 426 is mounted between the adjustment ring 417 and the plate spring 404 so that this acts as a blocking device or brake for the adjustment device 416 at least when the maximum permissible disengagement path is exceeded. An inadmissible adjustment in the friction clutch 401 can thereby be prevented even in the event where the normal disengagement path has been greatly exceeded and/or in the case of axially vibrating components.

The additional spring 426 designed as a plate spring is mounted between the adjustment ring 417 and the plate spring 404 so that from a certain disengagement path it is tensioned between the adjustment ring 417 and the plate spring 404 so that the adjustment ring 417 is biased by the plate spring 426 against the operating plate spring 404.

Areas of the adjustment ring 417 are thus in practice clamped or tensioned between the two plate springs 426 and 404. It is thereby ensured that from a certain disengagement path the adjustment ring 417 is secured against turning, The additional plate spring 426 has a ring-like base member 427 which has on the outer circumference radial extension arms 428 which are spread out in the circumferential direction of the base member 427 and engage in a radial groove 417a of the adjustment ring 417, as is apparent from Figures 12 and 13. The additional plate spring 426 has stop contours 422 which are formed by radial tongues 422 moulded on the inner circumference of the base member 427.

These stop contours 422 interact with counter stop contours 430 provided on the plate spring 404. The counter stop contours 430 are formed in the illustrated embodiment by heads of rivet elements 431 which are provided in the area of the tongues 404b of the plate spring 404. Instead of the rivet elements 431 it is also possible to use tongues which are integral with the additional plate spring 426 and interact with this in the same way as the tongues 122 of Figure 7 interact with the corresponding plate spring 104.

The distance 432 between the stop contours 422 and the counter stop contours 430 is measured in the engaged state of the friction clutch so that at least over a part of the clutch disengagement path there is no contact between the contours 422 and the counter contours 430. The stop contours 422 preferably come to adjoin the counter stop contours 430 only when the release point (37 in Figure 4) of the friction clutch 401 is exceeded. When this specific disengagement path part is exceeded the additional plate spring 426 is tensioned with the main plate spring 404 whereby as already mentioned the adjustment ring 417 is clamped against the main plate spring 404 and is prevented from turning as a result of the circumferential force applied by the coil spring 420.

The adjustment ring 417 has a groove 417a on its radially inner contour and approximately in the middle of its axial extension for mounting and swivel supporting the additional plate spring 426 on the adjustment ring 417.

As can be seen from Figures 12 and 13, the groove 417a is opened or interrupted in parts, viewed over its circumference and axially towards the plate spring 404. To this end the adjustment ring 417 has axially aligned radial indentations 440 which interconnect with the circumferentially aligned groove 417a. The adjustment ring 417 forms radially protruding cams 441 between the indentations 440 moulded on the inner circumference of the adjustment ring 417. The distribution of the indentations 440 over the circumference and their number correspond to the distribution and number of the extension arms 428 of the additional plate spring 426. The extension in the circumferential direction of an indentation 440 corresponds at least to the corresponding extension of an extension arm 428. The assembly between the additional plate spring 426 and the adjustment ring 417 can take place through axially inserting the extension arms 428 into the indentations 440 and through a following relative rotation between the two components 417 and 426. As a result of the said relative rotation between the two components 417 and 426 the extension arms 428 and the projections or cams 441 are superimposed at least in part axially as can be seen from Figure 12. As a result of this superimposition the tongues or extension arms 428 can, after the stop contours 422 have come to adjoin the counter stop contours 430, be tensioned against the radial projections or cams 441.

The additional plate spring 426 is designed and installed in the friction clutch 401 in such a way that during the disengagement phase, at least in the area of the disengagement path in which the stop contours 422 abut the counter stop contours 430, no relative turning of the adjustment ring 417 can take place relative to the additional plate spring 426. The contact between the stop contours 422 and counter stop contours 430 preferably takes place during a disengagement process only after the release point (point 37) in Figure 4). It is thereby ensured that the turning of the adjustment ring 417 which is necessary for compensating the wear on the friction linings 407 can take place. Relative rotation of the adjustment ring 417 opposite the additional plate spring 426 is preferably carried out in the practically relaxed state of the additional plate spring 427. As can be seen from Figure 12, the cams 441 of the adjustment ring 417 shift with increasing wear on the friction linings 407 to the right relative to the tongues 428. The circumferential extension of the cams 441 and the tongues 428 is measured so that at least over the overall permissible wear range of the clutch assembly - viewed axially - these projections 441, 428 are superposed on each other. The additional plate spring 426 can also be tensioned in the engaged state of the friction clutch 401. This can be carried out in that during engagement and shortly before the end of the engagement path the tongues 404b of the main plate spring come to adjoin the individual tongues 422a of the additional plate spring 426 whereby the additional plate spring 426 is tensioned at least slightly to the right in the direction of the clutch cover. The maximum contact pressure can thereby also be restricted in the engaged state of the clutch by the additional spring 426. The height of the contact pressure force restriction thereby depends on the characteristic line curve of the main plate spring 404 and of the additional plate spring 426.

In addition to the function of a brake for the adjustment ring 417 the additional plate spring 426 can also undertake the same function as the additional spring 126 of Figure 7.

It can likewise cause lifting of the disengagement force curve at least in the minimum area of the plate spring 404 whereby a more constant curve of the disengagement force can be ensured over the disengagement path. For this reference is made to the description of Figures 1 to 10.

The inventions are not restricted to the embodiments described and illustrated but can be used quite generally in the case of friction clutches, more particularly those having a device for compensating the wear on the friction linings. Furthermore the invention also includes variations which can be formed by a combination of individual features and elements described in connection with the various embodiments. Also individual features and methods of functioning in connection with those described in the drawings can stand alone to represent an independent invention. The applicant retains the right to claim further features only disclosed up until now in the description as being essential to the invention.

Claims

PATENT CLAIMS

1. Friction clutch for use with a clutch disc, more particularly for motor vehicles, wherein the clutch has a pivot support for a plate spring which biases an axially displaceable pressure plate towards the clutch disc, and a counter pressure plate, and has a device compensating the wear at least of the friction linings, characterised in that the force of a further energy accumulator is superimposed on the plate spring at least approximately from the path in the disengagement direction from which the clutch disc no longer or only insignificantly biases the pressure plate, and that the force path resulting therefrom is equalized relative to the force path of the plate spring.

2. Friction clutch according to claim 1, characterised in that the clutch disc is of the type having "sprung lining".

3. Friction clutch according to claim 1 or 2 characterised in that the further energy accumulator, a "compensation spring" is integrated in the clutch.

4. Friction clutch according to one of claims 1 to 3 characterised in that the compensation spring is a plate spring.

5. Friction clutch according to one of claims 1 to 4 characterised in that the compensation spring at least approximately from the path in the disengagement direction from which the clutch disc no longer or only insignificantly biases the pressure plate (release path), has a force path curve with a curvature differing from, such as directed against, that of the force path curve of the plate spring.

6. Friction clutch substantially as herein described with reference to the accompanying drawings.